HANDBOOK of LUBRICATION and TRIBOLOGY V O LU ME I Application and Maintenance S E C O N D
© 2006 by Taylor & Francis Group, LLC
E D I T I O N
HANDBOOK of LUBRICATION and TRIBOLOGY V O L U ME I Application and Maintenance S E C O N D
E D I T I O N
Edited by
George E. Totten
Boca Raton London New York
CRC is an imprint of the Taylor & Francis Group, an informa business
© 2006 by Taylor & Francis Group, LLC
Published in 2006 by CRC Press Taylor & Francis Group 6000 Broken Sound Parkway NW, Suite 300 Boca Raton, FL 33487-2742 © 2006 by Taylor & Francis Group, LLC CRC Press is an imprint of Taylor & Francis Group No claim to original U.S. Government works Printed in the United States of America on acid-free paper 10 9 8 7 6 5 4 3 2 1 International Standard Book Number-10: 0-8493-2095-X (Hardcover) International Standard Book Number-13: 978-0-8493-2095-8 (Hardcover) This book contains information obtained from authentic and highly regarded sources. Reprinted material is quoted with permission, and sources are indicated. A wide variety of references are listed. Reasonable efforts have been made to publish reliable data and information, but the author and the publisher cannot assume responsibility for the validity of all materials or for the consequences of their use. No part of this book may be reprinted, reproduced, transmitted, or utilized in any form by any electronic, mechanical, or other means, now known or hereafter invented, including photocopying, microfilming, and recording, or in any information storage or retrieval system, without written permission from the publishers. For permission to photocopy or use material electronically from this work, please access www.copyright.com (http://www.copyright.com/) or contact the Copyright Clearance Center, Inc. (CCC) 222 Rosewood Drive, Danvers, MA 01923, 978-750-8400. CCC is a not-for-profit organization that provides licenses and registration for a variety of users. For organizations that have been granted a photocopy license by the CCC, a separate system of payment has been arranged. Trademark Notice: Product or corporate names may be trademarks or registered trademarks, and are used only for identification and explanation without intent to infringe. Library of Congress Cataloging-in-Publication Data Catalog record is available from the Library of Congress
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and the CRC Press Web site at http://www.crcpress.com
This book is dedicated to my wife Alice (Ah Kum) for her continued support which allows me to pursue my passion in teaching through publication. It is her perseverance that literally defines the character of our family.
© 2006 by Taylor & Francis Group, LLC
STLE Preface
Handbook of Lubrication and Tribology: Volume I Application and Maintenance, Second Edition is sponsored and copublished by the Society of Tribologists and Lubrication Engineers (STLE). This book is one of a three volume series covering: in Volume II, Theory and Practice; in Volume III, Monitoring, Materials, Synthetic Lubricants, and Applications; and in this volume, Applications and Maintenance. The goal of this book is to provide an update to the first edition and to provide the latest information regarding application and maintenance in the broad field of Lubrication Engineering. The first edition was written over twenty years ago. In the intervening period, the science of tribology and the development of engineering best practices have evolved markedly. As a result, each of the chapters of the first edition has been rewritten and updated. A number of new chapters have been added to capture new information: In the section on Applications, the Hydraulics chapter was split into two chapters on pumps and fluids, and all new chapters on Tribology of Data Storage Devices, and Biotribology were added. In the section on Industrial Practices, a new chapter was added on Tribology of Metal Forming Processes. In the section on Maintenance, three new chapters were added on Lubricant Cleanliness, Environmental Implications of Lubricants, and Centralized Lubrication Systems — Theory and Practice. This volume was written by a peer recognized team of expert contributors from a wide variety of industry segments. Each chapter was written by an expert both knowledgeable and active in the subject area. Thanks go to these individuals; without their expertise and hard work this work would not have been possible. Thanks must also go to their employers for their support of this effort and contribution to our industry. Because of its emphasis on the practice of Lubrication Engineering, this book is an excellent reference for those preparing for STLE’s Certified Lubrication Specialist® Certification examination. As such, it has been recommended in the Body of Knowledge by STLE’s Certified Lubrication Specialist Certification Committee. This volume, like its predecessor, belongs in the reference library of all professionals in the field.
R.M. Gresham STLE Director of Professional Development
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© 2006 by Taylor & Francis Group, LLC
Preface
The first edition of the Handbook of Lubrication: Theory and Practice of Tribology — Volume I: Application and Maintenance was edited by E.R. Booser and was sponsored by the Society of Tribologists and Lubrication Engineers (STLE) to provide the latest information in the field. Volume I of the Handbook of Lubrication: Theory and Practice of Tribology covers Applications and Maintenance. Volume II covers theory and design and was published in 1984. Volume III which covers Monitoring, Materials, Synthetic Lubricants, and Applications was published in 1994 to extend the topical areas covered by Volume I and Volume II since their initial publication. Over 20 years have elapsed since the First Edition was published in 1983, and enormous changes continue to occur in the lubrication and tribology engineering sciences. Although Volume III did extend the areas covered, all of the areas initially covered in Volume I needed to be significantly updated. In view of these changes and the time that has elapsed since the appearance of the First Edition, STLE initiated the Second Edition of this invaluable text. The Second Edition of the Handbook of Lubrication and Tribology: Volume I Application and Maintenance, has been reorganized slightly to aid the reader in identifying chapters and topics of interest. All of the chapters from the First Edition, with the exception of the chapter on Marine Equipment, have been revised or completely rewritten. In addition, a number of new chapters have been added including: Biotribology, Tribology of Data Storage Devices, Tribology of Metal Forming Processes, and Environmental Implications of Lubricants. The chapter on Compressors and Vacuum Pumps was significantly expanded and the original chapter on Hydraulic Systems and Fluids was divided and expanded into two separate chapters: Hydraulic Pumps and Hydraulic Fluids. Altogether there are a total of 37 chapters much of which is either a totally new treatment of the subject or completely new information. This handbook provides the reader with an extensive reference to the most important and commonly encountered lubrication systems and fluids in industry. This text is of value to the practicing tribologist and lubrication engineer, mechanical or materials engineer, and failure analysis personnel. I am indebted to all of the contributing authors of the book for their tremendous effort and patience. Without their dedication and support, the successful completion of this text would not have been possible.
George E. Totten Seattle, WA
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© 2006 by Taylor & Francis Group, LLC
Acknowledgments
I wish to thank the Society of Tribologists and Lubrication Engineers (STLE) for their continued support throughout this project. Very special thanks to Robert Gresham and Barbara Rapacz, without whose support the successful completion of this project would not have been possible. I am especially indebted to Theresa Delforn and Shelley Kronzek of CRC Press, Inc. for their continued guidance and expert assistance throughout this process, from the beginning to the end. They have made a potentially difficult task into an absolute delight. Finally, and most importantly, I am especially indebted to my family, especially my wife Ah Kum for allowing me to be so totally involved in this project for such a long time. Their continued patience with my sometimes bad behavior is most especially appreciated.
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© 2006 by Taylor & Francis Group, LLC
The Editor
George E. Totten is President of G.E. Totten & Associates, LLC in Seattle, Washington, and a visiting professor of materials science at Portland State University. Dr Totten is coeditor of a number of books including Steel Heat Treatment Handbook, Handbook of Aluminum, Handbook of Hydraulic Fluid Technology, Mechanical Tribology, and Surface Modification and Mechanisms (all titles of CRC Press), as well as the author or coauthor of over 400 technical papers, patents, and books on lubrication, hydraulics, and thermal processing. Dr Totten is a Fellow of ASM International, SAE International and the International Federation for Heat Treatment and Surface Engineering (IFHTSE) and a member of other professional organizations including ACS, ASME, and ASTM. Dr Totten formerly served as president of IFHTSE. He received bachelor’s and master’s degrees from Fairleigh Dickinson University in Teaneck, New Jersey and a Ph.D. degree from New York University, New York.
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© 2006 by Taylor & Francis Group, LLC
Contributors
James R. Anglin
Dennis W. Brinkman
Sabrin Gebarin
Aluminum Company of America Alcoa Technical Center Alcoa Center, PA
Indiana Wesylan University Marion IN
Noria Corporation Tulsa, OK
José Castillo
O.-C. Göhler
Mark Barnes Noria Reliability Solutions Noria Corporation Tulsa, OK
D.J.W Barrell School of Mechanical Engineering The University of Leeds Leeds, UK
Edward P. Becker General Motors Powertrain
Bharat Bhushan Department of Mechanical Engineering Ohio State University Columbus, Ohio
Iomega Corporation Advance R&D Dept Roy, UT
Paul Conley Lincoln Industrial St. Louis, MO
J. Fisher School of Mechanical Engineering Institute of Medical and Biological Engineering University of Leeds Leeds, UK
Andy Hall Rolls Royce plc Customer Training Centre Derby, England
Hooshang Heshmat Mohawk Innovative Technology, Inc. Albany, New York
Malcolm F. Fox De Montfort University Leicester, UK
G.S. Fox-Rabinovich Department of Mechanical Engineering McMaster University Hamilton, Ontario, Canada
© 2006 by Taylor & Francis Group, LLC
Lincoln Industrial St. Louis, MO
Noria Corporation Tulsa, OK
Union Carbide Corporation Tarrytown, NY
EPRI/NMAC Consultant Lubricants of Lubrication San Rafael, CA
Ayzik Grach
James C. Fitch
Roland J. Bishop
Robert O. Bolt
Institute of Fluidpower Drives and Controls (IFAS) RWTH Aachen University Aachen, Germany
Arup Gangopadhyay Ford Research Laboratory Dearborn, MI
Emile van der Heide TNO Industrial Technology Eindhoven, The Netherlands
E. Ingham School of Biochemistry and Microbiology Institute of Medical and Biological Engineering University of Leeds Leeds, UK xv
xvi
Contributors
Douglas M. Jahn
Jude Liu
Farrukh Qureshi
Delphi Saginaw Steering Systems Saginaw, Michigan
Agriculture and Biosource Engineering Department University of Saskatchewan Saskatoon, Saskatchewan, Canada
The Lubrizol Corporation Wickliffe, OH
Mark J. Jansen NASA Glenn Research Center Tribology and Surface Science Branch Cleveland, OH
Z.M. Jin School of Mechanical Engineering University of Leeds Leeds, UK
Mike Johnson Noria Field Services Noria Corporation Tulsa, OK
Robert L. Johnson Noria Corporation Tulsa, OK
Michael L. McMillan General Motors R&D Center, Chemical and Environmental Science Laboratory Warren, MI
T. Meindorf Argo-Hytos GmbH Kraichtal, Germany
Hans M. Melief The Rexroth Corporation Industrial Hydraulics Division Bethlehem, PA
Paul W. Michael William R. Jones NASA Glenn Research Center Tribology and Surface Science Branch Cleveland, OH
Milwaukee School of Engineering Fluid Power Institute, Milwaukee, Wisconsin
H. Murrenhoff Rob Dwyer-Joyce Department of Mechanical Engineering University of Sheffield Sheffield, UK
Institute of Fluidpower Drives and Controls (IFAS) RWTH Aachen University Aachen, Germany
Barbara J. Parry T. Kazama Dept of Mechanical Systems Engineering Muroran Institute of Technology Hokkaido, Japan
Mohawk Lubricants North Vancouver, Canada
B.C. Pettinato Elliott Turbomachinery Co., Inc. Jeannette, PA
R. Lal Kushwaha Agriculture and Bioresource Engineering Department University of Saskatchewan Saskatoon, Saskatchewan, Canada
H.A. Poitz Air BP Lubricants Melbourne, Australia
M. Priest Roger Lewis Department of Mechanical Engineering University of Sheffield Sheffield, UK
© 2006 by Taylor & Francis Group, LLC
Jost Professor of Engineering Tribology School of Mechanical Engineering The University of Leeds Leeds, UK
Dirk Jan Schipper University of Twente Department of Mechanical Engineering Tribology Group Enschede, The Netherlands
Rick Schrama Dofasco Inc., General Maintenance Shops Hamilton, Ontario, Canada
Shirley E. Schwartz General Motors (retired)
Will Scott School of Mechanical, Manufacturing and Medical Engineering Queensland University of Technology Brisbane, Australia
Paul D. Seemuth Tribology Consulting, International LLC Hixson, TN
L.S. Shuster Department of Mechanical Engineering Ufa Aviation Institute Ufa, Russia
Jacek Stecki Subsea Engineering Research Group Department of Mechanical Engineering Monash University Melbourne, Australia
Richard K. Tessmann FES, Inc. Stillwater, OK
C.D. Tipton The Lubrizol Corporation Wickliffe, OH
Contributors
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Allison M. Toms
Drew D. Troyer
James F. Walton II
Condition Assessment Center GasTOPS Inc. Pensacola, FL
Noria Corporation Tulsa, OK
Mohawk Innovative Technology, Inc. Albany, New York
Larry A. Toms Technical Services Pensacola, FL
Simon C. Tung General Motors R&D Center Chemical and Environmental Science Laboratory Warren, MI
S.C. Veldhuis G.E. Totten Portland State University Department of Mechanical and Materials Engineering Portland, OR
© 2006 by Taylor & Francis Group, LLC
McMaster Manufacturing Research Institute (MMRI) Department of Mechanical Engineering (JHE-316) McMaster University Hamilton, Ontario, Canada
Martin Williamson Noria UK, Ltd. Chester Cheshire, UK
R.E. Yungk Air BP Lubricants Melbourne, Australia
Contents
SECTION I 1
Applications
Automotive Engine Oil
. . . . . . . . . . . . . . . . . . .
1-3
Simon C. Tung, Michael L. McMillan, Edward P. Becker, and Shirley E. Schwartz
2
Automatic Transmission Fluids . . . . . . . . . . . . . . . .
2-1
C.D. Tipton
3
Rear Axle Lubrication . . . . . . . . . . . . . . . . . . . .
3-1
Arup Gangopadhyay and Farrukh Qureshi
4
Automotive Chassis and Driveline Lubrication
. . . . . . . . .
4-1
. . . . . . . . . . . . . .
5-1
. . . . . . . . . . . . . . . . . . . .
6-1
Douglas M. Jahn and Simon C. Tung
5
Diesel, Dual-Fuel, and Gas Engines D.J.W Barrell and M. Priest
6
Aircraft Gas Turbines Andy Hall
7
Principles of Gas Turbine Bearing Lubrication and Design
. . . .
7-1
Steam Turbines . . . . . . . . . . . . . . . . . . . . . . .
8-1
Hooshang Heshmat and James F. Walton II
8
B.C. Pettinato
9
Compressors and Vacuum Pumps . . . . . . . . . . . . . . .
9-1
T. Kazama and G.E. Totten
10 Basic Hydraulic Pump and Circuit Design
. . . . . . . . . . . 10-1
Richard K. Tessmann, Hans M. Melief, and Roland J. Bishop xix
© 2006 by Taylor & Francis Group, LLC
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Contents
11 Hydraulic Fluids
. . . . . . . . . . . . . . . . . . . . . . 11-1
H. Murrenhoff, O.-C. Göhler, and T. Meindorf
12 Coolants and Lubricants in Metal Cutting
. . . . . . . . . . . 12-1
S.C. Veldhuis, G.S. Fox-Rabinovich, and L.S. Shuster
13 Lubricating Industrial Electric Motors . . . . . . . . . . . . . 13-1 Drew D. Troyer
14 Effects of Radiation on Lubricants . . . . . . . . . . . . . . . 14-1 Robert O. Bolt
15 Wire Rope and Chain
. . . . . . . . . . . . . . . . . . . . 15-1
Paul Conley
16 Tribology of Hard Disk Drives — Magnetic Data Storage Technology . . . . . . . . . . . . . . . . . . . 16-1 José Castillo and Bharat Bhushan
17 Biotribology: Material Design, Lubrication, and Wear in Artificial Hip Joints . . . . . . . . . . . . . . . . . . . . . . . . . 17-1 Z.M. Jin, J. Fisher, and E. Ingham
SECTION II
Industrial Lubrication Practices
18 Steel Industry . . . . . . . . . . . . . . . . . . . . . . . . 18-3 Rick Schrama
19 Aluminum Metalworking Lubricants
. . . . . . . . . . . . . 19-1
James R. Anglin
20 Mining Industry . . . . . . . . . . . . . . . . . . . . . . . 20-1 Will Scott
21 Farm and Construction Equipment
. . . . . . . . . . . . . . 21-1
R. Lal Kushwaha and Jude Liu
22 Industrial Lubrication Practice — Wheel/Rail Tribology
. . . . . 22-1
Roger Lewis and Rob Dwyer-Joyce
23 Lubrication in the Timber and Paper Industries . . . . . . . . . 23-1 Paul W. Michael
24 Textile Fibers/Fabrics
. . . . . . . . . . . . . . . . . . . . 24-1
Paul D. Seemuth
25 Food-Grade Lubricants and the Food Processing Industry . . . . . 25-1 James C. Fitch, Sabrin Gebarin, and Martin Williamson
26 Aviation Industry . . . . . . . . . . . . . . . . . . . . . . 26-1 H.A. Poitz and R.E. Yungk
© 2006 by Taylor & Francis Group, LLC
Contents
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27 Lubrication for Space Applications . . . . . . . . . . . . . . . 27-1 William R. Jones and Mark J. Jansen
28 Friction and Wear in Lubricated Sheet Metal Forming Processes . . 28-1 E. van der Heide and Dirk Jan Schipper
SECTION III
Maintenance
29 The Degradation of Lubricants in Service Use . . . . . . . . . . 29-3 Malcolm F. Fox
30 Lubricant Properties and Test Methods . . . . . . . . . . . . . 30-1 Larry A. Toms and Allison M. Toms
31 Contamination Control and Failure Analysis . . . . . . . . . . 31-1 Jacek Stecki
32 Environmental Implications and Sustainability Concepts for Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . 32-1 Malcolm F. Fox
33 Lubrication Program Development and Scheduling
. . . . . . . 33-1
Mike Johnson
34 Lubricant Storage, Handling, and Dispensing
. . . . . . . . . . 34-1
Mark Barnes
35 Conservation of Lubricants and Energy . . . . . . . . . . . . . 35-1 Robert L. Johnson and James C. Fitch
36 Centralized Lubrication Systems — Theory and Practice . . . . . 36-1 Paul Conley and Ayzik Grach
37 Used Oil Recycling and Environmental Considerations . . . . . . 37-1 Dennis W. Brinkman and Barbara J. Parry
SECTION
Appendices
Appendix 1
. . . . . . . . . . . . . . . . . . . . . . . . . . A1-3
Appendix 2
. . . . . . . . . . . . . . . . . . . . . . . . . . A2-5
© 2006 by Taylor & Francis Group, LLC
I Applications
© 2006 by Taylor & Francis Group, LLC
1 Automotive Engine Oil 1.1
Automotive Engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
1-4
Engine Operation • Crankshaft to Crankshaft Bearing • Piston Pin to Piston • Piston Skirt to Cylinder Block • Piston Rings to Cylinder Block • Camshaft to Cam Follower and Valve Train • Oil Pump • Oil Filter
1.2 1.3
Issues Related to Energy Consumption in an Engine: Service Effects . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Engine Oil . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
1-8 1-11
Characteristics of Engine Oil and Functions of Its Additives • Viscosity Effects • Engine Oil Quality and Oil Degradation During Vehicle Use • Fluid Film Lubrication • Future Concerns
1.4
Gasoline Engine Oil Performance Categories and Associated Test Methods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
1-16
Introduction • ILSAC GF-4 and API SM Standard Tests
1.5
Edward P. Becker General Motors Powertrain
Shirley E. Schwartz General Motors (retired)
1-19
Service Effects of Diesel Engines Related to Engine Oil Degradation • Examples of Test Methods for Diesel Engine Oils • A Model for the Rate of Engine Oil Degradation in Diesel Engines
Simon C. Tung and Michael L. McMillan General Motors R&D Center, Chemical and Environmental Science Laboratory
Diesel Engine Oil Performance Categories and Associated Test Methods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
1.6
Future Directions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
1-21
Concerns Related to Conservation of Fuel • Effects at the Molecular Level • Insights Gained from Tests with an Alternative Fuel • Prolonging the Working Life of Engine Oil • Minimizing Emissions and Pollutants and Ensuring Backward Compatibility • Future Investigations
Acknowledgments. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
1-25 1-25
This chapter describes the functions of typical engines (gasoline and diesel), engine oil characteristics, and test methods. Included are descriptions of the tribological concerns associated with various engine components, service effects on engine oil, standard tests for engine oil and the types of service they represent, and an overview of the issues that need to be addressed in the future. 1-3
© 2006 by Taylor & Francis Group, LLC
1-4
Handbook of Lubrication and Tribology Rockers Valve springs Piston rings
Camshaft Valve Piston
Oil filter Cylinder block Con rod Journal bearings Crankshaft Oil pump Oil
FIGURE 1.1
Oil consumption Sump
The main components in an internal combustion engine.
1.1 Automotive Engines 1.1.1 Engine Operation An internal combustion engine, such as illustrated in Figure 1.1, is the predominant power source for most types of cars and trucks [1]. Various conditions influence engine development, such as the desire for high power output, legislative requirements for reduced emissions, increased fuel economy, and minimal generation of hazardous substances. Many technical and environmental challenges await those who attempt to address these concerns. The following discussion contain examples of current conditions (and in some cases the evolution of current conditions) to provide insight into the types of issues that must be understood and actions that are desirable to meet future concerns successfully. Directions in which future developments may evolve are included. Passenger car engines in North America typically use a “four stroke” cycle, which represents the number of times a piston changes direction before the events in the process of powering the engine are repeated. Some diesel and spark-ignited engines use a “two stroke” cycle, but this is not common for passenger car applications because two stroke engines may provide higher emissions of unburned fuel. Some engines locate the camshaft and valves above the engine and others locate the valves within the engine block. Engines also differ with regard to the number of cylinders and the orientation of those cylinders, such as inline or V-shaped. Figure 1.2 provides an example of a typical V-6 (six-cylinder) engine. The working mechanism of a spark ignition engine is as follows: 1. Intake: One of the valves (the intake) in the cylinder head opens when the piston is near the top of the cylinder, and as the piston moves downward, air and fuel are injected and move downward with the cylinder. 2. Compression: When the piston begins to move upward again, both valves are closed, and the contents of the cylinder (vaporized fuel and air) are compressed. 3. Power: As the piston nears the top of the compression stroke, a spark plug fires and combustion of the fuel takes place. The burning fuel creates carbon dioxide, water vapor, and other compounds. As a consequence of this gas formation, pressure rises rapidly within the cylinder. The force of the combustion gases pushes the piston down again. 4. Exhaust: As the piston reaches the bottom of the power stroke, energy from the expanding gases has been transferred from the piston to the crankshaft via the connecting rod. At this point, the exhaust valve opens and the piston then rises and sweeps most of the combustion products out of
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Automotive Engine Oil
1-5
FIGURE 1.2 Cross-section view of a V-6 engine.
the cylinder. When the piston is near the top of the exhaust stroke, the exhaust valve closes and the intake valve opens. The sequence of events then repeats. Any given engine can have various numbers of cylinders and arrangements of those cylinders, but common arrangements are V-6, V-8, and inline 4. The “cam-in-block” engine type is becoming less prevalent than an overhead cam, since a cam-in-block engine requires stiffer springs and a higher load on the springs. Lighter loads on engine components tend to reduce both wear and energy consumption. Many diverse surfaces interact and have the potential to experience wear when converting the chemical energy of the fuel into the mechanical energy of the crankshaft. These interactions are described below.
1.1.2 Crankshaft to Crankshaft Bearing As the crankshaft turns, sliding occurs between the crankshaft and the engine block structure as well as between the crankshaft and the connecting rod. The load is transferred through journal bearings, which are designed to run primarily under hydrodynamic conditions. In this bearing interface, the engine oil acts mainly as a viscous fluid, and the friction in the bearings is directly related to the viscosity of the engine oil. Since the proper clearance between the shaft and the bearing is important for good engine performance, both the shaft and the bearings should be manufactured using strong and stiff materials, to minimize deformation. Vehicle engines, however, are often shut down for long periods of time. A shaft will then settle into contact with its bearing until the engine is started again. Also, solid particles (such as residues from manufacturing, contamination, wear, etc.) can be entrained in the engine oil, and these particles have
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
the potential to damage the shaft or bearing surface if the particles are larger than the minimum clearance between the shaft and bearing. Soft, compliant bearing surface materials minimize the sticking of the shaft to the bearing during shutdown, and such materials also can capture some small debris particles and remove them from circulation. This property of a bearing material to capture debris is called embedability [2]. To meet these contradictory requirements, crankshafts are usually made from a hard, stiff material such as cast iron or steel. The bearing is made using a steel backing (for strength and dimensional stability) and coated with a soft alloy (for embedability). For many years, lead-based alloys were used in crankshaft bearing applications. However, legislation now forbids the use of lead in many applications, and the low strength of the lead alloys limits the output of engines. Modern engine bearing coatings are usually made from aluminum-tin alloys, which are stronger but also have poorer embedability, so that engine and oil cleanliness become critical for long-term engine durability [2].
1.1.3 Piston Pin to Piston The piston pin transfers force from the piston to the connecting rod. The interface between the pin and the piston is also a type of journal bearing, but the motion in this case is not full rotation. In the fixed pin design, the pin is press-fit into the connecting rod, and the motion between the pin and the piston is fully reversed partial rotation. In the floating pin design, the pin is free to rotate within both the rod and the piston, and the motion is indeterminate. The floating pin has been shown to reduce the operating temperature of the piston pin boss and is therefore the preferred design [3]. In either case, the velocity of the pin is not sufficient to generate a full fluid film between the surfaces, and a condition of boundary lubrication results. The tribological properties in the pin to pin–bore interface are primarily controlled by the material properties of these parts. Automotive pistons are usually made from aluminum–silicon alloys. The piston pins are usually made from low or medium carbon steel, which is formed into a hollow cylinder and is then carburized. The carburization process results in very high hardness of the pin surface and helps minimize adhesion between the pin and piston. It has been demonstrated that increasing the oil supply to this interface reduces the tendency for scuffing [3].
1.1.4 Piston Skirt to Cylinder Block The piston skirt to cylinder block interface is one of the primary contributors to total engine friction [4]. The design challenge in this case is to maintain a small clearance between the piston and the block in order to avoid seizure, while minimizing noise and vibration [5]. The aluminum–silicon alloys used for most automotive pistons are lighter than the cast iron pistons of the past, which therefore reduces engine mass and vibration. Also, the higher thermal conductivity of aluminum helps prevent overheating of the top of the piston. Sometimes, however, the piston requires additional cooling, which is usually provided by adding devices to direct a jet of oil onto the underside of the piston. In this case, the engine oil is acting as a coolant. The cylinder bore is usually made from gray cast iron, which has a lower coefficient of thermal expansion than aluminum. This creates a design challenge, since a piston with adequate clearance at running temperature may be too loose (and hence noisy) at low temperature. To reduce friction and prevent scuffing of the piston, oil must be supplied to the cylinder bore walls. Nevertheless, the clearances are so tight that special coatings are applied to most pistons, such as nickel ceramic composites or molybdenum disulfide [6,7]. These coatings also reduce the friction in the interface of the piston rings and the piston skirt with the cylinder walls.
1.1.5 Piston Rings to Cylinder Block The piston rings function as a set of sliding seals that try to separate the combustion gases above the piston from the crankcase environment below. The most common arrangement is a set of three rings,
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Automotive Engine Oil
1-7
the upper compression ring, lower compression ring, and the oil control ring, as can be seen in Figure 1.1. The ring-block sliding interface has been estimated to account for 20% of the total engine mechanical friction [8]. Oil usually reaches the cylinder bore surface by being thrown from the crankshaft after flowing through the bearings. Some oil is necessary for the compression rings to function properly, but the oil that escapes past the compression rings is lost. The oil control ring ensures that only the necessary amount of oil reaches the compression rings. The upper compression ring experiences the highest loads and oil temperatures, and it must provide a good seal to the cylinder surface with very little engine oil. To provide acceptable durability, this ring is usually made either from nitrided stainless steel or from steel coated with molybdenum.
1.1.6 Camshaft to Cam Follower and Valve Train As the camshaft rotates, it presses against a flat or roller surface, which reciprocates to open and close the valves. The interface between the camshaft and follower is unidirectional sliding between nonconformal surfaces. Although engines are designed to provide oil to this interface, it is likely that oil will be scarce at times. For example, when starting a cold engine, the cams will begin turning before pressure is sufficient to pump oil to the top of the engine. Only a few material combinations are used successfully in this application, and even those wear sufficiently during the life of an engine to require periodic adjustment or the use of self-adjusting hydraulic elements. The severity of these various surface interactions is reduced by the presence of engine oil. The complex configuration of a typical valve train is illustrated in Figure 1.3. Since friction reduction is an important means for conserving energy and preserving nonrenewable fuel sources, techniques for reducing friction will continue to become increasingly important.
1.1.7 Oil Pump To assure an adequate distribution of lubricant through the engine and sufficient flow to maintain hydrodynamic conditions in engine bearings, automotive engines use a pressurized lubrication system. The oil flows in a circuit beginning with the sump, from which the oil is drawn into a pump. The pump then delivers pressurized oil through a filter, then to passages in the block and head, to the crankshaft and camshaft bearings, as well as to the hydraulic valve lifters in engines equipped thus. The oil is then thrown from the rotating components onto the cylinder walls, valve lifters, and other components. As the oil runs off these surfaces, gravity directs it back to the sump through passages in the head and block. Two types of pumps are commonly used, the spur gear pump and the gerotor, as shown in Figure 1.4. Examples of typical bearing and seal configurations are shown in Figure 1.5. Spur gear pumps are the older design and have the advantage of relatively quiet operation. However, the gerotor has the advantage of greater efficiency and can be made to take up less space in the engine compartment, so that most recent designs use the gerotor. The pump incorporates a relief valve for pressure regulation.
1.1.8 Oil Filter The oil filter is intended to remove potentially harmful particles from circulation. The filter element is usually either a pleated paper or metal mesh. Oil filters are rated by various tests, including industry standard methods (e.g., SAE 1858) and proprietary tests. The most commonly reported performance figures are for particle removal and flow restriction. It is desirable to have the highest level of particle removal with the lowest flow restriction. Since flow decreases as trapped material increases, filter systems generally have a bypass circuit included which opens when the pressure ahead of the filter reaches a predetermined level. This allows the oil to continue circulating if the filter becomes plugged, although contaminants are then allowed to circulate through the engine.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
(a) Valve keys Rocker arm
Valve spring retainer
Pushrod
Lifter/tappet Valve spring Camshaft
Valves
(b)
FIGURE 1.3 Valve train configuration and components: (a) valve train basic configuration; (b) valve train components — valves, seat, and guides.
1.2 Issues Related to Energy Consumption in an Engine: Service Effects The fuel provides the energy to maintain vehicle motion. However, the magnitude of the fuel consumption depends on a great number of factors. The extent to which energy is lost during operation of a given vehicle will vary with such characteristics as vehicle weight, engine type, component design, operating conditions, outside temperatures, type of terrain (flat or hilly), the number of below-freezing starts in which the engine oil never warms completely before shutdown, the viscosity of the engine oil at various
© 2006 by Taylor & Francis Group, LLC
Automotive Engine Oil
1-9 Suction
Suction
Output
Output
Spur gear
FIGURE 1.4
The two main types of pumps in internal combustion engines.
FIGURE 1.5
Engine bearings and seals.
Gerotor
operating temperatures, service history and age of the vehicle, and extent of component wear [9,10]. According to the information in Figure 1.6 (derived from an automotive database), friction in the engine, transmission, and axles represents approximately 11% of the energy consumed by a light-duty vehicle such as a gasoline-fueled passenger car. Within this 11% portion of energy usage, the piston skirt and piston rings contribute significantly to energy loss. Cooling and exhaust also represent a significant fraction of the energy loss. The severity of surface interactions between the moving components in an engine is reduced, to a greater or lesser extent, by the presence of engine oil. The oil provides different functions in different regions of the engine. Lubrication conditions are often subdivided into boundary, mixed, and hydrodynamic domains, according to the Stribeck curve (Figure 1.7), which also shows the lubrication regimes in which various engine components usually operate. Figure 1.7 indicates the relationship between the coefficient of friction (vertical axis) and a term consisting of the oil’s viscosity at a given operating temperature, multiplied by the relative difference in speed between the two surfaces, divided by the load that one surface exerts on the other. The range over which various engine components operate is indicated by the horizontal arrows. It should be noted that the vertical axis is drawn on a logarithmic scale, and the differences in friction would be greater if drawn on a linear scale. The low point on Figure 1.7 indicates the condition under which friction is a minimum
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
Distribution of energy losses in a typical light-duty vehicle
Exhaust 33%
Wheels 12%
Piston skirt friction 25%
Axle & transmission 22.5%
Air pumping 6% Braking & coasting 7.5% Engine friction 7.5% Axle & transmission friction 3% Accessories 4%
Crankshaft 5% Piston rings 19%
Valvetrain 6%
Cooling 29%
Bearings 22.5%
Research and development center
FIGURE 1.6
Typical values for energy loss in a light-duty vehicle. Boundary Coefficient of friction
1.
Mixed
Hydrodynamic
Piston rings
0.1
Piston skirt Valve train .01
Engine bearings
.001 Viscosity × Speed Unit load
FIGURE 1.7
Stribeck diagram, including the operating regions of several engine components.
(and thus fuel consumption will be minimized for a given vehicle). Engines do not operate at a constant temperature, vehicles sometimes drive on rough roads, and various additional conditions influence vehicle operation, so that Figure 1.7 represents a highly idealized assessment of friction effects. In the hydrodynamic region, the sliding surfaces are completely separated by an oil film, and friction is essentially due to shearing of the fluid. As the sliding speed and viscosity of the engine oil decrease and loads increase, the two opposing solid surfaces begin to interact. Moving to the left on the Stribeck curve, the coefficient of friction rises sharply as the load is shared between the fluid and the solid surfaces in the region identified as mixed lubrication. At some sufficiently low value of viscosity and component speed and at a sufficiently high load, the contact zone moves into the domain of boundary lubrication [1,2].
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Automotive Engine Oil
1-11
1.3 Engine Oil 1.3.1 Characteristics of Engine Oil and Functions of Its Additives The major component of engine oil is its base stock (i.e., the oil itself). Various additives are added to the engine oil, each of which provides a highly specific mode of action to protect the engine and reduce the rate at which the engine oil degrades. Zinc dialkyl-dithiophosphate (known as ZDP) is an essential additive in engine oil, and it has two functions: inhibition of oil oxidation and protection against wear. To protect the oil against oxidation, ZDP tends to react faster with oxygen than the rate of attack by oxygen on the oil base stock. In this way, the oil base stock and its other additives are less likely to be oxidized. In addition, ZDP reacts with iron on an engine’s surface (particularly in a heavily loaded contact) by laying down a phosphorus and sulfur coating that is resistant to wear. The phosphorous in ZDP can poison catalytic converters, which has contributed to a trend in recent years to reduce ZDP concentrations in engine oil. Thus, additional compounds that provide supplemental oxidation protection are generally also incorporated into an engine oil formulation. A detergent in the engine oil behaves somewhat like a soap, in that it reduces the tendency of partially oxidized oil to form tar-like deposits on a hot surface. A dispersant helps keep degraded oil from coagulating, so that the coagulated oil will not be able to block narrow lubricant passageways. A pour-point depressant allows the oil to flow at low temperature.
1.3.2 Viscosity Effects Appropriate engine oil viscosity is essential for satisfactory engine performance, but maintaining suitable viscosity over a temperature range that can extend well below 0◦ C and well above 100◦ C requires an additive in the engine oil (a viscosity index improver, typically called a “VI” improver) that helps to minimize the adverse consequences of large temperature fluctuations. A VI improver is a long-chain polymer that is less soluble in cold oil but more soluble in warm oil. When cold, the VI improver folds in upon itself and offers less resistance to oil flow. Thus, the VI improver facilitates cold starting of an engine. When the oil is hot, the VI improver expands into a loose coil, so that the viscosity of the engine oil increases over what it would otherwise be at an elevated temperature. This expansion and contraction effect may diminish as the VI improver ages and is broken down by high-shear conditions, which are likely to be experienced whenever engine oil passes through narrow, hot contact points such as in a heavily loaded bearing or underneath the piston rings. The oil container will display a term such as SAE 5W-30, in which the “5W” signifies the oil viscosity when the oil is cold; the “30” indicates the viscosity at normal operating temperatures. Viscosity requirements under various shear conditions for the different viscosity grades are established by the Society of Automotive Engineers (SAE) and are included in SAE J300, “Engine Oil Viscosity Classification.” The latest requirements are summarized in Table 1.1. Oil containers usually also display something relating to the performance capabilities of the oil. The two most common symbols indicating that engine oil satisfies a particular performance standard for gasoline engines in the United States are the API Certification Mark (starburst) and the API Service Symbol (donut).
1.3.3 Engine Oil Quality and Oil Degradation During Vehicle Use A container of engine oil, such as one would buy in a store or at a service station, should have a symbol that indicates whether the oil meets current standards (as indicated above). Unfortunately, some stores also carry engine oils that do not have a current designation, and an uninformed purchaser is at risk of buying an inappropriate grade of engine oil. In addition, overly degraded engine oil puts an engine at risk of damage. Examples of oil analysis tests that are helpful in determining the extent of engine oil degradation during use include changes in viscosity as described in various standard tests such as ASTM D 445, D 446, D 4683, and D 4684. (Note: all American Society for Testing and Materials [ASTM] standards cited in this
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1-12 TABLE 1.1
Handbook of Lubrication and Tribology SAE Viscosity Grades for Engine Oilsa
Low-temperature ◦ C cranking viscosityb (mPa sec, max)
Low-temperature ◦ C pumping viscosityc (mPa sec, max with no yield stress)
Low-shear-rate kinematic viscosityd (−mm2 /sec at 100◦ C, min)
Low-shear-rate kinematic viscosityd (mm2 /sec at 100◦ C, max)
High-shear-rate kinematic viscositye (mPa sec at 150◦ C, min)
6,200 at −35 6,600 at −30 7,000 at −25 7,000 at −20 9,500 at −15 13,000 at −10 — — —
60,000 at −40 60,000 at −35 60,000 at −30 60,000 at −25 60,000 at −20 60,000 at −15 — — —
3.8 3.8 4.1 5.6 5.6 9.3 5.6 9.3 12.5
— — — — — — <9.3 <12.5 <16.3
40
—
—
12.5
<16.3
50 60
— —
— —
16.3 21.9
<21.9 <26.1
— — — — — — 2.6 2.9 2.9 (0W-40, 5W-40, and 10W-40 grades) 3.7 (15W-40, 20W-40, 25W-40, 40 grades) 3.7 3.7
SAE viscosity grade 0W 5W 10W 15W 20W 25W 20 30 40
a 1 mPa sec = 1 cP; 1 mm2 /sec = 1 cSt. All values are critical specifications as defined by ASTM D 3244 (see text, Section 1.3). b ASTM D 5293. c ASTM D 4684: note that the presence of any yield stress detectable by this method constitutes a failure regardless of viscosity. d ASTM D 445. e ASTM D 4683, CEC L-36-A-90 (ASTM D 4741), or ASTM D 5481.
paper can be found in the Annual Book of ASTM Standards, available from ASTM International, West Conshohocken, PA.) Loss of antioxidant/antiwear protection as a consequence of exposure of engine oil to the exceedingly high heat and pressure of combustion processes can be measured using ASTM D 5483. Remaining oil alkalinity (including remaining corrosion protection) can be measured using ASTM D 2896 or D 4739. (Note: the authors have observed that ASTM D 2896 is particularly useful when attempting to compare the rates of engine oil degradation in a broad range of service types.) Accumulation of acids in the engine oil (due to incomplete combustion of the fuel or oil oxidation in hot spots) can be measured using ASTM D 664. The extent to which wear debris or corrosion residues have entered the engine oil (e.g., measurements of iron, copper, lead, aluminum) can be determined via analysis of the chemical elements in the oil. This type of analysis is particularly beneficial when one uses a nontraditional fuel or a new material in an engine application. That is, whenever engines are modified to meet new conditions, it becomes important to determine whether the engine oil degrades differently than was the case before the modification. If there are different mechanisms of oil degradation, tactics will have to be developed to understand those mechanisms and to find techniques to reduce any adverse consequences. Engine operating conditions can be particularly harsh on the engine oil compared with conditions experienced by most other types of automotive lubricants. An operating engine produces carbon dioxide, water vapor, nitrogen compounds, and partially burned oil or fuel, which then become pollutants. The effects of pollutants can be minimized by exhaust after-treatment. In reality, service effects can be more complex than the above description. The complexities arise from differences between one type of engine and another, the type of service that the engine is experiencing at a given instant, the nature of the fuel, the severity of the terrain, the weather conditions during use, and the extent to which the driver uses rapid accelerations. Service effects can be roughly categorized under four headings: easy freeway, high-temperature high-load service, taxi service, and extreme short-trip service at low outside temperatures [9,10]. Even though one or another of these service conditions may predominate for a given driver, in reality, most vehicles, at some point, are driven under each of these conditions.
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Automotive Engine Oil
1-13
The sequence of events that engine oil experiences in freeway service is approximately as follows. An adequate supply of oil is pulled up from the oil pan, filtered through an oil filter, and distributed to those engine components that require lubrication. Sufficient oil pressure needs to be available to permit bearings to ride on a fluid film. Engine oil is moved along an engine cylinder bore by the motion of the piston rings. The conditions within the cylinder are extremely harsh, since the fuel typically explodes several times a second, creating extreme heat and high pressures, and the fuel sometimes creates corrosive chemicals whenever combustion is not complete. The heat of combustion would be strong enough to partially oxidize the engine oil and form organic acids (i.e., oxidized hydrocarbons), except that the oil’s antioxidant (as long as it has not degraded) blocks oxidation and acid formation [11–14]. Once an antioxidant has degraded, the second line of defense against acid attack is the detergent, which is an alkaline agent that neutralizes acids that form during exposure of the engine oil to heat in the presence of oxygen. Thus, the detergent reduces the rate at which polar reaction products accumulate on engine surfaces, and therefore the detergent helps keep the engine’s surfaces clean. When the detergent has become inactivated during long-term vehicle use without an oil change and without oil refreshing via addition of make-up oil, engine corrosion (from acid attack) becomes more likely, and excessive deposits may form on engine surfaces. Since the oil’s antioxidant is also its antiwear agent, once the antioxidant is largely degraded, the wear rate in an engine can then accelerate. These harmful effects are highly unlikely in freeway service, except under conditions in which the engine oil has not been changed for far greater distances than recommended in owner’s manuals or when a driver has installed an engine oil of inferior quality. In many ways high-load service is similar to freeway service, except that oil temperatures and engine speeds are higher, and thus the rate of engine oil aging will be faster. In addition, oil viscosity is lower at high temperature, which means weaker oil films are present in high-temperature contact points. An increase in engine speed also means that a flame front will impinge on the cylinder bore more frequently than during service at slower engine speeds. Thus, higher engine speeds promote higher oil temperatures, faster oil degradation, and increased stress on an oil film [15]. These effects accelerate the wear rate, though use of an oil cooler can reduce some of the adverse consequences of high-temperature service. If the engine oil has not been changed soon enough in high-load service, so that the oil’s protective additives have become overly degraded and ineffective, various ill effects are likely to be observed. For example, the lighter ends of the engine oil will boil away faster than in freeway service, so that the engine oil’s viscosity increases. The antioxidant/antiwear agent in the oil will experience faster thermal degradation, which means that an increase in the engine wear rate may occur sooner once the oil’s antiwear agent is no longer effective. Oil thickening due to chemical interactions within the oil will increase. If high-load service continues long enough without an oil change or if an inappropriate engine oil has been used (e.g., an engine oil that has not passed the standard tests, but which is easily available to the public), the oil can become viscous enough that engine failure becomes a concern [15]. Chemical and physical effects of city driving (such as taxi service where the oil is completely warm but severe accelerations may occur) differ from those found during freeway or high-load service. For example, if a taxi is attempting to move through a large, congested city during rush hour, the taxi’s engine will be idling when the vehicle is at a red light. When the light turns green, the taxi will accelerate, but will move slowly whenever traffic is again blocked. A sudden acceleration tends to produce incomplete combustion of the fuel and formation of both organic acids and additional harsh chemicals. Periods at idle or at slow vehicle speed (so that the engine receives very little cooling from the wind) also promote formation of aggressive chemicals that can condense in or be formed in the engine oil. The effects resulting from extreme short-trip service during cold weather (e.g., all trips lasting only 5 or 10 min with outside temperatures below freezing) can be particularly harsh [15]. On start-up, the oil and the engine are cold. Once the engine has started, fuel condenses in the engine oil. Even if the weather warms or a longer trip is taken, so that the lighter ends of the fuel evaporate, the heavier ends of the fuel are likely to remain in the oil and cause the oil to have a lower-than-normal viscosity. Partially burned fuel (including organic acids and other reactive chemicals) also condenses in the oil. These fuelderived agents can degrade protective oil additives, attack the oil’s base stock, and attack some engine
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Handbook of Lubrication and Tribology
materials [11]. For example, organic acids derived from the fuel begin to neutralize the detergent and cause the antioxidant to become less effective. Acids can also modify an engine’s surface properties, so that engine materials are more easily removed by mechanical action (i.e., rubbing) than would normally be the case. Thus, these corrosive wear conditions accelerate removal of metal from rubbing surfaces during short-trip winter driving. Water-in-oil emulsions form a “white sludge” that also contains fuel, partially burned fuel, and oil additives. If an engine has not been turned on for several weeks in winter (after having been used in extended short-trip service), a layer of “white sludge” (containing measurable amounts of polar oil additives and water) may form and drop to the bottom of an oil pan. This condition of water in oil occurred during a test in which vehicles that had been driven exclusively on 3-km trips for 2 years (without having changed the engine oil) were left parked for 2 weeks in midwinter [12,13]. By the end of the 2 weeks, water and polar oil additives had settled to the bottom of the oil pan and had frozen. The oil uptake in the oil pan was totally blocked by the frozen water. Thus, the engine oil did not flow when the engine was started. The driver immediately turned off the engine when he noted a lack of oil pressure. Warmth applied to the bottom of the oil pan resolved the problem. However, such a scenario poses a risk of severe engine problems if a vehicle is driven very far under conditions in which the engine oil cannot flow up from the oil pan to lubricate the engine.
1.3.4 Fluid Film Lubrication The thickness of a fluid film, in many cases, plays a crucial role in the durability of an engine component. The film thickness and the oil film’s capacity to protect against wear, corrosion, and excessive friction are related to the viscosity of the bulk of the oil at the operating temperature, the oil temperature increase in a heavily loaded contact, the roughness of each of the surfaces in the contact zone, the speed at which one surface is moving relative to the adjacent surface, the presence or absence of any debris (such as honing burrs on the cylinder of a newly manufactured engine or dust in the engine oil), and the extent to which the lubricant has degraded. The “fluid film ratio” is a useful measurement determined by comparing the oil-film thickness to a term representing the roughness of each of the mating surfaces. If the oil-film thickness is of the order of magnitude of the surface roughness, undesirable wear may result. To avoid this possibility, a designer can modify the surface roughness, increase the viscosity of the lubricant, change engine design to supply more oil to the contact zone, reduce load, or upgrade the oil’s additive package (all of which may be difficult to modify if at the same time acceptable lubrication is still to be provided to other regions in the complex engine structure). This assessment assumes that the lubricant wets the surface of interest. If a surface material and its lubricant are not mutually attractive (i.e., the fluid beads up on a surface rather than spreads spontaneously over it), a mating surface has the potential to wipe away the lubricant rather than use the lubricant to form a lubricating film. An accelerated wear rate then occurs. The following simple test can determine the wetting characteristics of a surface and any fluid that may come into contact with that surface. A drop of lubricant is placed on a level portion of the surface. If the drop beads up, especially if the edge of the drop forms a steep angle with the surface (such as 80 or 90◦ ), one can assume that the contact is nonwetting, or nearly so. An even more surprising effect related to a nonwetting surface can sometimes be identified. A drop of lubricant is smeared to form a thin film over a clean sample of the surface of interest. Next, the point of a pin is lightly dragged a distance of a centimeter or more over the middle of the thin layer of fluid on the surface. If the fluid is not attracted to the surface, the fluid can recoil from the path over which the tip of the pin has moved, so that a hole forms in the fluid film. This nonwetting condition is not desirable in a lubricated contact. Such concerns must be addressed whenever one considers utilizing fuels other than hydrocarbons. Simple compatibility tests can often be conducted in advance to determine whether alternative fuels may influence the durability of engine components (seals in particular) or diminish the ability of the engine oil to provide suitable lubricating films. These examples illustrate that it is not always easy to replace one automotive material or fluid with an alternative substance. Unexpected consequences may result.
© 2006 by Taylor & Francis Group, LLC
Automotive Engine Oil
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1.3.5 Future Concerns Ongoing engine-related concerns include such issues as reduction in exhaust emissions, expense and effort required in the development of standard tests for engine oils, and creativity needed to envision zeropollution vehicles of the future. Examples are shown in Figures 1.8–1.10. Figure 1.8 illustrates the current emission standards that designate the allowed values for nitrogen oxide emissions up to the year 2007. Limits on particulate emissions are also shown in Figure 1.8. The metric used in the vertical scale of Figure 1.8 is grams per brake horsepower-hour (g-bhp/hr). Figure 1.9 describes lower nitrogen oxide emissions and better volumetric efficiency can be achieved by using a cooled exhaust gas recirculation
Allowed values, g-bhp/h
7
NOx
6
Particulates
5 4 3 2 1 0 1988
1991
1994
1998
2002
2007
Year
FIGURE 1.8 Emissions requirements for diesel engines. (a)
EGR recirculation duct
Intake manifold
Exhaust manifold
(b) 15% cooled exhaust gas
Potential corrosion 25–35% more heat to coolant Potential bearing wear
FIGURE 1.9 Lower nitrogen oxide emissions and better volumetric efficiency when using a cooled exhaust gas recirculation system: (a) exhaust gas recirculation system; (b) exhaust gas recirculation system wear control challenges.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology Number of tests to qualify products 16
15 12
12 ing reas
8 4 0 API:
FIGURE 1.10
y
ualit
oil q
8
Inc
5
5
CE
CF-4
2
CD
CG-4
CH-4
CI-4
Improved engine protection when using exhaust gas recirculation in engines with CI-4 lubricants.
system. However, as shown in Figure 1.9(b), it also describes some of the consequences of using exhaust gas recirculation. Figure 1.10 shows the increase in the number of required standard tests for certifying performance of engine oils in diesel engines. Each additional test represents increased expense to the lubricant manufacturer. In addition to efforts to make engine oils more environmentally friendly, many automobile manufacturers have now incorporated devices into their vehicles that indicate the point at which engine oil should be changed. In some cases, sensors of various types are used (e.g., acid, base, or temperature sensors, along with a computer assessment of the point at which a given reading has become excessive). In other cases, a micro processor determines the rate at which the engine oil has degraded, signal is given to the driver to “change oil soon,” and at a slightly later date the driver will be provided with a “change oil now” warning. Such systems can both reduce the chance of harm to the engine and, for a driver who spends most of the time on the freeway, greatly extend the point at which the engine oil needs to be changed. Examples presented in this section indicate the complexity of simultaneously attempting to reduce energy consumption, reduce friction, prolong the life of the engine oil, minimize wear and corrosion in an engine, identify environmentally friendly power sources, reduce the amount of polluting substances that enter the environment after lubricant disposal, but at the same time still retain personal mobility. Insight, continuing effort, and creativity are required in these domains. One can assume that efforts will continue to be aimed at enhancing fuel economy, reducing pollution, and exploring various alternatives to gasoline and diesel fuel.
1.4 Gasoline Engine Oil Performance Categories and Associated Test Methods 1.4.1 Introduction In the early years of automobile use, engine oil had to be added or changed after an exceedingly brief interval. In addition, oils were not standardized. Individual vehicle operators could be at the mercy of their intuition with regard to the purchase of an appropriate automotive lubricant for their cars. With time, people began to realize that something had to be done to avoid the possibility of serious adverse consequences if a driver had used the wrong kind of oil for an engine. Thus, there was strong motivation to look for chemical agents that both provided protection to an engine and promoted long life of the engine oil. This trend (toward improving engine and oil durability and using additives in the oil to provide specific beneficial attributes) is ongoing. Promoting environmental acceptability has also become an essential ingredient of responsible engine oil formulation. Throughout these developments, it has been essential to create and conduct appropriate engine oil test methods that ensure oils available to the public produce appropriate engine protection. Various classes of standard tests are available to confirm that current automotive engine oils provide the desired protection, including long oil life, corrosion and wear protection, resistance to the formation of
© 2006 by Taylor & Francis Group, LLC
Automotive Engine Oil
1-17
sludge and deposits, ability to remain within an appropriate viscosity range, etc. ILSAC (the International Lubricant Standardization and Approval Committee) and API (the American Petroleum Institute) are two organizations that play a major role in overseeing the availability of standard engine-oil-related test methods. The American Society for Testing and Materials (ASTM) is typical of the organizations that publish standard procedures to be used when conducting an automotive test. Such standard tests are prepared in painstaking detail, so that there will essentially be no chance of conducting a standard test incorrectly if one has followed the written directions. Automotive companies tend to be the developers of such tests. In general, different tests are used for gasoline-fueled engines than are used for diesel engines. Tests typically need to be updated periodically, for various reasons. In some cases, test components, such as a specific type of engine, may no longer be available. Changes in the chemical nature of the fuel, such as the transition from leaded to unleaded fuel, may mean that a former test is no longer pertinent to current engine wear, corrosion, and sludge characteristics. Future engine designs that differ from current test engines mean that standard tests will have to be created using the newer types of engines, since the older engines may not be predictive of current performance. If it is possible to legitimately substitute a bench test for an engine test (such that the fundamental mechanisms of oil and engine damage correlate strongly with the results from the bench test), the bench test becomes far less labor intensive and expensive. A brief overview of the evolution of standard engine oil test methods and the status of current automotive engine oil test method development is provided in the following paragraphs. Early test methods for engine oils were far less sophisticated and less specialized than the tests of today. It can be anticipated that the tests of the future will be even more specialized. Wherever possible, bench tests will be substituted for engine dynamometer tests, such as was the case in the development of the Ball Rust Test, a bench test that replaced the Sequence IID (i.e., Sequence 2D) engine test, which measured the ability of an engine oil to protect against the kind of corrosive damage that can occur during extended short-trip winter service in which water and corrosive chemicals (derived from the partial combustion of the fuel) enter and remain in the engine oil for extended periods and cause engine corrosion. At the fundamental level, oil analyses can determine whether a given engine oil has all the required additives in its formulation (and thus is not deficient, such as an “SA” quality oil would be). Such information can be pertinent to engine durability field problems, since most vehicle warranties are invalidated if the wrong grade of engine oil has been used in an engine. As of early 2004, the only two designations widely used to describe light-duty, gasoline engine oil performance were API SL and ILSAC GF-3. Later in 2004, API SM and ILSAC GF-4 oils became available in the marketplace. The engine test and bench test performance requirements for API SL are similar to those for ILSAC GF-3, but, in addition, ILSAC GF-3 oils must also meet energy conserving requirements. Similarly, API SM requirements as well as energy conserving requirements must be passed before an engine oil can be designated as ILSAC GF-4. The test methods for engine oils must be in accordance with the requirements outlined in the American Chemistry Council (ACC) Product Approval Code of Practice. These requirements include registration of all tests, use of only calibrated equipment and facilities, and guidelines for acceptable modifications during program development. These requirements were implemented when the API SH and ILSAC GF-1 designations for engine oil were adopted in 1993, and the requirements have been continued as new performance categories have evolved.
1.4.2 ILSAC GF-4 and API SM Standard Tests In January 2004, ILSAC issued its latest Minimum Performance Standard for Engine Oils, ILSAC GF-4. Compared with GF-3 (the previous engine oil category), oils meeting GF-4 requirements provide improved oxidation resistance, improved high-temperature deposit control, better cam and lifter wear discrimination, improved low-temperature wear protection, and improved low-temperature used-oil pumpability. ILSAC GF-4 oils also have reduced phosphorus and sulfur contents to provide enhanced emissions system protection and to help vehicles meet the stringent Tier 2 Bin 5 emissions standards, which require, among other things, that vehicles emit no more than 0.07 g/miles (0.045 g/km) of nitrogen oxides over
© 2006 by Taylor & Francis Group, LLC
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120,000 miles (190,000 km) of driving. GF-4 oils also provide improved fuel efficiency for both new and used oils, compared with GF-3 oils. GF-4 oils began to be marketed during the second half of 2004, and all oils licensed to display the API Certification Mark (starburst) must meet GF-4 requirements by April 30, 2005. The companion S category to GF-4 engine oils, designated API SM, was defined by the API Lubricants Committee. The API SM category includes the same performance requirements (except for fuel efficiency) as ILSAC GF-4, for those viscosity grades defined by GF-4 (i.e., SAE 0W-20, SAE 5W-20, SAE 5W-30, and SAE 10W-30). For other non-ILSAC viscosity grades, some other differences between API SM and ILSAC GF-4 requirements exist, as outlined in API 1509, “Engine Oil Licensing and Certification System,” latest edition. Descriptions of the standard tests for ILSAC GF-4 and API SM engine oils follow. Although the performance limits in many of the engine and bench tests in ILSAC GF-4, as well as the chemical compositional requirements, were modified to achieve the benefits described previously, there was only one new engine performance test developed for GF-4 (the Sequence IIIG Test, which replaced the Sequence IIIF Test, ASTM D 6984). The IIIG Test utilizes the same General Motors 3800 Series II engine used in the IIIF Test, but the IIIG Test has different operating conditions and uses retrofitted valve train metallurgy. The measured parameters in the IIIG Test include average cam plus lifter wear, end-of-test kinematic viscosity increase, and a composite assessment of piston deposits. An end-of-test oil sample from the IIIG Test is also evaluated for its low-temperature engine oil pumpability characteristics (ASTM D 4684). In addition, the test length was increased to 100 h (from 80 h in the IIIF Test), engine load was increased from 200 to 250 Nm, and sampling and additions of make-up oil were minimized to increase the severity of the IIIG Test. Oil sump temperature was actually decreased from 155 to 150◦ C in IIIG (a decrease in test severity), because of concerns over abnormal depletion (degradation) of the engine oil’s antioxidant/antiwear agent, ZDP, at temperatures above 150◦ C. The Sequence IIIG Test retains the same alloy-cast-iron lifters used in the IIIF Test, but in the Sequence IIIG Test, the camshaft is phosphated (with a manganese phosphate coating) to minimize scuffing during break-in of the test engines. Thus, the Sequence IIIG Test addresses the issues that were of concern at the time of its inception. As conditions and issues evolve, it can be anticipated that this test (and other test methods) will evolve to meet future needs. The Sequence IVA (i.e., 4A) Test mimics city service and determines whether the engine oil provides sufficient wear protection to an overhead cam and slider followers. The Sequence VIII (i.e., 8) Test measures the extent of shear of the viscosity index improver. In addition, the Sequence VIII Test determines whether the engine oil provides sufficient protection to copper-lead bearings when using unleaded fuel. The previously available test (L-38) used leaded fuel, and thus the L-38 test is no longer appropriate for vehicles using the current unleaded fuels. The Sequence VG Test (i.e., 5G) addresses some of the issues related to partial replacement for the Sequence VE Test (ASTM D 5302). Sequence VG measures the sludge and deposit control tendency of engine oils under engine conditions that simulate stop-and-go city service in vehicles. The Sequence VIB (i.e., 6B) Test replaces the Sequence VIA Test (ASTM D 6202) for measuring the fuelefficient properties of an engine oil. Like its predecessor, the Sequence VIB Test measures the improvement in fuel efficiency of a test oil compared with an ASTM standard reference oil. Unlike its predecessor, however, the Sequence VIB Test not only measures the fuel efficiency of the oil when it is relatively new (after only 16 h of aging in the engine), but also the fuel efficiency after 96 h of aging, which corresponds to about 4000 to 5000 miles (6400 to 8000 km) of vehicle operation. Different levels of fuel efficiency improvement are required, depending upon the SAE viscosity grade of the engine oil (the same groupings of viscosity grade as were defined in the ILSAC GF-2 requirements for the Sequence VIA Test). Sequence VIB fuel efficiency requirements apply only to ILSAC GF-3 and GF-4 oils, not to API SL or SM oils. The Ball Rust Test is a bench test that mimics the effects of extreme short-trip winter driving. It replaced a previously used engine dynamometer test, and thereby saves considerable expense, time, and effort in the testing process. In the test, an engine component (ball) is immersed in a fluid that contains engine oil to which has been added the kinds of corrosive chemicals that are generated from incomplete combustion of the fuel when the oil and the engine are very cold (e.g., organic acids and other oxidized compounds). At the end of the test the extent of rust formation is evaluated electronically.
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As can be seen, test methods for engine oils can be complex, time consuming to develop, and expensive to run, and they need to be revised whenever engine designs have changed (e.g., as a consequence of environmental issues, including modifications to fuels, lubricants, or engine materials). Thus, the upgrading of standard tests is an essential and ongoing effort to ensure that new materials, engine design, fuels, and government mandates related to vehicle operation are adequately addressed.
1.5 Diesel Engine Oil Performance Categories and Associated Test Methods 1.5.1 Service Effects of Diesel Engines Related to Engine Oil Degradation An overview of service effects on engine oils used in diesel engines includes the following. Under freeway driving conditions, in the absence of hilly terrain, a modern diesel engine operating under a light load will move along the freeway with no (or very little) visible evidence of generation of soot. The rate of engine oil oxidation is influenced primarily by the engine oil volume, oil temperature, engine characteristics, engine speed, and load. Thus, under light-duty freeway conditions, engine oil life will be a maximum for a given engine design, oil sump volume, and engine oil temperature, as was the case for gasoline-fueled vehicles. In contrast, if the same diesel vehicle is heavily loaded and driving up a long, steep incline (such as can be found in mountainous regions) or experiencing stop-and-go city driving with frequent stops followed by heavy accelerations, the rate and extent of engine oil degradation will increase, and soot may form during an acceleration. The harder and hotter an engine works, the faster the engine oil’s antioxidant degrades, the engine oil alkalinity decreases, and the engine oil acidity increases. City driving produces lower vehicle speeds with frequent stops and starts, which generate the potential for both soot formation and an increased rate of engine oil degradation. Thus, two major differences between service with gasoline and service with diesel fuel are the generation of soot in diesel engines and the fact that diesel engines in North America do not normally experience short-trip cold-start driving [16,17]. Mandates for the reduction of sulfur in diesel fuel, starting in 2006, should help promote longer engine oil life, since fewer acidic reaction products should be generated from the fuel.
1.5.2 Examples of Test Methods for Diesel Engine Oils Both the physical and chemical properties of diesel engine oil change during use. Standard tests for engine oils used in diesel service focus on those conditions that may produce damage to the engine or the oil, or may cause oil-related engine failure, as was also the case with gasoline fuel. The various active diesel engine oil performance categories include at least two (and typically more than two) engine performance tests that must be conducted to demonstrate compliance with category requirements. Various bench tests are also included. A description of the standard tests for diesel engine oils can be obtained from such organizations as the ASTM or the API. Examples of test methods that determine the physical and chemical properties of engine oils used in diesel applications include the following. The volatility of the engine oil should not exceed 15% in the Noack Volatility Test (ASTM D 5800). Shear stability of the engine oil is measured using ASTM D 6278, and the limiting values for shear stability depend on the initial viscosity designation of the engine oil (e.g., a 15W-30 engine oil will be compared with 15W-30 a standard oil, and a 15W-40 engine oil will be compared with a 15W-40 standard oil). The low-temperature pumpability of used diesel engine oil is measured in a mini-rotary viscometer at −20◦ C, using a test that is a modified version of ASTM D 4684. High-temperature, high-shear characteristics of the engine oil can be determined using ASTM D 4683. The ability of the engine oil to resist forming foam is determined by ASTM D 892. The oil’s capability to control aeration (i.e., ability to allow bubbles in the oil to escape at a sufficiently fast rate) is confirmed with the EOAT (Engine Oil Aeration Test). The extent of engine oil thickening due to the accumulation of soot in the engine oil can be measured by the Mack T-8E Test, ASTM D 5967. The engine oil’s capability
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to inhibit corrosion of bearings is measured by ASTM D 6594, in which metals of interest include copper, tin, and lead. Bench tests that are much simpler than standard engine tests can provide useful insights into the nature of interactions between contaminated engine oil and wear of engine surfaces. For example, bench wear tests have documented the role that diesel soot can play in increasing the wear rate of engine materials [16]. Since vehicle service modifies the properties of engine oil in a variety of ways, bench tests with fresh oil do not necessarily provide useful information about the characteristics of used engine oil. Measurements of the changes in physical and chemical characteristics of diesel engine oil tend to be a reflection of both the nature of the service as well as the characteristics of the engine. Engine tests related to performance of engine oils in diesel service look for stress or failure of the engine or its components, such as extent of protection against wear in both sliding and rolling contacts, the extent to which corrosive wear occurs, and unacceptable changes in oil properties during severe service such as soot loading, acid formation, and viscosity increase. Several brief descriptions of standard tests that have been used in currently active API performance categories for diesel engine oils (i.e., API CG-4, CH-4, CI-4, CF, and CF-2) are as follows. The Detroit Diesel 6V92TA Test lasts 100 h and measures oil volatility, wear, and protection to the engine. Various 8-h segments are interspersed with 3-h shutdowns. The engine speed and load are specified for the different segments of the test. Scuffing of cylinder liners is noted, and pistons and rings are inspected for wear or other forms of damage. Limiting values for various measurements are provided, to determine whether the engine oil has exceeded its ability to adequately lubricate and protect the engine. For example, the maximum allowed port plugging in a given cylinder is 5%, and the overall average port plugging should be no more than 2%. Required engine oil analyses include a measurement of wear metals as well as additive metals. If the concentration of additive metals has increased, this provides a direct measure of the amount of oil volatility that has taken place. That is, if a metallic element such as calcium, which is part of the engine oil formulation, has increased in concentration by 5%, that means approximately 5% of the oil’s base stock has evaporated. The Caterpillar 1M-PC Test is conducted using a Caterpillar 1Y73 single cylinder indirect injection engine and measures scuffing of rings, pistons, and cylinder liner, piston deposit formation, and piston ring sticking. The test is part of the requirements for API CF and API CF-2. The Cummins M11 EGR Test is part of the API CI-4 category and measures ring and overhead wear, extent of filter plugging, and sludge formation, as related to exhaust gas recirculation (i.e., EGR). Fiftyhour segments of the test are conducted overfueled (i.e., using more fuel than would be required during the service conditions of the test) at an engine speed of 1600 rpm and alternated with 50-h segments at 1800 rpm overfueled with retarded timing. The total test duration is 300 h. Engine measurements include ring and overhead wear, filter plugging, sludge formation, weight loss of engine crossheads, and sludge formation on the engine valve covers and in the oil pan. Oil analysis measurements include viscosity, base number and acid number, concentration of additive elements in the engine oil, and accumulation of wear metals in the engine oil. For example, under high-temperature operation zinc concentration in the engine oil (from ZDP) will increase when the more volatile engine oil hydrocarbons have evaporated. Under severe service conditions, once the engine oil’s additives are no longer effective, concentration of wear metals such as iron will also increase. The Engine Oil Aeration Test uses a 1994 7.3 L V-8 engine. The test is conducted at 215 brake horsepower at 3000 rpm. The amount of air in the engine oil is determined at 1, 5, and 20 h during the test. Wear metals are determined at the start of the test and at 20 h. No more than 10% air is allowed for API CG-4 oils. Additional standard tests measure characteristics such as roller follower wear and the influence of soot in the engine oil on engine wear. The Roller Follower Wear Test (for categories API CG-4, CH-4, and CI-4) uses a General Motors 6.5 L, indirect injected diesel engine. The engine speed during the test is 1000 rpm at near-maximum load. The test duration is 50 h, and make-up oil is added at 25 h. Oil gallery and coolant-out temperatures are controlled at 120◦ C. New roller followers are installed at the beginning of each test. At the end of the test, roller followers are removed and the extent of wear is measured. Oil analyses are conducted at
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40 and 100◦ C. The alkalinity (total base number) of the engine oil is determined, as are wear metals and additive elements. The maximum allowed wear on the roller follower axle is 11.4 µm for CG-4 oils and 7.6 µm for CH-4 and CI-4 engine oils. Various additional tests for engine oils used in diesel service are also available. The above discussion documents most of the test methods currently deemed necessary for ensuring acceptable engine oil performance in diesel engines. Proposed future regulations regarding emissions of nitrogen oxides and particulates (such as soot) for future diesel engines mean that current engines will need to be modified to meet requirements for the year 2007 as shown in Figure 1.8. The necessary modifications may include catalysts and particulate traps. A very low sulfur concentration in the fuel will also be required. An exhaust gas recirculation system, as shown in Figure 1.9, has the potential to increase the temperature of an engine and its engine oil. Higher engine and engine oil temperatures mean greater susceptibility to increased wear and a faster rate of engine oil degradation. Faster oil degradation has the potential to accelerate the reduction of the corrosion–inhibition capabilities of the engine oil. In addition, engine oils that provide benefits for 2007 model engines may not necessarily provide the desired benefits for previous engines. Thus, it becomes worthwhile to explore any areas in which uncertainties exist.
1.5.3 A Model for the Rate of Engine Oil Degradation in Diesel Engines To investigate the fundamental mechanisms of engine oil degradation in diesel engines, a number of tests were conducted with a vehicle having a medium-duty diesel engine, covering a full range of service conditions. An extensive number of computerized measurements were taken, and oil samples were collected and analyzed, including soot generation, remaining antioxidant in the engine oil, and oil acidity and alkalinity. Then, when attempting to generate a mathematical model for the rate of engine oil degradation in diesel engines, it became necessary to include a term for both soot generation and a term for the rate of degradation of the engine oil’s antioxidant, because these two measurements interacted with each other, according to the results of the testing that had been conducted. Thus, to provide the desired statistical correlation, the two terms had to be multiplied together, which strongly confirmed an interaction between the effects of soot generation and engine oil degradation. The results of this testing are described in U.S. Patent 6327900 [17].
1.6 Future Directions 1.6.1 Concerns Related to Conservation of Fuel In several countries, Government regulations mandate improvements in fuel economy. Vehicle weight reduction is one way to address this mandate. If aluminum components are used in an engine or in any automotive wearing surface, it becomes important to identify a design such that aluminum oxide (which can form on an aluminum surface) is not rubbed off and allowed to enter the lubricant, because aluminum oxide is abrasive and can act as a severe polishing agent, which will increase the wear rate. Lubricant additives that are optimal for use with an iron surface do not necessarily provide the same wear and scuffing protection for an aluminum surface. Aluminum engine blocks with cast iron cylinder bores must be designed such that differences in the rate and magnitude of thermal expansion of these metals do not cause unacceptable gaps to open between contacting surfaces.
1.6.2 Effects at the Molecular Level When considering interactions between one material and another, or between materials and their lubricants, conventional lubrication wisdom does not necessarily provide a complete understanding of interactions at the molecular level. Investigators believe that it may be feasible to create a number of beneficial effects if nano-materials can be optimized for specific applications. Desired applications include
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materials optimized for such attributes as friction reduction, optimum hardness, scuffing resistance, enhanced strength, and thermal stability.
1.6.3 Insights Gained from Tests with an Alternative Fuel Hydrocarbon fuel supplies from fossil sources are finite. To ensure future mobility, various alternatives to hydrocarbon fuels have been tested, such as solar power, batteries, hydrogen, alcohol fuel (such as methanol or ethanol), and “flexible fuels,” which may contain up to 85% alcohol, but which also incorporate enough hydrocarbon (at least 15%) to permit a cold start. Suitable performance has been obtained in most cases. However, each type of alternative fuel may also have its unique disadvantages. In addition, the transition from one fuel to another often requires important modifications to any automotive materials that touch the fuel or its reaction products, in order to avoid incompatibility problems. Thus, utilization of alternatives to hydrocarbon fuel typically requires a significant development effort to ensure appropriate durability of materials. A number of years ago, at a time of heightened interest in finding a substitute for gasoline, experiments using methanol-containing fuel were carried out by various car companies. The fuel (termed M85) consisted of 85% methanol and 15% unleaded gasoline. City and freeway driving tests were conducted on a chassis dynamometer, so that each vehicle on test experienced exactly the same road and weather conditions as the other test vehicles [18]. A flexible-fuel vehicle that ran on gasoline was also part of the test, so that a quantitative assessment of the effects caused solely by differences in the fuel could be obtained. The findings from that test were as follows. In freeway service there was no difference in the rate of engine oil degradation or engine damage between the methanol fuel and unleaded gasoline. This suggested that under conditions in which the combustion of the fuel is complete, the nature of the fuel was not a factor in the extent of engine oil degradation. In city service, the engine oil used with gasoline degraded approximately 2.5 times faster than oil in the methanol-fueled vehicles that had been tested under identical conditions. That is, methanol was significantly milder to the oil than gasoline during city driving. The reason for this difference in severity was because the molecular weight of methanol is considerably lower than that of gasoline, so that the products of partial combustion of methanol boiled out of the oil and thus were not available to inactivate the engine oil additives. The investigators were surprised by this result, since at that time it was assumed that alcohol effects on engine oil would be more aggressive than those of gasoline under all types of service conditions. In extremely cold short-trip winter service, methanol was harsher to the oil than was gasoline, since combustion of methanol produces approximately twice as much water per kilometer of service than does gasoline. Toward the end of short-trip testing with methanol fuel, the engines were being lubricated with a mixture that contained less than 50% engine oil and slightly more than 50% contamination (water, fuel, and fuel reaction products). When gasoline fuel entered the engine oil during short-trip winter service, the viscosity of the engine oil decreased. When methanol fuel entered the engine oil, the methanol formed an emulsion, which caused the viscosity of the oil to increase. Even though extreme short-trip driving in a winter climate is not representative of most trips, it is desirable to learn about any potential problems before the vehicles are in the hands of the general public. These results indicate that when using a non-traditional fuel, an investigator must confirm that engine materials are not endangered. An additional concern that needs to be explored when using alternatives to gasoline is to make sure that engine and seal materials are compatible with the alternative fuel of interest. For example, in studies conducted by exposing polyacrylate, silicone, and nitrile seals to methanol fuel, it was found that methanol was able to extract beneficial protective additives out of some seals, so that the susceptible seals might become prone to hardening [19]. To conduct such a seal test, a thin segment of the elastomer of interest is immersed in the desired test fluid (here, M85 fuel) and allowed to remain in contact with the fluid at a temperature of interest for whatever duration the investigator deems important. Analyses can determine whether beneficial additives have been extracted out of the elastomer by the test fuel. In addition, an alternative fuel may enter a seal and soften it. Thus, any seal that had become softened by an alternative
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fuel would be at risk of accelerated wear or failure. Even though some types of seals had become soft in the short-trip study with methanol fuel, seal failure did not occur. Solubility relationships between engine materials and engine fluids can usually be identified in handbooks or tables incorporating titles such as“solubility parameters”or“cohesion parameters.” The solubility parameters include three numerical terms. One of the terms indicates the extent to which the substance of interest is soluble in hydrocarbons such as gasoline. A second term indicates the extent to which the substance of interest is soluble in mildly polar materials such as a compound that has a chain of several carbon atoms linked to a polar atom such as chlorine at one end of the hydrocarbon chain. A third term indicates the extent to which the molecule of interest is soluble in a hydrogen-bonding fluid such as water. If any engine material interacts adversely with a fuel of interest, either the engine material needs to be replaced or the material needs to be coated or in some way protected. Use of these solubility parameters to determine material compatibility before attempting a vehicle test with an alternative fuel can sometimes allow an investigator to avoid an engine failure. If piston rings have a molybdenum fill (or another kind of potentially removable fill), it is worthwhile to conduct a bench test to determine whether the fuel interacts with the fill material. In the case study cited above, M85 fuel was capable of removing the molybdenum fill in piston rings, which caused the rings to contact the cylinder bore on the sharp edges at the center of the rings. The sharp edges of the rings (where the molybdenum was removed) contacted and abraded the cylinder bore. It is also important to determine that bearings are not damaged in the presence of an alternative fuel. A simple immersion of the material of interest in the alternative fuel, followed by analysis of the elements that have entered the fuel by being extracted from the engine component, provides an assessment of compatibility. In addition, the amount of fuel in an elastomer sample should be measured. If analysis indicates a problem, it becomes prudent to identify an alternative material composition for the component that exhibited the incompatibility. These driving tests with an alternative fuel provided results that were sometimes at odds with conventional automotive wisdom. Because of this divergence, the investigators were able to gain an enhanced understanding of several mechanisms of oil degradation and their relationship to engine wear and corrosion for both gasoline and alcohol fuels. In addition, these results highlighted the fact that it may be risky to make assumptions about durability of engine materials in both city and short-trip service, when using an alternative fuel, since the alternative fuel may behave differently than hydrocarbons when in contact with engine materials. The kinds of issues related to alternative fuels described above need to be investigated, understood, and resolved before alternative fuels and the engines in which they are used are placed in the public domain. Useful information in this regard can sometimes be gained from bench tests (e.g., wear tests), but unanticipated interactions between the fuel and engine oil may occur under actual operating conditions that may be outside the domain of a simplified or single parameter bench test. In addition, valuable insights regarding fundamental causes of various engine effects may be derived when comparing test results from different kinds of fuels. These tests indicate that, by paying attention to material compatibility and recognizing the needs that are specific to a given fuel, alternatives to gasoline can become highly successful automotive power sources. If at some future date supplies of hydrocarbon fuels begin to dwindle significantly, many major automobile companies already have test results in their archives that will permit them to successfully utilize alternative fuels. Ethanol fuel is currently in use in Brazil.
1.6.4 Prolonging the Working Life of Engine Oil A decade or two ago, many North American vehicles owner’s manuals indicated that a driver should change engine oil at 5,000 km (3,000 miles) under most driving conditions other than freeway service, 12,000 km (7,500 miles) was a recommended North American oil-change interval for freeway or longerdistance service, but possibly only 10 or 20% of the drivers drove under conditions that met the criteria for the longer oil-change interval. Oil-change intervals for vehicles that were developed to use “synthetic”
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engine oil were approximately 2 times longer during freeway or autobahn service than was the case for a normal mineral oil. Even though it was well known that oil quality and service type greatly influenced the rate of engine oil degradation, it has only been within the last two or three decades that automobile manufacturers have begun to provide an in-vehicle warning to the driver that an oil change is needed. Such a warning system is now available on a significant fraction of current-model vehicles. Some of these warning systems include a sensor that may determine whether an engine oil has become excessively acidic or whether the engine or oil has exceeded a reasonable value for some other parameter such as oil temperature or viscosity. Other systems incorporate a computer model that calculates the rate of engine oil degradation based on measurements that the vehicle manufacturer believes are important. Examples include such values as the temperature of the engine oil and the number of times the oil has been exposed to a combustion event. No matter which technique has been used as part of an oil-change indicator system, the end result is that, in general, the oil-change interval indicated by the warning system often is longer than the values that had previously been listed in an owner’s manual, since the presence of a sensor or a model eliminates much of the uncertainty that a driver might have regarding the appropriate point at which to change engine oil. It also reduces the probability that a driver may completely neglect to change the engine oil [15,17]. Computer monitoring of all aspects of vehicle operation is likely to continue expanding in the future and is expected to extend oil-drain intervals and prolong the availability of oil supplies.
1.6.5 Minimizing Emissions and Pollutants and Ensuring Backward Compatibility Two important approaches toward minimizing automotive emissions are (1) improving the efficiency of the engine so that fewer pollutants are produced during the combustion process and (2) reducing the amount and type of chemicals in the fuel, engine, and engine oil that can adversely influence the effectiveness of a catalytic converter [20]. A tactic that is useful to improve the efficiency of an engine is to adjust engine parameters on an instantaneous basis during vehicle operation, so that combustion will be optimized immediately and engine efficiency will be maximized [20]. Another issue of great importance for engine oils is the concept of backward compatibility. Current engine oil formulations must provide adequate protection to older engines for which previous engine oil formulations were designed; otherwise there is potential for generating engine problems and causing customer confusion.
1.6.6 Future Investigations Solar energy, batteries, and various types of alternative fuels have been developed to power vehicles, but only a few alternatives to hydrocarbon-fueled vehicles have remained available to the public for more than a relatively short interval of time. Hybrid-electric vehicles represent a positive step toward reducing consumption of hydrocarbon fuels, but such vehicles do not fall within the domain of “completely renewable.” Addressing mobility issues, including reliability and utilization of renewable resources (from manufacture, to use, to environmentally acceptable recycling or disposal), will provide challenges to future generations of lubrication engineers. For example, automotive lubricants in the future may have to differ measurably from those of today to meet the demands of advanced vehicles. The information needed to lubricate hybrid vehicles, fuel cell-powered vehicles, or whatever other type of automobile will be in use in the year 2050 or 2100 will have to be gained experimentally. Automotive engine oils will have to be formulated using base oils and additives which do not cause deterioration of emission control system components, since it is important to move in the direction of reducing pollution as far as possible. At the same time, customers are demanding more maintenance-free vehicles, so that fill-for-life lubricant systems will be a preferred development. These desires will require novel approaches to lubricant formulation and revolutionary, as opposed to evolutionary, advances in additive chemistry. Whereas we are currently pursuing low-SAPS (sulfated ash, phosphorus, sulfur) oils to enable current and near-term emission requirements to be met,
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zero-SAPS oils will likely be required in the longer term. Innovation in both the lubricant industry and the automotive sector is desirable, such as designing lubricating systems that automatically sense the amount and condition of the lubricant, adjust fluid levels accordingly, replenish additives when necessary, and regenerate the oil by removing contaminants. Such innovations represent significant challenges to industries that have become accustomed to small, stepwise increases in performance over the years.
Acknowledgments The authors thank James Spearot and Frank Caracciolo of General Motors Research and Development Center, Robert Olree of General Motors Powertrain, Thomas Boschert of Afton Chemical Corporation, Larry Smith of Infineum, Ben Weber of Southwest Research Institute, Ewa Bardasz of Lubrizol, and George Schwartz of Electromechanical Associates for providing information relevant to the completion of this chapter.
References [1] C.M. Taylor, The Internal Combustion Engine in Theory and Practice, Cambridge, MA: MIT Press (1985). [2] K.C. Ludema, Friction, Wear, Lubrication — A Textbook in Tribology, Boca Raton, FL: CRC Press (1996). [3] M. Takiguchi, M. Oguri, and T. Someya, “A Study of Rotating Motion of Piston Pin in Gasoline Engine,” SAE International Paper No. 938142, Warrendale, PA (1992). [4] P.K. Goenka, R.S. Paranjpe, and Y.R. Jeng, “FLARE: An Integrated Software Package for Friction and Lubrication Analysis of Automotive Engines, Part 1: Overview and Applications,” SAE International Paper No. 920487, Warrendale, PA (1992). [5] H. Kageyama, T. Suzuki, and T. Ochia, “Numerical Study on the Three Dimensional Contact Pressure and Deformation of Piston Skirt,” SAE International Paper No. 2001-08-0082, Warrendale, PA (2001). [6] K. Funatani, K. Kurowawa, P.A. Fabiyi, and F.M. Puz, “Improved Engine Performance Via Use of Nickel Ceramic Composite Coatings (NCC Coat),” SAE International Paper No. 940852, Warrendale, PA (1994). [7] V.D.N. Rao, D.M. Kabat, D. Yeager, and B. Lizzote, “Engine Studies of Solid Film Lubricant Coated Pistons,” SAE International Paper No. 970009, Warrendale, PA (1997). [8] L.L. Ting, “A Review of Present Information on Piston Ring Tribology,” SAE International Paper No. 852355, Warrendale, PA (1985). [9] S.E. Schwartz, S.C. Tung, and Michael L. McMillan, “Automotive Lubricants,” ASTM Manual 37 on Fuels and Lubricants, chap. 17, pp. 465–495, West Conshohoken, PA (ASTM Headquarters) (2003). [10] A. Kapoor et al., Modern Tribology Handbook, Vol. 2, Materials, Coatings, and Industrial Applications, chap. 32, pp. 1187–1229, Boca Raton, FL: CRC Press (2003). [11] S.E. Schwartz and D.J. Smolenski, “Development of an Automatic Engine Oil-Change Indicator System,” SAE International Paper No. 870403, Warrendale, PA (1987). [12] S.E. Schwartz, “A Model for the Loss of Oxidative Stability of Engine Oil during Long-Trip Service. Part 1. Theoretical Considerations,” STLE Tribology Transactions 35: 235–244 (1992). [13] S.E. Schwartz, “A Model for the Loss of Oxidative Stability of Engine Oil during Long-Trip Service. Part 2. Vehicle Measurements,” STLE Tribology Transactions 35: 307–244 (1992). [14] P.J. Younggren and S.E. Schwartz, “The Effects of Trip Length and Oil Type (Synthetic Versus Mineral Oil) on Engine Damage and Engine-Oil Degradation in a Driving Test of a Vehicle with a 5.7L Engine,” SAE International Paper No. 932838, Warrendale, PA (1993). [15] D.J. Smolenski and S.E. Schwartz, “Automotive Engine Oil Condition Monitoring,” Lubrication Engineering 50: 716–722 (1994).
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[16] F.G. Rounds, “Effect of Lubricant Additives on Pro-Wear Characteristics of Synthetic Diesel Soots,” Lubrication Engineering 43: 273–282 (1987). [17] J.E. McDonald et al., “Oil Life Monitor for Diesel Engines,” United States Patent 6327900, December 11, 2001. [18] S.E. Schwartz, “An Analysis of Upper-Cylinder Wear with Fuels Containing Methanol,” Lubrication Engineering 42: 292–299 (1986). [19] S.E. Schwartz, “Effects of Methanol, Water, and Engine Oil on Engine Lubrication System Elastomers,” Lubrication Engineering 44: 201–205 (1986). [20] N. Canter, “Development of a Lean, Green Automobile,” Tribology and Lubrication Technology 60: 15–16 (2004).
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2 Automatic Transmission Fluids 2.1 2.2
Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Physical Properties . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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Viscosity • Other Physical Properties
C.D. Tipton The Lubrizol Corporation
2.3 Frictional Properties . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.4 Oxidation Stability . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.5 Foam . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.6 Compatibility with Organic Materials . . . . . . . . . . . . . . . 2.7 Corrosion Protection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.8 Antiwear . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.9 Composition . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.10 Specifications and Testing Requirements . . . . . . . . . . . . 2.11 Timeline of ATF Specifications . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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2.1 Introduction Automatic transmission fluids (ATFs) are highly specialized lubricants designed to function in automatic transmissions found in mobile equipment exposed to extremes of ambient environmental temperatures. They function as heat transfer fluids, lubricants, hydrostatic hydraulic fluids, and power transfer fluids. To effectively perform these functions ATFs must have a carefully designed set of both physical properties and performance attributes. Viscosity is one of the key physical properties of ATFs. The viscosity of a modern ATF, might be compared to a 0W-20 engine oil with a kinematic viscosity at 100◦ C ranging between 5.5 and 8.0 cSt. The low temperature fluidity of a modern ATF is also a critical design element and is generally under 20,000 cP at −40◦ C. Essential performance elements include friction and friction durability, oxidation resistance, and EP/antiwear properties. Other performance properties, also included in specifications, establish how compatible the ATF is with seals, metal alloys, plastic components such as thrust washers and gears, electrical insulation, resistance to foaming, and resistance to viscosity shear down. To provide all these properties there may be 10 to 20 different additive components required, making ATF one of the most complex lubricants in the industry. Additionally, ATF is generally formulated from a very high quality mineral or synthetic hydrocarbon base fluid, and the additive content is generally between 10 and 20%. More recently, alternative types of automatic transmissions other than clutch-type step-shift transmissions have been developed, which has resulted in several specialized types of fluids in the
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marketplace that are similar to ATF but differ in specific properties. These transmission types include continuously variable transmissions (CVTs) of both the push-belt-type and the chain pull-belt-type, dual clutch automatic transmissions (DCTs), and infinitely variable transmissions (IVTs), or traction drives. Despite the differences in fluid performance requirements of these transmissions the physical properties of the fluids remain similar to those of conventional ATE.
2.2 Physical Properties The physical properties of an automatic transmission fluid are of great importance to the transmission design engineer. These properties and how they vary with operating conditions to a large extent determine the selection and design of other transmission components. This interrelationship has become very critical as transmission designs are optimized for cost, weight, and fuel economy and has caused transmission design engineers to ever more stringently specify the properties of automatic transmission fluid. Some selected physical properties and their importance are discussed here.
2.2.1 Viscosity Viscosity is one of the most important physical properties of a lubricant and is a key parameter in the function of the transmission hydraulic system. From an engineering standpoint, all fluid flow calculations in a transmission require viscosity as a function of temperature. In the dynamic mechanical and thermooxidative environment of the transmission, viscosity does not remain constant with time and many specification requirements for ATF are designed to minimize the change in viscosity of an ATF during use. Viscosity is a measure of a fluid’s resistance to flow. The basic unit of absolute viscosity is the pascalsecond (Pa sec) or the force required, in dynes, to move a surface one square centimeter in area past a parallel surface at a speed of one centimeter per second, the surfaces being separated by a fluid film one centimeter thick. The centipoise (cP) is the common unit of absolute viscosity and is equal to 0.001 Pa sec [1]. Kinematic viscosity is the measure of viscosity where the flow of a fluid is due to the force of gravity. The common unit for kinematic viscosity is the centistoke (cSt) and is equal to 1.0 × 10−6 m2 sec−1 . The absolute viscosity divided by the density of the fluid equals the kinematic viscosity. Automatic transmission fluid viscosity is typically specified in terms of kinematic viscosity in centistokes at 100◦ C (ASTM D445) on the higher temperature end of the scale and in terms of Brookfield viscosity centipoise at −40◦ C (ASTM D2983) units at the low temperature ranges. Fluid viscosity affects engine cranking torque, internal hydraulic pressure and clutch shift times, and internal leakage [2]. Furthermore, the viscosity of the ATF is a function of temperature. ATF viscosity also varies with length of use in a transmission because of permanent shear of polymers used as additives to control the viscosity of the fluid. A variety of tests have been used to measure permanent shear including the Kurt–Ohrbahn Fuel Injector (D6278), the Fuel Injector Shear Stability Test (ASTM D5275), Sonic Shear (ASTM D2603), and the Tapered Roller Bearing or KRL Shear Test (CEC DIN 51350 Part 6) [3]. Typical plots for kinematic viscosity vs. temperature of some automatic transmission fluids are given in Figure 2.1. Early automatic transmission fluid specifications were more concerned with the minimum viscosity attained during service in the field to ensure adequate pumpability, sealing, and film thickness performance [4–6]. Modern transmissions equipped with solenoid valves and smaller cross sectioned flow passages are, however, more dependent on constant values of viscosity to ensure maintenance of original factory shift calibrations. For these reasons, some manufacturers are specifying ATFs with much less viscosity change with shear [7]. Some specifications are also reducing the kinematic viscosity at 100◦ C from 7.0–8.0 cSt range down to 5.4–6.0 cSt at 100◦ C in order to provide better fuel economy and less viscosity change during service due to the lower polymer contents. To meet low temperature Brookfield viscosity [8] requirements as low as 9000 cP at −40◦ C, automatic transmission fluids are formulated with combinations of mineral or synthetic base fluids to give a base blend in the 3.0 to 4.0 cSt viscosity range. To this is added a viscosity modifier with the desired shear
© 2006 by Taylor & Francis Group, LLC
Automatic Transmission Fluids
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100,000
Viscosity (cSt)
10,000
1,000
100
10
1 –50
FIGURE 2.1
0
50 100 Temperature (°C)
150
Kinematic viscosity vs. temperature for typical ATFs.
stability depending on the specification requirements to increase the viscosity to the desired value at 100◦ C. Current specifications limit the choice of base fluids to those with low volatility, very good low temperature viscosity, and good oxidation resistance. Table 2.1 gives a listing of the development of viscosity requirements in some ATF specifications in North America.
2.2.2 Other Physical Properties There are other physical properties of automatic transmission fluids that are useful to the designer of automatic transmissions. Among these are thermal conductivity, volumetric thermal expansion, heat capacity, and specific gravity. In Table 2.2 are shown results for two typical ATFs formulated from Group I [9] basestocks. The values are fairly consistent, so large variations from these values for other ATFs are not anticipated. A more detailed description of the properties of a particular ATF can be found in Reference 1.
2.3 Frictional Properties Friction is not a property inherent to a particular ATF. Rather, the friction that is observed is the result of the interaction of a three part tribological system, two solid rubbing surfaces which may be dissimilar and an automatic transmission fluid. The friction results obtained in a given situation depend on the nature of these three elements and the conditions of temperature, pressure, speed, and other factors. In an automatic transmission, the rubbing friction elements are generally a plate or a band-type clutch which may operate as a holding element, a shifting element, or a continuously slipping element. Continuously slipping clutches are often found to torque converters to enhance fuel economy. In a clutch system, one rubbing element is usually a porous composite material and the other rubbing element is usually a steel reaction member. The ATF functions in the interface to lubricate the elements, carry away heat, and to mediate the coefficient of friction vs. speed, temperature and load. Composite friction materials used in automatic transmissions are usually composed of a fibrous material such as cellulose, or a synthetic such a Kevlar®. Asbestos fibers have not been used since the 1970s due health concerns with asbestos fibers. Fillers are used to control wear and friction of the material and may include diatomaceous earth, graphite, or even metal particles. All this is bound together, typically with a phenolic resin, and bonded to a steel backing. Other types of friction plates may include sintered bronze and fluorocarbon elastomers. Reaction members may be a steel plate or a cylindrical drum in the case
© 2006 by Taylor & Francis Group, LLC
2-4 TABLE 2.1
Handbook of Lubrication and Tribology History of the Specification of ATF Viscosity in North America
Specification
Kinematic viscositya (ASTM D445 at 100◦ C)
General motors Type A
49 SUS min. at 210◦ F
Low temperature viscosityb
Minimum viscosity 46.5 SUS at 210◦ F during and after cycling and performance tests 46.5 SUS at 210◦ F during and after cycling and performance tests 5.5 cSt at 210◦ F during and at end of specified tests
Type A, Suffix A
49 SUS min. at 210◦ F
DEXRON®
7.0 cSt at 210◦ C
DEXRON®-IIb
None specified
5.5 cSt at 210◦ F during and at end of specified tests
DEXRON®-IIE
None specified
>5.5 cSt at 100◦ C used from Oxidation Test >5.0 cSt at 100◦ C used from Cycling Test >5.5 cSt at 100◦ C used from Oxidation Test >5.0 cSt at 100◦ C used from Cycling Test >5.5 cSt at 100◦ C used from Oxidation Test >5.0 cSt at 100◦ C used from Cycling Test
DEXRON®-III
DEXRON®-IIIH
Ford Type F
None specified
None specified
fluid fluid fluid fluid fluid fluid
M2C138-CJ M2C166-H MERCON®
49 SUS min. (7.0 cSt) at 210◦ F 7.0 cSt min. at 98.9◦ C 6.8 cSt min. at 100◦ C 6.8 cSt min. at 100◦ C
46.5 SUS (6.2 cSt) min. after 8000 cycle WOT test 6.2 cSt min. after FTLM BJ 12-4 6.0 cSt min. after FTLM BJ 12-4 5.0 cSt min. in GM Cycling Test
MERCON® V
6.8 cSt min. at 100◦ C
6.0 cSt min. at 100◦ C after 20 h KRL Shear
DaimlerChrysler (Chrysler Group) MS-3256
49 SUS min. at 210◦ F
48 SUS min. at 210◦ F after Chrysler shear 461C-112
MS-4228
49 SUS min. at 210◦ F
48 SUS min. at 210◦ F after Chrysler shear 461C-112
MS-7176, Change G (ATF+3®) MS9602, Change F (ATF +4®)
7.4–7.7 cSt min. at 100◦ C 7.3 to 7.8 cSt at 100◦ C
6.5 cSt min. after 30 passes in ASTM D3945B 6.5 cSt min. after 20 h KRL Shear
7000 SUS max. (extrapolated from 210◦ F and 100◦ F values) 4500 cP max. at −10◦ F 64,000 cP max. at −40◦ F 4000 cP max. at −10◦ F 55,000 cP max. at −40◦ F in Brookfield viscometer 4000 cP max. at −10◦ F (−23.3◦ C) 50,000 cP max. at −40◦ F in Brookfield viscometer 1500 cP max. at −20◦ C 5000 cP max. at −30◦ C 20,000 cP max. at −40◦ C 1500 cP max. at −20◦ C 5000 cP max. at −30◦ C 20,000 cP max. at −40◦ C 1500 cP max. at −20◦ C 5000 cP max. at −30◦ C 20,000 cP max. at −40◦ C 1400 cP max. at −40◦ F in Ford test method BJ 3-2 1700 cP max. at −18◦ C 1700 cP max. at −18◦ C 1500 cP max. at −20◦ C 20,000 cP max. at −40◦ C 1500 cP max. at −20◦ C 9000 ± 4000 cP max. at −40◦ C 7000 cP max. at −20◦ F in Chrysler method 461C114 2300 cP max. at −20◦ F in Chrysler method 461C114 4500 cP max. at −28.9◦ C 20,000 cP max. at −40◦ C 3000 cP max. at −28.9◦ C 10,000 max. at −40◦ C
a Viscosity of ATF was originally specified in Saybolt Universal Seconds (SUS) at Fahrenheit temperatures until the specific-
ations underwent metrification and were modernized. b Generally, low temperature viscosity is by the ASTM D2983 Brookfield procedure unless otherwise stated; temperatures
are stated in either Fahrenheit or Celsius depending on the time period of the specification.
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Automatic Transmission Fluids TABLE 2.2
Fluid ATF #1
ATF #2
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Selected Physical Properties of Automatic Transmission Fluids
Temperature ◦ F/(◦ C)
Thermal conductivity Cal/sec cm2 (◦ C/cm) × 105
Volumetric thermal expansion ◦ C−1 × 103
Heat capacity Cal/g◦ C
Specific gravity
0/(−17.8) 100/(37.8) 200/(93.3) 300/(148.9) 0/(−17.8) 100/(37.8) 200/(93.3) 300/(148.9)
30.6a 30.4 30.2 30.0a 31.4a 31.2 31.0 30.8a
8.8 25.3 41.6 58.2 9.9 25.6 41.0 56.7
0.46a 0.50 0.55 0.60 0.43a 0.48 0.53 0.58
0.8655b 0.8584 0.8335 0.7986 0.8640b 0.8574 0.8314 0.7969
a Extrapolated values. b Extrapolated from 4◦ C.
of bands. Surface finish and hardness are critical factors of steel reaction members in maintaining proper friction and durability. A shifting clutch may experience very high temperatures, up to 600◦ C, in some cases, therefore flow of fluid through the clutch to provide cooling is very important. Often times, grooves are cut or embossed into the friction material to enhance lubrication and cooling. Composite plates are manufactured to provide compressibility and conformability to increase friction coefficient; porosity aids in both cooling and lubrication of the interface by allowing absorbed fluid to be squeezed out during engagements. With high interfacial temperatures being generated, it is important that the ATF have good thermal and oxidative stability and be able to retard buildup of decomposition products which may clog pore structures critical to the function of the friction material and “glaze” the surfaces. In a clutch engagement, the observed friction is a combination of hydrodynamic and asperity friction [10]. In the initial stage of engagement, there is mainly hydrodynamic friction brought about by shear of the fluid film. Later in an engagement one finds that asperity friction begins to take over. The automatic transmission fluid can mediate the friction coefficient observed in both phases of the engagement. Particularly, the coefficient of friction can be controlled in the very slow speed end of an engagement, where asperity contact is dominant, by the use of friction modifiers. The friction vs. speed curve (µ–v) can be controlled to having either a negative or a positive slope in this region. This is important because perceived shift quality or “shift-feel” and also the propensity for stick-slip or “shudder” to occur depends, to a large extent, on the slope of the µ–v curve of the ATF. Negative slope behavior can lead to shudder in torque converter clutches and harsh shifts in shifting clutches. The dynamic torque provided by an ATF is important to the determination of the torque capacity of the clutch unit and to the number and types of friction elements used. Static torque or friction coefficient is also important in that the static holding capacity of a clutch is dependent on this value and it is generally desired that the value be as high as practical in concert with good antishudder characteristics. Current North American ATF specifications such as DEXRON®-III (Revision H), MERCON® V, and MERCON® are for service-fill of General Motors and Ford vehicles, respectively. Factory-fill requirements generally reflect service-fill requirements but often with specific additional testing required. Friction tests used to determine the properties of an ATF are usually conducted in an SAE#2 electric motor driven friction tester using a clutch pack or band, or in a vehicle when antishudder and shift quality are of concern. SAE#2 Tests are generally of the engagement type where the clutch and fluid are cycled for up to 30,000 cycles with key parameters like dynamic torque are required to stay within narrow limits. Figure 2.2 is a diagram of a typical ATF SAE#2 engagement torque trace and the parameters used by various passenger car manufacturers to estimate performance. Slow speed or “static” friction measurements (Figure 2.3) are usually determined using an auxiliary slow speed motor to turn the plates at a constant slow speed, typically 0.72 rpm, while engagement pressure is maintained at a constant value.
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Handbook of Lubrication and Tribology
4000
300
3500
250
Apply pressure
200
3000 2500
Torque
2000
150
1500
100
1000
Speed
50
500 0 1500
0 0
Speed (rpm)
Torque (Nm) Temperature (°C) pressure (kPa)
Dynamic friction engagement
500
1000 Time (msec)
FIGURE 2.2
Dynamic friction engagement.
Static friction break-away 0.72 rpm motor speed 200
0.180
180
0.160
160
Torque (Nm)
120
Tt (mt) Torque 2 sec after the start of static trace (2175 msec)
100
0.120 0.100 0.080
80 0.060
60
Friction coefficient (Mu)
0.140 Ts (ms) Maximum torque immediately after start of static trace
140
0.040
40 5 Nm threshold start of static trace (175 msec)
20
0.020
0 0
500
1000
1500
2000
0.000 2500
Time (msec)
FIGURE 2.3
Static friction break-away.
2.4 Oxidation Stability The factory-fill automatic transmission fluid is now expected by automobile manufacturers to last for the useful life of the vehicle, a so-called fill-for-life fluid. The ATF and its frictional and viscosity characteristics are a key part of maintaining the designed shift quality and antishudder durability of the vehicle. Therefore, it is desired that the ATF not be changed such that there is little chance that a misapplication of the wrong type of ATF can occur. Drain plugs and dipsticks are being eliminated to reduce interference with the factory-fill fluid. In order to ensure that a fluid can last the life of the vehicle, good stability toward oxidation is essential. Oxidation of the fluid can (1) increase both the high temperature and the low temperature viscosity of the fluid; (2) cause the formation of sludge, varnish, and particulate matter which can interfere with the function of the friction materials and the other mechanical surfaces; (3) cause corrosion of bushings, thrust washers, and bearings; and (4) cause degradation of elastomeric seals. Oxidation of the fluid is also a function of the conditions that the fluid experiences during service in a vehicle with temperature being
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Automatic Transmission Fluids
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the key factor. Oxidation in vehicles in the field has been related to the length of service time the fluid has been at high temperature. These conditions typically occur during trailer towing or in traffic at engine idle speeds. Requirements for oxidation resistance in transmission fluid specifications have become ever more stringent as equipment makers strive to ensure fill-for-life. For example, General Motors uses a full-scale electric motor-driven transmission for its oxidation testing in its DEXRON® series of specifications. From the original DEXRON® to the current DEXRON-III (H revision), the length of the oxidation test has increased from 300 to 450 h at 163◦ C and the air flow rate has been changed from 30 to 90 cc/min. Additional requirements have been added besides sludge and varnish limitations including a 3.25 maximum change in total acid number, and an infrared absorbance maximum increase of 0.45. Other tests used to measure oxidation are the Aluminum Beaker Oxidation Test (ABOT) used by Ford in its MERCON® series of specifications, ISOT (Indiana Stirring Oxidation Test) used primarily in Japan, and the DKA oxidation test (CEC L-48 METHOD) used in Europe.
2.5 Foam Excessive foam buildup in a transmission can be very detrimental to transmission function and to safety. Foaming can be induced by the churning of the rotating parts and gears in the fluid, air leakage on the suction side of the hydraulic pump, and other factors. Since one of the functions of an ATF is as a hydraulic fluid, foaming causes a large increase in the compressibility to the ATF, interfering with the pressure applied to clutches. In some cases this can lead to premature failure of the clutch due to lack of proper apply pressure. Foaming also causes the heat capacity of the ATF to be reduced, so it is less effective in removing heat from the clutches and other parts, and foaming also leads to an increase in the volume of the ATF. This latter effect can be very dangerous in vehicles having a dipstick, because if the volume of the ATF exceeds the free air space in the transmission, the ATF can be thrust up the dipstick tube and ejected.
2.6 Compatibility with Organic Materials Seals, engineering polymers, and wire insulation for electronic controls are some of the materials that are of concern when in contact with ATF. Seals are used as gaskets, shaft seals, and seals for the clutch pistons and drums that are hydraulically actuated. Deterioration of these elastomeric seals leads to either internal or external leakage in the transmission. Internal leakage can lead to the loss of adequate hydraulic pressure to critical clutches, and external leakage is a negative to the vehicle purchaser. A wide variety of different types of elastomers are used in transmission including silicone, nitrile, polyacrylate, fluorocarbon types (e.g., Viton®), and others. ATF can interact with a seal material in several ways which can be detrimental. It can either swell or it can shrink the elastomer. Too much shrinkage or too much swelling can lead to deterioration of dimensional and physical properties of the seal and cause leakage. For any ATF formulation that gives an inadequate amount of swelling (a particular problem with highly refined Group III and Group IV base fluids), chemical seal swell agents can be added to increase the positive volume change of elastomers, especially with nitrile-type elastomers. Chemical additives in the ATF can also react with elastomers, depending on the elastomer-type. For example, too much sulfur in an ATF can cause hardening of a nitrile rubber due to additional cross-linking of the polymer. Silicone rubber can be depolymerized by strong acids generated from oxidation of the fluid, especially at higher temperatures. Transmission fluids are generally tested for seal compatibility by soak tests such as the ASTM D471. A small amount of positive volume change is usually desirable in these tests, but excessive volume increase and shrinkage are to be avoided. In addition to the bench scale soak tests, parts from full-scale transmission tests are examined to determine whether each type of seal material is performing as well as they do with a good reference fluid.
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Handbook of Lubrication and Tribology
Other organic polymers are used for insulation of the electrical wires for electronic controls. These are usually tested by the wire suppliers against reference ATFs to ensure compatibility. Polymeric engineering materials found in transmissions include parts made from phenolic resins, polyesters, and nylon. Nylon has been used to fabricate gears and thrust washers used in automatic transmissions. Degradation of nylon parts can be a problem when in contact with ATF at higher temperatures [11] and either the ATF or the nylon formulation is improperly selected.
2.7 Corrosion Protection Automatic transmissions contain parts made from a variety of metals. The cases and valve bodies may be aluminum; shafts, gears, and clutch drums may be steel alloys; thrust washers and bushings can be made of various copper containing alloys; oil coolers may be made of brass; and bearings and thrust washers may be coated with tin or lead. In some heavy-duty transmissions, the clutch plates themselves may be made of sintered bronze. It is critical to the function of these types of parts that they not undergo deterioration through chemical attack from water, oxygen in the air, or from the ATF itself. Most ATFs contain specific anticorrosion agents to retard the rusting of ferrous alloys, and the corrosion of copper, zinc, lead, and tin. Among the tests used to measure anticorrosion for ATF are the ASTM D665 and the ASTM D1748 humidity cabinet for rust of ferrous alloys, and the ASTM D-130 for copper corrosion. In addition, full-scale transmission test parts are examined for corrosion and compared to good performing reference fluids. Ford uses copper and lead strips in its ABOT test and has a requirement on the maximum percentage weight loss of the lead strip.
2.8 Antiwear Automatic transmission fluid must be able to successfully lubricate and prevent wear in multiple mechanical parts in a transmission including vane and gear pumps, one-way clutches, planetary gearsets, bushings, bearings, thrust washers, and power transfer chains. Modern automatic transmissions have undergone an evolution in weight reduction and increased power density in an effort to save weight and reduce size for better vehicle fuel economy. These changes and also the desire for the transmission to last the lifetime of the vehicle with its original fluid have raised additional concern about the antiwear capability of the transmission fluid. Earlier specifications may have included only one specific wear test such as a 4-ball wear test [12,13] or a power steering vane pump wear test [4]. It is now typical for an ATF specification to include several wear tests, including a FZG four square gear test [14] and a test to measure wear of a sprag-type one way clutch. Transmissions run in dynamometer tests are also monitored for any signs of unusual wear when tested with candidate ATF fluids. Many earlier ATFs relied on zinc dithiophosphate antiwear technology, similar to that used in engine oils, for wear protection but few, if any, ATFs formulated after about 1980 in North America contain zinc dithiophosphate. ATFs nearly universally use metal-free antiwear agents employing a combination of phosphorus, boron, and sulfur materials.
2.9 Composition Automatic transmission fluids are composed of a base fluid plus a complex additive formulation intended to meet all of the physical and performance properties required of an ATF. The base fluid is usually a petroleum based or synthetic hydrocarbon mixture with a viscosity of between 3.0 and 4.5 cSt at 100◦ C. Viscosity at low temperature, volatility, and oxidation stability are important criteria in the selection of base fluid. The additive is required to help the ATF meet the physical and performance properties that are not easily met by the base fluid alone. The classes of additive components that can be used to make an ATF include
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Automatic Transmission Fluids
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viscosity modifiers, fluidity modifiers, pour point depressants, foam inhibitors, dispersants, detergents, antiwear agents, oxidation inhibitors, corrosion inhibitors, detergents, and friction modifiers. Often there is more than one representative of each class present, so the total number of additive components used to make an ATF is usually in the 10 to 20 range. Viscosity of an automatic transmission fluid is a combination of the contributions from the base fluid plus the additives. Viscosity modifiers are used to adjust the high temperature viscosity and also to control and reduce the low temperature viscosity. Typically, polymethacrylate and esterified maleic anhydride/styrene copolymers are used. These high molecular weight molecules are constructed using special alcohol mixtures in the ester functions to provide modification of wax structures that may come from the base fluid. This provides improved flowability at low temperatures as measured by the Brookfield viscosity test. The viscosity modifiers also typically incorporate dispersant functionality to aid in sludge dispersancy during oxidizing conditions. Low temperature viscosity is often further controlled by the addition of special low viscosity oils and by the use of pour depressants. These pour depressants are similar to the viscosity modifiers, but are specially designed to modify the formation of wax networks and the treatment levels are lower. Specialized dispersants, usually of the polyisobutenylsuccinimide-type, are another major component of an ATF additive system. Dispersants are truly multifunctional components which act to suspend dirt, sludge, and debris, they extend clutch frictional durability through their cleaning effect on composite friction materials, and they influence friction coefficients in relation to sliding speeds. Some dispersants can also act as carriers for other additive functionalities like boron, which can be active on antiwear and anticorrosion. Metallic detergents like calcium and barium overbased sulfonates and carboxylates were once extensively used in automatic transmission fluids to provide stable friction coefficients and to neutralize acidic products of oxidation. These detergents act in much the same way as a dispersant on composite friction materials to maintain cleanliness and retard buildup of deposits and glaze. Detergents also, through their ability to neutralize acidic materials, are able to protect elastomeric seal materials and composite friction materials from premature deterioriation [15,16]. The classic antiwear agents used in automatic transmission fluids were zinc dithiophosphates. However, in the mid-1980s with the increase in the severity of clutch friction testing, the thermal instability of most zinc dithiophosphates led to the discontinuation of use in most ATF formulations. Instead, phosphorus materials which had relatively low activation temperatures, such as trivalent phosphorus esters and phosphorous acid itself, came into use for antiwear. Active sulfur compounds are generally not used in ATF formulations because of corrosiveness toward copper-based alloys, attack of elastomers used for sealing, and deleterious interactions with composite friction materials. Boron often finds use in ATF formulations as a noncorrosive antiwear supplement to the phosphorus materials used in ATF. Oxidation inhibitors are important in maintaining the properties of the ATF for the life of the vehicle. Exposure to the oxygen in the air and long periods of time at elevated operating temperatures lead to attack on the molecules of the base fluid and additives by oxygen and subsequent decomposition to oxidation propagating peroxy radicals. Oxidation inhibitors act to intercept these radicals and render them less harmful so they do not continue to propagate creating acidity and polymeric degradation products [17]. Alkylated diaryl amines is the most ubiquitous class of oxidation inhibitors found in ATFs, but certain less active sulfur compounds and hindered phenols are also used. Again, these inhibitors are used either singly or in combinations to get the most effective combination for a given additive and base fluid combination. Great care must be exercised in the selection of oxidation inhibitors in order that a low level of undesirable corrosiveness, interaction with composite friction materials, and elastomeric seals is achieved. One of the key components in an ATF formulation is the friction modifying chemistry needed to achieve the proper levels of static and dynamic friction, µ/v slope, and long term antishudder durability [18]. In today’s vehicles, there is a need to have both a relatively high static coefficient of friction for clutch holding capacity and also a positive µ/v slope for the life of the fluid. Therefore, many older classic friction modifiers which were very effective at depressing static coefficients of friction are no longer as useful. In more modern fluids, thermal and oxidative stability of the friction modifier is an important
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Handbook of Lubrication and Tribology
selection criterion. In addition, friction modifiers are chosen which produce a smaller reduction in static coefficient values. The friction modifiers normally have a polar functionality in the molecule which interacts by surface adsorption on the composite friction material and the reaction member. The oleophilic remainder of the molecule extends into the fluid from the surface [19]. It is these surface films and their structure and lability that changes the friction when the oil film becomes thin and asperity contact of the rubbing elements becomes significant. Typical functional groups that interact with the surfaces are alcohols, amines, amides, carboxylic acids, imidazolines, and oxazolines. Typical carbon chain lengths for friction modifiers are from 10 to 24 carbon atoms. Corrosion inhibition and general compatibility with materials is required for the long-term durability and function of transmission components. For ferrous metals calcium overbased sulfonates, nonionic surfactants, and certain aminic surface-active agents are employed, if necessary to pass antirust requirements. Copper, tin, zinc, and lead alloys are often used for bushings, bearings, thrust washers, electronics, coolers, cooler lines, and braze joints. Typical inhibitors of corrosion of these materials are various commercially available benzotriazole derivatives and derivatives of 3,5-dimercaptothiadiazole. Calcium overbased sulfonates act to protect metals, elastomers, and engineering plastics from acidic attack. With the move to highly refined mineral oil and synthetic hydrocarbons for base fluids, there are fewer aromatic and naphthenic components in the base fluids to maintain the volume of the elastomeric seals. In recent years, increasing amounts of seal swelling additives have been required to maintain positive swell values required by transmission manufacturers in seal immersion testing. In the past, highly aromatic oils were often used as seal agents, but the use of these ceased when carcinogenicity questions arose in the use of these materials. Today, chemical seal swell agents are often used. These are materials that have an affinity for the elastomeric seal material and can be preferentially absorbed into the interstices of the elastomer matrix. Nitrile rubbers are often the most difficult to swell, so seal swell agent treatment levels are optimized to the requirements of this type of elastomer. Typically, pthtalate esters, sulfones, or other low molecular weight esters are used in ATFs as seal swell agents at the minimal levels required to provide the specified volume change. If a naphthenic oil with good oxidation stability and other properties is available, it may also be incorporated as a seal swell agent. Other types of seal materials that are used in transmissions include fluorocarbon, polyacrylate, and silicone based elastomers, but these types of elastomers generally have less need for seal swell agents than do nitrile elastomer-types. Antifoam additives are required to prevent excessive foaming of the fluid in the transmission. Siliconetypes are most commonly used including polydimethylsiloxanes, fluorosilicones, and functionalized siloxanes. Most of these liquid materials are insoluble in oil and are incorporated into automatic transmission fluids as very fine dispersions. The silicone droplets are surface active and act to destabilize the air/oil interface of a foam bubble through changing the surface tension. Most transmission fluids are required to be dyed red by manufacturer specifications in order to identify sources of leakage and, generally, to provide fluid identification and to distinguish ATF from undyed engine oils.
2.10 Specifications and Testing Requirements The first specification to standardize the requirements for a fluid for automatic transmissions was introduced in 1949 by General Motors and was designated “Type A.” The Type A specification also provided a process for qualification of service-fill ATF through a trademark and licensing procedure, a process that continues to the present day and serves to help police and enforce the quality of transmission fluids bearing the trademarks. Eventually it was found that some Type A qualified automatic transmission fluids showed deficiencies in oxidation resistance and this led to the introduction of the Type A, Suffix A specification in 1957. In 1967, General Motors introduced the DEXRON® trademark and specification to continue to upgrade the fluids on friction durability, oxidation stability, and viscosity properties. The DEXRON® specifications have continued to be upgraded in the ensuing years.
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Automatic Transmission Fluids
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In the late 1950s, Ford’s factory-fill fluids met specification M2C33-A-B which describe a fluid with similar characteristics to Type A, Suffix A. Ford introduced the M2C33-D specification in 1961. This change was driven by the need for better oxidation control, antiwear performance, and higher static capacity. Ford introduced a new type of specification for automatic transmission fluid in 1967. Ford had the objective of a “fill-for-life” fluid with improved antioxidation, wear, and friction performance. This new fluid was described by the M2C33-F specification and was conventionally called “Type F” [5]. It was also a service-fill specification and Ford granted approvals under the specification along with qualification numbers. This specification was similar to M2C-33D. Differences included a 6-pack clutch friction test which required a high static coefficient of friction. Ford’s use of an ATF with a friction curve similar to base oil would undergo less frictional change with time than a fluid that used a chemical friction modifier which might degrade, allowing the oil revert to base oil characteristics. Another driving force was to reduce the number of plates in the clutch pack to get a more consistent shift characteristic. Oxidation resistance was increased by raising the temperature of their 300-h transmission oxidation test from 300 to 325◦ F. Ford moved away from the Type F concept of a high static friction fluid with the introduction of the M2C138-CJ specification in 1974 [20], which called for a friction-modified fluid. This type of fluid was intended to alleviate difficulties with engineering for good gear engagement noise perception with the high static friction fluid. A new friction test was introduced which called for a narrow band of acceptability (0.9 to 1.0) in the ratio of static to dynamic friction in the M2C138-CJ fluids. This required that the static friction had to be less than the dynamic friction. This specification was followed by the M2C166-H [21] specification for factory-fill fluids requiring improved friction characteristics for lock-up torque converters for factory-fill fluids. In this specification, Ford introduced the ABOT to replace the 300 h oxidation test carried out in a motored transmission. The MERCON® specification was introduced by Ford in January 1987 [22], a trademarked fluid, for service-fill with procedures for qualification and licensing of fluids to ensure quality in the marketplace. The development of the modulated and continuously slipping clutch torque converters prompted the need to develop the MERCON®V specification. Requirements for verifying antiwear capabilities and antishudder characteristics were included in the specification. In the late 1960s Chrysler Corporation also upgraded its requirements for automatic transmission fluid with the MS-4228 factory-fill specification [6] replacing the older MS-3256 which had described fluid similar to Type A, Suffix A. Very good low temperature viscosity was a major objective in order to reduce the time for engine starting and the time for shifts to occur when the fluid was still cold. Requirements for oxidation stability were also upgraded through the use of a new bench oxidation test. The next major upgrade by Chrysler Corporation, now DaimlerChrysler, was the introduction of the MS-9602 specification in the late 1990s and the marketing of the new factory-fill fluid as ATF+4® [23]. A history of these specifications is given in Table 2.3.
2.11 Timeline of ATF Specifications ATF physical properties and tests were covered earlier in this chapter. Performance testing of ATFs to meet factory-fill and service-fill qualification requirements are handled by each individual OEM in the case of ATF. There is no overall specification for an ATF which can be used in any vehicle as exists for engine oils. Friction and friction durability, oxidation resistance, and antiwear characteristics are the three key areas at which most performance testing is directed. In the absence of an industry-standard specification, certain OEMs, for example, General Motors and Ford, have written their own service-fill specification and assigned a trademark designation to it, DEXRON® in the case of General Motors and MERCON® in the case of Ford Motor Company. It is through the use and licensing of the trademarks that the composition and quality of approved fluids is maintained. In each case specific tests are required that ensure suitability for service of current models and back application to previous transmission. Since not all transmission manufacturers use the same transmission designs or the same friction materials, the specifications
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Handbook of Lubrication and Tribology TABLE 2.3
Timeline of ATF Specificationsa
Specification Type A Type A, Suffix A Type F MS-3256 MS-4228 DEXRON® DEXRON®-IIb M2C138-CJ DEXRON®-II MS-7176, Change D (ATF +2®) M2C166-H MERCON® DEXRON®-IIE DEXRON®-IIE-Rev. DEXRON®-III MERCON® V DEXRON®-IIIGc MS-7176, Change G (ATF +3®) DEXRON®-IIIH MS-9602, Change F (ATF +4®)
Number
6137-M 6137-M
6137- M 6137- M 6297- M 6417- M
Year 1949 1957 1959 1964 1966 1967 1973 1974 1978, July 1980, May 1981, June 1987, January 1990, October 1992, August 1993, April 1996, July 1997, April 2001, June 2003, April 2003, July
Company GM GM Ford Chrysler Chrysler GM GM Ford GM Chrysler Ford Ford GM GM GM Ford GM DaimlerChrysler GM DaimlerChrysler
a Not all specification changes are included. b DEXRON®-II was originally released using “C” qualification numbers. In 1975 after suppliers
rolled over approvals because of a GM mandated fix for a cooler corrosion problem, the qualification numbers were preceded by a “D.” This led to referring to the fluids as DEXRON®II “C” or “D.” c Upgrade for ECCC vehicle test and sprag clutch wear.
may differ in test requirements. However, current automatic transmission fluid requirements are similar enough that it is common to have fluids approved for both a DEXRON®-type and a MERCON®-type requirement. This means that the fluid so designated has been through two separate qualification programs in order to bear both trademarks. Factory-fill fluids are usually similar to service-fill fluids in performance, but they have gone through additional testing at the manufacturer that has not been required for most service-fill fluids. Most manufacturers make their factory-fill fluid available for service use through their dealerships. Specifications for on-highway automatic transmissions generally call for friction-modified (low static coefficient of friction) fluids today. Ford-Type F fluid used prior to about 1980 was an older fluid that did not contain friction modifier. Fluids designed to meet this specification are still commercially available but are not suitable for modern Ford vehicles. Each transmission or vehicle manufacturer has a special set of requirements designed around the characteristics of the materials used in the manufacture of the transmission. Friction testing usually can be categorized into two types, initial friction characteristics tests and tests measuring durability characteristics. Both General Motors and Ford have long-term durability tests run in a SAE#2 clutch friction tester in their specifications for service-fill. These tests use flat disk-type clutch plates. The exact length of the test, the friction material used on the friction plate, the arrangement of the plates in the clutch pack, temperatures, and so on are unique to the specification involved. Also, the acceptable limits in the envelope of change of the test parameters allowed throughout the test vary. Table 2.4 shows a comparison of the test parameters for three major service-fill ATF specifications. Other long-term friction tests include a band friction durability test and the engine-driven 4L60transmission dynamometer-cycling test used by General Motors. Initial friction tests often include a shift quality test run in a vehicle and vehicle shudder testing to test compatibility of the fluid with torque converter clutch operation.
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Automatic Transmission Fluids TABLE 2.4
2-13
SAE#2 Test Parameters for Various ATF Friction Durability Tests
Composition plate Grooving Test length (h) Test length (cycles) Total energy (J) Clutch release (kPa) Clutch pack clearance (mm) Motor speed (rpm) Clutch apply (kPa) Fluid temp (◦ C) Fluid volume (ml) Cycle rate (cycles/min) Plate configuration Energy per engagement (J/cm2 )
DEXRON®-IIIH flat plate
Ford MERCON® 15k flat plate
Ford MERCON® V 20k flat plate
SD1777 None 100 18,000 27,000 150 1.02 ± 0.13 3600 345 140 600 3 S-F-S-S-F-S 85.6
SD1777 Cross hatch 63 15,000 20,740 ≥75 0.70 ± 0.13 3600 275 ± 5 115 ± 3 305 ± 5 4 S-F-S-S-F-S 88
SD1777 Cross hatch 83 20,000 20,740 ≥75 0.70 ± 0.13 3600 275 ± 5 135 ± 2 305 ± 5 4 S-F-S-S-F-S 88
Antiwear is another critical aspect of ATF performance for which there are a multitude of tests. Ford’s MERCON® specification requires an ASTM D2882 vane pump test, while the MERCON® V specification requires the same vane pump tests and a FZG Gear Wear Test (ASTM D5182), a Four-Ball Wear Test (ASTM D-4172), a Falex EP test (ASTM D3233), and a Timken Wear Test (Modified ASTM D2782). All these additional tests have a requirement to be run at 150◦ C. General Motors in its recent DEXRON® specifications requires a special sprag (one-way) clutch wear test, and the modified ASTM D2882 vane pump test requiring 15 mg maximum weight loss. Antiwear is further assessed through the inspection of the dynamometer-cycling test, where it is examined for any unusual wear vs. reference on the gears, bearing, etc. Resistance of the ATF to oxidation is important for fill-for-life considerations. The GM DEXRON® specification employs a full-scale 4L60 transmission driven by an electric motor at a 163◦ C sump temperature and 450 h to assess oxidation resistance. Requirements include a change in total acid number (TAN) of less than 3.25 mg KOH/g fluid, a sludge amount less than or equal to a reference oil, and a change in infrared carbonyl absorbance of 0.45 max. Ford MERCON®V testing uses a special bench ABOT to assess oxidation for 300 h at 155◦ C with a change in TAN of 3.5 max, pentane insolubles of less than 0.35%, and other criteria being required. Corrosion ratings of copper and lead strips suspended in the ATF during the ABOT procedure are also requirements. The lead strip weight loss cannot exceed 3.0% and the visual rating of the copper strip at 50 h and at 300 h is 3b maximum. Copper corrosion protection is usually determined by conventional ASTM D-130 testing which is provided for in both MERCON® and DEXRON® specification requirements with a 1b maximum rating at 3 h and 150◦ C. Note that this is a significantly higher temperature than the 121◦ C temperature conventionally used for gear lubricants. The ASTM D665 Turbine Oil Rust test is employed in both MERCON® and DEXRON® specifications to measure the ATF’s ability to protect against rust. Additionally, the DEXRON® specifications require the ASTM D1748 Humidity Cabinet Rust Test. Corrosion is also tested for in other tests like the MERCON® ABOT where the catalytic copper strip can have only a 3b maximum rating. In DEXRON® oxidation testing in the 4L60 transmission parts are monitored for any corrosion that exceeds the reference. Good antifoaming performance is also a key requirement for ATFs. Tests for antifoam performance include the ASTM D892 three-sequence foam test used by most manufactures. Additionally, Ford adds a fourth high temperature sequence, the ASTM D6082 test. GM also has a special foam test and an aeration test to measure air-entrainment, and it also limits foam levels in its oxidation and cycling tests run in full-scale transmissions. Future trends in ATF testing will most probably concentrate on two areas of performance. First, continued improvement in friction stability and oxidation resistance will be important for fill-for-life
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performance. Stability related features such as viscosity at all temperatures are becoming more important to maintain new transmission performance for the life of the vehicle. The second driving force in testing is the introduction of new technologies. Dual clutch transmissions, for example, may require durability testing with new friction materials using longer high-energy engagements not typical of conventional automatic transmission shifting clutches. Push-belt CVTs like those developed by Van Doorne Transmissie and adapted by OEMs in Europe, Japan, and the United States require automatic transmission-type fluids with the capability of operating with conventional clutches and also requiring high metal–metal friction for low slippage and good torque transfer in the belt to variator interface. Conventional automatic transmissions and CVT transmissions may be fitted with start-up clutches replacing the torque converter for better fuel efficiency. This again will require specialized testing to model the stresses on the fluid in this sort of application.
References [1] Kemp, S.P. and Linden, J.L., “Physical and Chemical Properties of a Typical Automatic Transmission Fluid,” SAE Paper Number 902148, International Fuels and Lubricants Meeting and Exposition, Tulsa, Oklahoma, October 22–25, 1990. [2] Watts, R.F. and Szykowski, J.P., “Formulating Automatic Transmission Fluids with Improved Low Temperature Fluidity,” SAE Paper Number 902144, International Fuels and Lubricants Meeting and Exposition, Tulsa, Oklahoma, October 22–25, 1990. [3] Sprys, J.W., Vaught, D.R., and Stephens, E.L., “Shear Viscosities of Automatic Transmission Fluids,” SAE Paper Number 941885, Fuels and Lubricants Meeting and Exposition, Baltimore, Maryland, October 17–20, 1994. [4] Haviland, M.L., Anderson, R.L., Davison, E.D., Goodwin, M.C., and Osborne, R.E., “Dexron-II Automatic Transmission Fluid Performance,” SAE Paper Number 740053, Automotive Engineering Congress, Detroit, Michigan, February 25–March 1, 1974. [5] Ross, W.D. and Pearson, B.A., “ATF-TYPE F Keeps Pace with Fill-for-Life Requirements,” SAE Paper Number 680037, Automotive Engineering Congress, Detroit, Michigan, January 8–12, 1968. [6] Kobe, R.A. and Wagner, J.C., “The Chrysler TorqueFlite and Automatic Transmission Fluid,” SAE Paper Number 680036, Automotive Engineering Congress, Detroit, Michigan, January 8–12, 1968. [7] DaimlerChrysler Material Standard, MS-9602, Change E, 2002. [8] ZF Factory Fill ATF Standard ZFN 13014, 2000-12. [9] API 1509, “Engine Oil Licensing and Certification System,” 15th ed., April 2002, American Petroleum Institute, Washington, DC. [10] Yang, Y., Lam, R., and Fuji, T., “Prediction of Torque Response During the Engagement of Wet Friction Clutch,” SAE Paper Number 981097, International Congress and Exposition, Detroit, Michigan, February 23–26, 1998. [11] Ward, W., Snyder, J., Lann, P., and Derevjanik, T., “ATF Nylon Degradation,” SAE Paper Number 971625, May 1997, International Spring fuels and Lubricants Meeting, Dearborn, Michigan, May 5–8, 1997. [12] Coleman, L., “Development of Type F Automatic Transmission Fluids,” SAE Paper Number 680039, January 1968, Automotive Engineering Congress, Detroit, Michigan, January 8–12, 1968. [13] “Automatic Transmission Fluid,” Ford Motor Co., Manufacturing Standards. M2C33 E-F, March 1, 1967. [14] Ford Motor Company, “A Specification for MERCON® V,” Revised October 1, 1998. [15] Riga, A., Patterson, G., Pistillo, W., Scharf, C., and Ward, W., “Automatic Transmission Fluid Compatibility with Nylon Components by Thermomechanical Analysis and Thermogravimetry,” Thermochimica Acta, 226, 1993, 363–368. [16] Tipton, C. and Schiferl, E., “Fundamental Studies on ATF Friction I,” SAE Paper Number 971621, International Spring Fuels and Lubricants Meeting, Dearborn, Michigan, May 5–8, 1997.
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[17] Johnson, M., Korcek, S., and Zinbo, M., “Inhibition of Oxidation by ZDTP and Ashless Antioxidants in the Presence of Hydroperoxides at 160◦ C — Part I,” SAE Paper Number 831684, Fuels and Lubricants Meeting, San Francisco, California, October 31–November 3, 1983. [18] Slough, C., Ohtani, H., Everson, M., and Melotik, D., “The Effect of Friction Modifiers on the LowSpeed Friction Characteristics of Automatic Transmission Fluids Observed with Scanning Force Microscopy,” SAE Paper Number 981099, February 1998. [19] Zhu, Y., Ohtani, H., Greenfield, M., Ruths, M., and Granick, S., “Modification of Boundary Lubrication by Oil-Soluble Friction Modifier Additives,” Tribology Letters, 15(2), 2003. [20] Ford Engineering Material Specification, “ESP-M2C138-CJ,” released September 1974. [21] Ford Engineering Material Specification, “ESP-M2C166-H,” released June 1981. [22] “A Specification for MERCON® Automatic Transmission Fluid Trademarked for Service in Vehicles Sold by The Ford Motor Company,” Revised April 1, 1992. [23] Florkowski, D. and King, T., Chrysler Corporation, Skroubul, A., Texaco Lubricants, and Sumiejski, J., Lubrizol Corporation, “Development and Introduction of Chrysler’s New Automatic Transmission Fluid,” SAE Paper Number 982674, International Fuels and Lubricants Meeting and Exposition, San Francisco, California, October 19–22, 1998.
© 2006 by Taylor & Francis Group, LLC
3 Rear Axle Lubrication 3.1 3.2 3.3
Rear Axle Lubrication . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Viscosity Classifications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Gear Oil Classification . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
3-2 3-2 3-3
Service Designations in Current Use • Service Designations Not in Current Use
3.4 3.5
Performance Requirements for Gear Oils . . . . . . . . . . . . Gear Oil Composition . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
3-4 3-5
Base Oil • Viscosity Modifiers and Pour Point Depressants • Performance Package
3.6 3.7
Arup Gangopadhyay Ford Motor Company
Farrukh Qureshi The Lubrizol Corporation
Issues and Challenges for Rear Axle Fluids. . . . . . . . . . . Fuel-Efficient Gear Lubricants . . . . . . . . . . . . . . . . . . . . . . . .
3-8 3-8
Axle Efficiency Tests • Vehicle Tests • Spin Loss Tests • Effect of Gear Surface Finish on Efficiency • Effect of Cold Start on Axle Efficiency • Limited Slip Differentials
3.8 Summary . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
3-19 3-20
This chapter will briefly discuss axle hardware, viscosity classifications, fluid requirements, and lubricant composition. The primary focus will be on development of fuel-efficient lubricants while maintaining axle durability. An automotive engine develops maximum power at a relatively high speed. The torque of the engine is modified in various stages until it becomes a propulsive force at an appropriate speed at the interface of tires and the road. The rear axle is the final stage in the drive train of a vehicle, which is responsible for appropriately transforming the engine power into useful propulsive force. The power from the engine and transmission is transferred to the rear axle through the driveshaft. The driveshaft is connected to the pinion gear inside the axle housing, which is partly filled with lubricant. Rotation of the driveshaft turns the pinion gear, which in turn rotates a contacting ring gear, as shown in Figure 3.1. The ring gear is connected to two shafts on both sides which extend outside the axle housing and connect to the wheels. The pinion gear is supported on two roller bearings and the two shafts connecting to the wheels are supported on two axle roller bearings mounted on the housing. Therefore, energy losses in the rear axle are due to (1) losses in the four bearings, (2) shearing of rear axle lubricant in the axle housing, and (3) frictional loss in pinion and ring gear contacts. The amount of preload on the pinion bearing also contributes to frictional loss. An important component of the rear axle system is the gear oil which is required to play a critical role in the efficient and durable operation of the rear axle. Axle durability is related to the pinion bearings and also depends on the wear of ring and pinion gears. The bearing durability is in turn related to lubricant temperature. Excessive gear wear results in gear whine, which causes customer dissatisfaction.
3-1
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Handbook of Lubrication and Tribology
Ring gear
Pinion gear
FIGURE 3.1 A view of ring and pinion gears inside the axle housing.
3.1 Rear Axle Lubrication Usually spiral-bevel or hypoid gears are used in rear axles. Fluids used in rear axles are required to reduce wear, pitting, spalling, scoring, other types of gear tooth distress to increase the life and reduce the downtime of the equipment. Additional requirements may also include protection against oxidation, rust, copper corrosion, and foaming. Since vehicles have to operate in diverse climates, viscometrics at both high and low temperatures must also be tailored to provide adequate fluid film for protecting the surfaces. Most of these requirements, developed over the years, can be satisfied by passing industry specified standard tests. These tests have been shown to define lubricants with adequate properties. Additional requirements such as improving fuel economy require close cooperation between equipment manufacturers and lubricant formulators so that desired fluid properties are tailored to axle design. Several of the standard requirements and specifications are discussed in the following sections, followed by a detailed discussion on improving the efficiency and durability of rear axle fluids.
3.2 Viscosity Classifications Gear oil viscosity is the most important parameter that governs the fluid film thickness between operating surfaces and along with chemical additives, technology determines the degree of protection available for gears and bearings in the rear axle system. The viscosity of a fluid is its resistance to shear deformation and is usually measured by shearing the fluid under controlled temperature and shear rates. Another important parameter is viscosity dependence on temperature, as represented by the viscosity index of fluid. A detailed description of standard test methods will not be included in this work, but references will be provided as appropriate. Fluid viscosity also depends on shear rate and ambient pressure. Automotive gear lubricant viscosities are defined in SAE J306. The SAE gear oil viscosity classifications are shown in Table 3.1. SAE J306 has been updated in October 2005 and new viscosity grades have been added. These designations are used to specify the viscosity requirements for manual transmission lubricants. Multigrade gear oils that can maintain their film-forming characteristics over a large temperature range are being increasingly used. In order to reduce the temperature dependence of viscosity these multigrade gear oils contain significant amount of polymers. However, polymers are also susceptible to shear degradation under service conditions. If these polymers shear in service, it will lead to a drop in viscosity, resulting in reduced film thickness, and may eventually lead to equipment failure. It is required that the lubricant remain in its viscosity grade
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Rear Axle Lubrication
3-3
TABLE 3.1 Automotive Gear Lubricant Viscosity Classification SAE viscosity grade 70W 75W 80W 85W 80 85 90 110 140 190 250
Max temperature for viscosity of 150,000 cP (◦ C)a,b
Kinematic viscosity at 100◦ C, cStc Minimumd
Kinematic viscosity at 100◦ C, cStc Maximum
−55e −40 −26 −12 — — — — — — —
4.1 4.1 7.0 11.0 7.0 11.0 13.5 18.5 24.0 32.5 41.0
— — — — <11.0 <13.5 <18.5 <24.0 <32.5 <41.0 —
Note: 1 cP = 1 mPa · sec; 1 cSt = 1 mm2 /sec. a Using ASTM D2983. b Additional low-temperature viscosity requirements may be appropriate for fluids inten-
ded for use in light duty synchronized manual transmissions. c Using ASTM D445. d Limit must also be met after testing in CEC L-45-A-99, Method C (20 h). The precision of ASTM Method D2983 has not been established for determinations made at temperatures below −40◦ C. This fact should be taken into consideration in any producer–consumer relationship.
following a 20-h shear stability test. The shear stability requirement provides assurance that the lubricant will maintain sufficient viscosity and oil film thickness to prevent premature failure of equipment. A set of standard viscosity identification and labeling guidelines was also provided in SAE J360. The guidelines provide a standard worldwide vocabulary to identify automotive gear lubricants. This should make it easier for consumers to select proper lubricant viscosity grades.
3.3 Gear Oil Classification Since a number of different gear designs and geometries are used in automotive applications, gear contacts are subjected to varying loads, speeds, and kinematics. Depending on gear design and operating environment, they require different gear lubricants for proper protection. In order to facilitate proper selection of gear oils for different applications, the American Petroleum Institute (API) has published gear oil designations. These designations are intended to assist manufacturers and users of automotive equipment in the selection of transmission, transaxle, and axle lubricants based on gear design and operating conditions. Some of these designations have been rendered obsolete but will be included in the discussion for the sake of completeness. A brief description of API service designations of gears oils follows.
3.3.1 Service Designations in Current Use API GL-1 denotes lubricants intended for manual transmissions operating under such mild conditions that straight petroleum oils may be used satisfactorily. Oxidation and rust inhibitors, defoamers, and pour depressants may be added to improve the characteristics of these lubricants. Friction modifiers and extreme pressure agents should not be used. Lubricants meeting MT-1 standards are an upgrade in performance over the lubricants meeting API GL-1 and are preferred by major commercial manual transmission vehicle manufacturers. API GL-4 denotes lubricants intended for axles with spiral bevel gears operating under moderate to severe conditions of speed and load or axles with hypoid gears operating under moderate speeds and
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loads. These oils may be used in selected manual transmission and transaxle applications where MT-1 lubricants are unsuitable. Although this service designation is still used commercially to describe lubricants, some test equipment used for performance verification is no longer available. API GL-5 denotes lubricants intended for gears, particularly hypoid gears, in axles operating under various combinations of high-speed shock loads and low-speed, high-torque conditions. Lubricants qualified under MIL-L-2105D satisfy the requirements of the API GL-5 specification, although the API designation does not require military approval. API MT-1 denotes lubricants intended for nonsynchronized manual transmission used in buses and heavy-duty trucks. Lubricants meeting the requirements of API MT-1 service provide protection against the combination of thermal degradation, component wear, and oil seal deterioration. API MT-1 does not address the performance requirements of synchronized transmissions and transaxles in passenger cars and heavy-duty applications.
3.3.2 Service Designations Not in Current Use API GL-2 denotes lubricants intended for automotive worm-gear axles operating under such conditions of load, temperature, and sliding velocities that lubricants satisfactory for API GL-1 service will not suffice. Products suited for this type of service contain antiwear additives for film-strength improvers specifically designed to protect worm gears. API GL-3 denotes lubricants intended for manual transmissions operating under moderate to severe conditions and spiral-bevel axles operating under mild to moderate conditions of speed and load. These service conditions require a lubricant with load-carrying capacities exceeding those satisfying API GL-1 service but below the requirements of lubricants satisfying API GL-4 service. Gear lubricants designated for API GL-3 service are not intended for axles with hypoid gears. API GL-6 denotes lubricants intended for gears designed with very high pinion offset. Such designs typically require protection from gear scoring in excess of that provided by API GL-5 gear oils. A shift to more modest pinion offsets and the obsolescence of original API GL-6 test equipment and procedures have greatly reduced commercial use of API GL-6 gear lubricants. Another relatively new performance classification is MIL-PRF-2105E. This combines the performance requirements of its predecessor (MIL-L-2105D) and API MT-1. This specification had not been widely adapted globally. The primary reason was that it had not been possible for oil blenders and marketers in non-NATO countries to obtain a formal approval under this specification. This has changed with the release of SAE J2360. In a follow-up of a government directive in 1991, The Society of Automotive Engineers (SAE) released the J2360 classification in 1998. Effective January 2004, this classification has replaced MIL-PRF-2105E. Qualification to the SAE J2360 standard will continue to require a review of test data and parts by the Performance Review Institute (PRI). This is a process unique to the gear lubricant industry where a panel of industry experts meets to review performance data and parts to qualify a standard. The PRI will administer the approval process and maintain a qualified products list on its web site. Under this classification, an oil marketer or blender anywhere in the world can now obtain a formal approval and have its name and the name of its approved lubricant published on the qualified products list. This enables it to demonstrate a measurable and recognized quality of performance for its lubricants and should lead to a greater return on investment in the development of these lubricants.
3.4 Performance Requirements for Gear Oils In order to meet performance requirements, gear oils are subjected to different tests and they must meet certain criteria to pass the tests. A list of performance requirements for popular designations of gear oils
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Rear Axle Lubrication TABLE 3.2
3-5
Performance Test Requirements
Test Procedure
API GL-5
MIL-L-2105D
API MT-1
MIL-PRF-2105E and SAE J2360
X X X X — — — — — X — —
X X X X — — — — X X X X
— — — — X X X X X X — —
X X X — X X X — X X X X
ASTM L-33 ASTM D 6121 (L-37) ASTM L-42 ASTM L-60 ASTM D5704 (L-60-1) ASTM D 5662 ASTM D 5579 ASTM D 5182 ASTM D 130 ASTM D 892 Storage stability and compatibility Controlled field test
TABLE 3.3 Automotive Gear Oil Tests Test ASTM L-33
Description
ASTM L-60 ASTM D 5704 ASTM D 5662 ASTM D 5579
Gear test using differential assembly Gear test using complete axle assembly Gear test using complete axle assembly Bench test using spur gears Bench test using spur gears Bench test Gear test
ASTM D 5182 ASTM D 130 ASTM D 892
Gear test Bench test Bench test
ASTM D 6121 ASTM L-42
Characteristics measured Resistance to corrosion in presence of moisture Resistance to gear distress under low-speed, high-torque conditions Resistance to gear distress (scoring) under high-speed, shock-load conditions Oxidative stability Thermal and oxidative stability and deposits Seal compatibility Transmission cyclic durability (waived for approved lubricants) Spur gear wear Stability in the presence of copper and copper alloys Foaming tendencies
is given in Table 3.2. Brief descriptions of these tests as well as the characteristics that are measured during each test are listed in Table 3.3.
3.5 Gear Oil Composition Automotive gear oil is composed of base oil, viscosity modifier, pour point depressant, and performance package. A brief description of individual gear oil components will be provided in the following sections.
3.5.1 Base Oil Base oil is a major constituent of any gear lubricant. Base oil acts as a lubricant, heat transfer medium, debris carrier, and carrier for the performance package. API has classified base oils into different groups based on the amount of saturates and sulfur level. API base oil groups are listed in Table 3.4. The base oils can be either processed from mineral oils (API Group I–III) or chemically synthesized. A careful cost/benefit analysis must be carried out before base oil selection. Solvent-refined mineral oils are cheapest and cost increases as more steps in refining are involved. Synthetic fluids such as polyalphaolefins (PAOs), polyolesters (POEs), or polyalkyleneglycols (PAGs) provide unique properties but are usually more expensive than mineral-based fluids.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology TABLE 3.4 API Classification of Base Stocks API Group I II III IV V
Sulfur
Saturates
>0.03% <0.03% <0.03% Polyalphaolefins Those base stocks not included above
<90% >90% >90%
Viscosity index 80–120 80–120 >120
Another class of base oils processed from natural gas known as Fischer–Tropsch fluids or gas to liquid (GTL) fluids is also emerging. GTLs are expected to be commercially available by 2006–2007. It is anticipated that these fluids will provide performance equivalent to or better than API Group III fluids at a competitive price.
3.5.2 Viscosity Modifiers and Pour Point Depressants Another major component of multigrade gear oils is viscosity modifier. Viscosity modifiers are simply polymeric molecules whose molecular conformation in oil solution is sensitive to temperature. At low temperatures, the polymer chains remain curled up and do not appreciably impact the fluid viscosity. At higher temperature, the polymeric molecular chains relax and open up, imparting an increase in viscosity to offset some loss of base oil viscosity with increase in temperature. Other performance characteristics of viscosity modifiers are their solubility and their ability to resist chain scission due to shear under high shear rates that may approach 10−7 sec. In order to ensure that the viscosity modifier remains intact under severe shear rates, SAE J306 specifies that the fluid should remain within 10% of its original 100◦ C viscosity after shear in a tapered roller bearing device (CEC- L-45-T-93). The choice of viscosity modifiers will impact the pressure viscosity behavior of formulated lubricants, which in turn impacts the film-forming characteristics of gear oil under elastohydrodynamic (EHD) lubrication conditions as well as its fuel economy characteristics. Pour point depressants are usually needed for mineral oils that have a tendency to form wax crystals at low temperature. Pour point depressants are high molecular weight polymers that tend to prevent the growth and aggregation of wax crystals.
3.5.3 Performance Package Since a gear oil is required to provide performance that cannot be delivered by a simple mix of base oil and viscosity modifier, a performance package is usually necessary in the gear oil. In general, gear oil consists of 5 to 15% of performance package, which may contain: • • • • •
Antiwear/extreme pressure additives Oxidation inhibitors Corrosion inhibitors Foam inhibitors Friction modifiers
3.5.3.1 Antiwear/Extreme Pressure Agents Antiwear/extreme pressure agents are organic compounds that may contain sulfur, phosphorus, boron, and zinc. These compounds react with metal surfaces to form protective films under boundary lubrication conditions. Antiwear additives, such as organophosphorus additives and their degradation products, are believed to form polar species that react with ferrous-based surfaces to form iron organo/inorganophosphorus films. These films are effective at low to moderate temperatures and loads.
© 2006 by Taylor & Francis Group, LLC
Rear Axle Lubrication
API GL-5 80W-90 Pass duration — 24 h
FIGURE 3.2
3-7
API GL-1 SAE 90 Failed at: 3 min
API GL-1 SAE 250 Failed at: 10 min
Gear test results from the L-37 test.
Extreme pressure agents are commonly used in gear oils; these are considered to be more effective under heavy load and high temperature conditions. The extreme pressure agents become active at the higher contact temperature that results from metal-to-metal sliding at the contact under heavy load. Extreme pressure additives react with gear surfaces to form protective films which prevent metal-to-metal contact. The presence of antiwear/extreme pressure additives is essential in gear oils, particularly with modern systems which operate under high contact pressure, higher temperature, smaller sump size, and longer drain intervals. The effectiveness of antiwear/extreme pressure additives is shown in Figure 3.2, which compares the results from an L-37 test for GL-1 oils without antiwear/extreme pressure additives and GL-5 gear oil (SAE 80W-90) that contains antiwear/extreme pressure additives. A proper balance of antiwear/extreme pressure additives is essential to increase the gear life under various operating conditions. Addition of extreme pressure additives alone does not guarantee surface protection. The fatigue life of gear components is slightly affected in the absence of extreme pressure additives; however, the absence of antiwear additives significantly reduces gear fatigue life. The effectiveness of antiwear/extreme pressure agents is determined using the L-37, L-42, and FZG scuffing and wear tests listed in Tables 3.2 and 3.3. 3.5.3.2 Oxidation Inhibitors Most lubricants are susceptible to oxidation since they are made of hydrocarbon base oils. Oxygen present in the atmosphere reacts with the hydrocarbon molecules at higher temperature resulting in peroxy or other radicals. Increase in temperature and presence of metallic wear particles accelerates the oxidation reactions. The oxidation of lubricants occurs via a free radical reaction, leading to base oil polymerization. The polymerization of base oil increases viscosity and impairs the flow characteristics of a lubricant, which is essential for adequate film formation. Due to the abrasive nature of some oxidation byproducts, elastomeric seals may also abrade and fail. Oxidation inhibitors can be classified as hydroperoxide decomposers and radical scavengers, depending upon the mode of action. Sulfur-containing compounds act as decomposers. These compounds react with hydroperoxide radicals and become oxidized to higher oxidation states. Nitrogen- and oxygen-containing compounds such as arylamines and phenols act as radical scavengers and render radicals innocuous, usually by an oxidation–reduction reaction. 3.5.3.3 Corrosion Inhibitors As a result of oxidative degradation of gear oils, acids may be introduced that chemically attack surfaces. If not properly formulated some extreme pressure agents may also attack surfaces. In order to prevent corrosion, certain inhibitors such as basic sulfonates and fatty amines may be used. The effectiveness of corrosion inhibitors is measured by the L-33 and ASTM D-130 tests.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology Light-duty vehicle horsepower trends 240 Horsepower
220 Cars
200
SUVs
180
Pickups
160
All light duty
140 120 1992
1994
1996 1998 2000 Model Year ('93–'03)
2002
2004
FIGURE 3.3 Vehicle horsepower trend over a 10-yr period.
3.5.3.4 Foam Inhibitors The churning action of gears in transmission and axles causes air entrainment or foaming, particularly at lower temperature. In order to avoid foaming, certain inhibitors such as polydimethylsiloxanes (“silicones”) and polyacrylates are used. 3.5.3.5 Friction Modifiers Friction modifiers are adsorbed on metal and other surfaces and result in reduced friction. Friction modifiers are particularly needed for modern axle designs equipped with limited slip capability and may also reduce the operating temperature due to reduced frictional heating at the contact.
3.6 Issues and Challenges for Rear Axle Fluids Over the years, light trucks, minivans, and sports utility vehicles (SUVs) have gained in popularity. In the years 2002–2003, vehicle production and sales in North America for the light truck segment exceeded those for the car segment for the first time in history [1]. The increase in the number of vehicles in the light truck segment has technical implications for delivering performance economically and efficiently. There have been significant improvements in drive train technology, which have resulted in improvement in performance, efficiency, and durability. These gains have been offset by consumer craving for higher performance. Technology improvements are being used to deliver greater power and performance, while the components that deliver performance have either remained virtually the same size or have become smaller. Increasing use of independent rear suspension has changed vehicle aerodynamics. The reduction in sump size, longer drain intervals, and restricted airflow for convective heat transfer have resulted in a challenging atmosphere of higher axle operating temperature. Figure 3.3 indicates that there has been an increase in horsepower of about 34% over a 10 yr period. There has been similar increase in power density (measured as a ratio of horsepower to vehicle weight) of about 20% during the same time period.
3.7 Fuel-Efficient Gear Lubricants Over the past few years, customer preference for light-duty trucks (including minivans and SUVs) has increased significantly. However, a major point of customer dissatisfaction for light duty truck owners has been their low fuel economy. Additionally, government regulatory agencies have continued to push for further increases in corporate average fuel economy (CAFE) requirements for cars and light-duty trucks. The CAFE standards will rise in the year 2005 and beyond. Therefore, automakers are considering several new engine and transmission concepts to improve fuel economy. The automotive rear axle also contributes to frictional loss, but less than major powertrain systems. The automakers are closely scrutinizing all
© 2006 by Taylor & Francis Group, LLC
Rear Axle Lubrication
3-9
systems and exploring where further frictional losses could be reduced. As a result, even inherently efficient systems such as gearboxes and axles have not escaped close scrutiny. It is anticipated that frictional losses in rear axles can be reduced through the use of lower viscosity rear axle lubricant, novel gear surface finishes, improved gear design, lower friction bearings and so on. During the last few decades, a number of well-documented attempts have been made to develop fuel-efficient gear lubricants primarily through use of lower viscosity and multigrade viscosity grades. Naman [2] examined the effect of 1% addition of MoS2 in a 75W viscosity grade lubricant and observed no fuel economy benefit in a fully warmed-up vehicle. O’Connor et al. [3] studied the effect of synthetic base content in axle lubricant on efficiency. They showed that the type of synthetic base oil had a relatively insignificant impact on axle efficiency, but a bigger effect was observed when the lubricant composition was changed from part synthetic to full synthetic. Law [4] showed that use of synthetic base stocks resulted in improved efficiency in automotive hypoid gears. Watts and Willette [5] measured axle efficiency and truck fuel economy with a few different viscosity grade lubricants. SAE 75W-90 and SAE 75W-140 viscosity grade lubricants improved axle efficiency over a SAE 90 grade lubricant by about 1.3 and 0.55%, respectively. This translated into about 1.2 to 0.7% improvement in fuel economy at 65 mph speed. Willermet and Dixon [6] evaluated the effect of lubricant viscosity on axle efficiency and based on these results predicted about 1% improvement in fuel economy by adjusting low- and high-temperature viscosities. The additive package also plays an important role in improving axle efficiency. Adams et al. [7] evaluated the effect of two additive packages on truck fuel economy, one containing a conventional S-P additive package and the other containing a potassium triborate and zinc dialkyldithiophosphate-based package. The triborate additive package showed about 1.1% fuel economy improvement over the conventional axle lubricant. Also, the lubricant temperature was 30◦ C lower than that observed with the conventional lubricant. Greene and Risdon [8] looked into the effect of molybdenum-based friction modifiers on axle efficiency. Molybdenum-based friction modifiers improved axle efficiency and the degree of improvement depended on the amount of Mo present in the lubricant. Increasing the Mo content from 0.1 to 0.3% increased axle efficiency from 1–1.8% to 2.6–2.9% under the low-speed/high-torque condition. Further improvement in axle efficiency was observed under a high-speed/low-torque condition. Recently, Bala et al. [9] evaluated the effect of lubricant viscosity, additive package, and lubricant viscosity index (VI) on axle efficiency under simulated city and highway driving conditions using an axle efficiency rig. They observed that lubricant viscosity and VI affected axle efficiency under different load speed conditions and sometimes in opposite directions, but generally lubricants with high VI appeared to show higher axle efficiency. Vinci et al. [10] formulated axle lubricants which provided a balance between temperature control and efficiency through proper selection of base oils, thickening agents, and performance additives. In order to improve axle efficiency, lubrication regimes must be considered, factors leading to energy dissipation in each regime be identified, and such energy dissipation be minimized using appropriate measures. Gear contacts operate in multiple lubrication regimes, resulting in energy dissipation both inside and outside the contact region. Energy dissipation outside the contact leads to load-independent losses and use of lower-viscosity fluids is of help. However, at higher operating temperature or under realistic operating conditions, lower-viscosity fluids may not provide fluid films that are thick enough to protect surfaces. This may lead to reduced bearing and gear life and premature failure. Gear contacts are heavily loaded and operate in EHD lubrication, as well as in mixed/boundary lubrication regimes. In EHD lubrication regime energy dissipation as well as the film formation characteristics of fluids depend on their pressure-viscosity behavior, in addition to nominal fluid viscosity. It has been shown that the pressure-viscosity coefficient of a fluid is directly proportional to average friction in a rolling/sliding EHD contact. The average EHD friction in a rolling/sliding contact can be directly measured and is designated as the traction coefficient. The traction coefficient of gear oils depends directly on the type of base oil and viscosity modifier. The traction coefficients of different base oils with similar nominal viscosity at 100◦ C are compared in Figure 3.4. Clearly, synthetic fluids such as PAOs exhibit lower traction coefficients than mineral oils. Viscosity modifiers are used in modern multigrade gear oils to reduce their viscosity–temperature sensitivity. They
© 2006 by Taylor & Francis Group, LLC
3-10
Handbook of Lubrication and Tribology Traction coefficient 100°C at 1.25 GPa, 2.5 m/sec
Traction coefficient
0.06
Group I
Group II
Group III
PAO-4
0.05 0.04 0.03 0.02 0.01 0 0
10
20
30
Slide to roll ratio
FIGURE 3.4
Traction coefficient of base oils at 100◦ C.
100 90 Efficiency, %
80 70 60 50 40 EPA FE range
Performance range
30 20 0
500 1000 1500 Output torque, N m
2000
FIGURE 3.5 A general axle efficiency curve. (Adapted from Gangopadhyay et al. SAE 2002-01-2821, with permission.)
also profoundly impact the pressure–viscosity behavior and traction coefficient of gear oils. It is desirable to reduce EHD friction to improve the efficiency and reduce the operating temperature of axle fluids. As the EHD film starts to become thin due to operating conditions, contact transitions to boundary lubrication regime and surface friction start to contribute to energy dissipation.
3.7.1 Axle Efficiency Tests Generally, improvement in axle efficiency is measured as a function of lubricant chemistry using a production rear axle which is coupled with a transmission and fired engine, as shown in Figure 3.5. The torque input to the axle is measured by an in-line torque meter and the output torques are measured by the load cells connected to a dynamometer. Axle efficiency is calculated from these measurements using a known gear ratio. It should be noted that a test axle is used for multiple runs. One should be cognizant of the fact that the axle performance continues to change as it accumulates test runs. It is essential that the axle is periodically referenced for its performance. A frame work for effectively considering the hardware changes during the life of an axle, as outlined by Akucewich et al. [11] provides valuable guidance. Figure 3.5 shows a typical axle efficiency curve, which consists of two parts: the low-torque range, which is related to Environmental Protection Agency (EPA) fuel economy cycle, and the high-torque range, which represents axle performance and durability (particularly pinion bearing) under trailer/tow
© 2006 by Taylor & Francis Group, LLC
Rear Axle Lubrication
3-11
FIGURE 3.6 A picture of experimental setup to measure axle efficiency. (Adapted from Gangopadhyay et al. SAE 2002-01-2821, with permission.)
condition. The test is conducted under a variety of load and speed conditions reflecting idling, city driving, highway driving, and trailer/tow conditions. In this test large-capacity torque meters are used to cover the entire range, and therefore, there are some uncertainties in the accuracy of efficiency measurement in the low-torque region. Although efficiency is higher than 95% in the high-torque region, in the low-torque region efficiency is much lower and, therefore, provides an opportunity for improvement. In the hightorque region, efficiency does not vary a lot, but lower lubricant temperature is desired for maintaining bearing durability. For axle efficiency improvement studies, it is convenient to run tests under low-torque conditions to screen potential candidates and then run tests under high-torque conditions for durability. Figure 3.6 shows another setup to evaluate axle efficiency under low torque conditions. This setup utilized a 223 mm (8.8 in.) diameter ring gear set and 3.55 axle ratio from a full size pickup truck. The pinion was driven by one motor while the energy output on the other side was absorbed by another motor. The electric drive controlled the speed on the input motor and the torque on the output motor. The output shaft on one side of the axle was shortened and prevented from rotation by welding a pair of side gear teeth. Two kinds of experiments were conducted: (1) where the lubricant was allowed to heat naturally due to friction and (2) where the temperature of the lubricant was held constant. For the latter, the axle housing was connected to an external oil sump where the lubricant was heated to the desired temperature by a heater and pumped in and out of the axle housing through a heat exchanger. The lubricant level was always maintained as specified in the owner’s manual. The input and output torques were measured by the electric drive system within ±5% error. The lubricant temperature was recorded by inserting a thermocouple in the axle housing. 3.7.1.1 Effect of Axle Speed and Lubricant Temperature on Axle Efficiency The effect of axle speed on efficiency is shown in Figure 3.7 during natural heating of axle. Generally, efficiency increases with applied torque because the lubricant temperature also increases with torque, as shown in Figure 3.8. The lubricant temperature generally keeps rising with continued operation due to constant frictional heat input. Therefore, axle efficiency was calculated when the temperature stabilized and did not vary more than 1◦ C for about 15 min. The increased temperature decreases lubricant viscosity, and therefore, decreases hydrodynamic drag, resulting in increased efficiency. Also, the higher the speed, the higher the efficiency because the film thickness is higher between gear surfaces, resulting in less severe asperity contact, and asperity friction is higher than fluid friction. The lubricant temperature is also higher at higher speed, leading to reduced lubricant viscosity, which will increase asperity friction. But the film
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology 1.00 0.95
Efficiency
0.90 0.85 0.80
1000 rpm 2000 rpm
0.75
2750 rpm 0.70 0
100
200
300
400
500
600
500
600
Output torque, N m
FIGURE 3.7
The effect of speed on axle efficiency. 160 1000 rpm
Oil temperature, °C
140
2000 rpm 2750 rpm
120 100 80 60 40 0
100
200
300
400
Output torque, N m
FIGURE 3.8
The effect of torque on lubricant temperature.
thickness increase due to increased speed must be greater than the film thickness decrease due to increased temperature. Figure 3.9 shows axle efficiency as a function of speed when tests were conducted at constant lubricant temperature of 113◦ C. The efficiency decreased with increasing speed at lower loads, which is opposite to what was observed when the lubricant was heated naturally. At a constant temperature, the lubricant viscosity is constant and, therefore, at higher speed the film thickness is higher than at lower speed. Therefore, at low torque, efficiency decreased due to increase shear loss, but at high torque, efficiency improved due to increase in load-dependent losses as compared with load-independent losses. At a given speed, the increase in efficiency due to increased temperature is shown in Figure 3.10, where the efficiency increased with increase in temperature at low torque due to decreased shear loss. 3.7.1.2 Effect of Lubricant Formulations on Axle Efficiency and Temperature Lubricant formulation plays a significant role in controlling axle efficiency and durability. In particular, the lubricant viscosity plays a more significant role in improving axle efficiency under low torque while the additive package plays a significant role in durability, in particular clutch pack for limited slip differential.
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Rear Axle Lubrication
3-13 1.00
113°C lubricant temperature
0.95
Efficiency
0.90 0.85 0.80 0.75 1000 rpm 0.70
2000 rpm
0.65
2750 rpm
0.60 0
100
200 300 400 Output torque, N m
500
600
FIGURE 3.9 Axle efficiency as a function of torque at different speeds at a constant lubricant temperature of 113◦ C.
1.00 2000 rpm
0.95
Efficiency
0.90 0.85 0.80 Temp. 75°C
0.75
Temp. 91°C 0.70
Temp. 113°C
0.65 0.60
0
100
200
300
400
500
600
Output torque, N m
FIGURE 3.10 Axle efficiency as a function of torque at a constant speed.
Figure 3.11 shows axle efficiency under various torque conditions at 1000 rpm for various lubricant formulations, mostly on 75W-90 and a few 75W-140 viscosity grades. The viscosities of the lubricants used in this test program are shown in Table 3.5. Lube A has the highest viscosity at 40◦ C and showed lowest efficiency, whereas Lube D has the lowest viscosity at 40◦ C and exhibited the highest efficiency at the lowest torque of 112 N m. The lubricant temperature at corresponding torques is shown in Figure 3.12. Although Lube B and Lube E have similar viscosities, and although the lubricant temperatures at lower torques are similar, they showed a difference in efficiency, probably highlighting the impact of chemistry. A similar effect could be observed for Lube A and Lube C , both of which are in the 75W-140 viscosity grade. The axle efficiency curves can be better understood when the results are plotted in the form of Stribeck curves, as shown in Figure 3.13, which shows that the test axle operates in both mixed and hydrodynamic lubrication regimes. The higher-viscosity Lube A operates primarily in the hydrodynamic regime at 1000 rpm, and the frictional loss is higher than any of the other lubricants. The torque loss of 75W-90 viscosity grade lubricants is comparable in the hydrodynamic regime, but they distinguish themselves in the mixed lubrication regime. In the mixed lubrication regime, Lube A and Lube B shows the highest
© 2006 by Taylor & Francis Group, LLC
3-14
Handbook of Lubrication and Tribology 1.00 Lube A 75W-140 Lube B 75W-90 Lube C 75W-90 Lube D 75W-90 Lube E 75W-90 Lube C9 75W-140
0.95
Efficiency
0.90
1000 rpm
0.85 0.80
0.75 0.70 112
171
229
340
450
568
Output torque, N m
FIGURE 3.11 Axle efficiencies of various lubricants. TABLE 3.5 Viscositites of Lubricants Investigated Lubricants
Viscosity, cSt
Lube A 75W-140 Lube B 75W-90 Lube C 75W-90 Lube C 75W-140 Lube D 75W-90 Lube E 75W-90
40◦ C
100◦ C
Viscosity index
192.2 110.4 136.7 170.6 64.2 116.1
24.3 17.7 22.8 25.7 14.5 18.1
172 177 190 227 168
Source: Adapted from Gangopadhyay et al. SAE 2002-01-2821, with permission.
120
Axle temperature, °C
110 100
Lube A 75W-140 Lube B 75W-90 Lube C 75W-90 Lube D 75W-90 Lube E 75W-90 Lube C9 75W-140
1000 rpm
90 80 70 60 50 112
FIGURE 3.12
171
229 340 Output torque, N m
The axle sump temperature for various lubricants.
© 2006 by Taylor & Francis Group, LLC
450
568
Rear Axle Lubrication
3-15 50 1000 rpm
Torque loss, N m
45 40 35 Lube A 75W-140 Lube B 75W-90 Lube C 75W-90 Lube D 75W-90 Lube E 75W-90
30 25 20
0
100
200
300
400
500
Viscosity × speed/torque
FIGURE 3.13 Stribeck curves for lubricants tested at 1000 rpm. (Adapted from Gangopadhyay et al. SAE 2002-012821, with permission.) 1.6
5.4 L navigator
% FE improvement
1.4
4.6 L F-150
1.2 1 0.8 0.6 0.4 0.2 0 Lube B
Lube C
Lube D
Lube B
FIGURE 3.14 Vehicle fuel economy obtained with various lubricants. (Adapted from Gangopadhyay et al. SAE 2002-01-2821, with permission.)
torque loss, whereas Lube E shows the lowest. This provides an opportunity for Lube A and Lube B to improve efficiency further by reducing losses in the mixed lubrication regime.
3.7.2 Vehicle Tests Figure 3.14 shows the improvement in fuel economy over Lube A with different 75W-90 lubricants, within a 90% confidence interval. These results were obtained in chassis roll dynamometer tests under Federal Test Protocol (FTP) metro/highway cycles. The results show significant fuel economy improvement and the level of improvement also depended on the vehicle being tested. Figure 3.15 shows axle temperature profile during chassis roll dynamometer tests in a Lincoln Navigator model year 2001, equipped with 5.4-L four-valve engine. Again, the dependence of lubricant viscosity on lubricant temperature could be observed. The highest-viscosity lubricant, Lube A, showed the highest lubricant temperature, whereas the lowest-viscosity lubricant, Lube C, showed the lowest.
3.7.3 Spin Loss Tests Spin loss tests are another way of evaluating the fuel economy improvement potential of rear axle lubricants. Such tests measure the power required to drive the axle in coast down mode. Specially equipped
© 2006 by Taylor & Francis Group, LLC
3-16
Handbook of Lubrication and Tribology
Axle temperature, °C
80
Lube A 75W-140 Lube B 75W-90 Lube C 75W-90 Lube D 75W-90
70 60 50 40 30 20
0
1000
2000
3000
4000
Test time, sec
FIGURE 3.15 Axle lubricant temperature during chassis roll tests.
dynamometers with high-precision torque transducers are used to measure torque loss when the axle is rotated either through the wheel or through the pinion from 70 to 0 mph. The measured torque loss is converted into horsepower (hp) loss. Reduced hp loss translates into improved fuel economy. Bjornen et al. [12] demonstrated that a lower-viscosity grade 75W-90 lubricant exhibits lower spin loss, which correlates with higher estimated fuel economy.
3.7.4 Effect of Gear Surface Finish on Efficiency It is observed that the ring and pinion gear contact operates in mixed and hydrodynamic lubrication regimes, the relative portion of which depends on operating conditions including surface roughness and lubricant viscosity. Some of the 75W-90 lubricants pushed the contact well into mixed lubrication regimes. Therefore, it is expected that further reduction in friction in the mixed lubrication regime could be possible by reducing the surface roughness of gears. The ring and pinion gear surfaces were superfinished using a chemo-mechanical technique. The isotropic finishing process is a two-step chemo-mechanical process. The first step is called the refining step, where the parts are put into a slowly rotating standard vibratory device containing a mild acidic solution and ceramic media. The acidic solution reacts with the metal surface and leaves a very thin soft film, which is then removed by abrasion with the ceramic media. The surface film is continuously reformed and removed, resulting in material removal from the surface. The second stage of the process is called the burnishing stage, where a basic solution is added in the vibratory bowl to neutralize the refining solution and also to remove any soft film remaining on the surface. This process results in a very smooth surface, on the order of 0.07 µm Ra . The typical production surface finish is 1 to 1.5 µm Ra . Figure 3.16 compares the axle efficiency measured at 1000 rpm between isotropic finish gears and production gears. The results demonstrated significant improvement in axle efficiency as well as lower axle operating temperatures with isotropic finish gears compared to production gears as shown in Figure 3.17. The axle efficiency improvement resulted in about 0.5% improvement in fuel economy in chassis roll dynamometer tests under metro/highway cycles. Ring and pinion gear finish also plays an important role in durability. Generally, the gears are phosphated to enhance wear resistance and assist in running-in of surfaces. These coatings have excellent lubricity and scuffing resistance under lubricated conditions; however, the coating itself is not very lubricous. Due to its porous nature, the coating provides lubricity by absorbing a large amount of lubricant in its porous structure. It has been claimed that the supply of lubricant from these pores helps to prevent severe damage to gear surfaces [13,14]. The coating morphology plays a critical role in its ability to absorb lubricant. The coating morphology depends primarily on coating process as well as on underlying roughness of gear surfaces. The surface roughness of the gear surface also plays an important role in gear wear. A high
© 2006 by Taylor & Francis Group, LLC
Rear Axle Lubrication
3-17 1.00
Production 1000 rpm
Isotropic finish
0.95
Efficiency
0.90 0.85 0.80 0.75 0.70 112
171
229
340
450
568
Output torque, N m
FIGURE 3.16 Axle efficiency of isotropic finish ring and pinion gears in contact with Lube B. 120 1000 rpm
Temperature, °C
110 Production Isotropic finish
100 90 80 70 60 50 112
171
229 340 450 Output torque, N m
568
FIGURE 3.17 Axle lubricant temperature with isotropic finish ring and pinion gear set at 1000 rpm.
surface roughness will probably push the components to more boundary and mixed lubrication regimes, leading to wear. On the other hand, a smoother surface finish could potentially increase cost. Therefore, an optimum surface roughness needs to defined based on cost and performance.
3.7.5 Effect of Cold Start on Axle Efficiency Fuel economy tests according to FTP cycles are conducted under a controlled room temperature. However, customers drive under a variety of temperatures depending on their geographic location. Generally, the axle warms up quite fast and, therefore, actual axle lubricant temperature during driving may not vary much whereas vehicle-starting temperature varies a lot depending on locations. A lower starting temperature would result in higher lubricant viscosity, which may result in lower initial fuel economy until the axle is fully warmed up. Therefore, it is of interest to know the effect of cold start temperature on axle efficiency. In order to assess the effect of cold start temperature on axle efficiency, the axle housing was wrapped around with nylon tubes through which liquid nitrogen was passed, as shown in Figure 3.18. The ring and
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
FIGURE 3.18 A close-up view of the arrangement to cool axle lubricant.
pinion gears were rotated slowly (around 10 to 15 rpm) to equilibrate the temperature of the components inside the axle housing without generating any significant frictional heat. When the desired temperature was reached, the liquid nitrogen flow was stopped and the test started at the required torque and speed condition. Figure 3.19 shows the axle efficiency and lubricant temperature due to cold start and compares them with an ambient start under various torque and speed conditions. At any given condition, the efficiency was calculated when the lubricant temperature stabilized. Under ambient conditions, the initial efficiency is about 71% under 68 N m torque and 532 rpm speed and it increased gradually to about 73% when the lubricant temperature stabilized. For the cold start, it took about 40 min to cool the lubricant to about −14◦ C. Upon starting, the efficiency was only about 41%, significantly lower than for the ambient start. The efficiency increased rapidly with the increase of lubricant temperature and finally leveled at about 67%, less than that observed with an ambient start although the stabilized lubricant temperature was similar, 42◦ C compared with 41◦ C. As the torque was increased to 611 N m, the axle efficiency and the lubricant temperature increased but the initial efficiency difference between the ambient start and the cold start diminished to 4.6% (Figure 3.19[b]). The stabilized lubricant temperatures for ambient and cold starts are about the same but the difference in respective efficiencies was only less than 0.5%. At 68 N m torque, as the speed is increased from 532 to 1775 rpm, both the axle efficiency and the lubricant temperature increased. The initial difference in axle efficiency between ambient and cold starts narrowed to 13.5% and, after stabilization of lubricant temperature, the difference was less than 1%. At the highest torque and speed condition investigated (Figure 3.19[d]), the initial efficiency difference widened to 11.5% and, after the lubricant temperature stabilized, the difference reduced to 6.6%. The larger difference in efficiency under these conditions could be due to the lower initial temperature of −22◦ C compared with about −13◦ C for the other starts. The results demonstrated significant efficiency loss at the beginning of a cold start, but the loss diminished as the lubricant warmed up. However, a significant efficiency loss was observed even after the lubricant temperature stabilized under low-torque, low-speed conditions.
3.7.6 Limited Slip Differentials More vehicles are being equipped with limited slip differentials for improved handling. When one wheel begins to spin, the limited slip differential provides more torque to the other wheel to move the vehicle. In other words, it limits the amount of differential action between the wheels. This differential action is obtained through the use of clutch packs, each consisting of a combination of steel plates and friction plates. One of the issues with limited slip differential is shudder, which translates into chatter to the driver. The shudder is related to stick slip motion between the clutch plate friction material and the steel plate. Generally, a friction modifier is added with the axle lubricant to reduce stick slip. The effectiveness of the friction modifier in reducing stick slip can be evaluated in a specially equipped rig where a friction material
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Rear Axle Lubrication
3-19 (b) 140
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100 0.80 0.70
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–20
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Efficiency
(a) 140
0.50 –20 0
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Cold sump
Cold efficiency
Amb. sump
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FIGURE 3.19 Comparison of axle efficiency between ambient start and cold start under (a) 532 rpm and 68 N m torque; (b) 532 rpm and 611 N m torque; (c) 1775 rpm and 68 N m torque; (d) 1775 rpm and 611 N m torque.
is loaded against a steel plate under a given load and lubricant temperature. The friction coefficient if measured as a function of speed. A positive slope of friction coefficient vs. speed curve is indicative of resistance to shudder. Another method to evaluate the effectiveness of a friction modifier on shudder is to load the rotating friction material against the steel plate under a given load and then measure the torque required to stop the rotation. The load, temperature, and speed of rotation are selected based on field correlation. In this test configuration, the stick slip phenomenon is reflected as oscillations in the friction coefficient vs. time curve. Another issue is the limited life of friction modifiers. The effectiveness of friction modifier decreases with aging. Therefore, it is very important to evaluate friction characteristics of lubricant both when fresh and used. Phosphorous-type friction modifiers are commonly used in gear oils, which can be acid phosphates, amine phosphates, or hydrogen phosphates. Okazaki et al. [15] investigated the structural change in friction modifiers in laboratory aging tests and blended a series of gear oils based on the compounds formed during aging. The evaluation of friction characteristics of these blends led to the development of a long-life friction modifier. Generally, the friction modifier acts on the clutch surface and does not necessarily impact friction at the ring and pinion gear surface [16].
3.8 Summary A critical component in rear axle systems is gear lubricant, and it plays a very important role in the efficient and durable operation of the system. Traditionally, the lubricant is required to provide thermal
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and oxidative stability, seal compatibility, adequate antifoaming tendencies, resistance to corrosion, and so on. However, modern vehicle designs and regulatory pressures are demanding additional performance from lubricants. The proliferation of higher horsepower, smaller sump size, longer drain intervals, and independent rear suspensions is pushing the lubricant temperature higher, leading to decreased film thickness while demand for increased fuel economy requires lower-viscosity lubricants. In contrast, higher towing capacity requires adequate durability, which requires higher viscosity lubricant to provide increased film thickness to prevent metal-to-metal contact. These conflicting requirements are making lubricant formulation very challenging. Increased use of limited slip differentials makes lubricant formulation to make a vehicle chatter-free during turns even more complex. These issues will continue to be significant as the CAFE requirements rise over coming years. The challenge to develop lubricants which improve the efficiency of rear axle systems while maintaining durability will continue to be faced by both the equipment manufacturers and fluid developers. A successful resolution to these issues will require close cooperation between equipment manufacturers and fluid formulators to use concurrent engineering techniques to optimize fluids for hardware with improved surface finish, optimum coating on gears, improved friction materials, and improved design of axle housing.
References [1] 2004 Market Data Book, Automotive News, May 2, 2004. [2] T. Naman, “Automotive fuel economy — potential improvement through selected engine and differential gear lubricants,” Society of Automotive Engineers Paper No. 800438, 1980, SAE International. [3] B.M. O’Connor, L.F. Schiemann, and R.L. Johnson, “Axle efficiency — response to synthetic lubricant components,” Society of Automotive Engineers Paper No. 821181, 1982, SAE International. [4] D.A. Law and C.N. Rowe, “The design of fuel efficient automotive hypoid gear lubricants,” Journal of Synthetic Lubrication, 11(1), 3–15, 1993. [5] R.F. Watts and G.L. Willette, “Newtonian multigrade gear lubricants: formulation and performance testing,” Society of Automotive Engineers Paper No. 821180, 1982, SAE International. [6] P.A. Willermet and L.T. Dixon, “Fuel economy — contribution of the rear axle lubricant,” Society of Automotive Engineers Paper No. 770835, 1977, SAE International. [7] J.H. Adams, K.A. Frost, L.M. Hartman, and L.J. Painter,“The effect of gear lubricant on fuel economy as measured in a line haul truck fleet,” Society of Automotive Engineers Paper No. 810179, 1981, SAE International. [8] A.B. Greene and T.J. Risdon, “The effect of molybdenum containing, oil-soluble friction modifiers on engine fuel economy and gear oil efficiency,” Society of Automotive Engineers Paper No. 811187, 1981, SAE International. [9] V. Bala, G. Brandt, and D.K. Walters, “Fuel economy of multigrade gear lubricants,” Technische Akademic Esslingen, Tribology 2000-Plus, 12th International Colloquim, January 11–13, 2000, Ed. W.J. Bartz. [10] J.N. Vinci, E.S. Akucewich, R.S. Cain, and F.S. Qureshi, “Developing next generation axle fluids: part II — systematic formulating approach,” Society of Automotive Engineers Paper No. 2002-011692, 2002, SAE International. [11] E.S. Akucewich, J.N. Vinci, F.S. Qureshi, and R.W. Cain, “Developing next generation axle fluids: part I — test methodology to measure durability and temperature reduction properties of axle gear oils,” Society of Automotive Engineers Paper No. 2002-01-1691, 2002, SAE International. [12] K.K. Bjornen, H. Chambers, and D. Degonia,“Development of a fuel efficient multipurpose 75W-90 gear lubricant,” Society of Automotive Engineers Paper No. 2003-0260, 2003, SAE International. [13] P. Hivart, B. Hauw, J.P. Bricout, and J. Oudin, “Seizure behavior of manganese phosphate coatings according to the process conditions,” Tribology International, 30(8), 561–570, 1997.
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[14] T. Oyamada and Y. Inoue, “Evaluation of the wear process of cast iron coated with manganese phosphate,” Tribology Transactions, 46(1), 95–100, 2003. [15] K. Okazaki, K. Noguchi, K. Motoyama, and T. Wakizono, “A study of friction characteristics and durability of LSD oils,” Society of Automotive Engineers Paper No. 932786, 1993, SAE International. [16] A.K. Gangopadhyay, S. Asaro, M. Schroder, J. Sorab, and R. Jensen, “Fuel economy improvement through frictional loss reduction in light duty truck rear axle,” Society of Automotive Engineers Paper No. 2002-01-2821, 2002, SAE International.
© 2006 by Taylor & Francis Group, LLC
4 Automotive Chassis and Driveline Lubrication 4.1 4.2
Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Grease Technology Overview . . . . . . . . . . . . . . . . . . . . . . . . .
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Standardized Testing • Classification Systems for Automotive Chassis and Driveline Lubricants • Chassis Grease Performance Specification and Classification — ASTM D4950 • ASTM D4950 Specification Summary
4.3
Lubrication of Chassis Systems . . . . . . . . . . . . . . . . . . . . . . .
4-5
Suspension and Steering Linkage System • Steering System • Steering Gears • Factors Affecting Lubrication of Steering Gears • Intermediate Shaft • Lubrication of I-Shaft Assemblies • Steering Column (see Figure 4.8[a] and [b]) • Overview of Grease Types Used for Steering Columns • Power Steering Hydraulic Fluid
4.4
Brake System Design and Lubricant Characteristics
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Brake Lubrication • External Brake Lubrication
4.5
Shock Absorber Lubrication . . . . . . . . . . . . . . . . . . . . . . . . . .
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Shock Absorber Fluid • Magnetorheological and Electrorheological fluids
4.6
Halfshaft Assemblies . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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Halfshaft Application Overview • Halfshaft Lubrication Requirements • Ball-Type CVJ • Trilobular Type CVJ • Plunging Joint Frictional Requirements • General Lubricant Selection — All Halfshaft Joint Types • Typical Lubricant Product Usage for Halfshaft Applications • Ongoing Developments in Halfshaft Lubricants • Summary of Halfshaft Lubricant Selection
4.7
Propeller Shaft Joint Lubrication . . . . . . . . . . . . . . . . . . . . .
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Simple Universal Joint • Cardan Joint (Double Hookes Universal Joint) • Constant Velocity Joint
4.8
Drive Axle Lubrication . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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Drive Axles and Limited Slip Differential System Lubrication • Lubricants for Traction Enhanced Differentials
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Douglas M. Jahn Delphi Corporation
Simon C. Tung General Motors R&D Center
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Automotive Wheel Bearings Lubrication. . . . . . . . . . . . .
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Wheel Bearing Overview • Wheel Bearing Grease Selection
4.10 A Vision of the Future . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Acknowledgments. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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4.1 Introduction Effective automotive chassis lubrication requires the understanding of many different tribological scenarios and is complicated by a vehicle’s exposure to a diverse and everchanging set of operating conditions. Exposure to environmental conditions such as temperature, mud, rain, dust, road salt and snow, as well as a variety of driver-controlled vehicle operating conditions will influence the chassis system’s performance and lubrication needs [1]. Ongoing advances in materials, processes, and surface engineering techniques [2] have significantly impacted automotive chassis tribological systems and thus, have impacted how we specify lubricants for those systems; the ultimate intent being that of achieving reliability and robustness in those systems. In addition, the expected useful life of vehicles and their subsystems has increased by as much as 100% in the last 20 yrs, with vehicle OEM (Original Equipment Manufacturer) warranties extending out to, and in some cases beyond, 150,000 miles. This chapter begins with a brief overview of basic grease technology [3], attempting to familiarize the reader with the general terminology, design, and attributes of lubricating greases. A short commentary then follows regarding lubricant pricing vs. performance; an area that application engineers responsible for specifying lubricants should be conscious of as they struggle with the everpresent balancing act of “performance vs. cost” and its impact on optimization and marketability of their product. Presented next is an overview of the standard classification systems for lubricants as they relate to automotive chassis and axle gear lube applications, with the intent of providing the reader with some working knowledge of the marketing and performance benchmarks used to evaluate and rate these lubricants against standardized performance tests critical to the intended applications. The remaining sections present overviews of automotive chassis and drivetrain applications [3] and the considerations that should be understood when making lubricant recommendations for each. Where applicable, these system applications are discussed component-by-component to provide a better understanding of the specific lubricant requirements. The chassis system, as discussed in this chapter, consists of three major components (1) Suspension and Steering Linkage System, (2) Steering System, and (3) Brake System. The Driveline System discussed includes the following four major components (1) Halfshaft assemblies, (2) Propeller Shaft, (3) Rear axle gear set and limited slip differential system, and (4) Wheel bearings. Lubrication of engines and transmissions, obviously a part of the drivetrain system, are discussed elsewhere in this text.
4.2 Grease Technology Overview Grease is a solid or semisolid lubricant that is used whenever a lubricant must remain in place. Grease is composed of oil or fluid, thickener, and additives. A typical composition consists of petroleum oil thickened with soap, with inclusion of engineered additives as needed to impart the desired chemical and physical properties. Grease is used in more locations in a vehicle than any other lubricant, including such diverse applications as door locks, remote control mirror assemblies, seat adjustment gears, window levers, windshield wiper gears, water pump bearings, horn, and other electrical contacts, brake systems, heating, and cooling systems, wheel bearings, U-joints, and many other places in the vehicle’s powertrain, chassis, body, and electrical systems [5]. Effective grease will resist oxidation and degradation, provide an antiwear fluid film, and prevent metal-to-metal contact. In addition, greases are often required to provide environmental protection such as
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Automotive Chassis and Driveline Lubrication
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corrosion and rust inhibition and oxidation resistance. Often greases are called upon to keep contaminants away, assist in the sealing process, and resist being washed or squeezed out of their desired location. The base stock in lubricating grease typically comprises 85 to 95% (v/v) of the grease composition and may be composed of either synthetically manufactured or refined mineral oil stocks of varying viscosities. Thickeners or other gelling agents generally contribute 5–12%, with performance additives typically added in the 1 to 5% range. Speciality lubricants designed for specific applications can differ substantially from the above values, depending on the materials used and the performance targets established.
4.2.1 Standardized Testing Standardized tests for greases typically fall into two primary categories: (1) physical property testing and (2) performance property testing. The most common physical property testing includes such attributes as hardness or consistency, high and low temperature fluidity/mobility, dropping point, oil separation tendencies, and shear stability. Performance property testing can vary widely depending on the intended application. Some of the more common tests include load carrying capacity or friction or both and wear testing (4-Ball Wear or Weld Load, or both, SRV, Timken E.P. Tester, Cameron-Plinth. etc.), oxidative stability, corrosion control properties, long-term shear stability, fretting control, electrical path resistance, chemical interaction, and effect on contacting materials (yellow metals, polymers, other lubricants, electronic components, and so on), water resistance and water stability testing. Additionally, chemical spectral analysis utilizing Fourier Transform Infrared Spectroscopy (FTIR) and elemental analysis using either Inductively Coupled Plasma Photospectrometry (ICP) or Atomic Absorption Spectroscopy (AA) methods are often conducted, either as quality control tools or as problem solving and R&D techniques. It is worth noting that in response to increasing Original Equipment Manufacturer (OEM) performance requirements over the last several years, driven by longer vehicle life expectations, increasingly severe operating conditions, and escalating consumer expectations of engineered products, the use of synthetic oils, nonsoap thickeners, and unique additive chemistries has increased significantly. Historically, these “speciality” lubricants were reserved for much more demanding applications, including those in the aeronautical, aerospace, defense, and electronics industries. However, as designs become more precise, tolerances become tighter, customers demand higher performance, and operating environments become broader, the use of the more costly “speciality” lubes offering extended service ranges has become justified. The increased performance of these lubricants, combined with optimization of the mechanical design, may be the only method of achieving the desired performance at an acceptable price.
4.2.2 Classification Systems for Automotive Chassis and Driveline Lubricants Prior to discussing lubrication of automotive components within the chassis and drivetrain systems, it is helpful to understand the standardized classification systems used to describe those lubricants. Just as many other universal applications employ the use of non-OEM classification systems for describing and rating their materials, so do some of the automotive applications discussed in this chapter. Specifically, chassis greases are characterized and rated by utilization of the American Society for Testing of Materials (ASTM) D4950 Specification, Standard Classification & Specification for Automotive and Service Greases. Likewise, a universal non-OEM-specific classification system is used for rear axle gear lubricants. Both of these classification systems, as well as the supporting National Lubricating Grease Institute (NLGI) Certification program are discussed below.
4.2.3 Chassis Grease Performance Specification and Classification — ASTM D4950 In an effort to provide some common ground for evaluation and performance ratings of chassis greases, ASTM D4950 was developed. ASTM D4950, titled Standard Classification & Specification for Automotive and Service Greases, describes categories and sets minimum performance requirements of lubricating greases for automotive service-fill applications for passenger cars, trucks, and other vehicles operating
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under various service conditions. Within this specification, two primary grease categories have been created: • General chassis grease (“L” category) • Wheel bearing grease (“G” category) Each of these two categories is further subdivided into groups based on severity of service. These categories and respective subcategories include: • General Chassis (L) 1. LA Classification: Applications include chassis components and universal joints under mild duty conditions including a. Frequent relubrication b. Noncritical applications 2. LB classification: applications include chassis components and universal joints under mild to severe duty conditions including a. Prolonged relubrication intervals b. High loads c. Severe vibration d. Exposure to water or other contaminants • Wheel bearing (G) 1. GA Classification a. Service typical of wheel bearings operating under mild duty b. Frequent relubrication c. Noncritical applications 2. GB classification a. Service typical of wheel bearings operating under mild/moderate duty b. Normal urban, highway, and off-highway service 3. GC classification a. Service typical of wheel bearings operating under mild/severe duty b. High temperatures c. Frequent stop and go service (buses, taxis, police, etc.) d. Severe braking service (trailer towing, heavy towing, mountain driving)
4.2.4 ASTM D4950 Specification Summary Table 4.1 summarizes the properties specified for each of the general chassis (LA and LB) and wheel bearing (GA, GB, and GC) NLGI Grease Classifications [5,6]. The specific test procedures and corresponding acceptance values for each grease are specified in detail in the ASTM D4950 specification and will not be further detailed here. Those wishing to obtain that detail are urged to review that specification. In brief, both the “L” and the “G” categories share many common requirements [7]. However each of the categories has additional, specialized requirements oriented to a specific chassis need. Simply stated, chassis greases (LA and LB) are more oriented toward general-purpose lubrication. The LB classification includes a fretting wear requirement unlike any of the other categories. The wheel bearing grease categories (primarily the GB and GC classifications) include requirements that are more conducive to wheel bearing environments, including water washout, oil separation, high temperature life, and leakage. In further support of industry commonization, NLGI has taken on the responsibility of independently certifying grease products to the D4950 specification. The intent of this effort is to establish a centralized coordinating entity that would provide review of performance data on candidate greases against the D4950 requirements. For those candidates that pass the requirements established in the specification, NLGI issues licenses for the use of the appropriate NLGI Certification mark for inclusion on the product
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TABLE 4.1 Specification Requirements: ASTM D4950 Automotive Service Grease Categories Service grease categories Chassis ASTM test method D 217 D 566 D 1264 D 1742 D 1743 D 2266 D 2596 D 3527 D 4170 D 4289 D 4290 D 4693
Wheel bearing
Description
LA
LB
GA
GB
GC
Penetration Dropping point Water washout Oil separation Rust protection 4-ball wear 4-ball EP High temperature life Fretting wear Elastomer compatibility Leakage Low temperature torgue
× × — — — × — — — × — —
× × — × × × × — × × — ×
× × — — — — — — — — — ×
× × × × × × — × — × × ×
× × × × × × × × — × × ×
label. The presence of an NLGI Certification Mark for a given product advertises the fact that the product meets the established minimum requirements set forth for that grease category, thereby providing a level of confidence in the product’s performance for both the retail consumer as well as the OEM [7]. It should be noted that a grease product can receive more than one certification licence, as indicated by the “GC-LB” Certification Mark below. This indicates that the grease product meets the requirement for both categories. This is advantageous for marketing and product deproliferation efforts since one product may be used for both chassis and wheel bearing applications. The NLGI Certification Marks are shown below. Reference to these certification categories (LB, GC, and GC-LB) is made several times later in this chapter when discussing potential greases for chassis applications. NLGI Certification Marks for Automotive Service Greases
NATIONAL LUBRICATING GREASE INSTITUTE
NATIONAL LUBRICATING GREASE INSTITUTE
NATIONAL LUBRICATING GREASE INSTITUTE
NLGI
NLGI
NLGI
AUTOMOTIVE WHEEL BEARING LUBRICANT
AUTOMOTIVE CHASSIS LUBRICANT
AUTOMOTIVE WHEEL BEARING & CHASSIS LUBRICANT
GC
LB
GC-LB
4.3 Lubrication of Chassis Systems For purposes of discussion in this text, a chassis system as shown in Figure 4.1 includes the following major components (1) suspension and steering linkage system, (2) steering system, and (3) brake system [5].
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Handbook of Lubrication and Tribology Rear upper control arm Ball joints Rear lower control arm
Rear differential
Anti-roll bar Steering column
Steering rack Wheel bearing
Caliper
Transmission (rear transaxle)
Brake rotor
FIGURE 4.1 Automotive chassis (typical).
In general, the functional requirements of effective chassis greases includes: • • • • • • • • •
Noise prevention in ball joints Corrosion prevention by both fresh and salt water Compatibility with polymeric materials used in the system Torque stability in shearing applications Wear prevention — loaded applications with repetitive motion Fretting prevention where grease redistribution is minimal Friction reduction Oxidative stability Extended lubrication intervals (recommended)
In addition, selection of chassis grease often includes consideration of acceptable performance in the following properties as these may have a significant effect on overall functional performance, depending on the application: • • • • • • • •
Evaporation Feedability Extreme (high and low) temperature stability Mechanical stability Oil separation Pumpability Water resistance and washout characteristics Water stability
4.3.1 Suspension and Steering Linkage System In a chassis suspension system as shown in Figure 4.2, the antiroll bar and cross-bar are the major supporting frames which are connected to the ball joint and shock absorber for the purpose of reducing vibration or noise. The suspension system is also connected to the steering column through the steering gear and intermediate shaft assembly. Most often, suspension system and steering linkage pivot points as shown in Figure 4.1 and Figure 4.2 utilize a traditional grease product as the primary lubricant. Greases selected for these chassis applications can vary in composition quite dramatically, depending on the functional demands of the system. However, the overwhelming majority of greases used for these applications to date include refined mineral base oil combined with a lithium soap or lithium complex thickener. This is due primarily to the effective
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Upper control arm
Shock absorber Lower control arm
Ball joint Cross member
Anti-roll bar
Ball joint
FIGURE 4.2
Front suspension (typical).
cost-to-performance relationship of these greases. Generally, NLGI “LB” classification greases provide a good starting point for overall nominal performance; “GC” rated greases are typically not required to meet the performance objectives of these assemblies. However, it should be noted that, depending on the requirements of application, “performance” additives such as friction reducers, antioxidants, antiwear agents, seal compatibility agents, viscosity improvers, extreme pressure additives, and corrosion inhibitors may be utilized to meet customer or system-specific objectives. Additionally, some steering linkage designs utilize secondary lubricants in the form of solid particulate materials that have been applied through deposition processes on critical surfaces. Alternately, polymeric sleeves or bushings are sometimes incorporated into the design to lower frictional coefficients. Solid lubricant additives and deposition techniques often include materials such as molybdenum disulfide, graphite, polytetrafluoroethylene (PTFE), or mineral-based chemicals to improve load-carrying ability, frictional coefficient, or wear properties. Polymeric materials utilized in the mechanical design of the component to improve performance often include polyoxymethylene (i.e., acetal™), polyamide (i.e., nylon™), PTFE or polyethylene based materials. Use of such materials in these components is usually in response to up-scale performance requirements such as durability, wear, or friction related issues.
4.3.2 Steering System In an automotive steering system as shown in Figure 4.3, the steering column is connected to the steering gear assembly by an intermediate shaft. The steering gear assembly in turn is connected to the steering linkage and ultimately to the steerable wheel assemblies via tie-rod, center link, ball joint, and ball stud components (see Figure 4.4[a] and [b]). The three primary components of the steering system, namely, (1) steering gear, (2) intermediate shaft, and (3) steering column are discussed individually in the following sections.
4.3.3 Steering Gears Two basic types of steering gears are utilized in automotive applications as follows. Rack & Pinion steering gears — Generally, R&P steering gears are the smaller and lighter of the two designs and are used in the vast majority of passenger cars and light duty truck applications today
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Handbook of Lubrication and Tribology
Ball joint Intermediate shaft
Steering column Ball joint Steering rack
FIGURE 4.3
Steering system primary mechanical components (typical).
(a)
Ball stud
Outer tie rod Seal Rubber bushing
Seat
Spring End cap
Grease zerk
(b)
Ball stud
Seal
Retainer
Seat
Tie rod housing
End cap (not shown)
FIGURE 4.4 (a) Tie rod and ball stud assembly — typical (exploded view). (b) Ball stud assembly — typical (exploded view).
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(Figure 4.5[a] and [b]). This design consists of teeth on a steering rack meshing with those on an opposing pinion shaft as they are aligned and assembled into a gear housing. Rotation of the pinion shaft via handwheel (steering wheel) movement actuates the steering rack left and right, thereby providing directional control to the steerable wheels through the tie-rod/ball-joint/ball-stud/linkage components. Recirculating ball steering gears — This design configuration is primarily used in larger light-duty truck and passenger car applications as well as for almost all medium and heavy-duty truck and off-road vehicles (Figure 4.6). In this design, a worm drives a ball nut or rack piston through a series of steel balls. The path of the balls includes an external guide tube that permits them to recirculate freely in either direction. This arrangement provides low friction between the steering shaft and ball nut, helping to keep steering effort low. This design usually incorporates one of several types of worm gears acting against short levers or a gear segment connected to pitman shaft. In cam and lever type steering gears, the worm is referred to as the cam. It meshes with pins on a lever attached to the pitman shaft. Both primary gear types as noted above are manufactured in either manual or power-assist versions. Generally speaking, the manual version of the gear is lubricated by applying grease to the internal components and providing a reservoir of grease internally within the gear during final assembly. The internal components rely on the actuation motion of the gear to redistribute grease to all necessary components. Components outside of the actuation path must rely on either initial greasing during assembly and regreasing during periodic maintenance. For the power-assisted versions, the hydraulic fluid itself lubricates internal components that are in contact with the fluid used for actuation assist. Those parts not exposed to the hydraulic fluid flow, but still requiring lubrication, are generally greased with high quality chassis grease. Power assisted gears require a supporting hydraulic system for assist in actuating the components. Although a variety of system configurations are used, the base system consists of a vane style pump, a fluid reservoir, steering gear, and hydraulic supply and return hoses. In most applications, this system is dedicated to steering system assist, although alternate system designs are sometimes used which provide hydraulic power for other vehicle functions such as suspension or brake actuation.
4.3.4 Factors Affecting Lubrication of Steering Gears In either gear design, the primary drive components of the gears must be capable of operating under very high dynamic loads due to the effort required to steer a vehicle under a variety of conditions. High contact stresses are often generated at specific component interfaces within the gear due to the inherent mechanical advantage provided by the gear design. In addition, the gear will commonly experience shock loading on the primary drive gear teeth and support surfaces in response to road inconsistencies encountered during vehicle travel. Under high load conditions, a boundary lubrication regime may be present resulting in asperity contact and subsequent thermal loading, thereby raising the internal gear and lubricant temperature. Churning and agitation of the lubricant by the gear set may also result in heat generation with a subsequent increase in the bulk temperature of the assembly. In addition to the internally generated heat common with hydraulic systems, the steering gear often is placed adjacent to engine or transmission components that create an undesirable thermal environment. Raising the temperature of the lubricant system increases the rate of oxidation. In addition, sludge or deposits formed on internal components as a result of oil oxidation may act as insulating films, thereby reducing the rate of heat transfer away from the component. This decrease in cooling effectiveness may promote increased bulk temperatures, further accelerating the rate of lubricant oxidation. As a result of the above noted conditions, a lubricant that provides a fairly high level of load carrying ability is required on the primary drive components. For rack and pinion gears, this is the pinion-to-rack tooth interface as well as the rack support mechanism, both of which generally fall outside of the hydraulic fluid path on power-assist versions. For recirculating-ball type gears, the highly loaded surfaces are again the tooth mesh areas located on the pitman shaft and mating ball-nut/rack piston nut.
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Handbook of Lubrication and Tribology (a) 1 2
3
5
6
4 7
14 13
15 8
9 11
6
10 12
1.
Ball stud
6.
Inner tie rod asm.
11.
Rack bearing
2.
Tie rod end
7.
Valve asm.
12.
Gear housing
3.
Boot clamp
8.
Pinion
13.
Mounting bushings
4.
Steering rack
9.
Valve housing
14.
Rack bushing
5.
Hydraulic lines
10.
Boot
15.
Upper pinion bearing
Cylinder lines
(b)
Flow out Flow in Pinion
Stub shaft
Valve Housing
Rack
FIGURE 4.5
Rack & Pinion steering. (a) Gear design (exploded view) and (b) cut-away view of valve (typical).
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Pitman shaft
Hydraulic fluid supply and return ports
Rack piston
Stub shaft Valve asm. Housing
Recirculating balls
Worm shaft
FIGURE 4.6 Recirculating ball steering gear — typical (cut-away view).
To meet these demands, most steering gear designs utilize a good quality chassis or extreme pressure (EP) grease at these contact points. Good load-carrying capability is achieved through the use of specific EP additives or by utilizing base oils of high enough viscosity to maintain sufficient film thickness, thereby avoiding asperity contact, and metal-to-metal wear. Other non-heavily loaded areas of steering gears may use lubricants targeted at specific performance criteria such as reduction in water or debris intrusion, ease of actuation at cold temperatures, compatibility with exposed elastomers, or ease of assembly. Although most steering gear applications for the general consumer automobile market utilize greases and the circulating hydraulic fluid for lubrication, other approaches are noted. Specifically, some gears utilize multi-purpose gear lubricants of API (American Petroleum Institute) gear oil classification GL-4 or GL-5 exhibiting a viscosity equivalent of SAE 80W or 90 (ISO grades 68 or 150). For tractor equipment, the lubricant may be a multi-purpose lubricant designed for use in tractor hydraulic systems, transmissions, and final drives, appropriately referred to “tractor hydraulic fluid.” In selecting the lubricant for a steering gear application, the following factors of design and operation should be considered: • Gear type • Gear speed • Operating temperature 1. Cold — actuation efficiency and lubricant mobility 2. Hot — lubricant viscosity affecting wear and leakage • Oxidative stability/thermal environment • Gear reduction ratio — loading/contact stress conditions • Surface finishes of tribological pairs • Potential for water intrusion/contamination • Lubricant leakage • Compatibility with seal materials • Good understanding of the shear stability properties of the lubricant to provide long term predictability • Lubricant cost vs. performance
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Handbook of Lubrication and Tribology Joint
Seal
Isolation element Joint
Slip shaft
FIGURE 4.7
• • • • •
Steering intermediate shaft — typical.
Dampening requirements for noise and vibration Fretting tendencies of the assembly Lubricant effect on assembly operation Employee safety Dispensing
4.3.5 Intermediate Shaft An intermediate shaft assembly, or I-shaft, connects the steering column to the steering gear, allowing nonlinear placement of the input and output shafts. In addition to providing transmission of torque, the intermediate shaft may also provide noise and vibration isolation, torsional stiffness, and collapsibility. Many design variations exist for intermediate shafts, but most utilize four primary functional components in their design (see Figure 4.7). These four components consist of (1) joints, (2) slip-shaft, (3) isolator, and (4) seal. 1. Joints. Most I-shaft designs utilize a Cardan joint (or double Hookes Universal Joint) on each end of the primary shaft. The lower and upper joints provide a flexible means to fasten the lower end of the I-shaft to the steering gear and upper end to the steering column. Clamp yokes with pinch bolts provide a specified clamping force to keep the connections secure. 2. Slip-shaft. The shaft provides the means to connect the lower joint to the upper joint and transmit torque. The shaft can also provide axial movement for ease of assembly and crash-worthiness. The shaft is generally one of three types: either slip-joint, energy absorbing collapsible, or solid. 3. Isolator. An isolator is sometimes provided to reduce both chassis vibration and noise. Different types of isolators and couplings have unique abilities for absorbing torsional and axial vibration. The isolator dampens axial vibration while maintaining relatively stiff torsional compliance. 4. Seal. A seal provides a means to protect the I-shaft from an unfavorable environment. Seal types commonly used include convoluted boot seals, bearing face seals, stone shields, front-of-dash seals, and pot-coupling seals.
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4.3.6 Lubrication of I-Shaft Assemblies As is the case for many other components, I-shaft designs can vary substantially, depending on vehicle dynamics, suspension design, mating component location, and performance targets. As such, the lubricants chosen for I-shafts may also vary greatly. The following discussion includes a general overview of each of the major lubricant areas within the I-shaft and is intended to provide basic understanding of I-shaft lubrication. When selecting lubricants for an I-shaft application, consideration should always be given to the application parameters as stated above for the specific vehicle in question. For all I-shaft applications, also see “General Grease Selection” discussion below. Joints. The joints located on either end of the shaft assembly are usually of the Cardan type, utilizing a double set of spiders with roller element needle bearings, held in place by cap assemblies. Utilization of “LB” grade chassis grease is typically a good starting point for grease selection. Slip-shaft. Since shafts are generally of two types, either slip-joint or energy absorbing collapsible for which the lubricating considerations may be different. In addition, slip-load requirements may vary substantially, depending on vehicle dynamics and target performance. Significant consideration should be given to the required frictional properties of this slip interface, as this will impact in-vehicle transmission of “noise, vibration, and harshness” (NVH) into the area of the steering column. From a loading, wear, and corrosion perspective, utilization of “LB” grade chassis grease is typically a good starting point for grease selection. Shaft to seal area. Depending on the design, grease may be required where the shaft contacts the respective seals. Typical examples include dash seals (I-shaft rotation on stationary dash seal) and slipshaft interface (outer tube interface with inner shaft — sliding, axial motion). Utilization of “LB” grade chassis grease is typically a good starting point for grease selection. General grease selection. As noted above, individual vehicle designs and dynamics play a significant role in I-shaft grease selection. For all applications, the specific parameters as listed below should be taken into consideration when selecting greases for I-shafts: • Thermal environment — most I-shaft assemblies are located adjacent to engine exhaust manifold or other thermal generators • Oxidative stability • Frictional requirements at the slip-shaft joint • Vehicle and suspension dynamics causing or contributing to NVH in the steering column/ I-shaft area • Serviceability of the I-shaft for regreasing • Compatibility with contacting polymeric materials
4.3.7 Steering Column (see Figure 4.8[a] and [b]) A vehicle steering column assembly is a complex collection of components, materials, and dynamics. Components range from small to large, fully mechanical to fully electrical, stationary to fully dynamic, sliding to rolling, heavily to lightly loaded, etc. The result of this component diversity is a corresponding diversity in lubricant use for steering columns’ applications [5]. The following is a list of typical componentry requiring lubrication within a steering column. This list is not all-inclusive, nor is it applicable to all designs. However, it does provide the reader with a good sense of the diversity noted above. Items typically lubricated within a steering column may include the following: slip-shaft, turn signal cancel cam, electrical contacts (multiple), tilt/centering sphere, upper & lower shaft bearings, switch assemblies, wash/wipe switch, turn signal switch, hi-lo headlight beam switch, brake-transmission interlock switch, friction enhancing devices, housing pivot pins, lock cylinders, shift assembly, actuators (variety), Supplement Inflation Restraint (SIR) (air bag) coils, park-lock assembly, teletube (telescoping columns only), and cable assemblies (multiple).
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(a)
(b)
FIGURE 4.8
(a) Steering column — typical (external view). (b) Steering column — typical (exploded view).
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When selecting lubricants for steering columns, many of the same criteria as previously discussed apply. However, it should be noted that temperatures and environmental conditions are not as variable for components within the passenger compartment as noted for under-hood applications. Also, the functional requirements of many components within a steering column include much lower actuation loads and contact stresses than noted in suspension or drive line components. As a result, cold temperature grease mobility plays a greater role in actuation of small, lightly loaded, or actuated parts. Steering columns — service conditions to consider when selecting grease: • • • • • • • •
Component loading ranges from low to high (design and component dependent) Temperatures from −40 to +85◦ C Require quiet operation Low grease bleed to avoid oil separation and migration Dampening properties in some applications Low grease volume per part Lubed-for-life (generally not serviced during life of vehicle) Compatibility with 1. Other greases 2. Plastics and rubber
4.3.8 Overview of Grease Types Used for Steering Columns • Wide variety — very application-specific • Base oils 1. All Types 2. Be cautious when using ester types with polymeric materials • Thickeners 1. All types (lithium, calcium, polyurea, PTFE, silica, clay, etc.) • Additives 1. All general types (antioxidants, viscosity index [VI] improvers, E.P. agents, friction modifiers, etc.) 2. Tackifiers (methacrylates, polyisobutylene, high molecular weight polyalphaolefin synthetic stocks, etc.) 3. PTFE 4. Very limited use of MoS2 (molybdenum disulfide)
4.3.9 Power Steering Hydraulic Fluid In general, automotive power steering systems exhibit functional similarities to other hydraulic systems used in manufacturing equipment (see Figure 4.9). However, greater care is taken in preserving a minimum performance level due to the fact that automotive hydraulic steering systems play a significant role in assisting with vehicle directional control. Therefore, automotive power steering systems are generally designed with expectations of operating under a greater safety margin that that of a manufacturing system. In support of this, the system components, including the hydraulic fluid, must be robust and capable of operating effectively under varying environments. Due to the “fill-for-life” aspect of most steering systems, the fluid must also maintain its performance over the life of the vehicle. Hydraulic fluids used in automotive systems can vary significantly, depending on the equipment design, intended function, and designer preferences. Some vehicles use a fluid designed specifically for steering system optimization while others utilize the vehicle’s automatic transmission fluid (ATF) in an effort to minimize part proliferation within the vehicle, at the assembly plant, and at dealer locations.
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Handbook of Lubrication and Tribology Fluid reservoir
Pump Return hose
Pressure hose
Valve assembly
Gear assembly
FIGURE 4.9
Hydraulic assist power steering system — typical.
The ability to use alternate fluids, such as the vehicle’s ATF, in a steering system is primarily dependant on two issues: • Steering system performance requirements and the specific testing regime used to validate that performance • The chemical make-up of the fluid and it’s effect on system elastomers System performance testing varies from manufacturer to manufacturer and is established by taking many things into consideration including, but not limited to, functionality of other components within the system, overall vehicle steering dynamics, thermal and oxidative environment, and viscosity shearing tendencies of the steering system itself. When ATF is considered for use within the steering system, the designer is faced with additional issues that are not prevalent when designing a fluid specifically for the steering system, such as frictional properties and additional elastomer compatibility requirements. Frictional properties — When ATFs are designed specifically for transmissions, the coefficient of friction (CoF) is optimized for transmission shift performance, taking into consideration other contributing factors within the transmission. Specifically, frictional coefficients provided by the performance additives in the fluid are best held at a specified value or range of values to optimize shift performance. Coefficients of friction that extend outside of those established ranges will result in undesirable transmission function. Notably, CoF values that are too low may result in sloppy shift engagement or in excessive transmission slippage during normal driving conditions. Conversely, power steering systems tend to run quieter, cooler, longer, more efficiently, and with less wear when utilizing fluids having lower bulk coefficients of friction. This is most evident in the vane pump-rotating group, where severe tribological conditions exist during “pressure relief ” cycling of the pump. Additionally, the desired CoF S/D (static-to-dynamic) ratios of the fluid may vary substantially for optimum steering system performance than those targeted for automatic transmission function. Elastomer compatibility — A wide variety of elastomers are utilized in both steering systems and transmissions. Fluids used for fill of these components must exhibit acceptable compatibility with elastomers in both systems. Critical attributes for elastomers include, but are not limited to, pliability, volume change tendencies (either shrinking or swelling), hardness changes, tensile, elongation and modulus properties,
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cold temperature flexibility, and tear resistance. All these properties can be readily affected by exposure to base stocks and additives potentially used in engineered fluids. Design of a fluid to minimize impact on polymeric materials, including elastomers, thermoplastic and thermoset polymers, as well as other gasketing materials is a balancing act that is often difficult when the inclusion of chemically active additive systems in the fluid are necessary to achieve the desired performance targets. When developing a fluid for use in both transmissions and steering systems, the composite list of polymeric materials present in both systems presents a larger matrix of compatibility issues that the designer must consider. In general, most power steering systems operate with bulk system continuous temperatures at or below 150◦ C. Excursion temperatures up to 250◦ C may be produced in specific, localized regions within the system under certain conditions, however, these are generally of short duration and are not routinely generated during normal driving conditions. System temperatures as noted above dictate the need for fluids with fairly robust oxidative stability. Thermal exposure, aggravated by the presence of entrained air, can lead to oxidation of the petroleum base stocks and additive chemicals. Oxidation by-products include acidic compounds which tend to promote additive insolubility, sludge, and varnish deposits, thereby increasing the tendency for corrosion and accelerated attack on polymeric materials. In summary, oxidative stability is a key factor in hightemperature hydraulic systems. Fluid viscosity plays a major role in steering system performance. Heavier base stocks provide a more robust fluid film conducive to preventing metal-to-metal contact. Additionally, fluid viscosity is critical in wear prevention during high-speed, high load conditions where hydrodynamic and elastohydrodynamic (EHD) lubrication play significant roles. Maintaining adequate viscosity throughout the operating temperature range is important. Therefore, choosing a fluid base stock with a high viscosity index (VI) value and utilizing VI improvers to enhance the temperature/viscosity relationship is recommended. When VI improvers are utilized, shear stability of the fluid should be evaluated. VI improvers most commonly used for hydraulic systems have historically consisted of high molecular weight polymethacrylates, polyisobutenes, ethylene–propylene copolymers, styrene–butadiene copolymers, or their derivatives. Shear stability of these materials can vary considerably, depending on the specific additive chosen as well as the severity of the shear mechanism in the pump and gear components. Typically, VI improvers will experience some level of shear down over time, thereby losing their ability to provide high temperature viscosity enhancement. Once sheared, the fluid will exhibit lower viscosity at high temperatures, resulting in increased wear and fluid leakage. The most desirable approach to maintaining viscosity across a wide temperature range is to use high VI base stocks, of either mineral or synthetic origin. However, performance has its price, as these stocks are more costly. Again, balancing performance and cost is the ultimate goal in achieving a marketable system.
4.4 Brake System Design and Lubricant Characteristics A typical automotive hydraulic brake system configuration is shown in Figure 4.10. Application of force on the brake shoe produces the required frictional torque. The frictional torque depends on the CoF between the brake shoe and the drum/disc as well as on many supporting design parameters. In the case of drum brakes the mechanism of self-actuation results in a nonlinear relationship between the frictional torque and the CoF. To overcome this, the frictional performance of a brake is characterized by “brake effectiveness,” which is the ratio of the brake frictional torque to the applied force. Brake friction pairs, including pads/rotors and shoes/drums, encounter the following wear mechanisms: • • • • •
Thermal degradation Abrasive wear Adhesive wear-tearing Fatigue Macro-shear
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Handbook of Lubrication and Tribology
Rear wheel brake drums and shoes
Master cylinder Hydraulic lines
Front rotors
Front calipers
FIGURE 4.10
Hydraulic brake system — typical.
As mentioned earlier very high flash temperatures (up to 1100◦ C) are reached at asperity junctions causing the brake material to melt or thermally decompose, leading to loss of material. Oxidative wear also occurs near the edges. Abrasive wear occurs due to plowing by wear debris and hard contaminants. Adhesion and tearing occur due to brake material sticking to the counterface and tearing off. Fatigue wear occurs due to repeated asperity encounters, and macro-shear is the sudden failure of friction material that has been weakened by heat. For practical purposes, brake wear is generally defined in terms of the distance traveled or the usage time. It may also be measured in mass worn per unit of frictional work.
4.4.1 Brake Lubrication The primary hydraulic actuation fluid contained within a brake system also acts as an internal lubricant. The fluid provides hydraulic actuation as well as lubrication for the metallic and polymeric components, that is, potential wear points within the system. Additionally, it provides corrosion protection for those components subject to corrosion-promoting environmental conditions. In a brake system, the fluid may be exposed to moisture or water. To prevent moisture from collecting and freezing in cold weather or causing rust and corrosion which can create brake failure, brake fluids exhibiting good water miscibility should be chosen. Since petroleum-based fluids generally cannot meet these requirements, most brake fluids are based on either silicone or glycol chemistry. They must remain relatively stable during use, maintain appropriate alkalinity to resist becoming corrosive, remain fluid at low temperature, exhibit good compatibility with elastomers, and be water tolerant. However, since brake fluids tend to retain water once it is introduced, brake system fluid should be replaced once the water content becomes too high for proper braking performance. In an effort to provide a common level of understanding and to establish minimum performance criteria, automotive brake fluids are referenced against various industry standards, including: • • • • •
SAE J1703 — Motor Vehicle Brake Fluid SAE J1704 — Borate Ester Based Brake Fluids SAE J1705 — Low Water Tolerant Brake Fluids SAE J1706 — Production, Handling, and Dispensing of SAE J1703 Motor Vehicle Brake Fluid SAE J1707 — Service Maintenance of SAE J1703 Brake Fluids in Motor Vehicle Brake Systems
These SAE Standards, along with OEM-specific requirements, commonly define the performance and technical requirements of automotive brake fluids currently used. Review of these standards should be undertaken if a further understanding of brake fluid technology is needed.
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Caliper housing
Brake pads
Hydraulic piston
Retaining springs/clips
FIGURE 4.11
Hydraulic brake caliper assembly — typical (exploded view).
4.4.2 External Brake Lubrication Brake mechanical components are subject to severe operating conditions including moisture, external environmental contamination, extreme temperatures, and retention of brake component wear debris (see Figure 4.11 for a typical hydraulic brake caliper assembly). As such, these components within the brake system require external lubrication to guard against early failure. Due to the high temperature conditions associated with brake systems, a lubricant providing excellent oxidative and thermal stability along with low volatility, good wear protection, and moisture stability is required. Most often, silicone-based greases that contain some form of solid lubricant such PTFE are best suited for this application. Again, lubricant usage may vary, depending on system specifics and designer preference.
4.5 Shock Absorber Lubrication 4.5.1 Shock Absorber Fluid Shock absorber fluid is typically a fluid that provides hydraulic fluid-flow resistance inside the shock absorber assembly, thereby dampening the movement of the shock and the attached suspension components as the wheels of a vehicle travel over uneven surfaces. Shock absorber fluid needs to remain stable during use, not entrain an excessive amount of air, exhibit good sealing performance, and have the appropriate viscosity characteristics. The preferred properties for shock absorber fluids are often impacted by the mechanical and hydraulic design characteristics of the shock assembly itself. Fluid-flow characteristics of the design, along with the desired dampening effect under a variety of dynamic vehicle conditions, will help to establish the viscometric targets for the fluid. Elastomer compatibility and corrosion resistance of the fluid are a function of the performance targets set for the shock assembly. As in most engineered systems, multiple parts of a system must work in concert with each other to achieve the desired system performance (see Figure 4.12 for typical shock absorber design).
4.5.2 Magnetorheological and Electrorheological fluids Magnetorheological (MR) fluids are suspensions of soft-magnetic iron particles in a carrier fluid, typically lubricating oil. When subjected to a magnetic field, the iron particles react by preferentially orienting themselves to the field, thereby causing a dramatic change in the fluid’s apparent viscosity, transitioning
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Handbook of Lubrication and Tribology Piston/shaft movement due to suspension travel
Shaft Piston design controls fluid flow and dampening
Hydraulic fluid reservoir regions
FIGURE 4.12
Hydraulic shock absorber — typical (cut-away view).
from a pourable liquid to a thick gel virtually instantaneously, with reversion back to a liquid-like state when the magnetic field is removed. While MR fluids were first reported by Rabinow [8] more than 50 years ago, applications of MR fluids have only recently become popular. The majority of recent applications of MR fluids have been in the automotive industry, particularly in shock absorber systems for the provision of active suspension control [9,10]. In automotive active suspension applications, position sensors located at critical points on the vehicle suspension measure magnitude and speed of suspension travel relative to an established reference point. This information is then fed into a system controller, which analyses the data and controls downstream electrical currents being sent to the shock assemblies [9,10]. The entire sensory, control, and feedback process is completed in a timeframe of milliseconds, thereby allowing the system to actively dampen undesirable suspension travel. Advanced controller designs allow the design engineer, and even the vehicle driver, to “tune” the dampening characteristics of the system to provide a “driver preferred” feel to the vehicle for various driving conditions. An MR fluid controlled shock absorber is shown in Figure 4.13 [9]. In addition to dampening control, both MR and electrorheological (ER) fluid devices can be used for controlled torque transfer and controlled vibration isolation. MR-based devices are also available to aid in optical finishing (Kordonsky [11]), and as dampers for controlling the response of large structures during earthquakes [11].
4.6 Halfshaft Assemblies 4.6.1 Halfshaft Application Overview Halfshaft assemblies, consisting of an axle bar with constant velocity joints (CVJs) located at both ends, provide a means of transmitting torque output from a driveline transaxle or gear set assembly, usually located at approximately the vehicle centerline, to outboard drive wheels. A typical halfshaft design is shown in Figures 4.14(a) (external view), (b) (cut-away view), and (c) (exploded view). Constant velocity joint designs vary considerably, but can be grouped into two primary classifications: ball joints and trilobular (also known as tripot or tripod) joints. Ball type CVJs can be further subdivided into fixed-center joints (i.e., Rzeppa design) and plunging ball-joints (i.e., cross-groove and double-offset
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MR Shock Absorber Application Chassis suspension Piston Piston movement causes MR fluid to flow through the annular gap The coil is energized to provide a magnetic field of desired strength across the flow gap Damping force can thus be varied as required Rod and rod seal system Rod is connected to piston and to chassis suspension Rod movement is controlled by strength of magnetic field in piston Rod seal is used to prevent MR fluid leakage Nonvertical mounting of shock absorber gives rise to side load on rod seal Rod seal wear is a concern Controlled damping Wheel motion results in shock absorber motion (tube) Sensors and controller generate appropriate current to coil to generate magnetic field which controls piston and rod motion Overall effect: controlled smooth ride
Rod seal
Rod
Side load MR fluid
Piston
MR fluid
Coil
Wheel
FIGURE 4.13
MR shock absorber application.
designs) [12]. The overwhelming majority of trilobular design CVJs are of the plunging type. Examples of each of these four are shown in Figures 4.15(a)–(h). On a front-wheel drive vehicle, the torque output device is typically the transaxle, with halfshafts transmitting torque to the outboard-mounted front drive wheel assemblies. When halfshafts are utilized on the rear of a vehicle, the primary goal is to achieve a combination of rear-drive capability with independent rear suspension (IRS) and rear-steering. The output torque-device on these vehicles is typically either a hypoid gear set carrier assembly or a rear-mounted transaxle assembly. Additionally, the angulation and axial travel capabilities of the plunging type CV joints located on either end of the axle bar allow for suspension travel and steering capabilities. In most applications, the joint placed on the inboard end of the axle and interfacing with the transaxle or hypoid gear set is of the plunging type (i.e., tripot, cross-groove, or double-offset designs), thereby allowing for axial displacement. This joint configuration allows suspension movement in the up and down direction, translating to axial travel of the axle bar in and out of the inboard joint while maintaining rotational torque transfer. The joint placed on the outboard end of the axle is dependent on the specific application. Outboard applications requiring a high degree of angulation, like vehicle steering, typically employ a fixed-center, nonplunging design (i.e., Rzeppa design). Outboard applications with less angulation requirement and the desire for axial movement, typically utilize a plunging, ball type of joint.
4.6.2 Halfshaft Lubrication Requirements Halfshaft joints are exposed to extremely high loads and contact stresses. With engine outputs continually on the rise and vehicle towing capabilities increasing, halfshaft capabilities must also keep pace. Some current designs are able to handle torque loading in the range of 7800 Nm. However, as halfshaft contact stresses increase and consumers continue to demand less NVH (noise, vibration, and harshness) in their
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vehicles, halfshaft greases have, and will continue to play a significant role in helping vehicle OEMs in achieving these targets. A wide array of operating conditions and other requirements must be considered during selection or development of grease for CVJ applications. The most demanding and difficult of these include wear protection, load carrying capability, frictional properties, heat generation, and seal compatibility. Ball type of joints, both plunging and fixed center, exhibit a different set of operating conditions and lubrication requirements than those of trilobular, or Tripot style joints. Each will be briefly discussed.
4.6.3 Ball-Type CVJ Ball-type CVJs exhibit a wide variety of contact geometries, loading characteristics, and relative motion during their operation. They tend to exhibit very high component loading and contact stresses (up to 4.0 GPa), as well as relative internal component motion ranging from 100% rolling to 100% sliding. The majority of the motion during normal operating conditions consists of a blend of rolling, sliding, and rotation. These joints typically operate in a variety of tribological regimes ranging from boundary conditions with asperity contact to full EHD lubrication. Their design and method of operation generally results in very high contact stresses leading to a higher level of self-induced thermal loading. Often, for the inboard plunging type ball joints, their location is in close proximity to other heat generating sources such as exhaust manifolds, catalytic converters, and engine/transmission components. In light of these functional and environmental considerations, greases chosen for these applications should generally exhibit a high level of load-carrying capability through the use of EP additives. Additionally, oxidative resistance is a critical attribute due to the thermal conditions noted above. Grease (a)
Outboard joint assembly
Joint boot seals
Axle bar
Inboard joint assembly (b)
FIGURE 4.14 A typical Halfshaft design. (a) External view. (b) Cut-away view. (c) Exploded view.
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FIGURE 4.14 Continued.
performance leading to reduction in heat generation by the joint itself is also beneficial in limiting the bulk thermal exposure of the joint. Although not used exclusively, the use of molybdenum disulfide and graphite solid additives is commonly employed when heat reduction is a desired end result. These solid additives also provide some level of effective reduction in CoF for ball-type of CVJs.
4.6.4 Trilobular Type CVJ Trilobular or Tripot type CVJs also exhibit a wide variety of contact geometries, loading characteristics, and relative motion during their operation. Loading their internal components tends to be somewhat less than that exhibited by the ball-type CVJs. Relative internal motion ranges from 100% rolling to 100%
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sliding, with the majority of the motion during normal operating conditions consisting of a blend of rolling, sliding, and rotation. The components in these joints operate in a variety of tribological regimes ranging from boundary conditions to full EHD lubrication. The most critical aspects of tripot type joint lubrication and performance are generally considered to be in two areas of focus (1) contact rolling fatigue on the spider trunnion/needle roller area and (2) overall frictional performance of the joint components during axial plunging under high load/torque conditions.
(a)
(b)
Outer race Cage Inner race Balls (c)
(d)
Outer race
Cage Inner race Balls
(e)
(f)
Outer race Cage Inner race Balls
FIGURE 4.15 Fixed center ball type joint (REEPPA): (a) cut-away, (b) exploded; Plunging ball joint — cross groove type: (c) cut-away, (d) exploded; plunging ball joint — double offset type: (e) cut-away, (f) exploded; plunging trilobular type — tripot: (g) cut-away, (h) exploded.
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(h) Roller ball Needles Tripot spider
Tripot housing
FIGURE 4.15 Continued.
Relatively high loading is often encountered at the spider trunnion surface as driveline torque is transferred from the axle bar to the tripot housing via the spider, the needle rollers, and tripot ball. Improper component design, inadequate lubrication, or service conditions exceeding the design limits of the joint will often result in early component failure. Use of lubricants with inadequate load-carrying capacity (EP properties) will generally lead to subsurface initiated failures at the contact interface resulting in spalling of the spider trunnion or needle roller surfaces. Additionally, if surface lubrication is inadequate (i.e., lubricant film thickness), surface related failures such as adhesion and surface wear may develop.
4.6.5 Plunging Joint Frictional Requirements The inherent process of transferring rotational torque through a plunging joint at angles greater than 0◦ results in the generation of a force that attempts to push the axle and spider assembly out of the tripot housing in an axial direction. Higher CoF values, increased joint angles, and elevated torques all accentuate this phenomenon. This generated axial force, when significant enough, will often manifest itself as shudder and vibration in the driveline, resulting in vehicle noise, vibration, and harshness. The most significant and broadly used advancement in halfshaft grease over the last 15 years has been the development and use of specific low-friction additives to reduce friction within plunging-type joints, thereby reducing generated axial force and subsequent driveline NVH. Specifically, low-friction additives utilizing organo-molybdenum complexes that react to high contact stresses at the heavily loaded interfaces within the joint have played a significant role in helping to offset the higher powertrain output torques and more demanding chassis dynamics that typically promote more driveline NVH [13]. These organo-moly low-friction additives work in the following manner. Extremely high loads between contacting components result in a boundary lubrication condition between the mating surfaces. The metal-to-metal contact (asperity contact and micro-adhesion) between surfaces generates localized heating of the grease at the contact interface, thereby activating the “organomoly” additives which, in turn, react with the metal substrates to form intermediate metallic oxide films on the contact surfaces. This oxide film layer exhibits a very low CoF, thereby reducing the force necessary to allow axial or oscillatory motion within the joint. The deposited layer is sacrificial and self-replenishing. When the film layer wears down to the base metal substrate, the entire process repeats itself with eventual deposition of a refreshed, low CoF layer. When selecting grease for a tripot type of joint, a few general recommendations are worth noting. A good EP containing grease should be utilized. Lower contact stresses are generally encountered than those noted in ball-type CVJs, therefore, a milder EP performance is allowable. However, the loads encountered are high enough to require a moderate level of EP performance of the grease. Frictional requirements should be considered, especially if the application is for a vehicle that exhibits high engine output or high in-vehicle halfshaft angles, thereby promoting the generation of driveline NVH.
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Handbook of Lubrication and Tribology
4.6.6 General Lubricant Selection — All Halfshaft Joint Types When selecting grease for any type of halfshaft joint, a significant consideration should be given to its compatibility with the joint seal. Most joint designs use a convoluted boot configuration to provide the necessary axial movement and joint angulation. Halfshaft boots are put through a very dynamic range of motion at relatively high stress levels and speeds, while at the same time being exposed to environmental conditions consisting of road salt, water spray, stone impingement, extreme cold, and moderately hot temperatures. To insure the longevity of the seal, the grease must exhibit good compatibility with the seal material. In addition to benchtop testing, the severe nature of this seal application dictates that validation testing include grease exposure while under dynamic actuation and aging conditions. One of the more common modes of halfshaft failure in the field is that of boot-seal failure. This results in loss of lubricating grease and eventual joint failure by gross damage to the internal components. Additionally, the grease should be oxidatively stable, water resistant, and able to provide a relatively high level of corrosion resistance due to the potential for moisture intrusion into the joint from high-speed water impingement at the seal interface. Grease shear stability is also an important factor. Depending on the joint configuration, shear-down of the grease may be beneficial in helping to relubricate areas of lesser dimensional clearance, whereas a grease with more “body” or higher apparent viscosity may tend to stay in place more effectively under the high rotational speeds associated with halfshaft operation-evidence of varying lubrication requirements for different joint designs. The following is a summary of the primary attributes and operating parameters that should be considered when selecting a lubricant for any halfshaft joint application. They include: • Extremely high contact stresses (up to 4.0 GPa) • Temperature range of −40 to +150◦ C (dependent on location relative to heat generating components) • Varied joint angulation • Seal compatibility (very dynamic seal movement) • Moderate to high rotational speeds (centrifuging effects) • “Lubed-for-life” application • Frictional requirements • Heat generation • Wear • Generated axial force — NVH (Inboard joints) • Blend of lubrication regimes • Rolling • Sliding • Skidding
4.6.7 Typical Lubricant Product Usage for Halfshaft Applications 1. Base oils • Mineral oil most common due to cost vs. performance 2. Thickener systems • Urea derivatives and simple lithium soaps constitute the largest usage • Other general types include • Simple calcium soaps • Metal complexes (Li and Ca) • Synthetic thickeners and base oils a. Lower use volumes due to cost impact, used when application requirements (i.e., thermal and oxidative resistance) exceed performance offered by nonsynthetics
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3. Additives • Friction modifiers (MoS2 , organo-molys, etc.) • EP agents • Antioxidants, VI improvers, antiwear, etc. • Lead compound usage — historically greatest in Asian and European markets. Almost entirely avoided today due to environmental and health impacts as well as governmental restrictions.
4.6.8 Ongoing Developments in Halfshaft Lubricants Halfshaft manufacturers can often justify the use of more expensive low-friction CV joint greases for some vehicle applications due to: • • • • • •
Desire to reduce vehicle “NVH” Increased driveline loading (higher output engines) Increased underhood temperatures Desire for lower heat generation within halfshaft joints Performance requirements of new joint and vehicle designs Development of grease chemistries compatible with new boot seal materials
4.6.9 Summary of Halfshaft Lubricant Selection Halfshaft joints present a very unique and complex tribological system impacted by many other vehicle inputs such as drivetrain geometry and output, as well as chassis and body dynamics. CVJ assemblies exhibit a wide variety of lubrication and wear regimes that are difficult to predict with classical tribological tools and explanations [14]. More often than not, lubricants for these systems are initially chosen using the best engineering information available in combination with application of tribological theory and testing. Ultimately, halfshafts are then tested by the OEM with candidate greases to verify lubricant performance under a wide range of chassis and drivetrain parameters that are virtually impossible to duplicate with standardized benchtop test procedures. In summary, selection of greases for halfshaft applications are best delegated to the manufacturers of halfshafts in concert with their respective OEM customers on a vehicle-by-vehicle basis. This approach provides the most comprehensive selection of these lubricants to avoid unpredicted issues once the product is in service.
4.7 Propeller Shaft Joint Lubrication Automotive propeller (prop) shafts transmit power between driveline components, typically from the engine output shaft to the rear axle gear set on rear wheel drive vehicles. Prop shafts typically utilize one of three types of joints to allow transmission of power to axially misaligned components. These include: • Simple universal joint (Hookes joint) (Figure 4.16[a]) • Cardan joint (double Hookes universal joint) (Figure 4.16[b]) • True constant velocity joints (Figure 4.16[c])
4.7.1 Simple Universal Joint Typically, this joint consists of an array of rolling elements held by a retaining cage between two raceways. The entire assembly is packed with grease and sealed to retain the grease internally and to prevent outside contaminants from entering the joint. The contacts within the joint are between the rolling elements (balls, needles, or rollers), the inner spider trunnion, and the outer cup. The rolling elements are usually made from a high carbon steel, which is through hardened. The inner spider trunnion and outer cup are load-bearing areas and are generally induction hardened or carburized. This type of joint is not considered
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Handbook of Lubrication and Tribology (a)
Spider
Spider trunnion
Seal Needle rollers Bearing cup (b)
Connecting housing
Single universal joint ASM #1
Single universal joint ASM #2
(c)
FIGURE 4.16 (a) Simple universal joint (Hookes joint), (b) Cardan joint (double Hookes universal joint), (c) True constant velocity joints.
a true “constant velocity joint” as the input and output shafts do not maintain continuous constant velocity relative to each other. Successful lubrication of universal joints is usually accomplished with the application of a chassis grease meeting the requirements of ASTM D4950, LB classification.
4.7.2 Cardan Joint (Double Hookes Universal Joint) This joint configuration consists of two single universal (Hookes) joints held together by a connecting housing. This arrangement essentially doubles the degree of off-axis capability of the single universal
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joint. Like the single universal joint (Hookes), the Cardan joint is not considered a true constant velocity joint. Lubrication of this type of joint is primarily the same as that of the single universal joint discussed above.
4.7.3 Constant Velocity Joint Constant velocity joints as discussed in the “Halfshaft Lubrication” section earlier are often used in prop shaft applications, replacing the standard single and double universal joints. These joints are typically of the ball-type, often employing plunging capability in the form of either a cross-groove or double-offset joint configuration. These upscale performing joints offer smooth, true constant velocity, high angle power transmission in prop shaft applications, and typically provide a higher load rating and improved packaging envelope than the simple universal joints that they replace for a given application. Due to the ball-type design, potential plunging capability and the high rotational speeds of these applications, heat generation is often an issue when selecting a lubricant. As stated earlier, a thickener system with good oxidation and thermal stability, along with solid additives that offer friction and heat reduction are often chosen with good results. However, as with halfshaft grease selection, all parameters of the system should be considered including contact stresses, seal compatibility, centrifuging effects, and joint geometry.
4.8 Drive Axle Lubrication 4.8.1 Drive Axles and Limited Slip Differential System Lubrication Hypoid type rear axle designs, as shown in Figures 4.17(a) and (b), exhibit a high degree of sliding between the teeth of the ring and pinion gears. In the early development of axle lubricants for hypoid gears, high load-carrying capacity lubricants containing active sulfur were developed. However, the active sulfur components utilized for acceptable gear set performance were noted to be chemically aggressive, adversely affecting the life of bearing races, rollers, and cages, as well as bushings and seal materials. Hypoid gear sets require a lubricant with significant EP capability, plus good lubricity, especially during the break-in stage. The more recent developments of specialized additive chemistries such as zinc dialkyldithiophosphate (ZDDP), chlorinated hydrocarbons, and organo-molybdenum complexes have improved the antiwear performance and efficiency within drive axles. More recently, formulations have been developed to include sulfur/phosphorous additives derived from sulfurized olefins and phosphorus esters to provide improved anticorrosion and antiwear performance. In addition to antiwear and anticorrosion performance, these universal EP gear oil formulations also include cleanliness dispersants, seal compatibility agents, and antifoaming compounds [15]. As fuel economy and temperature sensitivity has become more important in gear oil performance, the use of synthetic base stocks has increased. Synthetic base stock gear oils, such as those typically used to achieve 75W-90 or 80W-140 viscosity grades, are often used in applications where consistent operation is required over extended temperature ranges. Cold temperature fluidity is critical for the lubricant to flow to the gear teeth and to be carried up by the ring gear in sufficient amount to lubricate the pinion shaft bearings. In addition, synthetic base oil gear lubricants, formulated with appropriate performance enhancing additives, offer an improved level of life expectancy in the vehicle, thereby helping to meet the OEM’s desire for extended service intervals and “lubed-for-life” expectations.
4.8.2 Lubricants for Traction Enhanced Differentials Conventional, or sometimes commonly referred to as “open,” differentials have a major disadvantage in their ability to deliver equal torque to both wheels in varying traction conditions. If one drive wheel is
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Handbook of Lubrication and Tribology (a) Heel
Concave side
Toe
Convex side
Hypoid gears (b)
Spiral bevel side gears
Axle tube
Axle tube
Hypoid ring gear
FIGURE 4.17
Hypoid drive pinion
Hypoid type rear axle designs.
exposed to a surface with low traction and the applied drive torque exceeds the traction coefficient, the wheel will break loose and increase in rotational speed until it is revolving at twice the speed of the ring gear. At this point, the opposite drive wheel (exposed to a higher traction surface) will stop revolving and no torque will be delivered to the high traction wheel. To improve traction performance, limited slip and locking type differentials of various configurations have been developed. All limited-slip differentials have friction clutches that come into action when a substantial traction difference exists between the two sides of the differential. In these cases, internal spiral-bevel gears within the hypoid gear housing, in conjunction with the clutch assembly, compensate, and transfer increased torque to the high traction wheel. Differential clutch systems can present unique lubrication needs. Frictional properties must be such that clutch engagement is sufficient to deliver torque transfer and subsequent traction enhancement, but yet allow clutch slippage during conditions of rotational speed bias from left to right wheels. The clutches
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must engage firmly so that proper torque is obtained. Conversely, the clutches must slip smoothly when differential action is required. If the clutches do not release properly or if they lock and release in a repetitive stick-slip manner, then chattering or driveline shudder will be generated. In order to eliminate noise and vibration shudder when the clutch mechanism engages, the lubricant used in limited-slip differentials must have a friction-modified formulation that provides the correct balance for the intended design. Additives that provide boundary lubrication are one of various approaches used to meet this need. Two general approaches have been used to satisfy the need of limited-slip differential lubricants. The first approach is to formulate a complete rear axle lubricant with built-in limited-slip additives. The second approach is to provide a separate additive to enhance frictional performance as a “top-treat” product that is introduced into standard hypoid lubricants when used in limited-slip applications. In these situations, care should be exercised in top-treat additions, as these friction-enhancing additives tend to be highly surface active and may interfere with the normal function of the primary performance additives in the lubricant, thereby leading to severe gear set or bearing degradation. As with many other vehicle systems, performance expectations continue to increase with higher engine outputs at the same time that vehicle size and weight requirements are becoming more restrictive. This combination inevitability leads to component downsizing with increase loading expectations. As a result, drivetrain lubricants and the additive systems that provide their performance enhancements will continue to play a vital role in vehicle design and marketability. Test methods for gear lubricants include such operating conditions as high torque at low speed and high speed with shock loading. Lubricant test measurements include oxidation resistance, protection against wear and material failure, corrosion protection, retention of desired friction characteristics, acceptable viscosity at both low and high temperature, and seal compatibility. Gear lubricants used in manual transmissions must have appropriate friction characteristics to ensure proper functioning of the transmission synchronizers to prevent gear clashing and allow smooth shifting. A standardized system for classifying, testing, and rating rear axle and gear box lubricants has been developed which utilizes several specifications or standards to establish requirements. As such, two parameters are used to define a gear lubricant, namely (1) SAE Viscosity Grade, and (2) API Service Designation. Axle and transmission lubricant viscosity is specified through the SAE J306 recommended practice document. The various viscosity levels exhibit distinct features for low and high-temperature performance. The values for each of these categories, including lubricant viscosity grades of 75, 80, 85, 90, 140, and 250W, may be further understood by reviewing SAE J306. American Petroleum Institute Service Classifications are the most widely accepted user language for defining gear lubricant performance. These classifications, namely GL-1 through GL-5 define gear lubricants by general type, severity of service, and application. The OEM vehicle manufacturer and military specifications may also be used to define oil quality levels. The MIL-L-2105 military specification for multipurpose gear oil, which was issued in 1976 and defines an API GL-5 level lubricant, is the specification under which oils are procured by the U.S. Military. This specification is also used by most equipment builders to define their axle and transmission performance requirements, particularly for service fill. The latest revision as of time of this text is MIL-L-2105E, which includes the following test requirements: • • • • • •
High speed shock loading test: CRC L-42 Low speed — high torque test: CRC – L37 Resistance to corrosion in the presence of moisture: CRC L-33 Thermal oxidation stability: CRC L-60 Foaming tendencies: ASTM D892 Stability in the presence of copper and copper alloys: ASTM D-130
Additional information on the applicable specifications can be obtained from Society of Automotive Engineers (SAE).
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Handbook of Lubrication and Tribology
Wheel flange
Outboard seal
Spindle
Inboard seal
Ball Hub
FIGURE 4.18
Sealed wheel bearing — typical (cut-away view).
4.9 Automotive Wheel Bearings Lubrication 4.9.1 Wheel Bearing Overview Wheel bearings for automotive applications utilize rolling element bearing designs and fall into one of two general categories: either sealed or serviceable. Sealed wheel bearing assemblies (Figure 4.18) are filled by the bearing manufacturer and are permanently sealed prior to installation in the vehicle. These assemblies are considered “filled-for-life” and do not require intermittent service regreasing. Serviceable wheel bearings (Figure 4.19) are lubricated prior to vehicle installation and require relubrication at regular intervals. The exception to this is for wheel bearings utilized in straight rear axle vehicles where the wheel bearing is located at the outboard end of the rear axle tube and is continuously lubricated by the hypoid gear oil contained in the rear axle assembly. Conventional rolling element bearings are used in the road wheel hubs. These standard manufactured components are hardened high carbon steel, machined and ground to a high tolerance and surface finish. Usually, deep groove bearings or a back-to-back angular contact bearing assembly is used. The bearings are packed with grease during assembly and sealed with elastomeric lip seals. One of the more common causes of wheel bearing failure is by damage to the hardened bearing surface caused during incorrect fitting or as a result of an impact load during vehicle operation. The surface damage acts as an initiation site for rolling contact fatigue failure.
4.9.2 Wheel Bearing Grease Selection Grease selection for any wheel bearing should start with the consideration of the NLGI-GC classification as a requirement. Such greases have proven performance levels per standardized tests established in ASTM D-4950 representative of typical wheel-bearing applications. Refer to the earlier section titled “Classification Systems for Automotive Chassis and Driveline Lubricants” for an overview of the properties and requirements for standard chassis grease classifications. Functional requirements of wheel-bearing greases can be summarized as follows: • Wear prevention in ball bearings • Rust prevention
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Automotive Chassis and Driveline Lubrication
Spindle
4-33
Inboard seal
Inboard cage & roller ASM
Retaining nut and cotter pin
Rotor
Outer bearing race
Outer seal/ dust cap
Outboard cage & roller ASM
FIGURE 4.19 Serviceable wheel bearing — typical (exploded view).
• • • • • •
Seal compatibility Low-temperature torque Wear prevention in roller bearings Fretting corrosion prevention Friction reduction Good oxidative stability
Typical product usage for wheel-bearing greases can be summarized as follows: • Mineral oil based, lithium complex thickened greases constitute the largest usage 1. Cost effectiveness vs. performance • Other types used include, but are not limited to: 1. Polyurea 2. Calcium complex 3. Anhydrous calcium • Synthetic base oil usage more common than for chassis greases due to: 1. Extended temperature ranges 2. “Lubed-for-life” service (sealed assemblies) • Solid Additives Occasionally Used (MoS2 , Polyethylene, etc.)
4.10 A Vision of the Future Clearly, chassis and driveline automotive lubricants in the future will have to differ dramatically from those of today to meet the demands of future automotive vehicle performance. Drivability and controlled capability will be critical for chassis and driveline component design. What lubricant technology will be necessary for hybrid, electric, or fuel cell powered vehicles in the years 2050 or 2100 is yet to be determined.
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In the relatively short term, automotive chassis lubricants will need to be formulated with base oils and additives which do not cause deterioration to chassis system or driveline components while, at the same time, providing high-performance enhancements in support of high-efficiency and NVH-free vehicles. Additionally, customers are demanding more maintenance-free vehicles, thereby promoting increased development and implementation of fill-for-life lubricant systems for chassis and driveline components. These demands will require new, novel approaches to advanced lubricant formulations and tribological solutions [16] for vibration- and noise-free chassis systems and to improve overall vehicle ride and drivability. Additionally, unique low-friction and wear-resistant lubricants [16] will be required to meet escalating performance requirements, both current and near-term. It is apparent that further advancements in near-zero NVH lubricants will be required as vehicle systems become more refined and customer expectations continue to increase. This will nurture significantly more innovation in the lubricant industry than we have seen in the past 50+ years. Such advancements, combined with innovation in the automotive component design arena, will drive further optimization of automotive system capability and integration. Foreseeably, the future will embrace design and implementation of “smart” lubricant systems [17] which will automatically control critical lubricant parameters through sensor feedback and control, thereby optimizing vehicle and subsystem performance. These significant challenges to our automotive and petroleum industries will be seriously considered as the primary drivers for our ongoing research and development efforts. The end goal is to provide the vehicle owner with a very positive ownership and driving experience, while at the same time making significant gains in vehicle safety, resource utilization, technology advancement, and environmental stewardship.
Acknowledgments The authors would like to thank Dr. James Spearot and Dr. Michael McMillan of General Motors Research and Development Center, as well as Mike Richardson and D. Craig Cook of Delphi Saginaw Steering Systems for permission to author this text and for their assistance in the editing process and providing useful comment. In addition, we would like to acknowledge the assistance of Chao-Fu Qin of General Motors Chassis Division, Vardarajan R. Iyengar and Alexander A. Alexandridis of Delphi Dynamics and Propulsion Innovation Center; Mark Sweigart, Mark De Groat, Ryan Pavlawk, Kirk Reinbold, and Jeff Ortiz of Delphi Saginaw Steering Systems Division; as well as Larry Baldwin, John Clifford, David Reuter, and Roger Sexton, all of Delphi Energy and Chassis Division, for providing us invaluable images for use in this chapter.
References [1] Polishuk, A.T., Automotive chassis lubrication, CRC Handbook of Lubrication — Theory and Practice of Tribology, Volume 1, Booser, R.E., Ed., CRC Press, Boca Raton, FL, 1983. [2] ASM Handbook, Friction, Lubrication, and Wear, Peter Blau, Ed., Volume 18, ASM Handbook Series, 2003, 3rd ed., 2003. [3] ASTM 2003 Annual Books of ASTM Standards, Petroleum Products, Lubricants, and Fossil Fuels, Section 5, 2002. [4] Jahn, D.M., Grease Applications — Automotive; presented as part of National Lubricating Grease Institute’s Basic Grease Education Course at NLGI 66th Annual Meeting, Tuscon, Arizona, 1999. [5] Technical Committee Overview Publication, National Lubricating Grease Institute website content, http://www.nlgi.org/technicalcom.htm. [6] NLGI Lubricating Grease Guide, 4th ed., NLGI, 4635 Wyandotte Street, Kansas City, Missouri, USA 64112, 1996. [7] Rabinow, J., Magnetic Fluid Torque and Force Transmitting Device, U.S. Patent No. 2575360, 1951. [8] Tung, S.C., Rule, D., Iyengar, V.R., and Alexandridis, A.A., Wear testing development and test procedures for evaluation of seal materials, STLE Tribology Transactions, T47, 7–16, 2004.
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[9] Carlson, J.D., Catanzarite, D.M., and Clair, K.A., Commercial magneto-rheological fluid devices, International Journal of Modern Physics B, 10, 2857, 1996. [10] Kordonsky, D., Elements and devices based on magnetorheological effect, Journal of Intelligent Materials, Systems and Structures, 4, 65, 1993. [11] Fish, Dr. G. and Cole, J., Tribology and Lubrication Requirements of Constant Velocity Joints; reprinted from: 1998 Transmission and Driveline Systems Symposium: New Developments and Advanced Concepts in Systems and Components (SP-1324); SAE International Congress and Exposition, February 23–26, 1998. [12] Fish, Dr. G. and Jisheng, Dr. E., The Effect of Friction Modifier Additives on CVJ Grease Performance, NLGI Technical Paper #0127, NLGI, 4635 Wyandotte Street, Kansas City, Missouri, USA. [13] Fish, Dr. Gareth, Constant Velocity Joint Greases, NLGI Technical Paper #9903, Presented at the 66th Annual Meeting National Lubricating Grease Institute, October 24–27, 1999. [14] Automotive Lubricants and Lubrication, The Society of Tribologists and Lubrications Engineers — Detroit Section, Seminar presented on November 5, 1992. [15] Tung, S.C. and McMillan, M.L., Automotive tribology overview of current advances and challenges for the future, Tribology International, 37, 517–536, 2004. [16] Hsu, S.M., Nano-lubrication: concept and design, Tribology International, 37, 537–545, 2004.
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5 Diesel, Dual-Fuel, and Gas Engines 5.1
Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
5-1
The Diesel and Otto Cycles • 2- and 4-Stroke Operation
5.2
Range of Engines and Applications . . . . . . . . . . . . . . . . . . .
5-3
Low Speed 2-Stroke Marine Diesel • Medium Speed 4-Stroke Marine Diesel • Medium Speed 4-Stroke Multi-Fuel/Diesel • 4-Stroke Dual-Fuel • 4-Stroke Medium Speed Gas
5.3 5.4
Cooling Systems. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Lubricating Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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Diesel Engine Lubricant Specifications • Trends in Diesel Lubricants • Stationary Gas Engine Lubrication
D.J.W. Barrell and M. Priest University of Leeds
5.5 Conclusion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Acknowledgment. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
5-11 5-11 5-11
5.1 Introduction This chapter introduces the wide range of large internal combustion engines that are used for marine propulsion and power generation. The latter can be smaller on-board engines or much larger stationary engines. In principle, they have many similarities to automotive engines, although the range of fuel quality and the different legislative limitations imposed on the emissions can be as influential as the effects of scale. Internal combustion engines can have a high specific power output and are thus favored for a wide range of vehicles. They are also useful in fixed plant, in electrical power generation, for example. The basic designs of all these different engines today are broadly similar to those from several decades ago, but improvements in lubricants and materials, as well as combustion, have not only increased efficiency but also reduced the output of undesirable emissions. The latter is beneficial and is nowadays a requirement by law since governments have been setting increasingly tighter demands. Large engines tend not to use gasoline as a fuel but rather heavier petroleum distillates (e.g., diesel oil) and gas. There are broad similarities among all the reciprocating piston engines and the choice of fuel generally corresponds to differences in fuel delivery and ignition method. There are many comparable features between these and the automotive engines discussed in Chapter 1, in terms of both component design and lubrication demands. These engines can be very large and an example is shown in Figure 5.1 [1], where the staircases illustrate the scale. The engines can employ the Diesel or Otto thermal cycle and
5-1
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FIGURE 5.1
Handbook of Lubrication and Tribology
12-cylinder Sulzer RTA96C giving 68,640 kW (93,360 BHP).
the 2- or 4-stroke cycle. They have a multitude of uses — diesel engines in particular, for propulsion of large vehicles, both on land and water.
5.1.1 The Diesel and Otto Cycles The Diesel cycle employs compression to ignite the air–fuel mixture. The piston compresses the air charge until the temperature exceeds 540◦ C. Near Top Dead Center (TDC), the fuel is injected as a mist and spontaneously ignites to provide a very large pressure to drive the piston. High cylinder pressures are used and the high compression ratio aids engine efficiency. The Otto cycle employs a spark to ignite the air–fuel mixture, as is common in gasoline automobile engines, and this is used in gas engines. To avoid early ignition of the fuel, a lower compression ratio is used, resulting in lower cylinder pressures before ignition. Rapid combustion leads to momentarily high cylinder pressures until the piston displacement results in an expanding cylinder volume.
5.1.2 2- and 4-Stroke Operation The 2-stroke cycle completes the intake-compression-power-exhaust stages in one engine revolution, scavenging the cylinder through cylinder ports. The more frequent combustion can result in higher engine output, but at the expense of lower peak engine speed, higher piston temperatures and lower engine efficiency. The 4-stroke cycle results in a more complete combustion of the fuel, crucial to minimising exhaust pollutants, and a more flexible efficient engine. Higher engine speeds are possible with the latter method, and this compensates for loss of a power stroke in comparison with the 2-stroke principle.
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5.2 Range of Engines and Applications This range of engines covers a broad spectrum of applications and fuels, for marine and land use, stationary and propulsion. There is a wide range of 2- and 4-stroke engines available covering power requirements from around 1 MW to 100 MW. Many of these engines are used for marine propulsion. Not only may there be problems due to variable quality of the fuel in this application, but the environment will not be kind to machinery.
5.2.1 Low Speed 2-Stroke Marine Diesel Large tankers and container ships require powerful engines that are very reliable, with several years between overhauls. Their engine speeds range from 120 rpm down to 60 rpm for the larger cylinders, where the engine stroke is around 3 m and the bore around 0.6 m. There may be 12 or more cylinders, each producing in up to 4 MW. Figure 5.2 [2] shows an section through an example of a 2-stroke marine diesel engine. Note the extreme engine height in comparison with the bore diameter.
FIGURE 5.2 Section through a crosshead-engine.
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FIGURE 5.3
Handbook of Lubrication and Tribology
Crosshead-engine connecting rods.
These engines have a long piston stroke and a crosshead is placed between the piston and the crankshaft, effectively creating a taller piston. Both the crankshaft and crosshead journal bearings will be traditional white-metal bearings as used on smaller engines. The crankshaft thrust bearing will be of the tilting-pad type. Typical crosshead connecting rods are shown in Figure 5.3 [3]. The crosshead bearing may be hydrostatically lubricated. The taller cylinder will need lubricant fed via quills through the cylinder wall as splash lubrication would not effectively reach the upper cylinder walls. The cylinder head valves may be operated hydraulically or by solenoids, removing the need for a traditional camshaft. Engine designs are aimed at economical operation, but the control of exhaust emissions is increasingly important. Marine diesel nitrogen oxides (NOx ) emissions need to comply with Annex VI of the International Convention for the Prevention of Pollution from Ships 1973/78, commonly known as the MARPOL Convention (www.imo.org) [4]. NOx levels must not exceed 9.8 g/kWh above 2000 rpm, rising to 17 g/kWh below 130 rpm. MARPOL was developed by the International Maritime Organisation (IMO). This requirement may be met by control of compression ratio, timing of the fuel injection and valves, plus injector nozzle optimization. Rail fuel-injection is commonly used as it allows greater control of fuel volume and timing, for improved efficiency and reliability. Nonmarine standards are tougher and all applications will probably facer lower limits in the future. Service intervals are largely determined by piston ring and bore wear. Liner material choice and ceramic coatings on piston rings are assisted by good design. A modern piston design is shown in Figure 5.4 [1]. Although the piston rings have similar designs to automotive rings, the piston is made from steel and the ring grooves may be chrome-plated. Careful liner honing and upper cylinder lubricant injection improve the extent of a good lubricant film. Reducing the piston-bore clearance above the top ring, which can be achieved by the use of a separate liner insert, minimizes the accumulation of carbonaceous deposits that can encourage bore polishing. Cooling of the piston crown is assisted by the delivery of lubricant
© 2006 by Taylor & Francis Group, LLC
Diesel, Dual-Fuel, and Gas Engines
5-5
FIGURE 5.4 Sulzer Tribopack.
FIGURE 5.5 Cooling bores below the piston crown.
through the connecting rod to bores within the underside of the crown. The inverted piston crown in Figure 5.5 [2], shows such cooling bores. Piston crowns may also be water cooled by the cocktail-shaker method, using water delivered to a partially air-filled space where it can freely splash. Good engine rigidity is important to ensure good alignment of all bearing surfaces. The cylinder jacket may be formed from a single cast iron unit or assembled from separate blocks. Access hatches must be provided for engine servicing.
© 2006 by Taylor & Francis Group, LLC
5-6
FIGURE 5.6
Handbook of Lubrication and Tribology
Composite piston with steel crown and cast iron skirt.
The combustion chamber design is critical to achieve low surface temperatures and good lubrication. With large volumes of injected fuel, impingement is to be avoided and efficient combustion requires good injector nozzle design. Controlled cooling of the liner is accompanied by the aforementioned oil-cooling of the underside of the piston crown.
5.2.2 Medium Speed 4-Stroke Marine Diesel To achieve a more compact engine than the low-speed marine diesel, it is wise to demand a shorter piston stroke, which also allows for higher engine speeds for a given maximum piston velocity. Using an almost square configuration, with bore and stroke around 0.5 m and an engine speed of 500 rpm, a cylinder output of 1 MW is possible using a 4-stroke design. Trunk-type pistons are used, which are much more compact than crosshead pistons. They are basically similar to smaller automotive diesel engine pistons. The piston traditionally has a steel crown and nodular cast iron piston skirt, as shown in Figure 5.6, with two compression rings and one oil scraper ring. Additional pressure lubrication of the skirt may be used in cases where severe running conditions are expected. Raising engine speed, rather than cylinder pressure, means that turbochargers can be used to maximize the gains of the Miller cycle, with which there can be gains in engine efficiency by leaving the intake valve open for part of the compression stroke, reducing the work done by the piston. This results in lower compression end temperatures, leading to reduced NOx emissions [5]. Higher engine speeds demand superior balancing of the crankshaft and adjustable balance weights are provided. The crankshaft bearings may have temperature sensors fitted. Three-piece connecting rods (see Figure 5.7 [4]) are used to facilitate overhaul of the piston without disturbing the big-end bearing, and vice versa.
5.2.3 Medium Speed 4-Stroke Multi-Fuel/Diesel Multi-fuel engines can run off gas oil, marine diesel fuel, intermediate fuel, heavy fuel, crude oil, and gas, allowing the most economical available fuel to be used. These fuels, however, may contain chemically
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Diesel, Dual-Fuel, and Gas Engines
5-7
FIGURE 5.7 Three-piece connecting rod.
aggressive components, such as vanadium, sodium, and sulfur. Low temperature corrosion may occur due to the condensation of sulfuric acid. High temperatures, exceeding 430◦ C, encourage the reactions of vanadium and sodium. The exhaust valves are particularly vulnerable to such temperatures. Piston ring and cylinder wear is a major issue and good lubrication is an obvious requirement, but also bore polishing must be avoided as this increases lubricant consumption. A cylinder insert to reduce piston–bore clearance at the top land can reduce this problem by limiting the formation of carbonaceous deposits. Wear in the cylinder head can be reduced by fitting rotators to both the inlet and exhaust valves, to positively encourage normal rotation and to distribute thermal and mechanical loading. The lobes on the camshaft will be designed to minimize vibration of the tappet rollers. Exhaust emissions have to be controlled and the reduction of NOx can be achieved by direct water injection, which uses latent heat to lower the gas temperature, and by controlling the air–fuel properties to optimize combustion timing. Modern engines increasingly uses common rail technology, and control and monitoring is heavily automated.
5.2.4 4-Stroke Dual-Fuel Dual-fuel engines run on natural gas or light fuel oil and may change from one fuel to the other during operation. These engines are used for propulsion as well as electrical power generation at 50 or 60 Hz. Lean-burn technology is available for these fuels, with more air made available than is necessary for complete combustion. This lowers peak temperature and so reduces the production of NOx . The combustion begins with normal ignition of pilot fuel (light fuel oil [LFO]) that can be followed by either fuel as the main charge. Individual cylinder control is important to optimize the combustion. When full gas running is in operation, the LFO pilot fuel contributes less than 1% of the fuel consumed. Air–fuel ratios around 2.0 to 2.2 allow high thermal efficiency without the risk of knocking. Pistons, connecting rods, and crankshafts are typically built to the same design as the diesel engines described above. However, the cleanliness of the gas combustion, when used, in comparison to diesel means that the operating conditions may be less detrimental to engine components and the demands on the lubricants less taxing.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
5.2.5 4-Stroke Medium Speed Gas Electrical power generators can make use of lean-burn, spark-ignited gas engines using the Otto cycle. With bores and strokes in the range of 300 to 400 mm and up to 20 cylinders in a V-configuration, electrical power generation can approach 10 MW. The generating efficiency lies around the 50% region, but the heat output is used in cogeneration to produce an overall efficiency exceeding 90%. This efficiency gain can impose greater load on the engine and thus raise the specification requirements of the lubricant.
5.3 Cooling Systems Two circuits may be required to maintain optimum engine operation, using two separate temperatures. A lower oil temperature is desirable in the sump, to minimize oil degradation, and the circuit may operate between 60 and 80◦ C. It is therefore beneficial to isolate the lubricant fed to the cylinder, where higher temperatures exist, and may even be desirable to reduce condensation and associated corrosion. The cylinder blocks will therefore be maintained at a higher temperature (e.g., 95◦ C). An even cooler circuit may be used to control the intake air temperature.
5.4 Lubricating Systems Lubrication systems in the small to medium range of engines will be similar to those in automotive engines, using splash lubrication from a wet sump and a single, pumped lubricant supply from the sump to the engine bearings. Larger engines will pump lubricant to the mid-stroke region of the cylinder via quills, particularly on long-stroke crosshead piston engines. Large crosshead engines have cylinders which are isolated from the crankcase, unlike in trunk-type engines, and blow-by gases do not tend to contaminate the sump oil, which would increase the rate of lubricant degradation. Using a separate system to lubricate the cylinder also has the advantage of isolating the sump lubricant from cylinder temperatures and gases that can promote lubricant degradation. The additional role of the lubricant in cooling the underside of the piston crown has been described earlier. The lubrication system can use automatic back-flushing to avoid the need for regular replacement of disposable filters. Spectroscopic analysis is a useful technique for monitoring the status of the lubricant and engine. Metallic species in the lubricant arise to some extent from the oil additives, but this is decreasingly the case. Sudden increases in certain metallic elements from wear can give warning of component failure and, in some cases, identification of which type of component due to differences in composition.
5.4.1 Diesel Engine Lubricant Specifications General heavy-duty diesel lubricants will typically be mineral oil SAE 10W-30 or 15W-40 multigrades, SAE 30, 40, or 50 monogrades, or a synthetic SAE 5W-40. The lubricants have to demonstrate high oxidation and corrosion resistance, high total base number (TBN), dispersancy and detergency, and they must have appropriate viscometric properties. The relevant test standards include MIL-L-2104 C/D/E (United States), API CD/CE/CF4, Defstan 91-22/3 (United Kingdom), GOST 12337 (Russia), GOST 1861, and BIS:13656 (India). Engine manufacturers will make specific recommendations as to the choice of lubricant. Modern lubricants must deal with higher levels of soot and higher temperature. Modern engines also show less oil consumption, and reduced topping-up means the additive content must be stable or remain sufficiently high to last between services. The choice of marine diesel engine lubricant may also depend on the level of sulfur present in the fuel. The essential features that will be sought are good thermal and oxidation stability, high temperature film strength, high shear stability, good detergency and dispersancy, TBN reserves, and corrosion resistance.
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Diesel, Dual-Fuel, and Gas Engines TABLE 5.1
5-9
Methods to Reduce Exhaust Emissions
Method
Type
Diesel oxidation catalyst (DOC)
Oxidation
Selective catalytic reduction (SCR) Lean NOx catalyst (LNC) NOx adsorbers, lean NOx trap (LNT)
Reduction by ammonia/urea Reduction by hydrocarbons Adsorption
Emissions lowered Carbon monoxide (CO), hydrocarbon (HC), particulate matter (PM), odor NOx NOx , CO, HC, PM NOx , CO, HC
Detergents are especially important given the varied quality of fuel available. Calcium salicylates are used for trunk diesel engines, as well as for gas engines, due to their strong cleaning and antioxidant properties. Phenates are used in both cylinder lubricants and trunk diesel engines. Sulphonates have a relatively high TBN are useful in cylinder lubricants due to their ability to neutralize acids.
5.4.2 Trends in Diesel Lubricants Increasingly tighter emissions legislation is driving good advances in engine design. These are most applicable to the automobile sector, but are becoming is increasingly applicable in all applications. Reductions in particulate and NOx emissions are required and are partly dealt with by exhaust after-treatment methods. These consequently require lower levels of sulfur in the fuel, to avoid impairment. In the United States, the Environmental Protection Agency (EPA) has proposed a new, much stricter limit of 15 ppm of sulfur in diesel fuel from 2006. Other countries have ultra-low-sulfur diesel (ULSD) fuels with maximum levels of 50 ppm sulfur. Worldwide marine legislation may have to attain such standards in time. The fuel treatments, hydro-processing, and hydro-treating may reduce the natural lubricity of the fuel provided by some of the polar compounds present, resulting in increased wear. Similarly, the reduction in natural antioxidants may lead to the greater formation of sludge. Reduction of emissions is achieved by converting the compounds to a preferred form or trapping the particles. The methods are summarized in Table 5.1 and they must be considered when formulating the lubricant for the engine. Lubricants in the future will need to be designed for engines that use fuel more efficiently, with longer, more economical service intervals, and with strictly controlled exhaust emissions. Phosphorous levels will need to be reduced and sulfur may need to be avoided altogether. In October 2000, a new lubricant specification for heavy-duty 4-stroke diesel engines was introduced by Japanese Automobile Standards Organization (JASO), called the DH-1. These lubricants must meet Standard JASO M 355:2000. The improvements relate chiefly to reduced abrasive wear, corrosion, oxidation, and soot control.
5.4.3 Stationary Gas Engine Lubrication Natural gas engines are used to drive compressors, pumps and generate electricity, both for emergency use and primary power. These engines generally produce lower NOx , CO and particulate emissions. In common with diesel engines, the engine options include: 2- and 4-stroke cycles, a wide range of power output (up to 10 MW), and single up to 20-cylinder layouts. Air may be supplied by natural aspiration or turbo-charging and with either a stoichiometric air/fuel ration or by employing lean-burn technology. Increasing use of these engines for power generation may lead them to account for up to 30% of worldwide electricity generation by 2010. Dry natural gas is typically more than 85% methane. Appropriate engine oils differ from those used in diesel engines due to the lack of soot contamination of the sump oil, lack of fuel dilution, the higher combustion temperature, and the constant speed operation, which tends to assist the accumulation of deposits. Unlike diesel engine lubricants, there is no specific American Petroleum Institute (API) minimum performance, and field reliability is often used to gauge performance. The viscosities used are usually
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
SAE 40, with some SAE 30 and multigrade 15W-40 available. The last option, however, may not be recommended by the engine manufacturer due to poor long-term operation in the governor. One of the main problems with these engines arises due to knocking, which can be increased by the presence of ash deposits in the combustion bowl. Thus, the use of low ash (<0.5%) or even ashless (<0.1%) lubricants is advised. Ash results from burning of metallic detergents (calcium and barium) and antiwear additives (zinc). The ash can be beneficial in maintaining the valves, and the detergents are needed to combat the formation of acids, so a balance must be reached. Excessive ash formation may lead to burnt or worn valves [6]. Some older 4-stroke engines even require high ash oils to provide a sacrificial layer on the valve to resist valve seat recession. Valve damage can result from both lubricant and mechanical sources, so care is needed in diagnosing the source of failure. Clearly, 2-stroke engines have less need for high ash oils. Higher ash oils with higher base numbers can also be used to deal with corrosive fuels. Oxidation and nitration both lead to oil degradation. The former is especially relevant to gas engines due to the unusual combustion conditions. NOx formation depends on air–fuel ratio, combustion temperature, and timing and engine speed. These gases can react with the oil film remaining on the cylinder wall or be transferred to the sump oil. Better exhaust scavenging reduces the opportunity of the gases to interact. Cylinder walls above 150◦ C also encourage the breakdown of the nitration products. Reducing the contamination of the sump oil can be achieved by reducing piston ring blow-by, ventilating the crankcase, and maintaining the sump oil temperature above 80◦ C. Base oils such as polyalphaolefins (PAO) synthetics and hydro-treated paraffinic oils appear to be less susceptible to nitration problems, but Group I mineral base stocks are generally preferred to Group IV synthetics due to cost at present. The effects of nitration on the lubricant will appear as an increase in viscosity, acidity, and the levels of insoluble compounds, which may lead to varnish and sludge formation. There are no specific antinitration additives available at present. Detergents are vital engine lubricant additives. Salicylates are the main components of the additives. Phenates excel at high temperatures for short periods, but the requirement in these engines is for lubricants to work at lower temperatures for long periods. The relative merits of these detergents are shown in Table 5.2 [6]. Oil changes are recommended after around 1000 h, depending on the operating conditions and quality of lubricant. Careful condition monitoring, using Fourier Transform Infrared (FTIR) and measuring acid and base number, can be use to extend service intervals. Analysis of the oil for metallic wear debris and other likely contaminants (water, glycol) is still valuable, but excessive delaying of oil changes can encourage deposit formation and rapid component wear, counteracting any previous savings. Oil filtration was traditionally provided by sock filters, but centrifugal filters are now being used in addition. Nonselective catalytic reduction systems are common on stoichiometric engines and selective systems on lean-burn engines. Oxidation catalyst systems are also used on lean-burn engines to lower HC and CO emissions. Most of these systems are prone to damage by phosphorous compounds, found in Zinc Dialkyldithiophosphate (ZDDP), and ash, which has a bearing on the requirements for oil formulation.
TABLE 5.2
Surfactant Composition Effects [6]
Detergent technology (type of surfactant) Use in gas engine oil (GEO) Deposit control (engine cleanliness) Oxidation control, bulk conditions (long duration, high temperature) Thin film conditions (short duration, very high temperature) Acid neutralization (corrosion control)
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Sulfonate
Phenate
Salicylate
Not desirable Medium Poor
Core of traditional GEOs Good Medium
Best for modern GEOs Good Excellent
Poor
Excellent
Medium
Good
Poor (requires rust inhibitor)
Good
Diesel, Dual-Fuel, and Gas Engines
5-11
5.5 Conclusion Medium and large engines share many of the features and issues of automotive engines. Fuel oil may be of lower purity than automotive fuels, yielding greater chemical demands for the lubricant, while dry natural gas can be less taxing. These slower running engines operate over a narrower range of engine speeds, with less cold starting, so lubricant performance need only be optimized for more limited conditions. However, these engines are more expensive to service, their lubricants may need to be stable for longer periods, and engine reliability in these applications is a major factor. Engine manufacturers advise on the choice of lubricant and there is less reliance on industrywide standards than in the automotive sector.
Acknowledgment The authors would like to thank the Wärtsilä Corporation, Finland, for the kind permission to use their photographs and diagrams.
References [1] [2] [3] [4]
Sulzer RTA-C Technology Review, Wärtsilä Corporation, 2004. Sulzer RT-flex60C Technology Review, Wärtsilä Corporation, 2003. Sulzer RTA-T Technology Review, Wärtsilä Corporation, 2002. L. Goldsworthy, Controlling Oxides of Nitrogen Emissions from Ship Engines. 2001 Marine Diesel Engine NOx Measurement and De-NOx Technologies Workshop (Taiwan), Taipei, Kaohsiung, 2001. [5] M. Rose, Infineum International Ltd, Gas Engine Oils — A Question of Balance. Machinery Lubrication Magazine, March 2004. [6] R. Scott, Noria Corporation. Stationary Natural Gas Engine Lubrication. Machinery Lubrication Magazine, September 2004.
© 2006 by Taylor & Francis Group, LLC
6 Aircraft Gas Turbines 6.1 6.2 6.3
The Gas Turbine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . General Oil System Description . . . . . . . . . . . . . . . . . . . . . . Lubrication Circulation System . . . . . . . . . . . . . . . . . . . . . . .
6-1 6-3 6-6
Lubrication Storage • Heat Management System • Oil Scavenge • Oil Pumps • Filtration • Materials
6.4
Airflow . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
6-11
System Pressurization and Sealing • Vent System • Air/Oil Separation • Oil Firing
6.5
Bearing, Seal, and Gearbox Lubrication . . . . . . . . . . . . . .
6-14
Bearing Lubrication • Gear Lubrication
6.6
Lubrication Selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
6-16
Oil Grades
6.7
Oil System Indications. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
6-17
Engine Measurements • Aircraft Indications
6.8
System Maintenance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
6-18
Maintenance Manual Procedures • Maintenance and Servicing
6.9
Andy Hall Rolls Royce plc
Engine Health Monitoring . . . . . . . . . . . . . . . . . . . . . . . . . . . .
6-20
Oil System Analysis Techniques — Magnetic Chip Detectors (MCDs)
Bibliography . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
6-21
6.1 The Gas Turbine The gas turbine in its very basic design is a type of engine that is used to generate various forms of power output. The power output produced can vary depending on how the gas turbine will be utilized. Some applications require the power as thrust, while others convert this energy into shaft horsepower to drive other methods of propulsion or ancillary equipment. The gas turbine has seen service as power for aircraft, helicopters, ships, power stations, pumping stations, and even motor vehicles. However, no matter how the gas turbine is utilized, the fundamental principles of operation remain the same (Figure 6.1 and Figure 6.2). For an aircraft gas turbine or aero engine, two types of power output are usually required. The turbofan or turbojet engine produces thrust to move the aircraft forward; where thrust is measured in pound-force (lbf) or kiloNewtons (kN). For a turboprop or turboshaft engine, the power required is shaft horsepower (SHP) as these types of engines drive propellers or rotor blades, respectively. The gas turbine is constructed of three main parts, a compression system, a combustion system, and a turbine system, and around these are installed the ancillary equipment which enables the gas turbine to 6-1
© 2006 by Taylor & Francis Group, LLC
6-2
Handbook of Lubrication and Tribology Single spool axial flow trubo-jet
Three spool high by-pass ratio turbo-fan
Two spool low by-pass turbo-fan
Single entry two-stage centrifugal turbo-prop
Two spool axial flow turbo-prop
FIGURE 6.1
Two spool turbo-shaft with a free-power turbine
Mechanical arrangement of aero engine gas turbines.
IP compressor HP compressor HP turbine
LP compressor or fan Combustor IP turbine
LP turbine
External gearbox
FIGURE 6.2
The mechanical arrangement of an aero engine.
operate as designed. The turbine of the engine drives the components in the compression system, while the turbine blades recover work out of the gas flow passing through. The compression system “does work” on the air induced into the gas turbine, and in today’s high bypass aero engine the pressure rise of the air at delivery to the combustion system when the aero engine is operating at take-off thrust can be as high as 40:1 or approximately 600 psia, with potential pressure rises of 50:1 or 55:1 (735 to 810 psia) being achievable in the future. This high level of pressure at delivery to the combustion system does not only
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Aircraft Gas Turbines
6-3
cause problems with the selection of materials as a result of the increased stresses and temperatures; it can also cause a change in the balance of forces that act on areas within the compressor/turbine unit. The combustion system is the section of the engine where the fuel is burnt and heat energy is released. This heat energy is added to the compressor delivery air to provide the power to drive the turbine and also produce the required output power. Because of the requirement to ensure that gas turbines are highly efficient, the combustion system on many engines will operate at the highest temperature possible; the main design restriction on achieving the maximum temperature will be the materials that are used for both the combustion system and the turbine system. The obvious target to achieve in the combustion section and produce the highest thermal efficiency would be the stoichiometric temperature of the fuel, but this would mean the hot section of the engine would have to be constructed of exotic materials to cope with this temperature or there would be a need to increase cooling for the hot section components, which would lower the fuel efficiency of the engine. For today’s civil high bypass turbofan engines, the combustion system can be operating at temperatures of 1850 K or 1577◦ C (3330◦ R or 2870◦ F); as such, many of the components within the turbine systems are cooled to protect them from the full heat of the gas flow. The turbine of the engine is connected to the compression system by a shaft; the combination of compressor, shaft, and turbine can also be identified as a spool. Depending on the type of aero engine, the gas turbine can have a single shaft, two shafts, or three shafts, each shaft or spool rotating independently of the others. Ideally, each turbine, compression system, and shaft should rotate at its optimum speed, which can vary from 2500 rpm for very large diameter low pressure systems or fan, to 40,000+ rpm for very small diameter high pressure systems. In all cases, each of the individual shaft is supported by a series of bearings to ensure that the shaft rotates smoothly and to ensure that the shaft is located correctly in both axial and radial directions. The bearings can be located in the engine so that they rotate between the shaft and a fixed part of the engine’s internal structure, or in some cases they can be located between two rotating assemblies. The bearings of a gas turbine generate a large amount of heat, which is caused by surface friction of either the rollers or balls in the bearing and contact with the bearing race. Oil is primarily supplied to the bearings to remove this heat and provide lubrication, thus reducing the friction as the bearing rotates.
6.2 General Oil System Description The oil system for an aero engine is a full recirculatory system and its main function is to supply oil to the engine’s internal drives, gears, and bearings, where the main rotating assemblies are (Figure 6.3). The oil is used to lubricate these locations and remove all unwanted heat throughout all operating conditions and power settings. The system is designed to ensure that the oil supplied to these drives, gears, and bearings is in the correct condition with regards to cleanliness, pressure, temperature, and quantity. The complete oil system for an aircraft gas turbine can be divided into three main subsystems: • Oil feed, lubrication, and cooling (pressure side) • Return oil (scavenge side) • Breather system (vent) Within these areas are the necessary components that ensure that each of the subsystems functions as designed (Figure 6.4). The architecture of the oil system is designed such that bearing chambers and gearboxes will be at different pressures relative to each other and, therefore, the flow from the pressure system will have to be controlled to ensure each area of the engine receives the correct oil flow. The pressure pump delivers the total oil flow for the whole engine at one pressure and this is distributed around the engine by a network of pipes. Restrictors are placed in oil supply lines upstream of the oil jets. The size of the restrictor is controlled so that the pressure drop across the oil jets is correctly maintained and within a specific range; this ensures adequate targeting and oil velocity.
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6-4
Quantity transmitter Oil tank FOHE
Pressure pump
Oil temp sensors
AOHE
L.P turbine bearing chamber
HP/IP turbine bearing chamber
Internal gearbox drive
LP/IP compressor bearings
Pressure filter
Scavenge filter
Master MCD
Pressure transmitter
Magnetic chip detector
Pressure switch
Scavenge return
Relief valve
FIGURE 6.3
HP feed Vent air/oil mist
Oil system schematic.
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Intermediate gearbox
Gearbox input drive assembly External gearbox Scavenge pumps Centrifugal breather
Handbook of Lubrication and Tribology
Strainer
Pump
Aircraft Gas Turbines
6-5 Pressure relief valve
De-aerator tray
Centrifugal breather Pressure filter Strainer
Oil pump pack From oil tank
Torquemeter pump
Air cooled oil cooler
FIGURE 6.4 Turboprop oil system.
Finally, before the oil enters into the oil jet, it can be passed through a “last chance” filter to ensure that no debris generated downstream of any other filtration system can block the oil jet. This type of filter (or strainer) is usually a screw thread filter and will be built into the engine upstream of the oil jet. Generally the circulation of oil is far in excess of the capacity within the system. For example, in a large turbofan engine of approximately 70,000 lbs of thrust, the nominal oil quantity contained in the oil tank is approximately 23 L (24.3 US quarts), of which approximately 18 L (19 US quarts) is usable. The complete oil system, including pipes, bearing chambers, and gearboxes, will have a capacity of approximately 36 L (38 US quarts) in total. To supply the various parts of the engine that need lubricating, the approximate flows required are as follows: • The oil supply to all the bearing chambers (including any internal gearboxes) is approximately 2800 L/h (or 2960 US quarts/h). • The oil supply to an external gearbox is approximately 210 L/h (or 220 US quarts/h). Therefore, the entire contents of the oil tank circulate around the oil system approximately every 29 sec. Because the design of the engine has different parts of the oil system operating at different oil pressures, all the areas that are supplied with oil will also have their own dedicated oil scavenge pump to recover the oil back to the tank. Generally, the oil scavenge system will be constructed so that there is only one pipe between the sump and each scavenge pump; this prevents recirculation and preferential scavenging. In an aero engine, there are bearing chambers that are required to be sealed using bleed air from the compressors. The air pressure outside the chamber is always greater than the oil pressure inside the chamber; thus the air is “allowed” to leak across seals into the bearing chamber and, as a consequence, these chambers will have to be ventilated. As this air is no longer required by the engine’s systems, it will be discarded overboard. However, because of the environment within these bearing chambers, the vent air will contain small amounts of oil as an air/oil mist. As the vent air will be exhausted to atmosphere, the oil that is suspended in the vent air needs to be recovered and returned to the tank; otherwise the oil consumption of the engine will increase dramatically as the oil will be wasted. To achieve
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6-6
Handbook of Lubrication and Tribology Gear shaft air outlet slots De-aerator segments
Air to atmosphere
Air/oil mist entry holes
Driven gears
Return oil to gearbox
FIGURE 6.5
Centrifugal breather.
this, the compartments in the oil system that require venting are connected to a centrifugal breather (Figure 6.5). There are two other separate oil systems on a civil aero engine, one for the engine’s starter and one for the integrated drive generator (IDG). The starter for many aero engines is an air turbine starter and, as the name suggests, this is the main component for starting the engine. The oil system on the start is self-contained and the operation of the starter is controlled by its duty cycle, which is defined in one of the operating manuals for the aircraft. The IDG is the main unit that supplies AC power to the aircraft. Whenever the engine is running, the IDG is usually running under load and generating a large amount of heat. Like the starter, the IDG has its own self-contained oil system; however, because of the time for which the IDG will be used, the oil system for the IDG will have its own dedicated oil cooler and filtration system.
6.3 Lubrication Circulation System 6.3.1 Lubrication Storage The oil tank is the main reservoir for the supply of oil to the pressure pump; it receives filtered oil from the scavenge pumps and is located in a position on the engine where easy access can be obtained. The capacity of the oil tank by design will always be far greater than that required for the oil; the main reason is that the oil tank must have some of its capacity reserved as air space. Usually, the air space is approximately 25% of total tank volume and is required because the scavenge oil that returns to the tank still contains small amounts of air which has been used to pressurize and seal some bearing chambers. In order to release this air, the scavenged oil will pass into a de-aerator in the oil tank, releasing any trapped air; the oil tank will be vented by connecting it to the centrifugal breather as this air is also an air/oil mist and the oil will need to be recovered.
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Aircraft Gas Turbines
6-7
The oil capacity of the tank will be determined by the following factors: • There must be sufficient oil in the tank for the aircraft to fly for its maximum duration at the maximum oil consumption. • There must be capacity for a reserve quantity of oil. • There will be a small amount of oil in the tank that is unusable because it is in the bottom of the tank and below the inlet of the feed pipe to the oil pump. The type of oil tank that will be installed on the engine will also depend on the type of engine. On some engines, it is better to make the oil tank part of the external gearbox; on high bypass engines, the oil tank is usually mounted somewhere on the engine’s fan case. For military engines, there are two other elements of an aircraft’s operation that need to be taken into consideration when designing the oil tank: negative “G” and inverted flight. Both negative “G” and inverted flight have the effect of causing the oil to be forced to the top of the tank and away from any feed pipe that is supplying the oil pressure pump. Most aero engines will be designed to allow them to run without a supply of oil to the bearings for a short period of time, but any prolonged running will cause the bearings to overheat and fail. Therefore, the oil tank on a military engine will be designed so that the oil feed pipe is always immersed in the oil, even when under negative “G” or when the aircraft is flying inverted. This can be achieved by allowing both the feed and vent pipes to rotate and follow the oil in the tank; alternatively, the feed pipe can be built into the wall of the tank and the whole tank is then allowed to rotate. Tank design is very important; it has to be strong yet lightweight and, most important, its needs to be fireproof. Because maintenance engineers need to have a visible indication of oil quantity in the tank, a sight glass will be installed on the side of the oil tank; this will also need to be fireproof.
6.3.2 Heat Management System The oil supplied to the engine’s bearings and gears is not only required to provide lubrication; the important job of the oil is to remove any waste heat generated by acting as a coolant. However, this heat will also need to be removed from the oil so that the oil remains within its operating limits and does not become stressed. Therefore, the gas turbine will have installed in the oil system a number of oil coolers that generally utilize one of two methods of cooling: fuel-cooled or air-cooled heat exchange. The fuel-cooled oil cooler or fuel/oil heat exchanger is a device that allows the fuel being supplied to the engine from the aircraft to be used to cool the engine’s oil prior to the fuel being used in the combustion process. This unit is designed and constructed to ensure that there is maximum transfer of heat from the oil to the fuel; to do this, the fuel is broken down into many small quantities. The inside of the unit is very similar to a boiler; the fuel passes vertically through a couple of hundred small-bore pipes of approximately 5 mm diameter. The oil enters the unit at the bottom of the heat exchanger; baffle plates cause the oil to “weave” to the top of the unit across all the fuel pipes and exit at the top. This design maximizes the time that the oil spends in contact with the fuel system and allows equal heat transfer into all the fuel. This is an ideal opportunity to make use of the waste heat in the oil system because it is advantageous on gas turbines for the fuel to be heated before it is circulated around the engine’s fuel system. Many civil aircraft spend the majority of a flight in cruise at high altitude, and the atmospheric conditions at these high altitudes are that the air is very cold: static air temperature above 36,000 feet (11,887 m) is −57◦ C or −70.6◦ F. The majority of the fuel for the engines is stored in the aircraft’s wings, and even allowing for some kinetic heating of the air as it passes over the wings, the fuel will be cold soaked during the cruise phase of the flight. The waste heat from the oil is also useful because aviation fuel is hygroscopic and may have water suspended in it. As the fuel cold soaks in the aircraft’s wing tanks, this water would turn into ice crystals. If this cooled fuel was allowed to pass into the engine’s fuel system without heating, the water crystals could block either fuel filters or the intricate parts of the fuel control system. Therefore, the fuel needs to be heated and the scavenge oil returning to the oil tank needs to be cooled; hence there is a trade,
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with the waste heat from the oil transferring into the fuel as it passes through the fuel/oil heat exchanger. Generally, the amount of heat transfer that takes place between the oil and the fuel is approximately 200 kW for a large high bypass aero engine. The air/oil heat exchanger can be used either in conjunction with or as an alternative to the fuel/oil heat exchanger on some gas turbines where the fuel flows are small and the heat transfer from the oil to the fuel would be inadequate to keep the oil temperature within operating limits. The design of this unit is very similar to a car radiator, except that in this installation, the air used to cool the oil will be diverted from one of the compressors, usually downstream of the LP compressor. On some engines such as turboprops, the only oil cooler may be an air/oil heat exchanger; for these engines, the airflow passing over the engine’s nacelle is diverted through the heat exchanger to cool down the oil. The air/oil heat exchange can also be installed on large turbofan gas turbines as a backup cooler to assist the main fuel/oil heat exchanger for the portions of the aircraft’s operation when the fuel flows are low (i.e., during the early stages of descent). However, generally in this example the cooler uses air from one of the compressors on the engine. This air has been worked by the compressor and in order to recover this lost energy, more fuel has to be put in the engine to maintain the required power output. Therefore, in this instance, there is a performance penalty on the engine every time this type of air/oil heat exchanger operates as it increases the fuel consumption of the engine. Installing a control valve between the compressor bleed inlet and the heat exchanger matrix will optimize the use of air for cooling the oil and, therefore, the valve will open only when necessary to keep the engine fuel efficient.
6.3.3 Oil Scavenge The scavenge system on the aero engine does a very important job in the oil system; it recovers the oil from the engine and returns it to the oil tank. Generally, each of the areas connected to the scavenge system will have a single pipe connecting it to a dedicated pump; this ensures that there is no preferential scavenging. Each of the scavenge pumps will be of the same design as the pressure pump and will have the same capacity, which for a bearing chamber is probably excessive. Usually, the scavenge pump will be connected to the area of the chamber where the oil needs recovering; this is usually at bottom dead center. All the scavenge pumps can be built into a single pump assembly with the pressure pump; there is then a single pipe that connects all the scavenge pumps to the oil tank. Built into the single return pipe and downstream of the pump assembly will be some of the maintenance and engine health monitoring items; these include the scavenge filter and the master magnetic chip detector (MCD). These two items have to be positioned correctly because both rely on collecting debris from the oil that has been generated by the engine. So the MCD will be installed upstream of the scavenge filter and pick up any ferrous debris before it becomes lodged in the filter. From the scavenge filter assembly, the oil enters the oil tank, where it passes into a de-aerator that separates the remaining air that has been collected by the scavenge system and releases it into the oil tank.
6.3.4 Oil Pumps Generally, gas turbine engines will utilize a series of pumps; one or more are used to supply a pressure feed and a number for oil scavenge to recover the oil from different areas of the engine and return it to the oil tank. Generally, the same type of oil pump will be used for both pressure feed and scavenge systems; these can be mounted together in a single pump assembly. The pumps used on the gas turbine and in particular the feed pump are positive displacement-type pumps driven via a quill shaft from the accessory gearbox (Figure 6.6). It is a misconception that the oil pump provides the oil pressure in the oil system; the positive displacement pump supplies a constant volume of oil for a given rotational speed into the feed system, and it is the resistance of the system downstream of the feed pump that creates the oil pressure. Various configurations of oil pumps can be found installed on aero engines; the most common in service today are the vane pump, the gear pump, and the gerotor pump.
© 2006 by Taylor & Francis Group, LLC
Aircraft Gas Turbines (a)
6-9 Trapped volume (oil in feed pump or air/oil in scavenge pump)
High pressure outlet
Low pressure inlet Gear pump (c) Trapped volume
(b)
Gerotor pump
Vane pump
FIGURE 6.6 Positive displacement oil pumps.
The vane pump is constructed of an eccentric rotor that has sliding vanes installed in it. As the rotor turns, the vanes make contact with the casing. However, because of the eccentricity of the rotor to the casing, the vanes then create a variable swept volume which “pumps” the oil. These pumps are generally lightweight, but they can be susceptible to ingestion of debris from upstream, which can cause the vanes to fail. The gear pump operates by moving the volume of oil between the gear teeth and the casing. These pumps can be assembled into a gear train that has the benefit of allowing the introduction of idler gears to seal the inlet from the outlet; it also has the benefit of increasing the oil flow capacity of the pump. Compared with the vane pump, the gear pump is, by design, more robust and more resistant to failure due to debris ingression. The gerotor pump is the latest type of pump to be used within a fluid system on a gas turbine. The main difference with this type of pump is that it has two rotors, an inner and an outer element. Either element can be driven. The gerotor name is derived from the phrase “generated rotor,” which defines the mathematical procedure that is used to determine the shape of the inner rotor relative to the outer rotor. As the pump elements rotate, one element drives the other, and because the outer element has one tooth more than the inner, one tooth volume of oil is swept on each rotation. Using a high number of teeth on both elements minimizes pressure pulsations within the oil flow. In order to prevent ingress of debris to whichever type of oil pump is used for pressure feed or scavenge, installing filter screens on the oil inlet can protect the pump.
6.3.5 Filtration The quality of oil circulated is important to the operation of the gas turbine, and one area that is paid particular attention is its cleanliness. The oil that flows around the engine acts as a flushing agent by
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(a)
Location or thrust bearing
(b)
Oil outlet
Oil jet Bypass valve
Oil outlet Inlet flow check valve Filter element
Thread type oil filter
Filter bowl
Retaining circlip
Pop out indicator Screw thread filter
FIGURE 6.7
Pressure filter assembly
Different types of oil filter assemblies.
removing unwanted foreign particles and debris that may be present in the engine’s oil system and effectively contaminating the oil. This contamination can be in various forms: it could be swarf from the engine’s last build, it could be carbon, or it could be metallic particles generated by the rotating assemblies in the engine. Whatever the contamination, these particles can cause additional problems within the oil system by blocking some of the small orifices, in particular the oil jets that direct the oil onto engine components such as bearings. To remove any contamination from the oil system, the oil is passed through different filtration systems (Figure 6.7). 6.3.5.1 Scavenge Filter The scavenge filter is fitted in the oil system between the scavenge pumps and the oil tank. The scavenge filter is a very fine filter, usually approximately 30 µm, because it has to ensure that the oil returning to the tank is clean. However, it is likely that contamination generated by the engine will find its way into the scavenge filter; therefore, this is most likely to block first. Because of the fineness and task that the scavenge filter has to do, once it has been removed from the engine, it cannot be reused. On removal from the engine, the used filter will be cleaned so that any contamination collected can be examined and a new filter will be installed. As there is a greater potential for this filter to block than the others on the engine, the filter assembly can be fitted with a bypass valve that allows contaminated oil to miss the filter and go direct to the oil tank, thus preserving the supply of oil. 6.3.5.2 Pressure Filter Some engines can be fitted with a filter in the pressure feed supply, usually downstream of the pressure pump assembly. This filter is generally a coarse filter, for example, 125 µm, and it is located in this position to protect the engine from a failure of the pressure pump; it will also filter out any contamination in the oil that may have come from the oil tank. The pressure filter element can be removed, cleaned, and refitted to the engine provided it cleans successfully. With some of today’s civil engines, each time the pressure filter is removed and cleaned the filter is etched with the number of times it has been cleaned. Once a predefined limit has been reached, for instance four uses, the filter will be removed, cleaned, and replaced with a new element.
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6.3.5.3 Oil Tank Screen If the scavenge filter on an engine has blocked and any bypass valve fitted allows contaminated oil into the oil tank, this contamination will pass into the pressure feed system, having gone through the pressure pump first. Any large particles of debris, such as parts of bearings or gear teeth, could damage the pressure pump and prevent the oil system from functioning. Therefore, to prevent any large debris from reentering the oil system from the oil tank, the tank can be fitted with a filter screen around the entry to the feed pipe. 6.3.5.4 Screw Thread Filters There is still the potential for debris at build to enter the oil system between the pressure filter and the internal areas of the engine. Screw thread filters can be assembled into the supply lines that feed the oil jets. These screw thread filters also become known as “last chance” filters because this is the last chance to remove coarse contamination before the jets. This type of filter is usually identified as a strainer because of the fact that it is built into the engine and can only be cleaned when the engine is overhauled. Screw thread filters work by allowing oil to flow into the grooves before being forced through the screw threads, which strain out any coarse contamination. Generally there are three types of screw thread filters in use today, tapered groove, straight groove, and slotted.
6.3.6 Materials Because of the arduous conditions that occur in all areas of an aero engine, the construction of the engine has become ever more reliant on specialized materials, anything from titanium and steel to plastics and composites. The oil as it passes around the various parts of the engine will have to come into contact with different material types and in all cases, contact between them and the oil must remain inert. The types of materials that can be found inside a gas turbine’s bearing chambers and gearboxes are either corrosion or creep resistant steels and noncorrosion resistant steels. Therefore, in the case of the latter material, corrosion additives within the oil will help to maintain the internal components of the engine in good condition.
6.4 Airflow 6.4.1 System Pressurization and Sealing There are various types of fixtures that are utilized for sealing rotating assemblies around the engine in both the oil and air systems; these are used to control air pressures internal to the engine and to control the areas in the engine where oil is allowed to go (Figure 6.8). Examples of the kind of seals that can be used on a gas turbine are described here. 6.4.1.1 Carbon Seals The carbon seal tends to rely on close contact to maintain the sealing effect, so the leakage flow across seals will be very small. The contact with the rotating assembly is usually maintained because the carbon seal is spring loaded to allow it to follow the rotating part. However, this type of seal is prone to wear and thus, when the seal does start to wear, the leakage flow rate increases significantly and the seal’s effectiveness reduces because it is no longer able to follow the rotating assembly. Because a good carbon seal is fitted with zero clearance it will provide a seal even if there is a reverse pressure difference across it. However, one main disadvantage of the carbon seal is its replacement cost: it is significantly more expensive as a unit than most other types of seal. 6.4.1.2 Brush Seals The brush seal has had its problem in the past, but development of the unit and the materials it is constructed of has now made it a very effective sealing arrangement. The main problem with firstgeneration brush seals was that the bristles were made of carbon; they were installed in a ring and were
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Handbook of Lubrication and Tribology (b)
Static part
(c)
Honeycomb section liner Static part
Static part
Wearaway liner
Part that turns Carbon seal
Part that turns
Carbon seal (airseal)
Stepped seal (airseal)
(d)
Grooved seal (airseal) (e)
Static part Grooved labyrinth seal
Static part
Part that turns
IP compressor shaft
Honeycomb section liner Oil filled hydraulic seal Oil seals
Double stepped seal (airseal)
FIGURE 6.8 Aero engine — air and oil seals.
angled in the same direction as the normal rotation for the assembly they were brushing against. However, this caused the problem that in order to protect the seal from damage as a result of broken bristles, the engine was supposed to be allowed to rotate in one direction only. Unfortunately, aero engines have a tendency to windmill in either direction when standing idle due to the prevailing wind, and this caused all the bristles in the brush seal to break in half if the engine rotated in the reverse direction. With broken bristles, a brush seal is virtually ineffective; thus to restore the effectiveness of the seal would require a rework of the engine. The main change that enabled a brush seal to be a more reliable method of sealing internal engine compartments was when the bristles were changed from carbon to metal; this enabled the gas turbine shafts to rotate in either direction without damaging the bristles. The brush seal is a cost-effective method of providing sealing within a gas turbine and it has advantages over other methods. However, brush seals are prone to leak if a reverse pressure is applied across the brush. The brush can still wear; this increases leakage across the seal, making it less effective. Finally, brush seals are susceptible to contamination by oil, which also can affect their ability to function properly. 6.4.1.3 Labyrinth Seals The labyrinth seal is one of the cheapest and most common seals to be installed in a gas turbine. The use of a labyrinth seal tends to be a compromise between cost and performance. The seal rotates against an abradable lining which allows the seal to “cut” its own path in the lining when the engine first runs. The control that the labyrinth seal imposes on the internal parts of the engine relies on a positive pressure drop; the higher the pressure drop across the seal, the more effective the seal. However, because this seal is using air supplied by the engine’s compressors that will eventually be discharged overboard by the centrifugal
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breather, any increase of pressure drop across the seal will also mean an increase in airflow taken off the engine’s compressor. If more air is taken off the compressors to supply the seals, the effect on the engine will be an increase in fuel flow and temperatures within the turbine system. Other variations of the labyrinth seal are used on a gas turbine to control the flow of air around the internal structure; examples of these are stepped seals, double stepped seals, and air-blown or grooved seals. 6.4.1.4 Ring Seals Although ring seals are becoming less popular, many of today’s gas turbines still use them for sealing rotating assemblies. The ring seal has a ring which is allowed to float in its housing so that it can follow any orbiting of the seal runner and maintain a very tight seal. However, the design of these types of seals is one of the main disadvantages for certain uses. Operating under large pressure drops causes the ring seal to seize in position. Using them in the hot sections of the engine also has a tendency to cause the seals to seize owing to carbon buildup. Once a ring seal has seized into position, damage to any rotating assembly is inevitable. 6.4.1.5 Hydraulic Seals The hydraulic seal is another effective method of sealing the oil into a compartment within the gas turbine, but because of its design, it needs the components to be rotating to make an effective seal. The hydraulic seal relies on a trough to be filled with oil; the seal fins on the opposing component are dipped into the oil and thus create the seal. The pressure drop that can be maintained by this type of seal depends on the size of the trough and the quantity of oil within. Therefore, if effectiveness is dependent on the size of the trough, the size has an impact on the areas of the engine in which this seal can be used. Hydraulic seals are used on low-speed areas of the engine. The oil that flows into and out of the trough must also act as a coolant so that any heat generated by windage is dissipated. The temperature of operation for a hydraulic seal is always going to be limited by either the flash point or the auto-ignition temperature of the oil.
6.4.2 Vent System The vent system will be connected to bearing chambers and oil tanks so that air can be removed out of these areas. In the case of the engine’s bearing chambers, these, as we have seen above, use air from the compressors to pressurize and seal. The process allows air to bleed over the seals and into the bearing chamber, thus preventing oil coming out of the chamber. It is then necessary to remove this air; otherwise; the seals would fail to work effectively and the chamber could start to overheat, leading to conditions which could cause an oil fire. Therefore, the vent system will be connected to the centrifugal breather because the air that is removed from the bearing chamber is an air/oil mist and the oil needs to be recovered into the oil system. The vent system can also be connected to the oil tank because the returning scavenge oil supply will contain small amounts of air trapped in the oil. This is released in many oil tanks by passing through the de-aerator, which releases the air into the tank. The vent from the tank should also be connected to the centrifugal breather because it too will be an air/oil mist.
6.4.3 Air/Oil Separation As seen above, the two areas in the oil system that require air from the air/oil mist to be removed are the bearing chambers and the oil tank; however, there is usually only one unit that separates the air and oil, the centrifugal breather. The breather uses centrifugal force to separate the oil and air; the oil is recovered by scavenging it back into the return supply going to the oil tank through its own scavenge pump. The air within the breather passes into the impeller and is pumped via an exit pipe to atmosphere. Allowing the oil/air mist from the vent to be discarded overboard without recovering the oil would cause visible oil emissions and oil staining of the engine’s nacelle, as well as increasing the oil consumption of the engine.
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The important consideration when designing the breather system is how the air is exhausted. Some engines will vent the air from the breather via an exit in the wall of the nacelle. The main problem with this type of exit is that the air still has a very tiny amount of oil in it, which could be picked up by the turbulent boundary air passing over the nacelle and leave an ugly stain on the nacelle. To ensure that the air exiting the breather is exhausted into the free stream, a breather mast can be installed on the nacelle, and this moves the exhaust out of the nacelle’s boundary layer airflow. Other engine types take a different approach to venting the breather air overboard by allowing it to feed into the main exhaust of the engine. The outlet from the breather feeds down the center of the engine’s low-pressure shaft and out through an exhaust pipe that is located at the center of the turbine’s exhaust.
6.4.4 Oil Firing When designing bearing chambers on a gas turbine, it is important to understand at what internal conditions the engine is going to operate. Many engines run with a number of bearing chambers supplied with cooling and sealing air, which is designed to enter the bearing chamber as part of the sealing process. Other bearing chambers are simply sealed by some device or operation and no air enters these chambers. In those chambers that have to be ventilated, it is important to ensure that the oil and air conditions within the chamber remain balanced; any reduction in the removal of vent air because of a blocked vent pipe or a reduction in the quantity of oil scavenged from the chamber could increase the possibility of oil firing. Therefore, if the vent from a bearing chamber becomes blocked, the oil will become stressed owing to the increase in temperature and the anti-oxidant additives will start to protect the base stock from oxidizing. Any reaction by the antioxidant should be recognized by maintenance personnel as a change in color of the oil and possibly a rancid smell. There is also a chance that the reduction in vent air within the bearing chamber will cause an increase in Total Acid Number (TAN) for the oil owing to the increased oxidation. Oil firing will occur only when the conditions inside the bearing chamber cause the oil to exceed its auto-ignition point; fortunately, this will be the highest ignition point for the oil and usually well above the maximum temperature limit for engine operation. Neither the flash point nor the flame point should cause oil firing because very rarely is there any source of ignition present; also the area is generally too oil rich to allow oil firing to happen.
6.5 Bearing, Seal, and Gearbox Lubrication 6.5.1 Bearing Lubrication The bearings in an aero engine play an important part in the safe running of the shafts. There are three main types of bearings available; these are ball bearings, roller bearings, and plain bearings, of which the latter will rarely find a use in today’s gas turbines. However, both ball and roller bearings play an important role in the smooth running of the engine’s shafts and gears. It is important to ensure that all the bearings within the engine are supplied with the correct quantity of oil because the oil acts as a coolant as well as a lubricating agent. 6.5.1.1 Ball Bearings Ball or thrust bearings are used primarily as a location bearing and can be found in different types and sizes in the aero engine. Their job in the engine is to ensure that the shafts and components remain in the correct axial location. When these types of bearings are used for the main shafts on an aero engine, they are subjected to large axial loads as a result of the reaction forces of the compressor and turbine airfoils and the differential pressure loads on the faces of the compressor and turbine discs. Ball bearings can be used to offset gearbox axial loads where bevel gears are used to transmit loads. The two main types of ball bearings found in aero engines are deep groove ball bearings and complex or duplex bearings. Composite or duplex bearings are designed so that the contact stresses from
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the interaction of the balls against the race are reduced; however, in order to function at their best, running under conditions of zero thrust load should be avoided. 6.5.1.2 Roller Bearings Roller bearings transmit only radial loads and will allow any axial movement in the component they are attached to. Any thrust loads that act on the roller bearing have a reactive force on the cage against the end faces of the rollers. 6.5.1.3 Bearing Oil Supply It is important that the oil supply to the bearings is maintained at all times the engine is running and that it is accurately targeted at the contact surfaces. The most common ways of ensuring that the correct oil flow to the bearings is maintained are: • Jet feed: the oil is targeted at the gap between the inner and outer races. Oil is thrown outward onto the outer race and exits through a gap between the outer race and the cage. • Undershaft feed: oil feeds from holes within the shaft into rotating weirs created by wings on the cage. This oil lubricates all the inner race surfaces by centrifugal force, causing it to exit the bearing through a gap between the outer race and the cage. • Splash feed: this is the most common method of lubricating bearings within gearboxes. The oil is collected by rotating weirs in the assembly and allowed to “splash” on to the bearings. As well as lubricating and cooling bearings and gears within the engine, the oil can be used as a vibration damper. This arrangement is known as a squeeze film bearing (Figure 6.9). The oil that is supplied to the bearing can be redirected to a gap on the outside of the bearing, between the outer race and the bearing housing. This layer of oil will dampen out any induced vibrations. Oil feed
Squeeze film
Bearing outer race
FIGURE 6.9 Squeeze film bearing.
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6.5.2 Gear Lubrication All gas turbine engines will include several gearboxes within the mechanical arrangement. These can be used for providing drive to rotating accessories, power transfer to offtake shafts, and as reduction gearboxes to drive propellers. By far the best way to provide lubrication to gears is by use of jets directed at the ingoing mesh. The oil will provide both lubrication to the gears and act as a coolant to remove any waste heat. The heat that is generated by gears is usually a loading loss, which is in the order of 1% of the power transmitted.
6.6 Lubrication Selection 6.6.1 Oil Grades Generally the oil that is required for an aero gas turbine will be one of the following: 3, 5, or 7.5 centiStokes (cSt). Each has an application for which it is better suited. The 7.5 cSt oil was developed for use in turboprop engines as the increase in viscosity gave the oil better load bearing properties, and this allowed it to be used within the reduction gearbox and torque control for the propeller. Any oil that is to be used in an aero engine normally has to be approved for that engine type before it can be used in service, and one of the relevant manuals for the engine type will contain a list of the approved oils. As well as getting approval from the engine manufacturer, the oil will have to meet internationally recognized specifications, as follows: • 3 cSt oil will be to MIL-PRF-7808 or DEF STAN 91–94, which is equivalent to SAE-5. • 5 cSt oil will be to MIL-PRF-23699, DEF STAN 91–101, and DERD 2497/2499, which is equivalent to SAE 5W10. • 7.5 cSt oil will be to DEF-STAN 91–98 and DERD 2487, which is equivalent to SAE 5W20. The most common oil used in service on today’s gas turbines is the 5 cSt oil since the properties of this oil allow it to cope with the demands of the engine’s operation. Normally aero engines will be certified to operate from ambient temperatures of −40 to above 40◦ C and, as part of this, the engine will be required to start at these ambient temperatures, even if the engine and all the systems including the oil have been soaked to the same temperature. Obviously, at very cold temperatures the oil will be very viscous, anywhere from 8,000 to 11,000 cSt; therefore the engine must be able to cope with this kind of viscosity level and start as normal when required. Pressure pumps can have pressure relief valves fitted to them so that the cold, viscous oil can be recirculated around the pump to speed up heating and prevent excessive oil pressure occurring in other parts of the engine. Normally in these circumstances, the person running the engine from the flight deck will be required to ensure that the oil temperature is above zero (usually 5◦ C) before the engine is accelerated above the idle power level. The oil for the gas turbine engine will also need to have other properties to suit the environment it is running in, some of which are dependent on its additives. The oil needs to have low volatility; carbon is likely to form only at higher temperatures than for normal running and only if the oil becomes stressed. The flash point, flame point, and auto-ignition point should also be at a high temperature that can cope with the extremes of normal running and above, and still have sufficient margin available to aid aircraft safety. Gas turbine engines are constructed from many different materials and it is therefore important that the oil used, should not cause any safe issues due to possible reactions, corrosion or failure of components within the oil system. In particular, it must have good elastomer compatibility to prevent damage to any rubber seals with which it comes into contact with. Today’s high bypass gas turbine engines will be expected to remain installed on an aircraft for at least 20,000+ hours or approximately 5 years, during which time the oil system will be serviced, mainly be replenishment. Good practice in servicing an aero engine is for an airline to operate their fleet on the same brand of oil, however, this may not always be the case. If an airline’s gas turbines need the oil system to be replenished at an airport away from their main base, their usually brand of oil may not be available. In this case, most gas turbine manufacturers
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will allow the mixing of oil brands provided that the alternative used is on the engine manufacturer’s approved list. When mixing of approved oil brands does occur, then the combined fluid must not affect the performance of the oil system, neither should the oil brands react with each other.
6.7 Oil System Indications 6.7.1 Engine Measurements It is important to inform the flight crew of an aircraft about the operating condition of the oil system during any flight. However, on some aircraft types, this information will not be normally displayed on the flight deck unless there is a problem or it is selected by one of the pilots. This is because many of today’s commercial aircraft have a two-man flight deck and it is important to give them enough information to make an accurate assessment of the situation without causing a large increase in workload. Therefore, for some information, the aircrew do not need to know in detail whether the engine or aircraft systems are operating normally. The main oil parameters that are measured and can be displayed on the flight deck are as follows: • • • •
Oil pressure indication Oil temperature Oil quantity Oil filter blockage
Oil pressure indication on the flight deck can be either an absolute or delta pressure indication. Low oil pressure can be an indication of developing problems within the oil system and can give advance warning to the flight crew as to what action to take; if it persists, it may be necessary to shut the engine down. The pressure indication will have a minimum value that can appear in red on some displays; the indication will pulse if the low limit is reached. Oil temperature will have an upper limit, which is indicated on the flight deck gauge; flight crews will become familiar with the normal indication of oil temperature for each type of engine they operate and therefore become less interested in absolute values. This is important as in some cases the oil temperature can be measured in different areas of the oil system; for example, measurement in the scavenge oil returning to the oil tank reflects the heat transferred from the engine components. Oil quantity is measured by two methods (1) an electronic measurement can be made in the oil tank that is indicated on the flight deck and (2) there is a graduated sight glass on the side of the oil tank for maintenance engineers when they are required to service the engine. Differential pressure switches that read both inlet and outlet pressures are used to indicate oil filter blockage; the indication can apply to one or more filters. An amber “Clog” indication can be given on the flight deck to show an impending problem in the oil supply. If there are two filters that can give an indication of blockage on the flight deck, then the flight crew does not need to know which one is starting to block, as either filter blocking will result in a reduction of oil flow. Maintenance personnel will be able to interrogate the aircraft’s maintenance system on landing for any fault messages to identify which of the two filters has initiated the “clog” message indication.
6.7.2 Aircraft Indications The display of information that is available to the flight deck on a civil passenger aircraft has changed with the advent of the two-man cockpit (Figure 6.10). Previous aircraft such as the Lockheed TriStar would have two pilots plus a flight engineer whose responsibility it was to manage the aircraft and engine systems. The flight engineer’s panel was usually located behind the copilot and was a complicated array of analog dials and switches. The benefit of having this extra person on the flight deck was that all of the systems on both the aircraft and the engines could be monitored continuously. Modern aircraft utilize more and more computers for control and operation, including FADEC (Full Authority Digital Engine Control) on aero engines, which has changed the environment on the
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Handbook of Lubrication and Tribology
Secondary engine parameter page (Airbus A330 lower ECAM screen)
Cruise engine parameter page (Airbus A330 lower ECAM page)
FIGURE 6.10 Aircraft indications.
flight deck. Now the systems on both the aircraft and the engine can be monitored by computers and information can be made available to the flight crew. Thus the role of the flight engineer disappeared has for all but a few older aircraft. The pilots are now required to fly the aircraft and monitor both aircraft and engine systems, so there has to be control of the workload on the flight deck. Thus, pilots tend to be given minimum information about the status of the aircraft and engine systems. If there is no problem, why tell them? The aircrew only need to be made aware of events that could eventually lead to a major problem or minor problems that could escalate and directly affect the safety of the aircraft. However, most flight deck indication systems allow pilots to “scan” through each of the system indication screens to “give them a warm feeling” that everything is still operating normally. For a military pilot, the cockpit has followed a similar philosophy to that of the civil airliner, but for a slightly different reason. Military pilots have the added problem that they may be called upon to enter a combat zone, where their workload outside the aircraft will be very high, and they will need to spend as much time looking outside, not down at their instruments. Again, as military aircraft become more reliant on computers, they can and will assist the pilot by taking charge of monitoring the aircraft and engine systems.
6.8 System Maintenance 6.8.1 Maintenance Manual Procedures The maintenance and servicing of the oil system for an installed aero gas turbine, have to be carried out in accordance with procedures in the approved manuals. The main manual for an installed engine on a civil
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Aircraft Gas Turbines
6-19
aircraft is the aircraft maintenance manual (AMM) and the procedures required are located in chapters and sections that conform to the ATA100 index (Air Transport Association). Chapter 12 is the part in the AMM that details the procedures for servicing of the whole aircraft, and Chapter 79 details procedures that relate directly to the oil system on the engine and its components. However, the ATA100 index is far more complex than this; each procedure is identified by a system of chapter/section/subject and numbers and the number of digits increases as the task becomes more specialized. The first element is the chapter, and this is the major system or topic; the second element is the section, which is the first major breakdown of the system; the third element is the subject and is a further breakdown, ultimately leading to the piece-part level. For example, Chapter 12 29 70 71 72 79
Servicing Hydraulic system Standard practices (separate manual) Power plant Engine Oil system
These chapter numbers are allocated by the ATA and are mandatory. Section 79-0X Oil — general 79-1X Storage 79-2X Distribution 79-3X Indicating The first section number digit is also mandatory. The second section number digit allows the section to be broken down into specific items or components, for example, 79-31 79-32 79-33 79-34
Oil quantity indicating Oil temperature indicating Oil pressure indicating Low oil pressure warning system
6.8.2 Maintenance and Servicing Servicing of the oil system for an installed aero gas turbine can be found in the AMM, usually in Section 12-13-79. This will detail the procedure for checking the oil level and filling the oil tank where necessary. The procedure will also detail the following additional information: • Minimum time after engine shutdown before checking the oil level (usually 10 min) • Maximum time after engine shutdown when the oil level can be checked without having to run the engine to idle power (usually 6 h) • What to do if the oil level is below a minimum quantity Maintenance of the oil system appears in Chapter 79 and subsequent sections. This chapter covers fault diagnostics if there is a problem, and replacement and inspection of the units within the oil system. The inspection of units relates to the removable filters and the MCDs, how they are inspected, and the time interval at which they must be inspected. For example, the routine inspection of a pressure filter is that it is removed, cleaned, and reinstalled at 1600 flight hour intervals (ATA Chapter 79-22-43); the scavenge filter is removed, cleaned and discarded and a replacement fitted every 1600 flight hours (ATA Chapter 79-22-45); and the MCDs are removed, inspected for ferrous contamination, and refitted with a new seal ring every 130 flight hours (ATA Chapter 79-20-00).
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Handbook of Lubrication and Tribology
Obviously, if there is a potential problem on the engine that may need monitoring, the pressure filter, the scavenge filter, and the MCDs can be inspected at much shorter time intervals. The originating equipment manufacturer (OEM) for the engine can also change the inspection period if it believes the scheduled interval is too long, by introducing a revised interval by service bulletin.
6.9 Engine Health Monitoring 6.9.1 Oil System Analysis Techniques — Magnetic Chip Detectors (MCDs) MCDs are among the simplest devices for monitoring problems associated with components within the oil system (Figure 6.11). As suggested, these devices are simply probes that contain a magnet at thier tip; this tip is then immersed into the oil flow. The location of MCDs in the oil system is within the scavenge return pipework, between the scavenge pump and the scavenge filter, so that they pick up any magnetic particles from the oil before it passes through the scavenge filter. Generally, when they are in this location, they are usually identified as the master MCD. MCDs are available in one of three different types: screw-in, bayonet fixing, and screw-in electric MCDs. The majority of today’s civil aero engines are usually fitted with either screw-in or electric MCDs. The bayonet-type fixture sees less use today as it is the least secure; wear on the pins that locate the MCD in the locked position can allow the MCD to move and create an oil leakage path overboard. Some large bypass turbofan engines can also have additional MCDs fitted; these are usually located in individual scavenge pipes that are taking oil from the engine to the scavenge pump assembly; they should be of the same type (i.e., screw-in or bayonet, but not electric MCD) as the master MCD. Inspection of the MCDs on multiengine aircraft needs to be carefully controlled because of previous problems caused by human errors. All MCDs are fitted to the engine with rubber “O” ring seals, which have to be replaced on every inspection. There have been several instances where all the engines on an aircraft have been inspected at the same time and all MCDs on all the engines have been refitted without their
(a)
Metallic fines
Metallic flakes
Build debris or “SWARF”
Typical magnetic chip detector contamination Seal ring
(b)
Magnetic probe
Seal ring Detector housing Washers 2 off Bolts 2 off Magnetic chip detector
FIGURE 6.11
Oil system condition monitoring.
© 2006 by Taylor & Francis Group, LLC
Metallic chips and gear teeth fragments
Aircraft Gas Turbines
6-21
“O” ring seals; the consequence is a very rapid oil loss on the next flight and generally an emergency landing. This can be particularly detrimental to an airline that operates aircraft on ETOPS (extended twin-engine operations), where problems caused by human factors, increased in-flight shutdowns and emergencies can lead to their local airworthiness authority removing their ETOPS approval and restricting the routes that the aircraft can fly. Therefore, to reduce any problems on multiengine aircraft and especially those flying ETOPS routes, the scheduled inspection of MCDs is usually staggered between engines, so that only one engine is inspected at any one time. However, where this is not possible, then other “best practices” can be put in place, for example, • The removal and installation of the MCDs is carried out by one person; the inspection and sign-off is carried out by independent people. • The removal and installation of the MCDs is carried out by a different crew on different engines. • After installation of the MCDs, a ground-run leak test is carried out to ensure that installation of the MCDs is correct. Regular MCD inspections can provide maintenance with the following benefits: • • • •
Detecting failure early Allowing an airline to schedule an engine removal, reducing airline costs Reducing secondary damage, leading to a reduction in shop visit costs and turn-around times Reducing of in-flight shutdowns and flight delays
Bibliography [1] Early detection of transmission failures through oil system debris assessment, Derby, Rolls-Royce plc, 2001 [2] United Kingdom. Ministry of Defence. Defence Standards. Defence Standard 91-101. Lubriucating oil, gas turbine engine, synthetic grade 5cSt, NATO code 0-156, joint services designation OX-27 and OX-28, issue 3. Published 31 August 2001. 27 July 2004
[3] Mobil Jet Oil 254. 3 June 2003. Exxon/Mobil. April 2004. http://www.exxonmobil.com/ukenglish/aviation/PDS/GLXXENAVIEMMobil_Jet_Oil_254.asp [4] Maintenance Planning Document (MPD) for the Airbus A330/Rolls-Royce Trent 700 Engines. Toulouse France. Airbus Industries, April 2004 [5] ADRES for the Airbus A330/Rolls-Royce Trent 700 Engines. Toulouse France. Airbus Industries. October 2003 [6] AirN@v for the Airbus A330/Rolls-Royce Trent 700 Engines. Toulouse France, Airbus Industries, June 2004 [7] Trent 700/A330 Operating Instructions. Derby. Rolls-Royce plc. June 2002 [8] Trent 700 Engine Manual. Derby. Rolls-Royce plc. June 2004 [9] Rolls-Royce Trent 700 Line and Base Maintenance Course Notes. Derby. Rolls-Royce. January 2004. [10] Rolls-Royce Trent 500 Line and Base Maintenance course notes. Derby. Rolls-Royce plc. May 2004 [11] Shaw, Kevin and Turner, Guy. Rolls-Royce Trent 700 UnderCowl Pocket Guide (Misc 30046). Derby. Rolls-Royce plc. May 2001 [12] Smith, Mark. Rolls-Royce Trent 500 Undercowl Pocket guide. Derby. Rolls-Royce plc. December 2004 [13] The Jet Engine. 5th Edition. Derby, Rolls-Royce plc. 1996
© 2006 by Taylor & Francis Group, LLC
7 Principles of Gas Turbine Bearing Lubrication and Design Nomenclature . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.2 Lubricants and Materials . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
7-1 7-4 7-5
Lubricants • Fluid Film Bearing Materials
7.3
General Bearing Principles . . . . . . . . . . . . . . . . . . . . . . . . . . . .
7-11
Hydrodynamic Bearings • Hydrostatic Bearings • Compliant Surface Bearings • Rolling Element Bearings • Magnetic Bearings
7.4
Conventional Fluid Film Bearings . . . . . . . . . . . . . . . . . . . .
7-24
Bearing Parameters • Journal Bearing Performance • Thrust Bearings
7.5
Low-Speed Bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
7-46
Regimes of Operation
7.6
High-Speed and High-Temperature Oil-Free Bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
7-50
Gas Bearings • Complaint Surface Foil Bearings (CFB) • Magnetic Bearings
7.7
Design Considerations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
7-84
Performance Parameters • Bearing Configuration • Qualitative Guidelines
Hooshang Heshmat and James F. Walton II Mohawk Innovative Technology, Inc.
7.8
Advanced Bearing and Seal Applications. . . . . . . . . . . . .
7-100
Rotor-Bearing Dynamics and Engine Integration
References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Additional References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
7-116 7-118
Nomenclature Unless noted otherwise, the following symbols are used throughout the text. A Area AMB Active magnetic bearing B Damping coefficient 7-1
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7-2
Handbook of Lubrication and Tribology
B C Cm CM CFB D E Fτ G G Gx Gz Gτ H K K KB L
MCR M CR N P
Q Qz Qz , Qτ Q1 , Qin Q2 Q2 P R R1 R2 R Re
REB S T T U W W WCR W CR b c
© 2006 by Taylor & Francis Group, LLC
ωB/2µ ND(L/D)(R/C)3 = π L(R/C)3 Radial clearance Smallest clearance for = 0 Largest clearance for = 0 Compliant surface foil bearing Diameter of bearing for journal Elastic modulus Frictional force Mass flow rate G/1/2ρa ULC Turbulence coefficient in x or θ Turbulence coefficient in z or r direction Turbulence coefficient for viscous shear Power loss Spring coefficient K /[2µ NL(R/C)3 Structural stiffness O.D. CFB Width of bearing - in the z direction in journal bearings - in the r direction in thrust bearings Critical mass for journal bearing stability (MCR N )/2µL(R/C)3 Revolutions per unit time Unit load - (W /LD) in journal bearing - (W /A) in thrust bearings Volumetric lubricant flow (Qz /Qz F ); Level of starvation Side leakage of lubricant (hydrodynamic) Flow in at leading edge Flow out at trailing edge Flow out at trailing edge due to pressure gradient Radius of bearing or journal Inner radius Outer radius Perfect gas constant Reynolds number - (ρRωh/µ) in journal bearings - (ρrωh/µ) in journal bearings Rolling element bearings Sommerfeld Number (LDµN /W ) (R/C)2 = 1/W Temperature Temperature rise (T − Ti ) Linear velocity Load - (W /LDµN )(C/R)2 in journal bearings - (W /R22 µω)h 2 L2 ) in thrust bearings Critical load for journal bearing stability (WCR /W )Cω2 /g ) Tapered fraction of thrust bearing Specific heat of lubricant
Principles of Gas Turbine Bearing Lubrication and Design
d e f g h hmin h11 hp h2 h
hN h∗ ∗ hN h1∗ h2∗ Io m n p pa ps r r s t w x, y z α β βp βs βp δ δ0 δτ m θ θS θp θE θ2 θmin µ ν ρ φ
Amount of center offset or preload Eccentricity Friction coefficient (F τ/W ) Gravitational constant Film thickness Minimum film thickness Value of (h1 − h2 ) at R1, 0 Film thickness over pivot Film thickness over pivot Dimensionless film thickness (h/C) for journal bearings (h/h2 ) for thrust bearings Normal film thickness (h/δ) (Hn/δ) (h1 /δ) (h2 /δ) Half length of bump in angular direction Preload, (d/C) Number of pads in bearing Pressure Ambient pressure Supply pressure Radial coordinate (r/R2 ) Distance between bumps (pitch) Time, foil thickness Specific weight Rectangular coordinates Axial coordinate 2pa s l0 3 2pa s l0 3 Bearing compliance; (l − v 2 ); (l − v 2 ) CE t h2 E t Angular extent of bearing pad Angular extent of pad from start to finish Fluid film arc, (θ2 − θ1 ) (βp /β) Taper Circumferential taper (h1 − h2 ) Radial taper at θ = 0, (h11 − h12 ) Eccentricity ratio (e/C) (e/Cm ) Angular coordinate Start of bearing pad Location of pivot End of bearing pad End of hydrodynamic film Location of hmin Lubricant viscosity Poisson’s Ratio Density of lubricant Attitude angle, (θmin − π )
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7-3
7-4
Handbook of Lubrication and Tribology
φL ω ωi ωn ωi Subscripts a E F R S xx xy yx yy τ 1, 2, 3
Load angle Angular velocity Threshold instability frequency Natural frequency of system (ωi /ω) 12π(µN /pa )(R/C)2 Ambient End Full film Ring Start Force in the x direction due to a displacement in the x direction Force in the x direction due to a displacement in the y direction Force in the y direction due to a displacement in the x direction Force in the y direction due to a displacement in the y direction Friction quad Lobe 1, 2 or 3
7.1 Introduction Gas turbines are used in a wide array of stationary, marine, and aerospace applications. Land, sea, and air-based generating systems produce anywhere from a few tens of kilowatts to multi-megawatts of power. Drives for natural gas pipeline compressors and land and marine propulsion systems produce thousands of shaft horsepower. In aircraft propulsion systems they produce a few pounds to thousands of pounds of thrust or shaft horsepower. The simplicity of the gas turbine, its efficiency, and ability to use many different fuels makes it attractive for these applications. The gas turbine operates on the principle of the Brayton cycle where compressed air is mixed with fuel in a combustor, burned under constant pressure conditions, and then expanded through a turbine to perform work. Due to more stringent environmental legislation, the viability of alternative forms of power generation, and demands for increased power density, improvements in gas turbine efficiency, emissions, and life cycle costs are necessary and are being implemented. Since these gains cannot be allowed to sacrifice operability or reliability, modern gas turbines require bearings and lubricants that can handle extreme speed, temperature, and other stress without breaking down. It has been apparent for some time and especially since the mid-1980s when the Department of Defense, NASA, and industry established the Integrated High Performance Turbine Engine Technology (IHPTET) that tribological limitations associated with the bearings, seals, and lubricants represent major obstacles to achieving significantly improved performance and life in advanced gas turbines. For example, rolling element bearings (REB) used in modern gas turbine engines may operate at DN values as high as 2.5 to 3.0 million (where D is bearing bore diameter in mm and N is shaft speed in rpm). One of the outcomes of the Tribological Limitations in Gas Turbine Engines: A Workshop to identify the Challenges and Set Future Directions [1] was projections that indicated that operating temperatures will continue to rise and bearing DN values above 4.0 million will be required in the coming decades. Present bearings even those using ceramic rolling elements materials and lubricants will be inadequate for the task at hand and mere advances to existing practices will not be sufficient to meet the life and performance improvement goals. Rather, entirely new lubricants, bearing materials, and bearing designs will need to be conceived, developed, and tested. Areas already under development include: • High-temperature liquid lubricants such as polyphenylethers and perfluoralkylethers • Hybrid ceramic–steel materials • Solid film lubricant coatings
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-5
• Advanced alternative lubrication systems such as vapor phase or powder lubricants • Magnetic and auxiliary bearings • High-performance gas bearings capable of operating at temperatures up to 1500◦ F either alone or in combination with magnetic bearings Industry and government identified tribological problem areas for existing systems included: • • • • • • • •
Corrosion of bearing materials due to oil breakdown Lubricant limitations in life and temperature capability Contamination of the lubricant and the bearings Breakdown in elastomeric seals and components Reliability and service life of the bearing, lubricant, and lubrication systems Lubricant bulk and spot temperature limitations Seal wear High thrust loads and the impact on bearing and lubricant life
7.2 Lubricants and Materials This section provides a fundamental discourse on state-of-the-art bearing lubricants and materials, providing the basic data required for use in the equations and tables given in subsequent sections. When choosing a liquid lubricant it is important to determine the operating conditions since different lubricants and classes of lubricants will respond differently to their environment. Table 7.1 identifies the important lubricant properties for different application environments.
7.2.1 Lubricants The role of a lubricant is to accomplish one or more of the following actions (a) separate bearing and rotor surfaces in order to prevent material damage; (b) remove heat generated in the loaded region; (c) offer low shear resistance; and (d) wet the surfaces during stops and starts when the interface or loaded region is minimally lubricated. Depending upon the film thickness between the two surfaces the lubrication regime can be identified as shown in the extended Stribeck curve (see Figure 7.1), which includes unlubricated, liquid, and powder lubricated conditions with dry, boundary, mixed, hydrodynamic, and limiting shear stress regimes [2]. In hydrodynamic fluid film bearings the most important property in fulfilling the above functions is lubricant viscosity. The level of generated hydrodynamic pressures, and hence the load capacity, as well as performance characteristics including temperature, flow, power loss, etc., all depend very strongly on TABLE 7.1
Key Lubricant Properties and Application Environment
Property Bulk modulus Corrosion resistance Flash point Foaming Gas solubility Heat capacity Oxidative stability Pour point Pressure viscosity Specific heat Thermal conductivity Viscosity Vapor pressure
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Application environment High pressure Metals High temperature Gears, dampers, and circulation system Boundary lubrication and cavitation Heat transfer Oxidizing atmosphere (high temperature air/gas) Cold starting Elastohydrodynamic lubrication High temperature Heat transfer Hydrodynamic lubrication Vacuum or high temperature
7-6
Handbook of Lubrication and Tribology (a)
(b) W
(d) W
U
(e) W
U
W U
U
Lubricant
Friction
1
0 (1)
10–1
(10–1)
(e) Limiting shear stress
10–2
1000
(c) Mixed Dry
Friction
Normalized wear
Film thickness/surface roughness
100
powder
(a) Dry friction
(d) Hydrodynamic lubrication
(b) Boundary lubrication
10
10–3
1
l= Film thickness/surface roughness
Wear
(10–3) ∞ Speed parameter, m0 (R/C)2 (U/W)
FIGURE 7.1
05-0017
Extended Stribeck curve. TABLE 7.2 Viscosity Conversion Factors Multiply Stokes (cm2 /sec) Poises Centistokes Centipoises Centipoises Centipoises Reyns (lb force sec/m2 ) Centipoises
By
To obtain
Density (g/cm3 ) 100 Density (g/cm3 ) 1.45 × 10−7 2.42 × 10−9 5.6 × 10−5 6.895 × 103 10−3
Poises (gm/cm sec) Centipoises Centipoises Reynes (lb force sec/in.2 ) (lb force min/in.2 ) (lb mass in. sec) Pascal sec (N sec/m2 ) Pascal sec (N sec/m2 )
lubricant viscosity. Since viscosities are listed in a great variety of units, Table 7.2 is provided to facilitate conversion from one system to any other. The oils most commonly used in turbines and gear sets are shown in Figure 7.2 along with other common SAE grade oils. Some of the more pertinent characteristics of these oils are as follows: Thermal conductivity Specific heat Heat of vaporization Specific gravity Flash point Pour point
0.08 BTU/(h ft2 ◦ F)/ft (0.14 J/(sec m ◦ C)] 0.4 BTU/(lb◦ F), [2.52 K J/Kg ◦ C)] 80 BTU/lb, (187 K J/Kg) 0.88 at 60◦ F, (16◦ C) 410–470◦ F (210–243◦ C) −10◦ F (−23◦ C)
If the operating temperatures exceed 300◦ F or fall below −10◦ F, a shift to synthetic lubricants may be required. Candidate liquid lubricants for high-temperature applications are listed in Table 7.3. To increase
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Principles of Gas Turbine Bearing Lubrication and Design
7-7
100,000 50,000 20,000 10,000 5,000 3,000 2,000
Kinematic viscosity (cSt)
1,000
SA E5 0 E 2 SAE 0 W 30 SA E1 0W SA
500 300 200 150 100 75 50 40 30
SA
E4
0
He
avy
ste
am
cyl
ind
Gr ad
e1
20
01
15
Lig
0j
et
ht
en
oil
Me
diu
m
sp
ind
le
gin
eo
10 9,0 8,0 7,0 6,0
er
il
oil
Lig
ht
tur
bin
tur
eo
il
bin
ea
nd
ele
ctr
ic
mo
tor
5,0
oil
4,0 3,0 –20
0
20
40
50
80
100
120
140
160
180
200
220
240
260
280
300
Temperature (°F)
FIGURE 7.2 Viscosity of petroleum oils.
the resistance and longevity of petroleum oils during prolonged usage, additives are often desirable. A list of oil additives for particular contingencies is given in Table 7.4. As oil is cycled through normal operating conditions varnish, sludge, and carbon deposits may form. To ensure adequate life of lubricants it is essential that the severity of the operating environment be understood from the bulk and hot-spot temperatures, expected thermal cycling, oil tank capacity, possible oil consumption, and whether the oil is drained and replaced periodically or only topped off as needed as is done with aircraft gas turbine engines. For example, in an aircraft gas turbine engine, lubricant severity is high because efforts are made to reduce oil consumption and the number of thermal cycles is high. Oil coking degradation can also occur when oil temperatures exceed approximately 430◦ F. “Coke” is the solid residue that remains when oils are subjected to extreme temperatures and undergo oxidative and thermal breakdown. The higher the temperature, the harder and more brittle the coke residue becomes. Positive washing of hot surfaces with high liquid flow rates will help reduce residence time and hence oil temperature rise. Flow restrictions caused by shedding of coke deposits or hot shut downs can cause blockages of drain lines which reduce flow rates, increase oil residence times, and exacerbate oil coking and degradation. Even without coking, oil will breakdown over time that can result in diminished corrosion protection. If bulk oil temperatures are expected to exceed approximately 450◦ F, alternative lubrication methods must be considered since this temperature exceeds the capability of ester-based lubricants. Alternative approaches may include the use of hydrostatic or hydrodynamic gas bearings, magnetic, or even powder lubricated bearings in extreme temperature environments.
7.2.2 Fluid Film Bearing Materials Selection of bearing materials for specific applications involves a scrutiny of the following characteristics: (1) (2) (3) (4)
Compatibility Embeddability and conformability Corrosion resistance Compressive and fatigue strength.
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7-8 TABLE 7.3
Handbook of Lubrication and Tribology List of Synthetic Lubricantsa
210◦ F
100◦ F
−65◦ F
Flash point ◦F
Diester Turbo oil 15
3.6
14.2
12,600
430
−90
$10.00
MIL-0-6085
3.5
13.5
10,000
450
−90
10.00
MIL-0-6387
4.6
15.8
5,000
410
< −80
—
3.8
30.7
—
465
—
3.60
Skydrole
3.85
15.5
>20,000
355
70
12.00
Pydraul F-9
5.8
54
—
430
+5
3.75
16
40
850
600
< −100
30.0
SF-96 (300)
122
300
7,000
605
< −55
30.00
SF-96 (1,000)
401
1,000
20,000
605
< −55
30.00
DC-710
40
275
—
575
−10
40.00
Silicate OS-45 Orsil BF-1
3.95 2.4
12.4 6.8
2,400 1,400
— 395
< −85 < −100
20.00
Wide-temperature-range aircraft hydraulic fluid
Polyglycol LB-140X LB-300X LB-650X
5.7 11.0 21.9
29.8 65.0 141.0
— — —
345 490 490
−50 −40 −20
2.40 2.40 2.40
2.4 11.5 72 —
8.9 60.6 433 666.3
— — — —
260 500 545 None
−85 −35 −25 −55
2.40 2.40 3.00 2.50
Water-insoluble oils used for internal combustion engines (Prestone Motor Oil), high-temperature hearings in ovens and furnaces and gears Water-soluble oils used in wire drawing, metal forming, and some machine tools Water–polyglycol mixture used as nonflammable hydraulic fluid in die easting and machine tool work
3.1 6.1
48 470
— —
380 None
20 50
2.30 2.30
Die-casting machines and high-pressure compressors
7.9 106 4,000
72 3,600 —
— — —
310 410 450
−40 10 70
— 1.05 1.40
Electrical oils, hydralic and shock absorbing fluids, kilns and ovens, refrigerator compressors
Kinematic viscosity, cSt Type
Phosphate Tricresyl phosphate
Silicon SF-96 (40)
50-HB-55 50-HB-280X 50-HB-2000 Hydrolube 300N
Chlorinated aromatics Aroclor 1248 Aroclor 1254 Polybutenes No. 8 No. 20 No. 128
Pour point ◦F
Approx. cost per gallon
Typical uses MIL-L-7808 high-load capacity, high-temperature jet engine oil Low volatility aircraft hydraulic and instrument oil Aircraft hydraulic fluid for alternator drives Low flammability hydraulic fluid for diecasting machines Nonflammable aircraft hydraulic oil Nonflammable hydraulic oil for diecasting machines, punch pressures etc. Low-torque aircraft oil bearings, air craft hydraulic, and damping fluid Heat transfer, hydraulic, and damping applications Heat transfer, hydraulic, and damping application Heat transfer, high-temperature trolley bearings
a For a more detailed discussion, the interested reader is directed to Zaretsky, E.V., Ed. “Tribology for Aerospace Application,”
STLE SP-37, 1997, pages 137–168.
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design TABLE 7.4
7-9
General Types of Additives With Typical Chemical Compositions
Function Oxidation inhibitor Detergent Rust inhibitor Wear preventive Bounday lubrication Viscosity index improver Pour-point depressant Defoaming agent
Typical Chemical Type Phenolics, dithiophosphate Calcium petroleum sulfonate Organic acids, sodium petroleum sulfonate Trieresyl phosphate Chlorinated naphthalene, sulfurized hydrocarbon Polyisobutylene Polymethnerylate Silicone oil
Source: Walton, J.F., and Heshmat, H., “Complaint Foil Bearings For Use in Cryogene Turbopumps,” Proceedings of Advanced Earth-to-Orbit Propulsion Technology Conference. Held at NASA/MSFC May 17–19, 1994, NASA CP3282, Vol. 1, Sept. 19, 1994, pp. 372–381.
In hydrodynamic bearings, the most relevant items are the allowable maximum pressures before the material begins to deform or flow, and the value of Tmax it can endure. For compressive strength, an alloy with intermediate strength is desirable; an alloy too low in strength is prone to extrude under load, while too strong a metal, being brittle, may crumble under impact loading. Fatigue strength is particularly important in applications with dynamic loading in order to prevent the formation of cracks or surface pits. The use of a thin soft layer bonded to a hard backing metal often gives the desired combination of fatigue and compressive strength; in such cases, however, the fatigue strength of the bond itself requires attention. When a material has low corrosion resistance, difficulties can be minimized by using oils with good oxidation inhibitors and by maintaining low bearing temperatures. 7.2.2.1 Babbitts The most common bearing materials are babbitts, either tin based or lead based. The detailed properties of babbitts are given in Table 7.5. Babbitts can operate under conditions of boundary lubrication or dirty operation. They have excellent compatibility and nonscoring characteristics and are outstanding in tolerating errors in construction and operation. Their deficiencies with regard to fatigue strength can be improved by using an intermediate layer of high-strength material between a steel backing and a thin babbitt layer. Many of these, known under the name of trimedal bearings, use the following construction (1) a low-carbon–steel back, (2) an intermediate layer of copper or bronze, and (3) an overlay of lead-base babbitt from 0.001 to 0.020 in. thick. The intermediate layers increase the mechanical strength of the babbitt bearing and also provide reasonable good bearing surfaces in cases where the thin babbitt surface layer is destroyed in operation. 7.2.2.2 Nonbabbitt Bearing Materials Other common bearing materials used, whenever babbitt cannot be employed are: Bronze: Bearing bronzes may be grouped into lead bronzes, tin bronzes, and high-strength bronzes. The strength and high-temperature properties generally improve as one proceeds from high-lead to hightin to various high-strength bronzes. However, there is a loss in the compatibility properties as the amount of lead decreases. For this reason it is generally advisable to use the highest lead content and the softest bronzes while still retaining the necessary strength and load-carrying capacity. Silver: Silver bearings normally consist of electro-deposited silver on steel backings with an overlay of 0.001 to 0.005 in. of lead. Indium is usually flashed on top of the lead overlay for corrosion protection. They have outstanding metallurgical uniformity, excellent fatigue resistance and thermal conductivity, can carry very high loads, and can be operated at high temperatures. Although the lead coating helps to relieve problems of poor embeddability and conformability, silver bearings are not recommended for applications where misalignment and dirt are present [30].
© 2006 by Taylor & Francis Group, LLC
7-10
TABLE 7.5
Composition and Physical Properties of Babbitts
Tin-base babbitts Yield pointa
Ultimate strengtha
psi
psi
Composition, % Alloy 1 2b 3b 4 5
Brinell hardness
Melting point
Complete liquefaction
Specific gravity
Cu
Sn
Sb
Pb
66◦ F
212◦ F
66◦ F
212◦ F
68◦ F
212◦ F
◦F
◦F
7.34 7.39 7.46 7.52 7.75
4.56 3.1 8.3 3.0 2.0
90.9 39.2 83.4 75.0 65.5
4.52 7.6 8.3 11.6 14.1
None 0.03 0.03 10.2 18.3
4400 6100 6800 5550 2150
2680 3000 3100 2150 2150
12,850 14,900 17,600 18,150 18,060
6050 8700 9900 8900 8750
17.0 24.5 27.0 34.5 22.5
8.0 12.0 14.5 12.0 10.0
433 466 464 363 358
700 669 792 583 565
Lead-base babbitts Yield pointa
Ultimate strengtha
psi
psi
Composition, %
Brinell hardness
Melting point
Complete liquefaction
Specific gravity
Cu
Sn
Sb
Pb
As max
66◦ F
212◦ F
66◦ F
212◦ F
68◦ F
212◦ F
◦F
◦F
6(e) 7(f) 8 10 11 12 15(g) 16(f) 19
9.33 9.73 10.04 10.07 10.28 10.67 10.05 9.88 10.50
1.5 0.50 0.50 0.50 0.50 0.50 0.5 0.5 0.50
20 10 5 5 — — 1 10 5
15 15 15 15 15 10 15 12.5 9
63.5 75 80 83 85 90 82 77 95
0.15 0.60 0.20 0.60 0.25 0.25 1.40 0.20 0.20
3,800 3,550 3,400 3,550 3,050 2,800
2,050 1,600 1,760 1,850 1,400 1,250
14,550 15,650 15,600 15,450 12,800 12,900
8,060 6,150 6,150 5,450 5,100 5,100
15,600
6,100
21.0 22.5 20.5 17.6 15.0 14.5 21.0 27.5 17.7
10.6 10.5 9.5 9.0 7.0 6.5 13.0 13.6 8.0
358 464 459 468 471 473 479 471 462
581 514 522 507 504 498 538 495 495
Note: a in composites. b Babbitts predominantely used by electric utilities (ASTM alloy B23).
© 2006 by Taylor & Francis Group, LLC
Handbook of Lubrication and Tribology
Alloy
Principles of Gas Turbine Bearing Lubrication and Design
7-11
Table 7.6 Approximate Temperature Limitations of Various Bearing Materials 0
200
400
600
Temperature °F 800
1000
1200
1400
Babbits Lead base Tin base Sintered Metals Bronze Composites Iron Aluminum alloys Copper–lead Bronzes Leaded Tin Aluminum Cast Iron Hardened steels Tool steels Carbongraphites (Untreated) Carbongraphites (Treated) Stellites Nickel-based superalloys Metal-bonded carbides Metal-bonded oxides Ceramics 0
200
400
600
Temperature °C
Source: Walton, J.F. and Heshmat, W., “Complaint Foil Bearnags For Use in Cryogenic Turbopumps,” Proceedings of Advanced Earth-to-Orbit Propulsion Technology Conference. Held at NASA/MSFC May 17–19, 1994, NASA CP 3282, Vol.1, Sept. 19, 1994, pp. 372–381.
Aluminum: Aluminum bearing alloys offer excellent resistance to corrosion by acidic oils, good loadcarrying capacity, superior fatigue resistance, and good thermal conductivity. A smooth machine finish of the running surface is recommended along with a clean lubricant, a shaft hardness of 300 Brinell or higher, and a large enough clearance to allow for the high thermal expansion of the aluminum. Sometime the aluminum is overlaid with a thin coating of lead babbitt. This overlay assists in making up for the otherwise poor embeddability and conformability characteristics of the aluminum. The range of temperatures that these various bearing materials, as well as some other materials, can endure is given in Table 7.6.
7.3 General Bearing Principles Five kinds of bearings can be considered candidates for use in gas turbines, each preferable for a particular set of operating conditions. From a tribological standpoint these can be grouped as follows: 1. Fluid film bearings: These bearings carry the imposed load on a fully developed fluid film, be it a liquid or gas. To this group belong the various hydrodynamic journal and thrust bearings that generate support pressures by virtue of the relative motions between the surfaces and hydrostatic bearings whose load capacity is provided by an externally supplied high-pressure fluid. 2. Elastohydrodynamic bearings: These bearings operate on a combination of hydrodynamic pressures and forces generated by deflections of the elastic bearing surfaces. Bearings in this category are the compliant surface or foil bearings (CFBs); and REBs. 3. Magnetic bearings: The third kind considered here is a unique support system, known as the magnetic bearing. In this device the load is carried on forces generated by an electromagnetic field in the bearing structure.
© 2006 by Taylor & Francis Group, LLC
7-12 (a)
Handbook of Lubrication and Tribology v
v Fluid
Restrictor
Pump Hydrodynamic bearing (b)
Hydrostatic bearing Inner race
v Fluid
v
Compliant surface Outer race Compliant surface bearing
FIGURE 7.3
Rolling element bearing
Four generic kinds of bearings. (a) Fluid film bearings. (b) Elastohydrodynamic bearings.
U z O
x
y h z = (L/2) z=0 x=0
FIGURE 7.4
z = (–L/2) x =B
Schematic of a hydrostatic bearing.
The basic elements of these various fluid film bearing designs are sketched in Figure 7.3. In the following subsections the theoretical principles of bearing operation will be sketched in as brief a manner as possible. Sections 7.3 through 7.8 will provide the practical aspects of bearing design and application.
7.3.1 Hydrodynamic Bearings Figure 7.4 shows a generalized sketch of a bearing consisting of a moving and a stationary surface. The two surfaces are separated by a fluid film of variable thickness. As the lubricant is sheared through the clearance from inlet to outlet its velocity, pressure, temperature, and viscosity undergo considerable variations that influence directly the bearing performance. a. Incompressible lubrication. For bearings using incompressible lubricants, the basic mathematical expression that relates performance to the bearing’s geometrical and operational parameters is the
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-13
Reynolds differential equation, given by ∂ ∂x
h 3 (∂p ) ∂ h 3 (∂p ) ∂h + = 6µU µ (∂x ) ∂z µ (∂z ) ∂x
(7.1)
When solved, the Reynolds equation yields the pressure field p (x, z ) as well as the pressure gradients (∂p /∂x ) and (∂p /∂z ) in the bearing. By integration one then obtains • Load capacity
L
W =
0
B
p (x, z )dx dz
(7.2)
0
• Frictional force and power loss Fτ =
0
L
0
B
h ± 2
L
Qx =
∂p ∂x
0
h3 12µ
µU + dx dz; h
∂p ∂x
+
H = Fτ U
hU dz 2
(7.3)
(7.4a)
• Component flows in the x and z directions Qz =
B
0
h3 12µ
∂p ∂z
dx
(7.4b)
b. Compressible lubrication. For a bearing lubricated by gas, the Reynolds equation is similar but because of gas incompressibility a density parameter now makes an appearance, namely ∂ ∂x
ρh 3 (∂p ) ∂ ρh 3 (∂p ) ∂ρh + = 6µU µ (∂x ) ∂z µ (∂z ) ∂x
(7.5)
The basic differences between gas and liquid lubricated bearings are: • The load capacity depends on the prevailing ambient pressure whereas it does not with incompressible fluids. • In liquid-lubricated bearings there is often cavitation at the end of the fluid film, but there is no cavitation in gas films. • Opposite to that of liquids, gas viscosity rises with a rise in temperature. • Film thicknesses in gas bearings are one or two orders of magnitude smaller than with liquidlubricated bearings; this is due to the much lower viscosity of gases, see Figure 7.5. For the same reason frictional losses in gas bearings are considerably lower. • No special arrangements for lubricant supply are required. c. Effects of turbulence. The above equations are all for laminar flow, that is, when the lubricant velocity is below a Reynolds number of about 1500, where the Reynolds number is given by Re =
ρRωh µ
(7.6)
Should the flow in any region of the film exceed the above value then lubricant flow would be reduced and both power loss and temperatures would rise. The Reynolds equation that defines turbulent bearing
© 2006 by Taylor & Francis Group, LLC
7-14
Handbook of Lubrication and Tribology 10–4
Absolute viscosity, reyns
10–5
Light machine oil (ASTM 150)
10–6
10–7
Water
10–8 Air Water vapor 14.7 psia 10–9 –200
FIGURE 7.5
0
200 400 600 800 1000 Temperature, °F
Comparative viscosity of various fluids. 1 0.5
G factor
Gz Gx 0.1
Gt
0.05
0.01 100
500 1,000
5,000 10,000
50,000 100,000
Reynolds number
FIGURE 7.6 Values of turbulence coefficients.
operation is given by ∂ ∂x
Gx h 3 (∂p) ∂ Gz h 3 (∂p) ∂h + = 6u µ (∂x) ∂z µ (∂z) ∂x
(7.7)
where the functions Gx and Gz are given in Figure 7.6 and are themselves functions of the Reynolds number that is of the level of turbulence. The function Gτ is a multiplying factor of the power loss given by Equation (7.3).
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-15
)
r ea
ing
(2
B
W x
, By B yy
Kyy, Kyx
1)
g(
in ar
Be
y
, Bx B xx
Kxx, Kxy
FIGURE 7.7 Dynamic coefficients of a journal bearing.
d. Stability. Rotordynamics plays a crucial part in determining the critical speeds, vibrational amplitudes, and possible instabilities in rotating machinery. Since the fluid film possesses both stiffness and damping properties, it affects considerably the overall characteristics of such a system. Moreover, each bearing possesses four spring and four damping coefficients, which must be known to determine the effect of bearings on a rotordynamic system. The origin of these eight coefficients derives from the fact that a load imparted either along x or z produces a journal displacement in two perpendicular directions; conversely a journal motion either along x or z generates incremental forces in two perpendicular directions. The four stiffness coefficients thus generated are defined by Kxx =
∂Fx ∂ex
Cxx =
∂Fx ∂ e˙x
Kxy =
∂Fx ∂ey
Cxy =
∂Fx ∂ e˙y
Kyx =
∂Fy ∂ex
Cxx =
∂Fy ∂ e˙x
Kyy =
∂Fy ∂ey
Cyy =
∂Fy ∂ e˙y
where e represents displacement and e˙ velocities. The subscripts xx and yy refer to the colinear and those with xy and yx to cross-coupling coefficients for a given differential applied force direction divided by the resulting displacement or velocity direction. These eight springs and dashpots are schematically represented in Figure 7.7. The dynamics for a rotor bearing system is then determined by the set of two dynamic equations Kxx x + Kxy y + Cxx x˙ + Cxy y˙ + M x¨ = 0
© 2006 by Taylor & Francis Group, LLC
(7.8a)
7-16
Handbook of Lubrication and Tribology
where the dotted quantities denote velocities and the double dot are accelerations. Kyx x + Kyy y + Cyx x˙ + Cyy y˙ + M y¨ = 0
(7.8b)
The solution of this set of equations yields two important quantities with regard to system stability. • Threshold instability frequency ωi • Critical mass, MCR The highest rotor mass M that a bearing will support before becoming unstable is then given M < MCR system stable M > MCR system unstable The threshold instability frequency ωi gives the vibrational frequency at the onset of stability. Given ωn as the natural frequency of the system, we then have If ωi < ωn , system is stable If ωi > ωn , system is unstable The two quantities MCR and ωi are determined from the eight coefficients as follows: M CR = (Z /ωi2 ) = Z¯ / 4(π ωi )2 Z= ωi ωi = = ω
(7.9a)
K xx C yy + K yy C xx − K xy C yx − K yx C xy
C xx + C yy K xx K yy + Z (K xx + K yy ) − K xy K yx + Z
(7.9b) 1/2
C xx C yy − C xy C yx
(7.9c)
7.3.2 Hydrostatic Bearings There are several distinct advantages to the use of hydrostatic bearings, namely: • They can operate at low or zero speeds without affecting the load capacity. They can therefore be easily started under load. • Due to the depth of the fluid pocket which is one or more orders of magnitude higher than in hydrodynamic films, they have very low drag losses. The above merits are counteracted by some inherent disadvantages: • They need an external pump or some other source to supply the pressurized fluid; thus even though the bearing power loss is low, the pumping power is part of the energy expenditure. • Due to the depth of the pocket they are prone to turbulence in the fluid. • They require a compensation system, that is a regulatory mechanism to adjust the fluid pressure in the pocket in conformity with changes in the imposed load. Since the pockets containing the pressurized fluid are very deep, there is usually no significant hydrodynamic effect. The film thickness can be considered constant and so (∂h/∂x) in Equation (7.1) becomes
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-17
R1 Pa
P0 R2
h
Pa P0
P0
P0 2r0 l
Orifice
Capillary
2a Ps
Ps
FIGURE 7.8 Elementary configuration of hydrostatic bearing.
zero. Thus the Reynolds equation for hydrostatic bearings reduces itself to a Laplace equation ∂ 2p ∂ 2p + =0 ∂x 2 ∂z 2 π h 3 (po − pa ) CD πa 2 ps − pa 1/2 = 2 2p 6µ ln(R2 /r1 )
(7.10) (7.11)
The proper boundary conditions are determined by the supply pressure and the form of compensation used in the system. The most common restrictor used for compensation is either an orifice or a capillary. Referring to Figure 7.8 the relation between the supply pressure ps , the pocket pressure po , and the ambient pressure pa is for an incompressible fluid given by • For orifice compensation • For capillary compensation Using po obtained from the above relations, one can subsequently calculate load, film thickness, and the flow in the bearing.
7.3.3 Compliant Surface Bearings The compliant bearing combines features of fluid dynamics and elastic response, which make for a number of unique features of this device. It also accounts for its greater analytical complexity since in addition to the Reynolds equation one must consider the corresponding mechanical response of its surfaces. Given the large number of possible configurations no single deflection equation can be written for them all;
© 2006 by Taylor & Francis Group, LLC
7-18
Handbook of Lubrication and Tribology Foil
v Shaft
FIGURE 7.9
Extension dominated CFB.
(a) Shaft v Bearing shell Foil (b) Moving surface
Uo
Foil lo
FIGURE 7.10
S
Bending dominated compliant bearing. (a) CS journal bearing. (b) Basic element.
rather each family of CFBs will have its own set of characteristic expressions, which together with the Reynolds equation will provide the necessary solution. There are multiple advantages to the use of compliant bearings. • Due to their ability to yield under load, they perform well at high speeds and at high temperatures. • By introducing friction between the compliant and rigid surfaces Coulomb damping can be introduced. • As compared to either fluid film or REBs they have very low power losses. • They can operate with either liquid or gas lubrication. By virtue of their ability to deflect the difference between performance with a liquid or a gas is not as pronounced as in rigid bearings. One of the major disadvantages is the high starting torque due to rubbing. This problem can be alleviated by applying coatings to the mating surfaces. Another disadvantage is that their installation requires more installation space than rigid bearings. The two major groups of compliant bearings are those in which the dominant mode of deformation is extension and those where bending is the main mode. Quite often both modes are present. Figure 7.9 shows a bearing belonging to the first category. The figure shows one segment of the surface consisting of a strip of foil supported on a compliant subsurface. The corrugated subsurface rests on the rigid bearing shell. A model of the tension dominated configuration is given in Figure 7.10 where a highly flexible foil is stretched between fixed fulcrums. Here the imposed bearing load is supported by the radial component of the tension forces; the load capacity can be further increased by imparting to the foil a preload tension.
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-19
Load v
Y O
6
b
5
6
b
5
Hertzian pressure curve
4
4 3
3 Pressure distribution
2
2 1
1 –2
–1
0
1 y/b
–2
Light load
–1
0 Heavy load
10
10
8
–1
8
Film thickness
6
–2
1 y/b
6
4
4
2
2 0
1 y/b
–2
–1
0
1 y/b
FIGURE 7.11 Elastohydrodynamic behavior of rolling elements.
7.3.4 Rolling Element Bearings Most of the literature, not to speak of handbooks, on ball and roller bearings is based on the premise that there is metal-to-metal contact between the races and the rolling elements. Under load the ball or roller is compressed to form a circular, elliptical, or rectangular contact area, generating conventional Hertzian stresses. In reality, however, the contact area is covered with a lubricant film, albeit several orders of magnitude smaller than in hydrodynamic bearings. Under these conditions, as was asserted in the opening paragraph of this chapter, there is superimposed onto the Hertzian stresses a hydrodynamic pressure field. These combined effects produce what is called an elastohydrodynamic pressure distribution along with a finite film thickness, whose basic shapes are given in Figure 7.11. The calculation of these pressures, film thicknesses, power loss, and so on, must, as in the case of compliant bearings, be based on the use of the Reynolds equation in conjunction with the elasticity equations as they pertain to the particular geometry of the bearing. The more significant attributes of REBs are as follows: • They can start under load and operate well at lower, including zero, speed. • When on the point of failure they will continue to operate for some time providing an opportunity to shut down the machine. • They can support both radial and axial loads.
© 2006 by Taylor & Francis Group, LLC
7-20
Handbook of Lubrication and Tribology 280 260 240 220
Bearing speed, rpm × 103
200 180 160 140 120 100 80 60 40 20
State of art
0 0
20 40 60 80 100 120 140 160 180 200 220 240 Bearing bore, mm
FIGURE 7.12
Limits on ball bearing DN values.
The main shortcoming of REBs is that, depending on speed and load, they have a finite life span. They must, therefore, be periodically replaced, unlike hydrodynamic bearings which can operate indefinitely. They also require more installation space and are noisy. Given the complexity of the elastohydrodynamic solutions, design and performance calculations of REB’s are based on empirical relations, related primarily to the fatigue properties of the structures involved. Since fatigue failure is a function of level of stress and number of cycles, the viability of an REB can thus be related to rotational speed of the machine, shaft size, and bearing loading. The relation between the first two parameters, known as the DN curve, is shown in Figure 7.12. This limit is imposed by the dynamics of the spinning balls or rollers and the rotating races and cages. Thus, regardless of load this DN limit ought not be exceeded. The longevity of a given bearing when operated under the limitations of Figure 7.12 does depend on the imposed load. This dependence is given by
L10
1.67 × 103 = N
WD W
(7.12)
where L10 is the predicted life of the bearing carrying a 90% probability of survival; N = shaft speed, rpm; W = load on the bearing; WD = the dynamic load a bearing will endure for one million revolutions; and γ = 3 for ball bearings or γ = 3.3 to 4.0 for roller bearings. The value of L10 is influenced by a host of additional factors such as size and number of rolling elements, kind of materials used, the bearing’s angle of contact, lubrication method, surface finish, and others.
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Principles of Gas Turbine Bearing Lubrication and Design (a)
Air gap
7-21
(b)
Stator Air gap
Rotor
Rotor
Stator
Flux path
Coil
Flux path
FIGURE 7.13 Basic designs of magnetic bearing. (a) Typical thrust active magnetic bearing. (b) Typical radial active magnetic bearing.
7.3.5 Magnetic Bearings Figure 7.13 shows basic designs of radial and thrust active magnetic bearing (AMBs). These contain a stator wound with coils to create the magnetic field and a rotor that has ferromagnetic laminations to interact with the magnetic field. In a radial bearing it is common to have two sets of opposing coils with a sensor to provide feedback of rotor position along two axes. In a thrust bearing, position is maintained by a set of stationary coils on either side of the runner. Since each pair of poles acts independently, it is sufficient to illustrate the action of one set to make clear the operation of the bearing as a whole. The attractive force on a rotor, as shown in Figure 7.14, is given by F ∝ (I 1/C )2 where F is the magnetic force and I the current. With attractive coils, as the rotor moves toward the coil the force on this face increases while on the opposite coil it decreases; thus without control the system is unstable leading to metal-to-metal contact. With a sensor and feedback system current is decreased on the approaching side and increased on the opposite to force the rotor back toward the center. Actually direct feedback control is nonlinear and therefore not responsive enough; this is remedied by the introduction of a large bias current compared to which the regulating current is small. The above outlined arrangement is, however, not sufficient to control resonances. For this the system must account not only for displacement but also for rotor velocities. Thus, a feedback for velocity sensor is also required or a provision to convert the recorded displacements into velocities. A major advantage of AMB’s is that stiffness and damping can be varied, unlike in any other bearings where they are fixed for a given design. With such control optimum performance can be obtained over a wide range of operating conditions. Moreover, this flexibility offers means for virtual rotor balancing to counteract residual unbalance in the system. The system can even account for pedestal vibrations once they are fed into the control circuit of the bearing. All the varied bearing designs are combinations of three principles of magnetic operation: • Attraction by the use of electromagnets • Repulsion by the use of permanent magnets • Reluctance using permanent magnets Figure 7.15 illustrates these principles. Part (a) shows the attraction system which has been applied in a number of designs employing one to five active servo controls (linear and angular). The repulsion principle illustrated in part (b) yields a radially stable bearing by means of passive magnets, but is axially unstable and requires at least one active control. In the reluctance design shown in part (c), the magnetic
© 2006 by Taylor & Francis Group, LLC
7-22
Handbook of Lubrication and Tribology Measured journal displacement Bias Ib1 adjustment
Power amplifier no.1 Ib1 – i Stator
Air gap Y Controller X Shaft –1 Journal
Stator Ib3 + i Bias Ib3 adjustment
FIGURE 7.14
Power amplifier no.2
Elements of AMB’s servo control.
circuit always seeks a geometry so as to align itself with the edges of the rectangular salient pole faces. In the configuration shown, axial stability is achieved at the expense of radial instability that must be overcome by an active control. Part (d) is shown not because it represents another principle but because using permanent magnets to energize the air gaps and electromagnets for control yields a metastable system that can operate at almost zero control power when disturbance loads are absent; and it produces a linear force versus current which simplifies the servo control design. 7.3.5.1 Attraction Electromagnets These are by far the most widely used. A straightforward approach is to use two sets of orthogonally disposed magnet pairs to control the rotating shaft, as shown in Figure 7.15(a). This approach controls x and y as well as the two angular directions. The axial direction must be controlled by either an active or passive bearing. With today’s materials these electromagnets can be quite compact. Unit loadings of 58 lb/in.2 per electromagnet pole face can be realized with silicon iron; and 116 lb/in. with iron–cobalt– vanadium alloys. As an example, to provide an active radial magnetic bearing capable of supporting a 100 lb load with a stiffness of 100,000 lb/in. on a 2 in. shaft diameter would call for a 2 in. long bearing. 7.3.5.2 Radial Repulsion Design One radially passive design is shown in Figure 7.16. Here multiples of radially polarized permanent magnets of alternating polarization are stacked axially with matching sets so that the disks are in repulsion with the housing. This provides a radially centering force. However, the bearing is axially unstable and requires a servo-controlled axial magnet. The bearing has limited stiffness; Figure 7.16 shows that for a stiffness of 100,000 lb/in. such a bearing would be 20 in. long. Likewise the bearing does not by itself develop much damping and the introduction of eddy currents is required to provide radial damping.
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design (a)
7-23
(b)
Permanent magnet rings radially polarized
Air gap Stator
Air gap
Actively controlled: Rotor
Radial bearing requires two servo axes,
Rotor
Radially stable axially unstable
Stator Flux path
Thrust bearing requires one servo axis Flux path
Coil Thrust bearing
(c)
Stator
Radial bearing
Permanent magnet ring axially polarized
S
(d)
N Ferromagnetic salient ring poles Stator PM Axially stable radially unstable
Rotor
Coil
N
Radially stable axially stable by servo control
Coil
S
Rotor
Rotor Rotor center line
FIGURE 7.15 The directional stabilities of magnetic bearings. (a) Attractive electromagnetic bearings. (b) Repulsive permanent magnetic radial bearings. (c) Reluctance centering thrust bearing. (d) Reluctance centering radial bearings and actively controlled thrust bearing with permanent magnet bias.
L
Typical data: t D
Permanent magnet rings radially polarized
FIGURE 7.16 The radial repulsive magnetic bearing.
© 2006 by Taylor & Francis Group, LLC
Kr = 10,000 1b/in. for D = 2.0 in. L = 2.0 in. t = 0.04 in. and B × H = 35 × 106 Gauss–Oersted
7-24
Handbook of Lubrication and Tribology
Two coils connected in series
Stator Permanent magnet
N
S
S
N
Rotor
Stator
FIGURE 7.17
The reluctance type magnetic bearing.
7.3.5.3 Reluctance Designs A radially passive bearing using multiple concentric rings on a stator and rotor is shown in Figure 7.17. The design uses permanent magnets to energize the radially passive gaps but electromagnetic power can be used as well. Here the ferromagnetic circuit will seek a geometric arrangement so as to minimize total magnetic “reluctance.” This design is radially stable and it will maintain a uniform centering force, provided the axial magnetic gaps have an active servo control. Its greatest disadvantage is that the maximum radial stiffness is low compared to the axially destabilizing stiffness. Consequently to achieve good radial stiffness large axial electromagnets are required.
7.4 Conventional Fluid Film Bearings The bearings described in the present chapter are the most commonly used bearing in rotating machinery. These are oil lubricated, mostly babbitted journal and thrust bearings, which when properly designed and maintained will last throughout the life of the machine. The parameters that determine the adequacy of a bearing design are described first, following which equations, tables, and charts supply the numerical values of these parameters for various bearing configurations and various sets of operating conditions.
7.4.1 Bearing Parameters The minimum film thickness: All discussion of bearing performance includes the term“load capacity”by which is usually meant the load a bearing can support at a given minimum film thickness. The minimum film thickness is important because it gives an indication of the following: • The likelihood of physical contact between the bearing and shaft surface, which may lead to failure. • The intensities of the peak pressures and temperatures which tend to rise steeply with a decrease in hmin . • The bearing safety margin and ability to accommodate an unexpected increase in load or sudden shaft excursion. There is no fixed value for a satisfactory hmin . It depends upon a number of factors including size of bearing, nature of application, operating conditions, degree of reliability required, and others. Naturally, in no case should hmin be smaller than the sum of the asperities of the two mating surfaces.
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-25
u=0 u W
Oil inlet
Oil inlet N o O O⬘
uS
u2
f
hmin
Hydrodynamic pressure
FIGURE 7.18 Conventional 2-axial-groove bearing.
The maximum temperature: The importance of knowing the value of the film maximum temperature, (Tmax ), is often on a par with that of the minimum film thickness (hmin ). While too small an hmin can cause damage by physical contact, excessive temperatures cause failure either by softening or melting the bearing surface, even when there is an ample film thickness. In aligned journal bearings, Tmax usually occurs near the trailing edge of the loaded bearing pad along the axial centerline; in thrust bearings, it occurs near the outer radius where both linear speed and the circumferential path of the lubricant are largest. Temperature rise: It is common practice in industry to use T as a criterion of bearing performance, this quantity being the difference between the bulk temperature of the oil discharging from the bearing and the oil supply temperature. While monitoring T may be helpful in spotting sudden changes in bearing behavior it is a poor indicator of the magnitude of Tmax . Two bearings may have the same T , yet two radically different values of Tmax , and if a bearing is to fail because of excessive temperature, it will be Tmax that would initiate the failure. Thus, while T remains a useful overall indicator of the amount of total heat generation and of any untoward changes in bearing behavior it cannot represent or replace the crucial quantity Tmax .
7.4.2 Journal Bearing Performance For a journal bearing with the nomenclature as sketched in Figure 7.18, the performance quantities of interest can be calculated as follows: • Film thickness. For an aligned journal it is h = (h/C) = 1 ε cos(θ − φ)
(7.13)
• Sommerfeld number and load parameter. The Sommerfeld number, given by S = (µN /P)(R/C)2
(7.14a)
has traditionally been the most important parameter. However, a more convenient quantity is the inverse of S, here called load parameter, given by ¯ = [P/(µn)](C/R)2 = [W /(LDµN )](C/R) W
© 2006 by Taylor & Francis Group, LLC
(7.14b)
7-26
Handbook of Lubrication and Tribology
where P = (W /LD) is the unit loading. What this parameter says is that any combination of P, µ, N , C, and R, such as to leave the value of W unchanged, would result in the same bearing eccentricity ratio , and attitude angle φ. • Minimum film thickness. This is the smallest distance between the journal and bearing surfaces and is given by hmin h¯ min = = (1 − ) c
(7.15)
What is normally referred to as load capacity related to the low W which this hmin can support. • Friction coefficient. This is the ratio between the frictional force and bearing load. It is normally expressed in the form of R (R/C)Fτ f = C W
(7.16)
• Power loss. This, of course, can be obtained from the value of Fτ , namely H = FT · Rω = f · W · Rω
(7.17)
However, the power loss is often given directly in the form of H¯ =
H [π 3 µN 2 LD 3 /C]
(7.18)
In Equation (7.18), the quantity by which H is normalized represents the power loss in an unloaded concentric journal bearing, that is, one in which = 0. It is known as the Petroff Equation. • Flow. In a bearing with a single or multipad arrangement, flow of lubricant Q1 enters at the leading edge; an amount of lubricant Qz is lost from the sides of the bearing; and flow Q2 leaves at the trailing end of the pad and adheres to the moving surface in most cases. The adhered flow, Q2 , is carried to the next pad and the bearing convergent section. Thus, the net amount of lubricant to be replenished is Qz . The latter is referred to as side leakage. Clearly we must always have Q1 = Q2 + Qz . All of these flows are given in dimensionless form as ¯ = Q/(π NDLC/2) Q
(7.19)
the denominator representing the flow in an unloaded, concentric bearing, that is, at = 0 (for which case Qz = 0 and Q1 = Q2 ). • Temperatures. For isothermal conditions, a bulk temperature rise can be estimated from the values of power loss and side leakage, namely T = (Tav − Ti ) =
H c wQz
(7.20)
The parameters , f , Q1 , Q2 , and Tmax , which serve to evaluate the various bearing performance items, are to be obtained from a solution of the aforementioned governing equations for the specific geometry and specific operating regimes of the bearing under consideration. • The problem of viscosity. In most of the expressions, µ appears either as a variable, or as one of the normalizing quantities. Since in practice viscosity varies throughout the fluid film, the question
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-27
210 A — correct Tav B — approximate Tav 200
190 Calculated Tav
Correct trial line 180
Trial 1
B Trial 2
2 point trial line
A
Trial 4 Trial 3
170
)
v
160
e
lin
ti
lu
So
150
on
ed
Ta
t
la
cu al
ed
=c
v
Ta
m
u ss
(A
150
160
170 180 Assumed Tav
190
200
210
FIGURE 7.19 Method of calculating temperature rise.
arises as to what value to assign to µ when quantitatively evaluating bearing performance. Without going into much detail three approaches are possible. 1. Inlet viscosity µ = µ1 = const : This is the approach most widely used because it is the simplest; it also introduces the largest errors. In this approach, a viscosity, µ1 , corresponding to the known inlet temperature, T1 , is used for all calculations. This may be acceptable for cases where the T is expected to be low. Even then, it should be kept in mind that in most applications, a given variation in T produces a much more pronounced variation in µ so that a small T may still produce appreciable variations in viscosity. 2. Average viscosity: = av = const : The overall aim in this approach is to assume an average temperature, Tav , which yields levels of power loss and side leakage such that when used in Equation (7.4a and b), the calculated Tav , will be the same as the assumed one. The steps involved here are as follows: Assume a Tav > T1 which then specifies µav . Using the above µav , calculate H and Qz . Determine Tav from Equation (7.20) If the assumed Tav from step 1 differs from the calculated Tav in Step 3, assume a different Tav and repeat procedure. 5. Continue until the assumed Tav equals the Tav calculated from Equation (7.20).
1. 2. 3. 4.
Figure 7.19 shows the procedure for arriving at a correct Tav . Plotting the assumed Tav vs. the Tav calculated from Equation (7.20) one can, after several trials, arrive at the correct result which is a point lying on the 45◦ line. While a correct convergence may require four or five trials, Figure 7.19 shows that drawing a straight line through the first two guesses may yield an approximate Tav sufficient for many applications. The use of µav based on the above approach is a very useful and efficacious means for calculating bearing performance, and it provides results of good accuracy.
© 2006 by Taylor & Francis Group, LLC
7-28
Handbook of Lubrication and Tribology (a)
R
R Oil in
Oil in
Axial groove
Overshot groove
(b) LD R
R + C0
L
FIGURE 7.20
The “Overshot” journal bearing. (a) Two-axial groove bearing. (b) Overshoot-groove bearing.
Types of bearing used: Large turbines employ at least half a dozen types of journal bearings. Essentially all of these designs consist of partial arc pads having a circular geometry. The differences are mainly in the number and arrangement of the pads and in whether or not the centers of curvature of the pads coincide with the geometric center of the assembled bearing. This will become clear during discussion of the individual bearing types.
7.4.2.1 Circular Bearings The most common journal bearing is what is often referred to as a “two axial-groove bearing,” portrayed in Figure 7.20. Its bore is circular and each of its two identical pads may span anywhere from 120 to 160◦ in angular extent. A variation of this bearing, shown in Figure 7.20(b), is called the “overshot-groove bearing.” The bottom pad is identical to that of part (a); however, the top pad is cut circumferentially by a deep channel. The rationale for the overshot groove is to increase the amount of oil flow. This it does, however, by flooding the upper half, it also increases the power losses. In general, since the load is carried primarily by the bottom pad, the overshot groove in circular bearings does not affect hmin or Tmax one way or the other. Circular bearings, made up of more than two pads, are called 3-, 4-, or 5-axial-groove bearings. Similar to the 2-axial-groove bearings, the individual pads in these designs are separated by oil grooves with the load passing midway through the bottom pad. Circular bearings are used extensively since they are easy to manufacture, to install, and to repair. Their performance is good as long as stability is not a problem. The angular extent of the pads can be reduced down to 120◦ without penalty, the load capacity being about the same for all 360◦ > β > 120◦ . If the
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design Table 7.7
7-29
Performance of Axial Groove Bearings Two-groove
Three-groove
Four-groove
L /D
φ
S
Qz
φ
S
Qz
φ
S
Qz
1–1/2
0.2 0.4 0.6 0.8 0.9
62 51 42 29 23
0.435 0.179 0.0943 0.037 0.0167
0.165 0.290 0.350 0.320 0.290
67 49 36 27 20
0.99 0.37 0.15 0.05 0.019
0.11 0.205 0.29 0.39 0.40
80 72 49 25 18
1.9 0.71 0.245 0.259 0.0222
0.081 0.16 0.29 0.38 0.37
1
0.2 0.4 0.6 0.8 0.9
64 53 45 28 22
0.714 0.275 0.125 0.041 0.019
0.27 0.43 0.56 0.46 0.43
69 50 37 26 20
1.33 0.48 0.187 0.059 0.022
0.16 0.30 0.435 0.059 0.022
79 72 49 25 19
2.04 0.87 0.30 0.07 0.025
0.12 0.39 0.51 0.53
1/2
0.2 0.4 0.6 0.8 0.9
64 53 42 24 22
2.10 0.80 0.31 0.082 0.031
0.32 0.56 0.715 0.725 0.695
72 54 39 27 20
3.00 1.06 0.385 0.103 0.033
0.245 0.47 0.68 0.92 1.04
80 72 49 25 19
4.23 1.7 0.55 0.11 0.035
0.20 0.42 0.67 0.84 0.90
1/4
0.2 0.4 0.6 0.8 0.9
70 57 43 27 20
7.94 2.86 1.06 0.256 0.074
0.36 0.67 0.86 0.93 0.87
74 56 41 28 20
9.61 3.30 1.17 0.27 0.0735
0.30 0.0625 0.90 1.17 1.29
79 72 49 25 19
11.6 4.5 1.3 0.23 0.73
0.272 0.56 0.89 1.15 1.22
bottom pad arc is reduced below 120◦ , a deterioration in load 1capacity sets in which accelerates rapidly when arcs of less than 80◦ are used. As is true for most bearings the smaller the (L /D) ratio the lower the load capacity, though a low L /D, by increasing eccentricity, improves a bearing from the standpoint of stability. Table 7.7 gives the characteristics of 2-, 3-, and 4-axial-groove bearings. 7.4.2.2 Elliptical Design A journal bearing which has an increased capability to suppress instability is the elliptical bearing shown in Figure 7.21. This bearing looks much like the 2-axial-groove bearing, but the two lobes are assembled so that their centers of curvature are not coincident. Each lobe has been displaced inward, this displacement, expressed as a percentage of the machined radial clearance, being denoted as ellipticity or “preload.” In effect, then, even for operating at the geometric bearing center, both lobes have eccentricities greater than zero. Larger amounts of ellipticity increase the pad eccentricities and thus provide more stable operation. The design is penalized by higher friction losses, higher flow rates, and reduced load capacity at low eccentricities. At high eccentricities, the behavior of the elliptical bearing approaches that of the circular design. The bearing is commonly used because it is effective and is easy to manufacture. With shims placed at the horizontal split, the bearing is given a circular bore and then the shims are removed yielding the desired “elliptical” geometry. In order to properly understand and handle the elliptical bearing, the following central fact must be stressed. While, as is customary, journal positions is referred to the geometric center of the bearing ( and φ), the quantities that count, that is, the parameters that determine load capacity, hmin , etc., are the eccentricities and attitude angles with respect to the lobe centers 01 and 02 .
© 2006 by Taylor & Francis Group, LLC
7-30
Handbook of Lubrication and Tribology (a)
Pad 2 R +C O1
Pad 1
Oil in
R + Cm O
Oil in
R
O2
Cm
C
m = (d/c) e1,2 = (e1,2/C)
(b)
O1 W f1 N
O1 e O 2 dO e 2 e1 O⬘
e m
Full oil film
e1
O
f
hmin e2
O⬘
O2 f2
FIGURE 7.21 Geometry and nomenclature in elliptical bearings. (a) Concentric journal position. (b) Eccentric journal position.
The relation between the bearing parameters and the two-lobe parameters are, referring to part (b) of Figure 7.21, as follows: 1 = ( 2 + m 2 + 2 cos φ)0.5 sin φ φ1 = sin−1 e1 2 = [ 2 + m 2 − 2m cos φ]0.5 sin φ φ1 = π − sin−2 2
(7.21a) (7.21b)
(7.22a) (7.22b)
where = e/C, 1 = e1 /C, 2 = e2 /C, m = d/C. Thus, for a bearing with a certain ellipticity ration m, when bearing eccentricity and attitude angle ( and φ) are given, all other quantities can be determined from the foregoing relationships.
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Principles of Gas Turbine Bearing Lubrication and Design
7-31
It should be noted here that in all the normalized quantities, the machined bearing clearance C is the relevant parameter. Also, the value of hmin , by virtue of what was said above, is determined not by and φ, but by the eccentricity of either the upper or lower lobe. If the shaft center is above the horizontal centerline (and, as will be seen later, this is possible) then hmin is located in lobe #2 and hmin is determined by hmin = C (1 − 2 )
(7.23a)
θmin = 3π/2 + φ2
(7.23b)
If, as happens in the majority of cases, the shaft center is below the horizontal centerline, then the relevant equations are hmin = C (1 − 1 )
(7.24a)
θmin = π + ϕ2
(7.24b)
A convenient quantity in dealing with elliptical bearings is to introduce a bearing eccentricity ratio based not on C, but on the minor clearance Cm . Thus, since C¯ m = (Cm /C ) = 1 − m
(7.25)
m = (e /Cm ) = /(1 − m )
(7.26)
we have for this new eccentricity ratio:
The usefulness of the above expression lies in the fact that as in circular bearings, the journal cannot travel in the vertical direction beyond m m = 1. (It can do so in other directions.) It is thus possible to plot the loci of journal center for all values of m against a common vertical scale, as shown in Figure 7.22 [23]. As seen, at low eccentricities, the shaft center is often above the horizontal centerline locating the bearing’s hmin in the upper lobe. The plot includes, for comparative purposes, m = 0, which corresponds to a circular bearing. The general performance characteristics of the elliptical bearing are given in Table 7.8 and Figure 7.23. These are given for the entire practical range of (L /D) ratios and for ellipticity ratios from 0.25 to 0.75. Since m = 0 represents a circular bearing and m = 1 is the limit for possible ellipticity ratios (at m = 1 there is contact between the surfaces), the plots cover the entire spectrum of elliptical bearing design. Unlike the case of circular bearings, the introduction of an overshot groove in elliptical designs has a telling effect on performance. In essence, an extensive overshot groove destroys the effectiveness of the bearing as an elliptical design, approximating it to a circular bearing of clearance C. 7.4.2.3 Three-Lobe Design The 3-lobe bearing (Figure 7.24) represents a further accentuation of the features that characterize the elliptical design. It is more stable than the circular and elliptical varieties, but due to its lower average clearance and smaller arc of bottom lobe, it shows higher losses and lower load capacities than the two other designs. As the name implies, the conventional 3-lobe design consists of thee pads each of about 80 to 120◦ in angular extent, with the bottom lobe placed symmetrically about the vertical load. The bearing is shown schematically in Figure 7.26. All remarks made in connection with the elliptical bearing regarding the relationships, usage, and importance of the bearing and lobe parameters hold here too, though clearly the quantitative relations will be somewhat different. Thus, the relation between the parameters of the
© 2006 by Taylor & Francis Group, LLC
7-32
Handbook of Lubrication and Tribology
110°
100° L/D = 1.0 L/D = 1.5 90°
0.2 80°
L/D = 0.5 0.4 em
f
=0
0.6
70°
.5
m = 0.2
5
m = 0.75
m
60°
0.8 m=
1.0
10°
0
50°
20°
30°
40°
0° b = 150° m = 0, 0.25, and 0.75 valid for all (L/D)
FIGURE 7.22
Locus of shaft center for elliptical bearings.
geometric center (, φ) and of the individual lobes are here as follows: 1 = [ 2 + m 2 + 2m cos φ]0.5 sin φ φ1 = sin−1 1 π
0.5 2 = 2 + m 2 − 2m cos φ +φ 3 2π sin(π/3 + φ) − sin−1 φ2 = 3 2 π
0.5 3 = 2 + m 2 − 2m cos φ +φ 3 sin(π/3 − φ) 4π + sin−1 φ3 = 3 3
© 2006 by Taylor & Francis Group, LLC
(7.27a) (7.27b)
(7.28a) (7.28b)
(7.29a) (7.29b)
Principles of Gas Turbine Bearing Lubrication and Design Table 7.8
7-33
Performance of Elliptical Bearings m = 1/4
m = 1/2
m = 34
L /D
m
φ
j,2
S
Q7
φ
1,2
S
Q
φ
1,2
S
1–1/2
0.2 0.4 0.6 0.8 0.9 1.0 1.2
110 90 62 38 32 — —
0.33 0.39 0.61 0.81 0.90 — —
0.572 0.185 0.090 0.0333 0.0107 — —
0.17 0.29 0.35 0.305 0.29 — —
98 98 92 70 53 — —
0.52 0.56 0.59 0.74 0.85 — —
0.625 0.313 0.156 0.051 0.024 — —
0.23 0.25 0.33 0.333 0.30 — —
87 75 75 75 — 75 70
0.75 0.178 0.82 0.835 0.84 0.85 0.895
0.25 0.11 0.045 0.034 0.03 0.029 0.019
0.27 0.275 0.295 0.30 0.30 0.30 0.30
1
0.2 0.4 0.6 0.8 0.9 1.2
105 90 68 35 30 —
0.32 0.39 0.59 0.81 0.90 —
0.834 0.305 0.120 0.040 0.019 —
0.225 0.412 0.546 0.436 0.430 —
95 90 87 67 48 —
0.52 0.54 0.60 0.75 0.80 —
0.415 0.274 0.155 0.056 0.024 —
0.322 0.37 0.44 0.47 0.43 —
85 80 81 79 78 65
0.75 0.77 0.79 0.81 0.94 0.90
0.285 0.143 0.087 0.0607 0.0155 0.0202
0.41 0.41 0.41 0.44 0.44 0.45
1/2
0.2 0.4 0.6 0.8 0.9 1.0 1.2
100 90 82 38 30 — —
0.31 0.39 0.51 0.81 0.90 — —
2.0 0.57 0.29 0.071 0.0305 — —
0.275 0.53 0.72 0.72 0.69 — —
90 85 83 65 48 — —
0.51 0.54 0.61 0.76 0.86 — —
1.3 0.67 0.32 0.11 0.040 — —
0.52 0.57 0.67 0.72 0.675 — —
80 80 75 75 75 75 67
0.76 0.78 0.82 0.83 0.84 0.85 0.90
0.53 0.26 0.085 0.067 0.058 0.053 0.030
0.68 0.69 0.72 0.73 0.73 0.74 0.76
1/4
0.2 0.4 0.6 0.8 0.9 1.0 1.2
100 95 62 30 28 — —
0.31 0.41 0.61 0.81 0.90 — —
7.15 3.33 0.96 0.24 0.074 — —
0.42 0.59 0.94 0.93 0.93 — —
90 85 80 70 45 — —
0.51 0.545 0.62 0.74 0.875 — —
5.55 2.32 0.945 0.377 0.080 — —
0.56 0.72 0.825 0.91 0.875 — —
75 75 75 75 75 75 68
0.76 0.78 0.80 0.83 0.84 0.85 0.88
1.07 0.715 0.488 0.260 0.170 0.130 0.091
0.93 0.96 0.97 0.98 0.98 1.00 1.00
Q7
Note: Wherever φ > 90, 1.2 ; otherwise it denotes 1 . Source: Pinkus, O. and Sternlicht, B., “Theory of Hydrodynamic Lubrication,” McGraw-Hill, 1961.
Another detail to be noted is that although C is the machined clearance in all three lobes; when assembled, this dimension does not physically appear in the bearing. The largest concentric clearance, which occurs at the junction of the three lobes, is less than the machined clearance C, and will here be denoted by CM . It is given by C¯ M = (CM /C ) = (1 − m /2)
(7.30)
whereas C¯ m , as before, is given by (1 − m ). Figure 7.25 shows the locus of shaft center in a 3-lobe bearing of 100◦ arc extent. Since, as seen here, the shaft never rises above the horizontal centerline, the minimum film thickness will always be located in the bottom pad, and is given by hmin (1 − 1 )
(7.31)
Table 7.9 and Figures 7.25 and Figure 7.26 give the performance characteristics of the 3-lobe bearing for three (L /D) ratios and two ellipticity ratios. Together with the 3-groove circular bearing given in Table 7.7, which represents the case m = 0, a three point variation in m is provided from which performance for intermediate values of m can be obtained by cross plotting [25].
© 2006 by Taylor & Francis Group, LLC
7-34
Handbook of Lubrication and Tribology 2.4 2.0
1 2 3 4
3 1
1.6
4
L/D = 1/4 L/D = 1/2 L/D = 1 L/D = 1 + 1/2
m=0
2
1.2 0.8 2.6 3 2.2
Power loss factor H
1.8
1
4
m = 0.25 2
1.4 1.0 2.0
1
1.8
2
1.6
m = 0.5
3 1.4
4
1.2
2.2 1 2.1
4
m = 0.75 2
2.0 3 0.01
FIGURE 7.23
0.1 1.0 (mN/P) (R/C)3
10.0
Power loss in elliptical bearings.
It was said previously that the 3-lobe bearing is usually chosen for its superior stability characteristics and following is a brief illustration of its features vis-a-vis the elliptical bearing. Figure 7.27 shows the stable and unstable regimes plotted against Sommerfeld number for the two bearing types. Each of the constant lines gives the locus of the operation of a bearing with a fixed geometry as its speed is varied. The bearing parameter, η, is given by η=
µLD 2π W
2 R W 0.5 C CM
which is independent of rotor speed and describes a certain bearing geometry. As speed is increased, the bearing will, via S, proceed along a line of constant η and eventually enter the unstable region. The
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-35
f1 3
O1
3
CM
2 m
e1
O1 O
d
f2
O
m
R+C
O2 R
2
O3
Cm
C
e O9
e2
O2
m
f
O3 e3
1
f3
1 e1,2,3 =(e1,2,3/C) e =(e/C) m =(d/C)
FIGURE 7.24 Geometry and nomeclature of 3-lobe bearings.
W
v 0
90°
80°
0.2
70°
0.4
f 1/3 60°
0.6
2/3 50°
0.8
40°
m=0 30°
1.0
10°
Clearance boundary
20°
FIGURE 7.25 Locus of shaft center for 3-lobe bearings. (Taken from Pinkus, O., Trans. ASME, 78, 1956, 965–973. With permission.)
parameter η is most sensitive to bearing diameter and lobe clearance and less so to the length, L, and viscosity, µ. A move to a higher value of η, that is, to a more stable region is accomplished by either increasing µ, increasing L, increasing D, or decreasing C. Because the ordinate parameter ω(MC/W )0.5 also depends upon the clearance, changes in C will follow the slightly inclined dashed lines — toward the left if C is increased, toward the right if C is reduced. At light loads and small clearances, the 3-lobe
© 2006 by Taylor & Francis Group, LLC
7-36
Handbook of Lubrication and Tribology
Table 7.9
Performance of Three-Lobe Bearings. m = 1/2
m = 2/3
L /D
m
φ
1
S
Q
φ
1
S
1
0.2 0.4 0.6 0.8 1.0
42 53 55 50 30
0.58 0.63 0.71 0.815 0.965
0.45 0.18 0.10 0.048 0.0063
0.125 0.185 0.20 0.235 0.28
50 50 50 45 40
0.71 0.75 0.81 0.85 0.945
0.21 0.12 0.071 0.039 0.0095
0.18 0.185 0.20 0.21 0.23
1/2
0.2 0.4 0.6 0.8 1.0
45 55 55 50 30
0.57 0.63 0.71 0.815 0.905
0.84 0.40 0.20 0.084 0.011
0.265 0.31 0.335 0.425 0.50
50 52 50 45 40
0.71 0.76 0.81 0.86 0.945
0.43 0.21 0.11 0.054 0.0125
0.355 0.35 0.38 0.41 0.465
1
0.2 0.4 0.6 0.8 1.0
45 45 45 40 30
0.575 0.65 0.745 0.845 0.965
2.5 1.0 0.41 0.13 0.021
0.37 0.51 0.54 0.61 0.75
62 55 58 52 44
0.70 0.75 0.80 0.86 0.945
1.16 0.59 0.33 0.12 0.025
0.53 0.53 0.53 0.574 0.58
Q
Source: Pinkus, O. and Sternlicht, B., “Theory of Hydrodynamic Lubrication,” McGraw-Hill, 1961.
b = 100°
bG = 20°
– Power loss, H
3
2
m = 2/3
m = 1/2 m=0 1 0
0.2
0.4
0.6
0.8
1.0
em
FIGURE 7.26
Power loss in three-lobe bearings.
bearing is better than the elliptical bearing. Under heavier loads, the elliptical design is the more stable bearing. It should, however, be kept in mind that it is precisely the low load range that is the troublesome region of bearing stability. When a bearing is absolutely stable, the whirl ratio approaches zero and Figure 7.27(b) shows the variation of the whirl ratio at the threshold of instability. The whirl ratios with the 3-lobe and the elliptical
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design (b) 01
(a) 103 0. 00
h ig ed e sp
0.7 Elliptical
0.6
1=
0. 00
1
Unstable
3-lobe
Half-frequency whirl
0.5 0.4
3-lobe
0.3 0.2
1=
0. 01
10 L arge clear ance
Whirl ratio, (vj/v)
H
1=
v(MRC/W)1/2
102
7-37
Stable Low speed 1 0.01 0.1
Elliptical
(mNLD/W)
1.0
Small clear ance 10.0
(R/C)2
3-lobe b = 100° m = 0.5 Elliptical b = 160°h = (mLD/2pW)(R/C)2(W/CM3)1/2
FIGURE 7.27 frequency.
0.1 0 0.01
0.1
1.0
10.0
(mNLD/W) (R/C)2 3-lobe b = 100° m = 0.5 Elliptical b = 160°
Stability characteristics of elliptical and 3-lobe bearings (a) Stability regimes. (b) Instability threshold
bearings share similar characteristics. They both rise sharply at S = 0.1 to 0.2 and remain fairly constant thereafter. The elliptical bearing which is the less desirable for lightly loaded applications is seen to have a whirl ratio in excess of 0.5, thus a worse ratio than the 3-lobe bearing. The whirl ratio is an important parameter in determining the stability threshold for a flexible rotor which is always lower than that for a rigid rotor. Thus, Figure 7.27(b) can be viewed as the highest possible stability that can be achieved with these bearings [23]. 7.4.2.4 Tilting Pad Bearings Unlike the previously considered designs, the tilting-pad bearing is a generic name that covers many permutations. Its primary characteristic is that the individual pads are not fixed in position, but are pivotsupported so that during operation not only does the shaft move in response to operational conditions, but so do the pads, and each pad in a different fashion. A general picture of a pivoted shoe bearing is shown in Figure 7.28. The complexity of the design is partly evidenced in the configuration of a single pad given in Figure 7.28(b). Several things ought to be noticed. In the first place, the criterion of hmin as a measure of load capacity loses its meaning somewhat here since this hmin is not a fixed distance; however, the film thickness over the pivot, hp , is a geometrically fixed point. Under excessive load, given that the pad at hmin can yield whereas at hp it cannot, failure is more likely to occur at hp . Thus, the critical quantity here is perhaps hp rather than hmin . The next thing to realize is that the center of curvature of the pad is not fixed in space; when the pad rocks above the pivot, its center of curvature moves either in a positive or negative angular direction, shown in Figure 7.28(b) by ±γ . Next, should the preload be too low, some of the pads on the top of the bearing may become unloaded, in which case, as shown in Figure 7.29, the fluid film frictional moment about the pivot will make the leading edge of the pad scrape against the journal and cause“flutter,” obviously an undesirable contingency. The condition that the top pads not be unloaded is dictated by the amount of preload and shaft position; the lower the preload and the higher the , it is more likely that some of the pads will become unloaded. Figure 7.30 gives a sample graph, in terms of m and m, when a pad is likely to be unloaded. In this respect, loading between the pads, when the journal may reach m > 1, is a more undesirable mode of operation [30].
© 2006 by Taylor & Francis Group, LLC
7-38
Handbook of Lubrication and Tribology (a)
Pivot Pivot clearance circle
Pad R-Cm e w
R
Shaft
b1
(b)
W
Cm
Concentric journal
Pad 1
us1
e1 hp1
bs1
b1
b
N
E O1 Oe
Pad 2
1
O
R
up1 hmin 1 Concentric pad uE1
FIGURE 7.28
Tilting-pad journal bearing. Three-pad tilting-pad bearing. Geometry of tilting pad.
The number of possible design parameters and operating modes in a tilting-pad bearing is large. Some of these design options are: • • • • • • •
Number of pads, 3 > n > 8 Angular extent of pads, including the option of a variation of β for various pads Load vector passing between pads or over a pad Central or eccentric pivot location, that is, a choice of value for (βp /β) (L /D ) ratio Preload, m = 1 − (Cm /C ) Pad inertia, which often determines the ability of the pad to follow or track journal motion.
7.4.3 Thrust Bearings Much of what has been said previously about journal bearings applies also to the behavior of thrust bearings. It thus remains only to point out the differences that arise due to the different geometry of a thrust bearing shown in its generic elements in Figure 7.18[a].
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-39
Frictional forces
Frictional moment
N O + Shaft center
FIGURE 7.29 Unloaded tilting pad.
0.6
0 0.3.4
2
4
0.
1
0.
0
0.
0.6
em
1.0 0.8
em All pads always loaded
1.2 1.0 0.8
For the given value of m maximum em at which all pads are loaded.
2
Load
0
0.4 0.3 0. 2 0. 1
0.6 0.5
0.7
m
0.8
0
0.
Load
Some pads always unloaded
FIGURE 7.30 Regime of unloaded pads in a 5-pad tilting pad bearing.
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Handbook of Lubrication and Tribology
Thrust bearings are simpler to handle in that no cavitation occurs and in that it is sufficient to solve only for one pad (for a parallel runner all pads are identical). The geometry of the film, on the other hand, is more varied. The film shape in journal bearings is more or less universal, namely that prevailing between two eccentric circular cylinders. In thrust bearings it can be anything — shapes with one or two directional tapers, with or without flats, crowned profiles, pocket bearings, and finally tilting-pad designs. Another simplification with thrust bearings is that no instability problems, such as those that occur with journal bearings, arise in their operation. There is, thus, no need to evaluate stiffness and damping. The Reynolds equation for thrust bearings has to be written in polar instead of rectangular coordinates. In parallel to journal bearings, turbulence is accounted for on a point-by-point basis, here a function of r as well as θ ∂ ∂r
3
rh Gz µ
∂p ∂r
3
1 ∂ h Gx ∂p r ∂h + · =6 r ∂θ µ ∂θ (L/R2 )2 ∂θ
(7.32)
where Gx , Gz are the turbulence coefficients in the θ and r directions respectively, both functions of the Reynolds number given by Re = ρrωh/µ = f (r, θ ) The expressions above differ from those for the journal bearing in that they have a dependence on r. This is due to the variation of the Couette flow with r and to the film thickness being a function of both coordinates, r and θ. 7.4.3.1 Tapered Land Bearings The simplest tapered land bearing is one which has a constant angular taper, or h(θ) − h2 + δθ (1 − θ/β)δθ = (h2 − h1 )
(7.33)
with its geometry as shown in Figure 7.31(a). This equation, independent of r, is valid for a bearing surface with a circumferential taper alone. As will be shown later, the exact shape of the fluid film between fixed
B A
A L
Load U
Runner
h1
du B
Cut A–A
FIGURE 7.31
Tapered land bearing.
© 2006 by Taylor & Francis Group, LLC
h2
Principles of Gas Turbine Bearing Lubrication and Design
7-41
values of h1 and h2 does not affect the results appreciably. Thus, by their simplicity, the on-dimensional taper solutions provide a useful key for evaluating the bearings in general. The several crucial parameters in journal bearings are β. (L /D ), and (e /C ). Parallel quantities appear in thrust bearings, namely, β, the angular extent of the pad; (L /R2 ); and (h2 /δθ ) with δθ (like C) being a geometric quantity and h2 being the trailing film thickness at which the bearing is run. It should be also noted here that hmin = h2
(7.34)
Solutions for the tapered land bearing are given in Table 7.10, where Q r = (Qr /π R2 NL δθ )
(7.35)
is the side leakage with the index R1 indicating the leakage along the inner radius and R2 indicating the leakage along the outer radius. The total side leakage is then Q r = [Qr |R1 + Q r |R2 ]π R2 NL δθ The leakage out the end of the pad, Q2 , is given by: Q2 = 0.5π NLh1 (R1 + R2 ) + Q2 p π R2 NL δθ
(7.36)
where the first right-hand term is the shear flow and does not involve any computer obtained coefficients. The value of Q 2p can be obtained from Table 7.10 by subtracting Q r from (Q r + Q 2p ). Table 7.11 shows the relative load capacities and friction of three different thrust bearing configurations. One is a plane slider, that is, an inclined rectangular block; the second, a slider with an exponential film profile; and the third is the tapered land geometry of Equation (7.33). As seen, the results for a given value of (h1 /h2 ) are nearly identical, confirming our assertion that once h1 and h2 are fixed, the exact variation in h between these values is not of great importance. In all of the above results, it should be noted that P is the unit pressure given by: P = W /Area = 360 WT /[n βπ L (R2 + R1 )] where β is in degrees. WT is the total load on the thrust bearing and n the number of pads. Also it should be noted that the data for flow and power loss in Table 7.10 are for a single pad so that the total flow and losses are QT = n Qpad
HT = n Hpad
7.4.3.2 Composite Tapered Land Bearings A more practical and preferred thrust bearing geometry is a tapered land bearing having tapers in both the circumferential and radial directions with a flat portion at the end of the film. Its advantages are (1) it has higher load capacity, (2) has lower side leakage and, (3) at low speed and during starts and stops it provides a flat surface for supporting the load, thus minimizing wear. The geometry of such a bearing is shown in Figure 7.32. Its film thickness is given by h = h11 −
h11 − h12 h11 − [h11 − h12 /L](R − r1 ) − h2 (r − R1 ) − θ L bβ for 0 < θ < bβ h = h2 , constant for bβ < θ , β
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(7.37)
7-42
Handbook of Lubrication and Tribology
Table 7.10
L/R2 1/3
Solutions for Tapered Land Thrust Bearings
b1 /Sθ 1
1/2
1/4
1/8
1/2
1
1/2
1/4
3/8
2/3
1
1/2
1/4
1/8
β (deg)
µN p
L δθ
2
r = (r − R1 )/Lθ = θ/β Center of pressure
Qr
H δθ
at R1
at R2
Q r + Q 2p
θ
r
π µN 2 R24
80 55 40 30 80 55 40 30 80 55 40 30 80 55 40 30
1.423 1.180 0.947 0.870 0.321 0.257 0.225 0.211 0.0855 0.714 0.0652 0.0635 0.0278 0.0247 0.0238 0.0242
0.34 0.32 0.28 0.235 0.35 0.32 0.29 0.245 0.35 0.32 0.29 0.235 0.36 0.33 0.29 0.25
0.40 0.44 0.81 0.75 0.47 0.44 0.40 0.36 0.47 0.44 0.41 0.36 0.48 0.45 0.41 0.37
0.87 0.84 0.81 0.75 0.87 0.84 0.79 0.74 0.87 0.83 0.78 0.70 0.85 0.81 0.75 0.67
0.64 0.025 0.61 0.605 0.71 0.69 0.67 0.66 0.78 0.76 0.74 0.73 0.83 0.815 0.795 0.78
0.37 0.45 0.49 0.51 0.37 0.47 0.50 0.51 0.41 0.45 0.505 0.52 0.465 0.50 0.51 0.565
2.44 1.685 1.26 0.95 3.94 2.70 2.00 1.57 5.96 4.25 3.23 2.54 8.51 6.23 4.88 3.91
80 55 40 30 80 55 40 30 80 55 40 30 80 55 40 30
1.72 1.494 1.435 1.489 0.402 0.3585 0.352 0.370 0.1138 0.1062 0.1080 0.1103 0.0402 0.0399 0.0423 0.0470
0.23 0.19 0.145 0.11 0.23 0.19 0.15 0.11 0.24 0.20 0.15 0.11 0.25 0.20 0.16 0.11
0.405 0.36 0.31 0.20 0.41 0.33 0.31 0.26 0.42 0.27 0.32 0.27 0.42 0.28 0.32 0.27
0.75 0.69 0.61 0.57 0.74 0.61 0.60 0.53 0.72 0.65 0.56 0.49 0.70 0.62 0.53 0.44
0.62 0.61 0.60 0.59 0.685 0.67 0.655 0.65 0.755 0.735 0.72 0.71 0.81 0.78 0.77 0.765
0.48 0.51 0.53 0.55 0.46 0.52 0.53 0.55 0.48 0.52 0.54 0.56 0.50 0.53 0.55 0.57
2.90 1.96 1.47 1.13 4.72 3.33 2.49 1.92 7.32 5.29 4.065 3.18 10.81 8.06 6.30 5.01
80 55 40 30 80 55 40 30 80 55 40 30 80 55 40 30
2.240 2.185 2.320 2.590 0.538 0.537 0.578 0.653 0.1598 0.1655 0.1820 0.2035 0.0599 0.0649 0.0737 0.0861
0.12 0.082 0.052 0.033 0.13 0.084 0.0535 0.034 0.13 0.087 0.055 0.035 0.14 0.09 0.056 0.036
0.35 0.295 0.245 0.200 0.35 0.30 0.25 0.20 0.36 0.30 0.25 0.21 0.365 0.31 0.25 0.21
0.60 0.53 0.48 0.44 0.58 0.51 0.45 0.40 0.56 0.46 0.40 0.38 0.53 0.44 0.35 0.29
0.61 0.60 0.59 0.59 0.67 0.66 0.65 0.645 0.735 0.72 0.71 0.705 0.79 0.78 0.765 0.75
0.50 0.55 0.58 0.61 0.51 0.56 0.59 0.61 0.53 0.57 0.60 0.62 0.55 0.58 0.61 0.63
3.06 2.12 1.57 1.20 5.07 3.59 2.70 2.07 8.00 5.70 4.43 3.46 12.07 8.98 6.94 5.47
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Principles of Gas Turbine Bearing Lubrication and Design
7-43
Table 7.11 Performance of Thrust Bearings With Various Film Configurations Plane slidera
α
Exponential sliderb PL 2 h22
P= 2.00 2.50 2.85
µωR24
0.0810 0.113 0.135
0.0819 0.1137 0.435 F=
2.00 2.50 3.04
Sector padc
0.0826 0.106 0.125
Fh2 µω24 R24
0.66 0.74 0.84
0.81 0.875 0.95
0.78 0.825 0.88
Note: a h = αx b h = k1 e k2 x c h = h2 + δ(1 − θ/β) Source: Pinkus, O. and Sternlicht, B., “Theory of Hydrodynamic Lubrication,” McGraw-Hill, 1961.
h12 u
(r,u)
h2
v
Ro
Fla
t
h11
L
bb
Ri
b v Flat at R1
2h2∗ h12
h11
h2
d21
Ro
Ro Ri
Ri
FIGURE 7.32 Composite tapered land bearing.
Normalizing all h’s by h2 and all radii by R2 , we have θ θ r = (L/R2 ) − 1 h − h 11 − (h − 1) − δr 1− bβ (L/R2 ) bβ for 0 < θ < bβ h = 1, for bβ < θ < β
© 2006 by Taylor & Francis Group, LLC
(7.38)
7-44
Handbook of Lubrication and Tribology Table 7.12 Rei 500
1500
3500
Composite Tapered Land Thrust Bearings Q/R12 h2 ω
Wh22
Hca h2
µR14 ω
µR14 ω2
at θ = 0
at R1
at R2
H W ωh2
0.192 0.182 0.175 0.168 0.151 0.337 0.307 0.289 0.275 0.240 0.567 0.499 0.463 0.435 0.369
3.92 3.88 3.78 3.68 3.42 8.45 7.93 7.59 7.31 6.63 15.6 14.0 13.2 1.25 11.2
1.58 1.58 1.59 1.59 1.60 1.61 1.62 1.62 1.63 1.64 1.63 1.64 1.64 1.65 1.66
0.296 0.294 0.293 0.292 0.290 0.324 0.321 0.319 0.318 0.314 0.339 0.334 0.331 0.329 0.325
0.437 0.445 0.451 0.456 0.469 0.455 0.465 0.471 0.476 0.490 0.462 0.475 0.482 0.488 0.504
17.6 18.3 18.5 18.7 19.1 21.5 21.6 21.8 21.9 22.3 23.2 23.2 23.3 23.4 23.7
Note: (R2 /R1 ) = 2; β = 40◦ , h11 = 3, δr = 0.5; b = 0.8 a H includes losses over a 10◦ oil groove. All results are per individual pad. c
The expression for h has, as seen, three arbitrary parameters: • h 11 = (h11 /h2 ) — the dimensionless maximum film thickness at the lower left corner. • δ r = (h11 − h12 )/h2 — the radial taper along the leading edge θ = 0. • b = the friction of β tapered. In an optimization study in which both load capacity and lower power losses were considered, the following desirable proportions for the above three parameters were arrived at: h = 3.0 δ r = 0.5
b = 0.8
Physically, the above numbers imply a maximum film thickness at (R1 , 0) of three times the one over the flat; an outward decrease in film thickness at the leading edge half that of h2 ; and a flat portion equal to 20% of the pad’s angular extent. The performance of such a bearing for the case of a 40◦ bearing pad an (OD/ID) ratio of 2 is given in Table 7.12. The table provides data for both turbulent and laminar operation. The following comments will, perhaps, be useful: • The values of the Reynolds number Re1 = ρR1 ωh2 /µ1 • The losses as represented by Hc in column four include the losses over a 10◦ oil groove; the losses H over the pad above can be obtained from the last column in Table 7.12. • The flow QIN represents the inflow at θ = 0. The outflow will, of course, be given by Q2 = QIN − (QR1 + QR2 ) • The lowest value of Re1 given is 500. This value is close to laminar operation. • For the total bearing, the values of W , Hc , Q, and H should all be multiplied by the number of pads.
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Principles of Gas Turbine Bearing Lubrication and Design
7-45
v
(a)
Pressure profile U
(b) Resultant v Original shape Thermally and elastically distorted pad
FIGURE 7.33 The hydrodynamics of a tilting pad thrust bearing. (a) Zero load capacity of centrally pivoted slider. (b) Generation of a hydrodynamic film due to thermal and elastic distortions.
7.4.3.3 Tilting Pad Bearings The comments made about the tilting pad journal bearing regarding its complexity and large number of parameters apply equally well to the thrust bearing. However, in the case of a pivoted thrust pad such as the one shown in Figure 7.33, an additional complication overshadows the other difficulties; that there is theoretically no solution to a planar centrally pivoted sector. This can be deduced from the pressure profile sketched in Figure 7.33(a). Such a profile must always be asymmetrical with respect to the center of the pad; an asymmetrical pressure profile would impose a moment about the pivot tending to align the pad parallel to the runner. However, a parallel pad produces no hydrodynamic pressures, thus making the working of such an arrangement impossible. Yet such centrally pivoted, plan surface thrust bearings are widely used and they perform exceedingly well. Various theories have been advanced and stratagems employed to explain the workings of these bearings and to obtain a solution to the problems. Among these are: • Thermal or density wedge — The variation in viscosity or density of the oil is often credited with generating hydrodynamic forces in the parallel film. At best, such effects produce forces which come nowhere near the heavy loading supported by such bearings. • Thermal and elastic distortion of the pad — As shown in Figure 7.33(b), thermal and elastic stresses may crown a pad, so that in essence it produces a convergent–divergent film. In that case, it is possible for the resultant load to pass through the pivot and the pad can support a load. However, such bending can occur either with very thin pads or extremely high-temperature gradients. Yet such bearings performance satisfactorily even with very thick pads and under conditions of minimal heat generation. • Incidental effects — There are a number of incidental features which may play a more important role than the above theoretical explanations. Among these are: (a) Machining inaccuracies on the faces of both runner and bearing and rounded off edge at entrance to the pad, which in effect constitute a built-in taper.
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Handbook of Lubrication and Tribology
(b) Misalignment between runner and pads during assembly or during operation. (c) Pivot location not exactly at 50% of pad angular extent. These factors would combine to generate hydrodynamic forces and they are perhaps the most likely explanation for the satisfactory working of tilting pad thrust bearing.
7.5 Low-Speed Bearings One of the requirements in the bearing described in the previous section is a proper lubrication system. This includes a pump delivering oil at 10 to 50 psi supply pressure with all the accompanying equipment such as oil tank, filters, piping, sump, and cooling arrangements. When the bearings run at relatively low speed involving low power dissipation and therefore low bearing temperatures — as in fans, blowers, and some compressors — one can simplify the system by employing oil-ring lubrication. This consists of a self-contained oil delivery package placed adjacent to the bearing which dispenses with all the auxiliary equipment required for a more demanding operation. Figure 7.34 shows the components of an oil-ring lubrication setup. The ring, riding on the top of the exposed shaft, is a sort of viscous drag device that lifts oil from the sump and deposits it on the shaft. It is clear that in comparison to a pressurized supply system where the oil is distributed along an axial groove, the amount of oil lifted is not sufficient to provide the bearing with a complete oil film, and therefore an important parameter in oil ring operation is the amount of oil the bearing receives relative to what it needs for a full film. This is called the starvation ratio and is given by ˆ z = Qz /QzF Q where Qz is the side leakage under starved conditions and QzF is the side leakage for a full film. Thrust bearing
Shaft
vring v
Journal bearing Oil flow
Sump oil
FIGURE 7.34
Oil-ring lubrication system.
© 2006 by Taylor & Francis Group, LLC
Oil ring
Cooling water
Sump
Principles of Gas Turbine Bearing Lubrication and Design
7-47
7.5.1 Regimes of Operation The value of starvation ratio depends on shaft and ring speeds. Due to the centrifugal effects of the rotating ring and attached oil one can distinguish four regimes of ring behavior, as shown generically in Figure 7.35. The characteristics of these regimes are as follows [15]: Regime I : At the low end of journal rotation there is contact between ring inner surface and the journal, and the linear speeds of the two mating surfaces are about the same. There is thus a linear rise in ring rpm with the rpm of the journal. The oil delivered by the ring rises throughout this regime. At the upper end of Regime l, ring speed and oil delivery reach a local maximum. Regime II: At the beginning of this regime, direct frictional drag yields to a state of boundary lubrication between ring and journal. Due to this, slippage occurs and ring speed drops. Since the speed has decreased, so too does the amount of oil delivered by the ring. However, with further rise in journal speed, a full hydrodynamic film is established between journal and ring. The reduced viscous friction (the friction coefficient may drop from 0.1 to 0.01) and the larger film between the mating surfaces bring about a rapid increase in both ring speed and oil delivery. Once again, at the upper end, a local maximum in ring speed is achieved. Oil flow, however, at the end of this regime is an absolute maximum and represents the highest possible oil delivery by the ring. Regime III: The drop in ring speed and oil delivery following Regime II is associated with the onset of ring oscillations in the plane of rotation. While small oscillations already appear during the trailing portion of Regime II, the values of σθ become, within a short span, very large and bring about a drastic reduction in oil delivery. While, due to these planar oscillations, the ring speed drops only slightly, the oil delivery is affected to such a point that at the end of this regime it approaches asymptotically zero.
QZF
NR
Regime I
Regime III
Regime IV
QR Regime II
sZ su
0 Journal speed
FIGURE 7.35 Oil ring behavior as function of journal speed.
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7-48
Handbook of Lubrication and Tribology u=0 u
W
N 0 eF 0°F e
u1F
u2F
0° f
fF u0F
u1 u2 u0 Extent of full film Extent of starved film
FIGURE 7.36
Full and starved fluid films.
The angular swing of the ring at its maximum values can be of the order of ±5 to 10◦ with an oscillatory frequency equal to that of ring rotational frequency. Regime IV : This regime, essentially beyond our interest, is characterized by conical and translatory vibrations of the ring. While the oscillatory motion abates and the ring speed once again tends to increase with journal speed, oil delivery is essentially zero. The frequency of both the conical and translatory (or axial) vibrations is that of ring rotational frequency. Starting with the oscillatory vibrations and proceeding through the two other modes of instability, the violent motion of the ring causes splash and a throw-off of oil from the surface of the ring, and partly also from the journal, so that little oil reaches the bearing. Figure 7.36 presents the hydrodynamics of a starved journal bearing vis-a-vis a full film or flooded condition. For a fixed load, speed, and supply oil temperature, the deviations of a starved bearing from one operating with a full fluid film are as follows: • The film starts later and terminates earlier, that is θ1 > θIF and θ2 < θLf , producing in essence something similar to a partial bearing, though the upstream, and sometimes also the downstream, boundary conditions are different. • The eccentricity increases, producing smaller values of hmin . • The attitude angle decreases, yielding a more vertical locus of shaft center. • Since oil supply is equivalent to side leakage, there is a reduction of bearing side leakage and consequently an increase in fluid film temperatures. This is somewhat mitigated by a reduction in power loss due to a shorter extent of the fluid film. Table 7.13 gives a set of solutions for a wide range of loads and levels of starvation [15]. It is seen that the effects of starvation are much more pronounced at light and moderate loads, which are due to the fact that, at heavy loads, the full film extends over a narrower arc and the pressure gradients at θ1 being higher, a lower amount of Q1 is required to form a film. The loci of shaft center, both for constant oil supply values Q1 and for constant loads, W , are given in Figure 7.37. It is seen that starvation displaces
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design Table 7.13
7-49
Theoretical Performance of Starved Journal Bearings
W
Qz %
C
φ, degs
Q1
Qz
θ1 , deg
βs , degs
0.491
0 0.9 3.1 12.0 29.0 53.0 100a
1.00 0.903 0.808 0.625 0.446 0.272 0.081
0 3 6 11 18 28 78
0. 0.100 0.200 0.400 0.600 0.800 1.03
0. 0.899 × 10−3 0.309 × 10−2 0.0122 0.0289 0.0534 0.100
180 175 171 163 155 144 105
0 13 22 40 60 86 150
1.965
0. 0.7 2.8 12.4 30.3 50.2 100a
1.00 0.906 0.818 0.657 0.516 0.395 0.285
0 5 9 17 26 37 57
0 0.100 0.200 0.40 0.600 0.800 1.03
0 0.208 × 10−2 0.0817 0.0357 0.0875 0.162 0.289
180 172 167 155 144 129 105
0 16 29 57 82 101 150
9.825
0. 1.5 6.4 15.6 28.0 60.0 100a
1.00 0.914 0.846 0.793 0.753 0.702 0.672
0 8 14 19 24 31 37
0. 0.100 0.200 0.300 0.400 0.600 0.809
0. 0.659 × 10−2 0.0291 0.0697 0.125 0.267 0.445
100 168 158 148 139 122 105
0 25 46 69 82 108 132
34.39
0. 4.8 19.4 38.7 61.3 100a
1.00 0.929 0.896 0.881 0.073 0.866
0 11 17 21 23 26
0 0.100 0.200 0.300 0.400 0.560
0 0.0192 0.0802 0.162 0.253 0413
180 160 144 131 120 105
0 38 73 79 86 110
98.25
0 12.9 38.7 66.1 93.8 100a
1.0 0.954 0.947 0.946 0.945 0.944
0 12 16 17 18 18
0. 0.100 0.200 0.300 0.400 0.423
0 0.0463 0.138 0.235 0.335 0.357
180 149 131 118 107 105
0 50 70 87 100 101
Note: µ = const; β = 150◦ ; (L/D) = 0.93 a Full fluid film. Source: Heshmat, H. and Pinkus, O. “Performance of Starved Journal Bearings with Oil Ring Lubrication,” J. Trib. Trans. ASME, 107, 1985, 23–32.
the locus of shaft center inward of the full film line, that is, toward higher eccentricities and lower values of attitude angle φ. From the above and the parametric studies described in Section 7.8 the following conclusions can be drawn: • Except at very low speeds, most oil ring bearings operate under starved conditions. • The load capacity of oil ring bearings, first increases then decreases with rising shaft speed. • The locus of shaft center of oil ring bearings is much closer to the vertical axis than in full film bearings. • An optimum in the (L/D) ratio exists in oil ring bearings which ranges from 0.6 to 0.8. • The effects of starvation are much more pronounced at low and intermediate loads than at high loadings.
© 2006 by Taylor & Francis Group, LLC
7-50
Handbook of Lubrication and Tribology v
0
90°
0.1 80°
0.2 W = 0.49
0.3
1.96
0.4
70°
Full film
Q1 = 0.6
0.5
60°
0.6
0.4 50°
0.7
9.82
0.8 0.9 1.0
f
0.2
40°
34.4
0.1
30° 20°
0°
Q1 = Const W = Const
10°
FIGURE 7.37 Locus of shaft center at different levels of starvation. (Taken from Heshmat, H., and Pinkus, O. J. Trib. Trans. ASME 107, 1985, 23–32. With permission.)
7.6 High-Speed and High-Temperature Oil-Free Bearings This section will consider bearings suitable for operation at extreme speed and temperature ranges. These two parameters may occur either together or independently, that is, although usually high speed implies high temperatures, the reverse is not always so. One can have lower or moderate velocities but the environment may be such that the bearing will be exposed to high temperatures and the fluid film, the lubricant, and bearing materials must be able to cope with it. The other important consideration in high-speed bearings is that of stability. As will be seen below, the likelihood of baring instability, known as half-frequency whirl, rises with rotational speed. This becomes particularly intense when speeds twice the natural frequency of the system are reached. The kinds of bearings that are possible candidates for such applications are gas bearings, either hydrodynamic or hydrostatic, compliant surface geometries, and magnetic bearings.
7.6.1 Gas Bearings The differential equation governing the behavior of gas bearings is that given by Equation (7.5). For liquids ρ is a constant and the density terms fall out of the equation. In gas bearings it varies with both pressure and temperature. In most cases the perfect gas equation is applicable, or P = RT ρ
© 2006 by Taylor & Francis Group, LLC
(7.39)
Principles of Gas Turbine Bearing Lubrication and Design
7-51
Two factors make the gas film in bearings isothermal; one is the low heat generation; the other is the high thermal capacity of the bearing shell as compared to the tiny volume of gas in the film. We therefore have ρ = RT p = const · p
(7.40)
and Equation (7.39) becomes ∂ ∂x
ph 3 µ
∂P ∂x
+
∂ ∂Z
ph 3 µ
∂p ∂z
= 6U
∂(ph ) ∂x
(7.41)
All the solutions given subsequently are based on this expression with the proper boundary conditions applied to each specific geometry. As was pointed out in Section 7.2, unlike with liquids, gas bearing behavior depends on the ambient pressure; thus load capacity, for example, rises with a rise in pa . A new dimensionless parameter now makes an appearance which governs gas bearing behavior given by: =
6µω pa
2 R C
(7.42)
called the Bearing number. Thus, along with geometry and such variables as (L /D ) ratio, load, speed, etc., the value of p, or, in dimensionless form, , constitutes now an additional input. 7.6.1.1 Full Circular Bearings These bearings are the most commonly used if for no other reason than that they are easy to manufacture requiring no grooves or holes for lubricant supply. In obtaining a solution it is only required that at the sides of the bearing the hydrodynamic pressures fall to ambient pressure p, in most cases the atmosphere. Figure 7.38 and Figure 7.39 give the load capacities and frictional losses for full (360◦ ) gas journal bearings for a range of (L /D ratios from 1/2 to 2) and for the entire spectrum of possible 7 values. It can be shown that when 7 6 0, the gas bearing solution approaches that of a liquid; thus the solutions in the figures comprise cases from liquid lubricants to gases of very high compressibility. The reason that the (L /D ) ratios range as high as 2 is to compensate for the inherently low load capacity for gas bearings. 7.6.1.2 Noncircular Geometries As was pointed out in Sections 7.4.2.2 and 7.4.2.3, elliptical and 3-lobe geometries are often resorted to because of their higher stability characteristics. This is particularly desirable with gas bearings which tend to become unstable at high speeds. The solutions given in Sections 7.4.2.2 and 7.4.2.3 for these bearings are for centrally loaded cases, that is when the load vector passes midway through the bottom lobe. However, this is not the optimum mode of loading; better results can be obtained when the bearing is so positioned in the housing that the load is made to pass through the bottom lobe at an angle φL , see Figure 7.40, called the load angle. 7.6.1.2.1 Load Capacity There is a rather large number of independent parameters when dealing with noncircular gas bearings. Assuming even, as was done here, that the space or slots between the individual lobes occupy a negligible portion of the arc, that is, assigning to the elliptical bearing two arcs of 180◦ span each, and to the 3-lobe bearing, three symmetrical (they do not have to be equal) arcs of 120◦ each, we are still left with five independent parameters, namely p = p[(L/D), m, β , αβ , ]
© 2006 by Taylor & Francis Group, LLC
7-52
Handbook of Lubrication and Tribology 50 L/D = 1 40
40 L/D = 1/2
30
e = 0.8
WG
WG
30
20
20
0.7
e = 0.8 0.6
10
0
10
0.4 0.2
0.1
1.0
0.6
0 0 0.2
0.6 Λ
0.2 0
0 0.2
0.6 0.5 0.4 0.3 0.2 0.6
1.0
Λ
1/Λ
50
0.2 0
1/Λ
50 L/D = 1 1/2
L/D = 2 e = 0.8
e = 0.8
40
30
40
10
0.5 0.4 0.3 0.2 0 0.2
0.6 Λ
1.0
WG
WG
0.6
0.7
30
0.7
20
0
0.6
0.6 20
0.5 0.4 0.3 0.2
10
0.6
0.2 0
0
0 0.2
0.6 Λ
1/Λ
1.0
0.6
0.2 0
1/Λ
WG = (6pw/paLRΛ); Λ = (6mv/pa) (R/C)2
FIGURE 7.38 Load capacity of full (360◦ ) gas journal bearing. (Taken from Pinkus, O. and Sternlicht, B., “Theory of Hydrodynamic Lubrication,” McGraw-Hill, 1961.)
A solution for any set of these five parameters will yield the load capacity in the form of the Sommerfeld number S, and the line of action of the resultant force, or load angle φL with respect to the geometry of the bearing. Tables 7.14 to 7.18 give the results with regard to load capacity and some of the other bearing performance characteristics. It should be kept in mind that since load capacity means the relation between Sommerfeld number and minimum film thickness, this hmin is provided not by the bearing eccentricity which is unrelated to the surface curvature, but by the value of one of the lobes, that is, hmin = C(1 − )
(7.43)
In order to obtain this hmin we must search from among the two or three lobes the maximum lobe eccentricity ratio. The values of listed in the tables and figures are these maximum eccentricity ratios. These most often occur in the bottom lobe. In the few cases when the maximum occurs in Lobe No. 2,
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Principles of Gas Turbine Bearing Lubrication and Design
7-53
2.60 2.40 e = 0.8 2.20 2.00 1.80 e = 0.6 FC pmv R2L
1.60 1.40 e = 0.4
1.20 1.00
e = 0.2 0.80 0.60
L/D = 2
0.40
L/D = 1
0.20
L/D = 1/2
0 0
0.2
0.6
1.0
0.6
0.2
Λ
0
1/Λ
FIGURE 7.39 Friction in full (360◦ ) gas journal bearing. (Taken from Pinkus, O. and Sternlicht, B., “Theory of Hydrodynamic Lubrication,” McGraw-Hill, 1961.)
a1 a1 2
a2
a2
2
3 e1
R
e
B
R d
U
e2 a3
e e8 1 e3 e2
–a3
U
1
1
fL y
aB
Y fL
FIGURE 7.40 Nomenclature for noncircular gas bearings. (a) The elliptical bearing. (b) 3-lobe bearing.
this can be identified in the Tables from the fact that for the elliptical bearing this would require αβ > 90◦ ; and for the 3-lobe bearing αβ > 60◦ . While in a circular ungrooved bearing the direction of load application is immaterial. This is not so in the case of noncircular designs. The most common mode of bearing operation is with the load vector parallel to the vertical line of symmetry. This is the natural way of mounting the bearing and it
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7-54
Handbook of Lubrication and Tribology Table 7.14
Centrally Located Elliptical Gas Bearings αB at
at
α at
B
1/2
1
3
1/2
1
3
1/2
1
3
0.1 0.2 0.3 0.4 0.45
75 75 70 50 30
60 60 50 30 —
30 27 20 — —
0.53 0.58 0.67 0.82 0.92
0.56 0.62 0.73 0.87 —
0.59 0.68 0.79 — —
10 19 25 22 14
9 16 18 13 —
5. 7.5 7.5 — —
S at B 0.1 0.2 0.3 0.4 0.45
G at
F =
1/2
1
3
1/2
1
3
1/2
1
3
0.283 0.135 0.0772 0.0341 0.0147
0.250 0.117 0.0625 0.0270 —
0.282 0.128 0.700 — —
0.079 0.0703 0.0455 −0.0204 0.0412
0.127 0.125 0.0620 0.0444 —
0.122 0.101 0.0717 — —
3.28 3.41 3.70 4.54 6.27
3.28 3.40 3.76 4.90 —
3.21 3037 3.86 — —
Note: L /D = 1 m = 1/2 Source: Pinkus, O., Trans. ASME, 80, 1958.
Table 7.15
Optimally Located Elliptical Gas Bearings
βB
B
α
S
φL
G
1/2
0.1 0.2 0.3 0.4 0.45
80 80 80 80 80
0.53 0.57 0.63 0.69 0.73
11 20 28 35 37.5
0.297 0.143 0.0879 0.0581 0.0474
−3 −4 −7 −10 −13
0.0798 0.0735 0.0608 0.0408 0.0279
1
0.085 0.10 0.175 0.20 0.29 0.40 0.475
80 80 75 75 75 80 75
0.52 0.53 0.57 0.575 0.64 0.69 0.77
9 11 17 20 26 35 36
0.362 0.308 0.160 0.149 0.0888 0.0619 0.0411
−12 −12 −9 −13 −12 −18 −19
0.137 0.131 0.124 0.120 0.0995 0.0753 0.0204
3
0.09 0.11 0.22
80 80 75
0.52 0.53 0.60
10 12 21
0.449 0.336 0.163
−23 −23 −21
0.198 0.196 0.168
Note: L /D = 1 m = 1/2 Source: Pinkus, O., Trans. ASME, 80, 1958.
is also useful in that it enables two-directional rotation. However, this is not necessarily the optimum arrangement. Applying the load at various angles to the vertical centerline would yield different values of bearing performance. Somewhere an optimum angle exists for the direction of load application and these are shown in Tables 7.15 and 7.17. In practice it means that depending on the parameters S and , the bearing should be rotated with respect to the load vector anywhere from a few degrees to as much as 25◦ and nearly always in the clockwise direction, in order to obtain the maximum load capacity. Figure 7.41 summarizes graphically some of the data contained in the Tables. Table 7.18 is a practical summary of the various implications contained in the previously discussed results. In practice one is usually confronted with the given requirements of speed, load, ambient conditions, etc. In other words, S and are fixed. Given these parameters, Table 7.18 shows what eccentricities one can obtain by using a circular or noncircular design.
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design Table 7.16
Centrally Located 3-Lobe Gas Bearings αB at
B 0.1 0.2 0.3 0.4
7-55
1/2
at
1
70 67 63 53
55 55 50 39
α at
3
1/2
1
3
1/2
1
3
44 42 37 —
0.57(2)a
0.56 0.64 0.73 0.85
0.58 0.66 0.76 —
112(2) 106(2) 99(2) 23
8.4 14.9 18.3 17.2
7.1 11.7 13.5 —
0.64(2) 0.71(2) 0.81
S at
G at
F at
B
1/2
1
3
1/2
1
3
1/2
1
3
0.1 0.2 0.3 0.4
0.293 0.134 0.0734 0.0370
0.301 0.135 0.0734 0.0352
0.437 0.194 0.106 —
0.0358 0.0368 0.0370 0.0314
0.0902 0.0860 0.0773 0.0550
0.0736 0.0675 0.0575 —
3.62 3.86 4.30 —
3.60 3.82 4.25 —
3.55 3.76 4.13 —
Note: L/D = 1 m = 1/2 a (2) indicates that minimum film thickness occurs in right-hand lobe.
Table 7.17
Optimally Loaded 3-Lobe Gas Bearings
B
αB
α
S
φL
G
1/2
0.1 0.2 0.3 0.4
60 60 60 60
0.56 0.62 0.70 0.78
9 16 22 26
0.284 0.128 0.0720 0.0418
9 7 3 −5
0.0358 0.0364 0.0362 0.0336
1
0.085 0.10 0.175 0.20 0.28 0.30 0.40 0.455
55 55 60 60 60 60 60 60
0.55 0.56 0.61 0.62 0.68 0.70 0.78 0.83
7 8 14.5 16 21 22 26 28
0.385 0.300 0.165 0.140 0.0893 0.0810 0.0477 0.0347
1.5 1 −4 −5 −6.5 −8 −13 −17
0.0580 0.0902 0.0564 0.0862 0.0521 0.0790 0.0425 0.0359
3
0.10 0.225 0.30
50 50 60
0.57 0.67 0.70
8 15 22
0.444 0.171 0.121
−6 −7 −17
0.117 0.0655 0.0578
Source: Pinkus, O., Trans. ASME, 80, 1958.
Table 7.18
Comparison of Load Capacity of Various Gas Bearings
S
m=0 circular
1
0.365 0.162 0.866 0.0381 0.365 0.162 0.0860
0.2 0.4 0.6 0.8 0.25 0.475 0.655
3
Ellipticala
m = 1/2
3-lobea m = 1/2
φt =0
Optimum
φ=0
Optimum
0.54(−170) 0.59(−48) 0.67(−12) 0.815(−2) 0.57(−132) 0.65(−37) 0.75(−14)
0.525(−162) 0.56(−40) 0.635(−6) 0.77(+4) 0.535(−114) 0.59(−24) 0.685(−4)
0.55(−175) 0.615(−54) 0.70(−17) 0.84(−5) 0.595(−38) 0.68(−43) 0.80(−22)
0.545(−172) 0.605(−51) 0.685(−14) 0.815(−2) 0.57(−128) 0.645(−36) 0.755(−15)
Note: (L/D) = 1, Values of β for given and S a Numbers in parentheses refer to percentage reduction in load capacity from a circular design.
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Handbook of Lubrication and Tribology
L/D = 1 M = 1/2 Circular Ellipitical 3-lobe S 1.0 8 6
Λ=3
4
2 10–1 8 6 4 Λ=1
2 10–2 0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
e
FIGURE 7.41 Load capacity of symmetrically loaded gas bearings. (Taken from Pinkus, O., Trans. ASME, 80, 1958. With permission.)
7.6.1.2.2 Stiffness Characteristics As was done with load capacity and friction, the stiffnesses of the various bearings will be considered at identical Sommerfeld numbers for all the designs. This means that they will be evaluated at different values of S. This is pertinent from a practical viewpoint since the designer wants to know what the stiffness of the bearing will be under the given conditions of load, speed, etc. regardless of where the journal positions itself in the bearing clearance as a consequence of the imposed operating conditions. Since shaft displacement in different directions produce different responses, here the displacement will be considered in the direction of the load vector. Thus here the definition of the spring constant is given by K = (dF /de ), where F is the response force to a displacement along the load line. Table 7.19 and Figure 7.42 give the results. We see immediately the profound improvement in stiffness in the noncircular over the circular design. In the region of low eccentricities where instability usually occurs (low load, high speed) the value of the spring constants for the elliptical and 3-lobe designs are nearly an order of magnitude higher. 7.6.1.3 Special Design Several unorthodox configurations which have in the past been used on high-speed equipment, including automotive gas turbines, are bearings with grooved surfaces and foil bearings. Figure 7.43(a) and (b) show a herringbone grooved journal bearing and two versions of spirally grooved thrust bearings. In both designs, the bearing or runner surface consists of a lattice of grooves and ridges. From a hydrodynamic point of view, the geometry essentially consists of a series of step bearings, though, unlike with conventional steps, these are at an angle to the direction of motion. One of the achievements of such a design is that the fluid is being driven away from the edges of the bearing, minimizing side leakage and raising load capacity. By a proper orientation of the grooves the fluid can be pumped away from either the inner or outer periphery or from both edges. The advantages of these bearings also lie in the fact that, whereas an ordinary geometry
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-57
Table 7.19 Values of K = K /2µLN (C /R )3 m = 1/2
S
m=0 circular
1
0.365 0.162 0.0866 0.0381 0.365 0.162
2.8 5.4 12.8 87.2 5.5 6.8
3
Elliptical
3-Lobe
φL = 0
Optimum, φL
φL = 0
26.2 29.4 42.4 96.0 20.0 25.8
10.2 — 26.4 71.6 26.4 31.2
22.4 28.6 37.4 115.6 16.4 46.2
Optimum, φL — 20.0 26.4 152.6 — 40.4
Note: For L /D = = 1 Source: Pinkus, O., Trans. ASME, 80, 1958.
100 8 6 4 E
6
K (C/R)3 2mLN
10 8
L
K=
2
4 L/D = 1, m = 1/2, Λ = 1 2
C = Circular E = Elliptical; L = 3-Lobe
C
1.0 0.01
2
3
4 5 6 7 8 90.1
2
3
4 5 6 7 8 9 1.0
FIGURE 7.42 Spring constants for symmetrically loaded gas bearings. (Taken from Pinkus, O., Trans. ASME, 80, 1958. With permission.)
has poor stability characteristics for a concentric shaft position ( = 0), the herringbone bearing is superior to a conventional bearing at low eccentricities. These designs can, of course, be used also with liquid lubricants. 7.6.1.4 Hydrostatic Bearings 7.6.1.4.1 Thrust Bearings Hydrostatic gas bearings are subject to an instability called pneumatic hammer. It is therefore necessary that their recess volume be kept to a minimum. Referring to Figure 7.44 this means that r1 and δ are small. The entrance flow area 2πr1 h from the recess into the clearance will then become more restrictive than the orifice area (πd 2 /4) in which case the bearing is said to be inherently compensated. With an inherently compensated bearing the recess pressure po will equal the supply pressure ps and the drop (ps − p1 ) across
© 2006 by Taylor & Francis Group, LLC
7-58
Handbook of Lubrication and Tribology (a)
(b) Ridge Groove Groove
Shaft
Ridge Pumping from the outside to the center
Ridge Groove Ridge
Groove
Pumping from center outwards
FIGURE 7.43 bearing.
Special bearing designs. (a) Herringbone grooved journal bearing. (b) Spirally grooved thrust
ds
Ps, Supply pressure P0, Recess pressure P1, Inlet pressure Ps, Ambient pressure
d
Ps
Pa
h P0 P1
r2
Velocity boundary layers
r1
Entrance throat
r
X
L
P0 P1
d Velocity profile r1 r2
FIGURE 7.44
Hydrostatic gas thrust bearing.
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Principles of Gas Turbine Bearing Lubrication and Design
7-59
10.0
d ke
5.0
ed ok
ho
h nc
C
U
P
s=
1/2
2.0
)
1.0 P
s=
0.50
P
(
(g–1)ℜT m CDA0PS 2g m9=
10
s
5
=3
.5
0.20
P
s
=2
.5
= Ps
0.10
0
2.
= Ps
0.05
5
1. = Ps
Ps = 1.75
1
1.
25
1.
= Ps
0.02 0.01 0.01
0.05
0.2
0.5
2.0
FIGURE 7.45
100
(r2–r1)/L
1/2
( )
12mA0CD 2gℜT B= g–1 h3Pa
20
5.0
In (r2/r1)/p
Mass flow rate in a hydrostatic thrust bearing.
the bearing film is given by the compressible equation m = CD A0 ps
2γ (γ − 1)RT
pa ps
2 γ
−
p1 ps
γ +1 γ
(7.44)
where m is the mass flow of the gas in kg/sec; Ao is the entrance throat area 2π r1 h in m 2 ; T is the gas temperature in ◦ K; and γ is the ratio of specific heats. To avoid pneumatic hammer, gas thrust bearings must be designed with inherent Compensation, for a design with an (r1 /r2 ) ratio of 0.1. Figure 7.45 and Figure 7.46 provide appropriate design charts, given in terms of a parameter B defined as 12µAo CD B= h 3 pa
2γ RT γ −1
1 2
ln(r1 /r2 ) π
(7.45)
The chart in Figure 7.45 presents the dimensionless flow, m , for various values of p¯ = (ps /pa ) while Figure 7.46 gives load capacity. Chart 5 to 10 offers values of bearing stiffness K . To use the charts for design purposes one first used the known value of ps to determine B for the maximum obtainable stiffness from Figure 7.47. From Figure 7.46 one then determines the r2 that will carry the required load W. The value of (h/r2 ) is set within the limits of 0.5 × 10−3 < (h/r2 ) < 2 × 10−3 . With r2 and h established, the actual dimensional values of stiffness k and flow rates can be calculated from Figures 7.45 and Figure 7.47. Refinements are possible by a few more iterations.
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Handbook of Lubrication and Tribology
ke d Un ch ok ed
10.0
W9R =W/(pr 22ps)
Ps = 10
Ch o
5.0
2.0
Ps = 5
1.0
Ps = 3.5 Ps = 2.5
0.50
Ps = 2.0 Ps =1.75
0.20
Ps =1.5 0.10 Ps =1.25 0.05 Ps =1.1 0.02 0.01 0.01 0.02
0.05 0.1
0.2
0.5
B=
FIGURE 7.46
1.0
2.0
5.0
20
10
50
100
12mA0CD In (r2/r1) 2gℜT g-1 ph3Ps
√
Load capacity of hydrostatic gas thrust bearing.
Optimum orifice compensation
Un
Ch
ch
ok ed
r1/r2 = 0.1
P
s=
2.0
10
Optimum inherent compensation 1.0
P
s=
5
0.50 .5 =3 Ps
.5 =2 .0 Ps = 2 .75 P s = 1 .5 Ps =1 Ps
0.20 0.10 0.05
5 .2
=1 Ps
Stiffness factor, K9= – (h dW/dh)/(pr
2 2 ps)
5.0
ok ed
10.0
= Ps
0.02
1
1.
0.01 0.01 0.02
0.05 0.1
0.2 B=
FIGURE 7.47
0.5
1.0
2.0
5.0
10
20
12mA0CD In (r2/r1) 2gℜT g –1 ph3Pa
√
Stiffness of orifice and inherently compensated hydrostatic thrust bearing.
© 2006 by Taylor & Francis Group, LLC
50
100
Principles of Gas Turbine Bearing Lubrication and Design (a)
7-61
(b)
FIGURE 7.48 Two hydrostatic gas journal bearing geometries. (a) Inherently compensated. (b) Orifice compensated.
LD
10–1
Stiffness K =
2
d C* (1+1+2/3d 2 ) (P –P ) s a
1
10–2
Ps/Pa 20.0 10.1 1.1 1.5
1.0& 2.0
1.5 1.1
10–3 10–3
10–2
10–1
Restrictor coefficient Λsj =
1
101
102
3mnd2 √ ℜT (L/D) 2PsC √1+ d2
FIGURE 7.49 Stiffness of a hydrostatic gas journal bearing. (Taken from Reddickoff, J.M. and Vohr, J.H.,“Hydrostatic Bearings for Cryogenic Rocket Engine Turbopumps,” J. Lubr. Technol., 1969.)
7.6.1.4.2 Journal Bearings Typical configurations for journal bearings are shown in Figure 7.48(a) and (b). The former is an inherently compensated design, the latter has an orifice restrictor. Here, too, the recess must be small, or in the order of 10% of an incompressible fluid pocket. Capillary restrictors are not used because they cause pneumatic hammer. The analysis of these bearings is very complex and possible only with numerical methods. A typical set of performance curves for an L /D = 1/2 is shown in Figure 7.49 and Figure 7.50. The first shows stiffness as a function of a restrictor coefficient s for various ratios of (ps /pa ). This s is similar to the B parameter in thrust bearings for it represents the ratio of fluid film resistance to the resistance of the restrictor. The parameter δ appearing in the coordinates is defined as
δ = (d 2 /4hr1 )
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(7.46)
7-62
Handbook of Lubrication and Tribology 1
Flow, m
6mℜg jm pP 2s C 3
Ps/Pa = 20 1.5 1.25 1.1 10–1
10–2
10–3
20.0 1.5 1.25 1.1
10–3
10–2
10–1
101
1
Restrictor coefficient Λsj =
102
2√ℜT
3mnd (L /D) 2PsC √1 + d2
FIGURE 7.50 Flow in a hydrostatic gas journal bearing. (Taken from Reddickoff, J.M. and Vohr, J.H., “Hydrostatic Bearings for Cryogenic Rocket Engine Turbopumps,” J. Lubr. Technol., 1969.)
which gives the ratio of the throat area for an orifice restrictor to the throat area represented by the restriction of the bearing film. Its inclusion in the figures permits one to use these charts for both modes of restriction. The K values presented are the center stiffness of the bearing ( = 0). Since the stiffness remains essentially constant up to = 1/2 load capacity of the bearing can be calculated from W = CK ; or since it is not recommended that the bearings operate at higher eccentricities, the load capacity is given by W = 0.5 CK where K is obtained from Figure 7.49.
7.6.2 Compliant Surface Foil Bearings (CFB) In high-speed, high-temperature applications the CFB’s have the advantage of developing higher load capacities than with fixed geometry gas bearings. Since CFB’s have a complex structure consisting of a spring-like substructure with an overlying flexible surface, there exists a great variety of permutations on any given design. The presentation here will start with two basis models of a journal and a thrust bearing, to be followed with some more elaborate geometries. In all cases the lubricant will be that of air. 7.6.2.1 Foil Journal Bearing The solution of this bearing is based on Equation (7.39), coupled with additional expressions accounting for the elastic behavior of the top and bottom surfaces. The journal bearing is portrayed in Figure 7.51 and its solution is based on the following postulates. • The stiffness of the foil is uniformly distributed around the circumference and is linear with the amount of deflection. • The foil is assumed not to “dag” between bumps but to follow the deflection of the bumps. • In response to the hydrodynamic pressures the deflections are local, that is, they depend only on the force acting directly over a particular point. Under the above conditions the variation in h is due to the eccentricity e and the deflection of the foils. We then have h = C + θ cos(θ − φo ) + K1 (p − pa )
© 2006 by Taylor & Francis Group, LLC
(7.47a)
Principles of Gas Turbine Bearing Lubrication and Design (a)
fL-O
w
7-63
w
f L fL
uE
u ud v
b O O1 C fo
uo
Top foil
(b) Spacer block
Shaft
U
S
h
Bump foil
F
(c)
P t t Io
fF
fF
S
FIGURE 7.51 Configuration of a foil journal bearing. (a) Nomenclature for foil journal bearing. (b) Basic elements of bearing. (c) Configuration of bump foil.
where K1 is a constant reflecting the structural rigidity of the bumps, given by K1 = (p/pa );
α=
2pa s CE
3 lo (l − v 2 ) t
(7.47b)
K1 is then the compliance of the bearing with the quantities, s, lo , and t portrayed in Figure 7.51(c). 7.6.2.1.1 The Nominal Film Thickness In rigid journal bearings, the minimum film thickness is a clear and fixed quantity. It occurs at the line of centers and its value is constant across the axial width of the bearing. Also, generally, the film thickness
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7-64
Handbook of Lubrication and Tribology –(L/2)
Z=0
Journal
(L/2) hmin
huN
huominal huo
u = uN
u = uO
FIGURE 7.52
Minimum and nominal film thicknesses.
b = 120°, e = 0.6, uo = 220° (L/D) = Λ = 1, a = 5 U
)
z=
(L/2
uE
z=
0 uS
FIGURE 7.53
Film thicknesses in a 120◦ bearing pad.
anywhere is constant in the z direction (Figure 7.52). Since in our case pressures cause proportional deflections of the bearing surface, the film thickness in the interior of the bearing, where pressures are highest, will be larger than at the edges (z = ±L /2); also since the maximum pressures occur near the line of center, the film thickness is the interior of the θ = θo , the film thicknesses are larger than along another angular position θ = θN = where, because the pressures are lower, the film thickness, on the average, is smaller than at the line of centers. Figure 7.53 shows a three-dimensional film thickness plot for a 120◦ pad in which, while film thickness at the edge (z = 1/2) is small over most of the pad area, the surface has been deflected into much larger values of h. For these reasons a nominal film thickness hN (Figure 7.54) will be defined as the minimum film thickness that occurs along the bearing centerline, that is, at z = 0 at various values of α. While hmin for the rigid case occurs at θ = 180◦ , with increasing values of α the value of this hmin , or our hN , shifts downstream and increases in value; at α = 5, it is twice the value of the rigid case and has shifted downstream by nearly 100◦ . This should be kept in mind later on, when load capacity, that is, the W — hN
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Principles of Gas Turbine Bearing Lubrication and Design
7-65
1.4 1.2
10
hN
5 1.0
hN 1
0.8 h
hN
a=0
0.6 hN 0.4 0.2 0 40
L/D = 1 u0 = 180° z=0 80
Λ = 1.0 b = 180° e = 0.6
120 160 200 240 280 320 360 u, degrees
FIGURE 7.54 Location of nominal film thickness. (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.)
relation is plotted; an increase in load while increasing may also produce an increase in the nominal film thickness. 7.6.2.1.2 Active and Effective Bearing Arc Compliant foil bearings suffer a penalty in their ability to generate hydrodynamic pressures whenever the pad arc commences in a diverging region. The effect can be seen in Table 7.20 which show that by shifting in a 360◦ bearing the line of centers from 180 to 270◦ , there was a loss in load capacity of nearly 30% as well as a reduction in hN . In designing a foil bearing, if the eccentricity if fixed for the particular application, it is best to start the bearing at θ1 = φ (for a vertical load); if the eccentricities are liable to vary, some compromise value of θ1 = 0 can be chosen. 7.6.2.1.3 Performance Characteristics There are six geometric, structural, and operational parameters relevant to a foil journal bearing. These are β, α, (L /D ), , φL , and number of pads. There is also the eccentricity ratio and the attitude angle φ, the latter tied to the load angle φL . A set of standard conditions consisting of (L/D) = = α = 1;
= 0.6
is used, and any parametric variation commences from this set of reference values. 7.6.2.1.4 The Full Bearing Table 7.21 gives a detailed listing of the performance of a vertically loaded (φL = 0) full 360◦ bearing as a function of (L/D), α, and . Note should be taken of the fact that the start of the bearing, that is θs is so chosen as to avoid idle (p = ps ) regions at the upstream portion of the bearing. In effect, this requires that θs = φ. The case of nonvertically loaded foil bearings φL = 0, is given in Table 7.20. Some of the noteworthy points emerging from these tabulations are: • Effect of α: While in terms of , there is a drastic drop in load capacity with a more compliant bearing, in terms of hN there is actually an increase in load capacity.
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7-66
Handbook of Lubrication and Tribology Table 7.20 θ (deg)
Effect of Load Angle in 360◦ Bearings
φL (deg)
φ (deg)
θ2 (deg)
Pmax
hN
W × 102
T × 10
— 163.1 141.6 129.8 107.2 33.1 −11.9 −41.9 −72.9
— 4 11.6 20.2 27.2 33.1 28.1 23.1 17.1
— 32 74 95 142 240 272 300 325
0 1.018 1.087 1.133 1.201 1.253 1.245 1.228 1.201
0.40 0.41 0.48 0.52 0.58 0.62 0.62 0.61 0.58
0 1.04 11.5 21.4 40.0 56.8 54.9 50.0 40.1
— 43.9 38.1 33.9 27.2 24.6 23.7 23.7 25.5
0 20 45 60 90 180 220 245 270
Note: = 0.6; (l /d ) = = α = 1; θ1 = 0, θE = 360◦ Source: Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.
At large values of α, α > 10, the load the bearing can support is low, due to the fact that the flexible foil deflects sufficiently to maintain high film thicknesses even at large eccentricities. Thus from a design standpoint, it may be advisable to use high-compliance bearings at low loads; high loads, however, can be supported only with bearings of low values of α. In highly compliant bearings (particularly at high L /D ratios) an increase in eccentricity may produce an increase in hN , a phenomenon opposite to rigid bearings where hmin is the inverse of . • Effect of : The performance of a foil bearing as a function of conforms to the familiar pattern ˜ with an increase in , the load capacity, of compressible lubrication. After an initial rise in W ˜ both in terms of an increase in W as well as a rise in hN , tends to flatten off and approach an asymptotic value. The torque, however, rises almost as a linear function of the increase in . The more compliant bearing shows lower power losses due to the prevailing higher film thickness. 7.6.2.1.5 The Multipad Bearing The 3-pad design consists of three 120◦ arcs and the 5-pad design has five 72◦ arcs. In each case the vertical line of symmetry dissects the bottom pad, so that φL = 0 represents a load passing through the midpoint of the bottom pad (Table 7.22). Tables 7.23 and 7.24 give a spectrum of solutions for the performance of the 3-pad bearing and these results show the following: • Variation with load angle: Because of the cyclic nature of this bearing (symmetry for each 120◦ ) ˜ or T with a shift in load angle. In particular, there is no there is much less variation in either W acute loss of load capacity when the line of centers passes between pads. The optimum load angle for α = 1 is φL = −10◦ and for α = 5 it is φL = −14◦ . The improvement in load capacity over that of central loading (φL = 0) is of the order of 10 to 15%. • Variation with number of pads: Figure 7.55 shows the variation of 1-, 2-, and 3-pad bearings as a function of load angle. The plot shows clearly a drop in load capacity with the number of pads, that is, with a drop in extent of bearing arc β. As seen, the optimum for the 360◦ bearing occurs at φL = 0, at which point the torque also reaches it minimum value. The 3-pad bearing, as said previously, reaches an optimum at φL = −10◦ ; whereas the 5-pad bearing reaches an optimum at φL = −15◦ . 7.6.2.1.6 Stiffness Table 7.24 gives the values of the four spring coefficients for two values of compliance, the limiting case of α = 0 and α = 1. The α = 0 case differs from a rigid gas bearing in that the subambient pressures are eliminated from the pressure profile. A comparative evaluation of the stability characteristics of the 1- and 3-pad bearings is, of course, best done in a study of a rotordynamic system, particularly when the
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design Table 7.21
7-67
Performance of a 360◦ Foil Journal Bearing φ = θs
θ2 − γ
P max
hN
W × 102
T × 10
L/D = 0.5 0.3 0 1 5 10 20 0.6 0 1 5 10 20 0.9 0 1 5 10 20
63.5 59.0 48.5 60.0 34.0 40.0 36.0 32.0 30.0 27.0 12.0 19.0 21.0 21.0 21.0
81.5 87.0 100.5 114.0 114.0 62.0 76.0 97.0 105.0 114.0 52.0 71.0 91.0 99.0 108.0
1.046 1.043 1.037 1.025 1.017 1.25 1.144 1.073 1.05 1.033 3.73 1.33 1.12 1.077 1.048
0.70 0.72 0.80 0.85 0.94 0.40 0.51 0.68 0.795 0.90 0.10 0.41 0.64 0.76 0.91
4.4 4.2 3.2 2.9 2.1 17.9 13.7 8.3 6.1 4.2 157.3 34.7 14.8 9.8 6.3
9.34 9.15 8.53 8.09 7.54 15.85 13.93 8.19 7.33 6.54 26.1 13.8 9.8 8.5 7.3
L/D = 1.0 0.3 0 1 5 10 20 0.6 0 1 5 10 20 0.9 0 1 5 10 20
37.0 49.0 36.0 28.0 20.0 36.0 33.0 28.5 25.0 20.0 13.0 21.0 23.0 21.5 19.0
97.0 104.0 117.0 120.0 132.0 77.0 95.0 112.0 117.0 120.0 59.0 86.0 94.0 108.5 127.0
1.173 1.107 1.061 1.041 1.025 1539 1.253 1.114 1.074 1.046 4.850 1.434 1.154 1.103 1.063
0.70 0.77 0.94 1.04 1.14 0.40 0.62 0.90 1.055 1.22 0.10 0.52 0.86 1.05 1.26
27.9 23.7 14.8 10.3 6.37 94.9 56.8 28.8 19.4 12.2 504.5 102.8 42.9 27.8 17.2
22.7 21.2 18.6 17.5 16.5 31.1 24.6 19.1 16.9 15.0 58.2 28.1 19.5 16.7 14.4
L/D = 1.5 0.3 0 1 5 10 0.6 0 1 5 10 0.9 0 1 5
52.0 43.0 29.0 21.0 35.0 32.0 26.0 22.0 14.0 23.0 23.0
103.0 113.0 119.0 141.0 88.0 104.0 120.0 137.0 68.0 95.0 112.0
1.218 1.152 1.076 1.048 1.731 1.311 1.135 1.084 5.300 1.485 1.184
0.70 0.82 1.03 1.13 0.40 0.68 1.00 1.18 0.10 0.56 0.96
70.0 53.2 28.5 18.3 208.9 112.0 52.0 34.1 298.9 179.7 74.2
33.4 30.4 26.4 24.8 45.6 34.3 26.3 23.1 85.1 39.2 26.7
α
Note: = 1; φL = C Source: Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.
cross coupling components vary not only in magnitude but also in sign. However, the following items can be deduced from the tabulated data: ˜ the Kyy ’s are about the same for both designs, whereas the Kxx ’s are lower • When plotted against W for the 3-pad configuration.
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7-68
Handbook of Lubrication and Tribology Table 7.22
P Performance of a 3-Pad Bearing (L/D) = = 1, B = 120◦ θs
θL
φ
P max
W × 102
T × 10
40 210 217 220 225 245 273
−140.0 7.2 −2.3 −2.3 −5.2 −10.0 −17.4
79.2 37.2 27.0 27.7 29.8 55.0 77.6
1.073 1.072 1.075 1.079 1.082 1.088 1.077
12.1 12.7 13.7 14.1 14.6 15.0 12.6
21.4 21.8 21.9 22.0 22.0 21.5 21.4
29 180 200 2202 245 270 15
−145.0 44.1 −2.6 −10.6 −17.0 −25.5 −145.0
69.2 44.1 30.6 29.4 48.0 64.5 50.9
1.188 1.133 1.197 1.215 1.284 1.185 1.497
24.0 25.2 37.2 29.8 34.9 28.7 59.3
27.8 18.6 27.7 28.8 29.0 27.4 62.9
194 210 230 245 260
0.0 −10.8 −14.1 −14.7 −26.2
16.3 19.2 333.9 50.5 53.8
1.340 1.375 1.572 1.412 1.463
69.5 76.9 74.3 52.4 55.3
34.7 45.7 71.8 77.3 56.6
38 210 214 220 225 245 275
143.0 5.1 0.0 −5.3 −8.6 −13.3 −11.1
74.9 35.1 34.3 34.7 36.4 51.7 73.9
1.049 1.038 1.040 1.043 1.045 1.053 1.050
0.6
12 180 205 220 245 270
145.0 55.0 3.6 −9.7 −16.3 −23.1
67.4 55.5 28.6 30.3 48.74 66.9
1.104 1.048 1.077 1.103 1.121 1.106
15.7 12.6 16.1 18.3 18.4 15.9
25.4 16.5 25.3 26.7 27.2 25.4
0.9
25 200 203 210 230 245 260
164.0 3.5 0.0 −4.3 −16.3 −18.9 −21.5
61.0 23.5 23.0 23.7 33.7 46.1 58.5
1.178 1.122 1.119 1.158 1.198 1.202 1.186
24.4 25.3 26.3 28.1 29.7 18.2 25.2
46.5 34.4 36.2 41.6 67.1 72.1 53.7
α=1 0.3
0.6
0.9
α=5 0.3
Source:
0.01 7.52 7.87 8.31 8.60 9.07 8.18
20.7 21.1 21.1 21.2 21.2 20.8 20.7
Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.
• With the more compliant case, the K ’s tend to level off with a rise in eccentricity, the values of the coefficients approaching the structural stiffness of the system. In general, the advantage of the compliant bearings in the area of stability lies in that levels of stiffness can be selected by the designer via a proper combination of structural and hydrodynamic stiffnesses. Thus, instead of making his inertias and supports suit the inherent stiffnesses of purely hydrodynamic bearings, the designer may try to tailer and adjust bearing stiffness to the demands of high rotordynamic system.
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Principles of Gas Turbine Bearing Lubrication and Design Table 7.23
7-69
Mode of Load of 3-Pad Bearing
α
φ
W
φL
φ
W
1 1 1 5 5 5
0.3 0.6 0.9 0.3 0.6 0.9
37.5 28.5 16.0 35.0 30.0 23.5
13.8 36.0 68.0 7.8 17.0 26.2
−10 −10 −10 −14 −14 −14
55.0 29.0 18.5 53.0 41.0 30.0
15.0 39.8 78.0 9.0 18.6 29.8
Note: (L/D) = λ = 1; β = 120◦ each Central loading φL = 0; Optimum loading Source: Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of GasLubricated Foil Journal Bearings,”ASME Paper 82-LUB-40, 1982.
Table 7.24 Values of Spring Coefficients
α
= 360◦ 0.6 0 0.75 0 0.9 0 0.6 1 0.75 1 0.9 1
φ
W
Kxx
Kxy
Kyx
35.7 24.1 12.8 32.1 26.3 21.4
0.951 1.894 5.055 0.568 0.7833 1.028
1.920 3.416 7.202 1.129 1.231 1.268
−0.125 −1.166 −6.024 0.174 0.0254 −0.098
−2.345 −3.989 10.151 −0.693 −0.686 −0.627
3.237 8.981 44.593 1.130 1.378 1.602
0.635 1.321 3.695 0.359 0.511 0.689
1.123 2.102 4.728 0.5702 0.673 0.759
−0.092 −0.752 −3.344 0.0451 −0.017 −0.057
−2.05 −3.710 −8.768 −0.758 −0.821 −0.855
2.635 7.432 37.103 0.801 1.051 1.274
3-pad-120◦ each 0.6 0 26.0 0.75 0 17.4 0.9 0 8.6 0.6 1 25.5 0.75 1 20.5 0.9 1 16.3
Kw
Note: (L/D) = ; φL = 0 Source: Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.
0.80
L/D = 1 a=1
0.70
Λ=1 e = 0.6 1 Pad —360° 3 Pads —120° 5 Pads —72°
0.60
hN
0.50 0.40 W 0.30
W T=10
0.20 W
0.10 0 180
140
100
60
20
–20
–60
–80
FIGURE 7.55 Performance of multipad bearings. (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.)
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7-70
Handbook of Lubrication and Tribology (a)
(b)
U h2 h1
U Pressure distribution
Original surface
h2
h1
R2
L R1
bb B u
hN
b
Deflected surface
FIGURE 7.56 The geometry of compliant surface thrust bearing. (a) Nomenclature of the thrust bearing. (b) The elastohydrodynamics of a complaint foil bearing. Table 7.25
Effect of β on Bearing Performance
h1
α◦
◦
β
n
W × 102
T × 102
p max
W TOT × 102
2.5
4
1.2
3.0
3-1/3
1.0
4.0
2.5
0.75
30 45 60 75 30 45 60 90 30 45 60 90
12 8 6 5 12 8 6 4 12 8 6 4
0.259 0.430 0.573 0.686 0.257 0.435 0.588 0.808 0.238 0.415 0.574 0.814
4.19 6.09 8.02 9.98 4.13 5.97 7.81 11.58 4.01 5.82 7.56 11.08
1.0288 1.0320 1.0329 1.0331 1.0293 1.0334 1.0349 1.0349 1.0285 1.0336 0.0359 1.0373
3.12 3.44 3.44 3.30 3.08 3.48 3.53 3.23 2.86 3.32 3.44 3.26
Note: (L/R2 ) = 0.5; b = 0.5 Source: Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.
7.6.2.2 Thrust Bearings Figure 7.56 shows the configuration of the thrust bearing considered next, which resembles that of a conventional tapered land design. All the postulates stated in connection with the journal bearing apply here as well, except, of course, that the film thickness is different. This is now given by h = h2 + g (r, θ ) + C(p − pa )
(7.48)
where (h1 − h2 )[1 − θ/bβ} g (r, θ) = 0
for 0 ≤ θ ≤ bβ
for bβ ≤ θ ≤ β
7.6.2.2.1 Geometric Optimization As shown in Table 7.25 the maximum pressure changes vary little with an increase or decrease in β or the value of h1 . The table also shows that when the total number of pads possible for a given value of β is accounted for in the calculation of the total load capacity, there is little difference in the choice of a
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Principles of Gas Turbine Bearing Lubrication and Design
7
(L/R2) = 0.5 Λ = 3.0b = 0.8 a = 1.0 h1 = 2.0 W =10
(b) 10
3.5
9
W TOTAL
3
8
6
2.5
W Pad 102
7 5
W
6
WToT =102
(a)
7-71
2
5 4
4
3 2
3
1 20° 30° 40° 50° 60° 70° 80° 90°
L/R2 = 0.5 b = 0.5 Λ = 1.33 a = 4.0 h1 = 2.5 hu =1.31= constant b.Degrees 30 40
50
60
70
80
90
FIGURE 7.57 Effect of β on pad and total load. (a) Effect of β in stiff bearings. (b) Effect of β in soft bearings. (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Thrust Bearings,” ASME Paper 82-LUB-39, 1982.)
particular arc length. The effect is shown in Figure 7.57 for both a relatively stiff and soft bearing. In both cases the nominal film thickness does not vary with β. The total load capacity in a full 2π thrust bearing shows a maximum somewhere between 45 and 50◦ . The effect of the proportion of ramp is shown in Figure 7.58. For a relatively stiff bearing the optimum is about 70%, which is close to the case of rigid bearings. For the more compliant case the optimum is 50%. For high values of , the optimum value of b recedes to values closer to 40%. It can be seen that near the highest values of W the nominal film thickness is also at its highest, thus reinforcing a general load capacity optimum for compliant bearings at about b = 0.5. ˜ changes little with a The effect of varying h1 is given in Table 7.26. It shows that while the value of W ˜ variation of h1 , the nominal film thickness goes up appreciably, doubling in value for a doubling of h1 . Thus, in terms of customary load capacity criteria, the highest values of h˜ 1 seem desirable. To summarize, the optimum geometry for a bearing with the common OD to ID ratio of 2 is β = 45◦ ,
b = 0.5,
h˜ 1 > 10
7.6.2.2.2 Performance Characteristics The performance of a CFB for the optimized parameters of β = 45◦ and b = 0.5 are given in Table 7.27 for a wide range of parameters h1 , α ◦ , and ◦ . In terms of h2 normalization the range of extends to nearly 1000 and that of α to over 60. Figure 7.60 is a performance plot in terms of the basic variables involved in the bearing. The drop of load capacity with an increase in film thickness and with a decrease in the value of are trends known from other studies of gas bearings. What is particularly noteworthy in Figure 7.59 is the effect of α on load capacity. While at moderate ’s high values of α yield the highest load capacity at high , the optimum α is some intermediate value, in our case α ◦ = 1. Note should ◦ are all normalized by the geometric ramp height δ; be taken that all quantities, that is, ◦ , α ◦ , and hN and that the Figure 7.59 plots contain implicitly various values of h1 . The relation between h˜ 1 and h˜ N is given in Figure 7.60. This graph can be used to determine various ramp heights for different points of Figure 7.59. The h˜ 1 − h˜ N graphs support the conclusion of the previous section as to the desirability of using high values of h˜ 1 since they yield high nominal film thicknesses. It also shows that the higher α ◦ and ◦ are, the higher the film thickness.
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7-72
Handbook of Lubrication and Tribology 10 (L/R2) = 0.5 b = 45° h1 = 2.0 a* = 1.0 Λ* = 3.0
9 8 7
T W × 103 T × 102
6
5 W 4 3
1.1
2 hN
hN 1.0
1 0
0.2
0.4
0.6
0.8
1.0
b
FIGURE 7.58 Effect of extent of ramp of performance of CS thrust bearing. (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Thrust Bearings,” ASME Paper 82-LUB-39, 1982.) Table 7.26
Effect of h 1
h1
W × 102
T × 10
hN
10 15 17 19 20
8.169 8.626 8.730 8.811 8.845
44.5 72.2 83.3 94.5 100
5.106 7.675 8.730 9.730 10.265
K × 103 114 116 116 117 122
Note: β = 45◦ , L /R2 = 0.5, b = 0.5, α a = 1.0, a = 10. Source: Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Thrust Bearings,” ASME Paper 82-LUB-39, 1982.
Finally, Figure 7.61 gives a plot of the spring constant for the bearing. The stiffness of the bearing is, of course, a function of both the structural stiffness as represented by KB and of the hydrodynamic film stiffness. Since they are in parallel, high loads would tend to flatten the values of K for the softer bearings, leaving essentially the structural stiffness KB as the dominating spring constant. 7.6.2.3 Advanced CFB Designs The construction of CFBs lends itself to a number of modifications that can enhance a particular performance characteristic in accordance with operational requirements. Due to their analytical complexity and space limitations they cannot be gone into to any extent but a mere listing of some of them will give an idea of the range of possibilities latent in this group of bearings.
© 2006 by Taylor & Francis Group, LLC
Compliant Foil Thrust Bearing Performance α=0 W × 102
a hN
h1 = 2 0.1 1.0 10.0 20.0 40.0
0.014 0.144 0.632 3.33 6.246
h1 = 5 0.3 1.0 10.0 20.0 40.0
α=1
T × 10
K × 103
W × 102
a hN
1.0 1.0 1.0 1.0 1.0
0.024 0.241 2.40 4.79 9.50
0.314 3.24 38.6 76.4 125.0
0.104 0.145 1.52 2.736 4.238
0.192 2.343 22.2 34.66 46.4
0.25 0.25 0.25 0.25 0.25
0.35 3.50 33.5 65.2 127.0
12.2 171.0 1230.0 1730.0 2060.0
h 1 = 10 0.1 1.0 10.0 20.0
0.66 8.818 60.38 86.84
0.111 0.111 0.111 0.111
1.63 16.1 150.0 292.0
h 1 = 20 0.1 1.0 10.0 20.0
2.084 22.35 134.8 185.5
0.0526 0.0526 0.0526 0.0526
6.60 64.6 605.0 1179.0
a
α=4
α = 20
T × 10
K × 103
W × 102
a hN
T × 10
K × 103
1.001 1.011 1.114 1.197 1.276
0.024 0.24 2.27 4.34 8.21
0.314 3.20 31.5 45.5 51.5
0.014 0.147 1.214 1.828 2.406
1.004 1.046 1.326 1.461 1.574
0.024 0.23 2.04 3.78 7.05
0.32 3.35 17.5 18.8 16.4
0.204 1.833 6.633 8.135 9.231
0.268 0.378 0.635 0.71 0.758
0.341 2.72 17.3 30.8 57.0
14.0 77.7 104.0 95.7 78.0
0.206 1.164 2.697 3.362 3.612
0.316 0.538 0.905 0.978 1.015
0.31 2.07 12.9 28.6 44.9
91 1,120 6,120 8,240
0.661 2.170 8.109 9.457
0.161 0.307 0.566 0.630
1.30 7.81 44.5 79.5
61.1 113.0 114.0 95.6
0.571 2.309 5.345 6.020
0.192 0.38 0.695 0.785
575 4,720 26,600 34,200
1.103 3.876 8.845 10.24
0.16 0.282 0.54 0.596
3.54 18.3 100.0 179.0
90.0 128.0 117.0 97.5
0.843 2.705 5.674 6.280
0.162 0.364 0.666 0.722
a
a
W × 102
a hN
T × 10
2.3 29.8 26.1 19.7 14.7
0.152 0.513 0.854
0.452 0.834 1.15
0.24 1.43 10.2
5.24 6.92 3.64
9.915
1.18
38.8
1.16
1.15 6.50 37.5 68.0
40.5 64.8 56.7 45.0
0.464 1.619 3.323 3.627
0.234 0.485 1.825 0.88
0.98 5.38 32.3 59.8
25.0 35.6 25.0 18.5
2.91 14.8 83.6 152.0
52.2 70.6 56.2 44.0
0.623 1.83 3.467 3.733
0.209 0.404 0.790 0.835
2.39 12.1 71.7 113.0
29.7 37.8 24.3 17.7
a
K × 103
Principles of Gas Turbine Bearing Lubrication and Design
Table 7.27
Note: β = 45◦ , b = 0.5, (L/R2 ) = 0.5
7-73
© 2006 by Taylor & Francis Group, LLC
7-74
Handbook of Lubrication and Tribology 100
b = 45* (L/R2) = 0.5 b = 0.5 a* = 0 a* = 1 a* = 4
10
— W × 102
Λ* = 20 1
Λ* = 1
0.1
Λ* = 0.1
0.01 0
0.2
0.4
0.6
0.8
1.0
1.2
1.4
h*N
FIGURE 7.59 Load capacity of CS thrust bearing. (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Thrust Bearings,” ASME Paper 82-LUB-39, 1982.)
15 b = 45* b =0.5
13
(L/R2) = 0.5
20 10
a* = 1 a* = 4
11
20
Nos. refer to Λ*
10
hN
9 1 7 1 5 0.1
3 a = 0 All Λ*
1 1
3
5
7
9
11
13
15
17
19
21
h1
FIGURE 7.60 Relation between h1 and hN . (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Thrust Bearings,” ASME Paper 82-LUB-39, 1982.)
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-75
10 b = 45* (L/R2) = 0.5 b = 0.5 a* = 0 a* = 1 a* = 4
5 2 1 5 2
Λ* = 20
10–1
0.1
5 Λ* =
2 –2
=1
10
Λ*
5 2 10–3 5 2 10–4
— W –4
10
10–3
10–2
10–1
1
FIGURE 7.61 Stiffness of CFB. (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Thrust Bearings,” ASME Paper 82-LUB-39, 1982.)
7.6.2.3.1 CFBs With Variable Direction Stiffness The bump foil design can be manipulated to provide a wide range of desirable dynamic properties. One such arrangement to vary the stiffness in both radial and circumferential directions is shown in Figure 7.62. The stiffness gradient permits the formation of a variable hydrodynamic wedge in accordance with variation in load or speed. As speed increases the particular arrangement can be made to increase film convergence which enhances stability precisely when it is needed. In one such application an advanced air-lubricated journal bearing reached speeds of 132,000 rpm carrying a unit load of close to 100 psi. Much of the above applies also to thrust bearings. A proper thrust bearing geometry is one that has a taper in the circumferential direction followed by a flat portion. A CFB can be constructed with a foil possessing stiffening elements at the trailing edge, as shown in Figure 7.62(b). The stiffening elements placed between the top and bump foils provide a variable stiffness gradient from the leading to trailing edge yielding the desired converging shape. 7.6.2.3.2 CFBs with Controlled Coulomb Damping A foil bearing can be constructed to improve internal damping and thus enhance its stability characteristics. This can be done by affecting the Coulomb damping due to the relative motion between the top and bump foil surfaces, as well as between the bump foil and the housing (see Figure 7.63). This relative motion occurs, of course, when the bearing is loaded and the foils are radially deflected. To improve the friction characteristics of this relative motion the rubbing surfaces are sputter coated with copper, silver, or some other high friction material. Sometimes the surfaces of the journal and mating top foil are coated with dry lubricants to minimize friction on start-up and shut down. To further enhance stability the bump foil
© 2006 by Taylor & Francis Group, LLC
7-76
Handbook of Lubrication and Tribology (a)
W
W + DW
(b)
Multilayered top foil
Runner
Variable pitch supports
FIGURE 7.62 Variable support stiffness in compliant bearings. (a) Variable radial stiffness. (b) Variable longitudinal stiffness.
Top foil
Dry film coating
Copper coating
Relative motion under detormation
Bump foil Housing
FIGURE 7.63
Mechanism for Coulomb damping in foil journal bearing.
can be circumferentially split along axial lines to improve alignment and axial compliance. A single pad design of this variety carried unit loadings up to 100 psi. 7.6.2.3.3 CFBs With Cantilevered Leaves A bending dominated foil bearing is the cantilevered leaf type design. A journal bearing of this type is shown in Figure 7.64(a). Here each leaf is preformed with a specific radius to induce a desirable film profile. The thrust bearing variety is shown in Part b. It is essentially a thin plate with springs mounted below it which is separated by a backing plate from a foil assembly plate to which the leaves are attached. The plates are in the form of annular rings and the leaves in the form of annular sectors. The leaves should be thin enough to permit considerable flexibility but still thicker than the hydrodynamic film.
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design (a)
Journal
(b)
7-77 Thrust Foil segment removed
Foil assembly
Foil housing Assembled foil segments
Blacking plate Foil segment removed
Spring assembly Foil
FIGURE 7.64
Foil
Garrett self acting CFB.
7.6.3 Magnetic Bearings To illustrate the potential of magnetic bearings, a specific design example will be followed through. It will provide orientation and a guide for general cases where their use is contemplated, particularly with regard to controlling instability and resonances at high speeds. 7.6.3.1 General Principles Practical MBs are mostly the attractive type. Radial AMBs generally adopt an 8-pole stator configuration as shown in Figure 7.65. Both the stator and journal consist of laminations of ferromagnetic material. The journal is shrunk on a shaft without windings. The use of laminations reduces eddy current, which not only causes power loss, but degrades the dynamic performance of the bearing. The stator poles are separated into four quadrants. In each quadrant, the electromagnetic windings are wound in such a way that the magnetic flux circulates mainly inside the quadrants so that each quadrant of poles can be controlled independently. Magnetic force is proportional to the ratio current to air-gap squared (Figure 7.66). To support a load in a controlled axis (Figure 7.67), unequal steady state or bias currents are induced in opposite pairs of poles, such that W = f (I12 − I32 )/C 2
(7.49)
where I1 > I3 f = a magnetic pole constant for a given number of windings. The bias current produces an I 2 R loss, which is a major power loss in an AMB. However, the total resistance in the current path is not large; the AMB loss is, in general, insignificant compared to fluid film bearings. The journal floating in the magnetic field due to its bias currents alone is stable. Linearized
© 2006 by Taylor & Francis Group, LLC
7-78
Handbook of Lubrication and Tribology A quadrant Current Windings
Flux path
Air gap Shaft
Laminated journal
Laminated stator
FIGURE 7.65 An 8-Pole configuration of an active magnetic bearing.
Mafnetic force
Air gap decreases
Designed air gap
Air gap increases
Linearized feedback current operating point
Bias current
FIGURE 7.66 Nonlinearity of magnetic force. (Taken from Allaire, P.E., Li, D.F., and Choy, K.C., J. Lubr. Technol., Trans. ASME, July 1980. With permission.)
feedback control is achieved by making the air gap large relative to the journal’s vibrations. For stable operation the journal motion must be sensed and corrected instantaneously and continuously by superimposing a small control current to each bias current. For example, as shown in Figure 7.67, when the journal moves up by a small displacement, the current in the top quadrant will be reduced to a small amount, i, and the bottom quadrant increased by i. The control currents produce a net downward force, F , which pulls the journal back to the center. From sensing Y to producing F , a series of AMB components are involved, namely sensor, controller, power amplifier, and electromagnets.
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-79
Measured journal displacement Bias Ib1 adjustment
Power amplifier No.1 Ib1 – i Stator
Air gap Y Controller X Shaft –1 Journal
Stator Bias Ih3 adjustment
Ib3 + i
Power amplifier No.2
FIGURE 7.67 An independently controlled axis.
7.6.3.2 AMB Components 7.6.3.2.1 Electromagnets For a given maximum load including static and dynamic loads, the AMB physical size is determined by the saturation flux density of the lamination material. For the 8-pole configuration, Fmax = 5.75 × 105 Ap Bs2
(7.50)
where, Fmax = maximum load, N Ap = surface area per pole, m2 Bs = saturation flux density, W/m2 The maximum value of the flux density in the linear range which is about 90% of the actual Bs value should be used in applying Equation 7.50. Choosing the axial length Lp , the circumferential pole width is Ap /Lp . The radial dimensions can be determined from a given shaft diameter at the AMB. For sizing, the following guidelines should be followed: • • • •
The cross-sectional area at any point of the flux path is not less than Ap . Adequate wiring space is provided. The axial length is no greater than the journal OD. As a rule, the air gap should be ten times the expected journal vibration.
The pole surfaces are the most effective areas for heat dissipation by convection. The ampere-turns per pole is fixed for a given ferromagnetic material; the optimal choice of winding turns, Mt is a trade-off of total current and inductance load, L, to the power amplifiers. The latter is proportional to Mt2 ApC which is a crucial parameter causing control delay and bearing instability. More than 8 poles can be designed for the stator, such as 16 or 24 evenly spaced. A large number of poles saves radial space because it localizes flux circulation. The coil pairs can be in series or parallel to a power amplifier with the same trade-off.
© 2006 by Taylor & Francis Group, LLC
7-80
Handbook of Lubrication and Tribology
7.6.3.2.2 Power Amplifiers Converting a low power control voltage signal to a high power control current and actuating the electromagnets requires power amplifiers. Two types are commercially available, the linear and the pulsewidth-modulation (PWM) type. The linear amplifier applies the control signal to a power transistor in an “active mode.” The transistor continuously regulates the current through the windings from a DC source, Vs , with the current directly proportional to the control signal. The PWM type applies the control signal to generate high voltage pulses at a fixed frequency above audible range. The on-time period of each pulse is proportional to the input signal. The voltage pulse train produces current to the windings. The PWM type is electrically noisy and needs its own filters. The power transistors operate in a “saturation mode” with much less power loss. There are three requirements for the power amplifiers. First, the control current, i, cannot be larger than the bias current. Second, the inductance of the electromagnets causes the control current to diminish and delay above a certain frequency (cut-off frequency). The PWM amplifiers usually apply their own current feedback to increase this frequency. Third, the value of Vs /L, called the current slew rate limit, is the maximum amperes per second that the amplifier can provide. 7.6.3.2.3 Sensors Three displacement sensors prove to be practical, the capacitance probe, inductance, and eddy current probe. Each varies with its advantages and disadvantages, but all relate the small distance between the stationary sensor and the rotating shaft to an output electrical signal in volts. A low-pass filter is usually included in the sensor conditioning device to eliminate high frequency noise, including its own FM carrier. This filter, similar to the power amplifier cut-off characteristics, may cause a significant time delay in the frequency range of interest. A phase-lead circuit implanted in series in the feedback loop can reduce the delay. An inexpensive and reliable sensor is not yet available for measuring the journal velocity. Different analog circuits, such as a differentiator with a low-pass filter, and phase-lead circuits have been used to produce a pseudo velocity from the displacement measurement. An analog surrogate called a Velocity Observer, instead of differentiating displacement, integrates journal force (equivalent to acceleration) to obtain velocity. The output of any pseudo velocity circuit is a combination of displacement and velocity signals. Thus, its feedback not only produces damping, but also contributes to the stiffness. 7.6.3.3 AMB Stiffness and Damping From the previous discussion, a practical single axis control can be represented by the block diagram of Figure 7.68. A radial AMB needs two independently controlled axes like this, while a thrust AMB needs only one. A second-order, low-pass filter was assumed to be part of the sensor though it could have been a fourthorder or some other type of filter. Gp is the sensor sensitivity; for example, 1000 V/in. (40 V/mm). A phase-lead circuit is applied in series here for compensating the time delay caused by the inductance loads to the power amplifiers. One may set the phase-lead parameter “a” to be equal to the amplifier cut-off frequency, ωn . Thus a system “zero” cancels a system “pole” in the 3-plane. This does not improve the current slew rate of the amplifier, but does increase the damping-to-stiffness ratio around ωn . The other phase-lead parameter “b” is set in the range of a ≤ b ≤ 10a. The AMB stiffness and damping of this controlled axis can be calculated by using the equations below with S equal to jω.
© 2006 by Taylor & Francis Group, LLC
−F /Y = K + jωb = −Ki (i/Y ) − Km
(7.51)
i/Y + (TS )(TC )(Tp )(Ta )
(7.52)
Principles of Gas Turbine Bearing Lubrication and Design
7-81 Journal motion at AMB CL
Adjustable PID gains Cd
Y
Gpvc2
Yp
S2+√2vcS + vc2 Measured Sensor journal displacement low-pass filter
vo S + vo
Q
S S + vv
Z
Km
–
Ce
– b S+a × a S+b –
E
Phase-lead circuit
Gava S + va Power amplifier
i
+
+ F
Ki
Electromagnets
Cv
Controller Rotor/AMB force coupling equation F = Kii + KmY AMB state equations i ’ + vai =GavaE E ’ + bE = – (b/a)[CdYp’ +CeQ’ +CvZ ’ +a(CdYp +CeQ +CvZ )] Q’ + voQ = voYp Z ’ + vvZ =Vp Yp’ =Vp Vp’+√2vcVp + vc2Yp =Gpvc2Y
FIGURE 7.68 A single-axis control diagram.
where ω is the excitation frequency and TS = Ga ωn2 /(S 2 −
√
2ωnS + ωn2 )
TC = −Cd − Ca ωo /(S + ωo ) − CV S /(S + ωV ) Tp = (b /a )(S + a )(S + b ) Ta = Ga ωn /(S + ωn ) The equations indicate that both K and B are functions of excitation frequency ω, not rotational speed. Numerical results are plotted in Figure 7.69 using the AMB data. The frequency axis in this plot is normalized with respect to 50 Hz which is the average of two rigid-body critical speeds of a rotor. The amplitude is normalized with respect to (Gp Ca Ki Cd − Km ). At the low-frequency range where the integral control dominates, the plot shows negative damping values. This should not cause alarm, however, since mechanical system resonances seldom exist in that low range. At the high-frequency range, especially where the first two bending criticals exist, negative damping can cause resonances. This is discussed later. 7.6.3.4 Rotor-AMB System Dynamics 7.6.3.4.1 System Design Guidelines Active Magnetic Bearings are generally less stiff than rolling element or hydrodynamic oil-film bearings. Therefore, the first two system criticals have relatively rigid mode shapes, and their vibrations are easily controlled. The third and fourth criticals with bending mode shapes must be given careful consideration in high-speed turbomachines.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
Lp = 2.0 in. (50.8 mm) D = 2.5 in. (36.5 mm) C = 0.020 in. (0.5 mm)
50,000
Ap = 1.25 Nt = 100 turns Fmax = 200 lb (890 N) I1 = 3.5 A I3 = 2.0 A
Ki = 80 lb/A (356 N/A) Km = 12,500 lb/in. (2.19 × 106 N/m) Cd = 0.26 Cv = 0.6 Cc = 1.0 Gp = 1000 V/in. Gz = 1 A/V vc = 5000 Hz vo = 1 Hz vv = 500 Hz va = 500 Hz
Dynamic stiffness (lb/in.)
in.2(8.06 × 10–4 mm2) √(K 2 + v2B2) K 0
vB
–50,000 1
10 100 Excitation frequency (Hz)
1,000
FIGURE 7.69 AMB stiffness and damping a numerical example.
Taking the rotor model in Figure 7.70 as an example, its critical speed map shows that the rotor operates between the third and the fourth criticals. Two identical 8-pole AMBs are chosen to support the rotor with dimensions in Figure 7.69. The first issue is finding the best method for determining the stiffnesses. In this case, the stiffness per bearing can be made 1000 lb/in. or 10,000 lb/in. (1.75 × 105 N/m or 1.75 × 106 ) N/m). The answer depends on rotor shock load. To take 1 g shock, this rotor of approximately 100 lbs (45 kg) moves radially 50 mi and 5 mi (1.25 mm and 0.125 mm) respectively, for the lower and higher stiffnesses. The catcher bearing is set at 10 mi (0.25 mm) away from the rotor for a designed air gap of 20 mi (.5 mm). To avoid pounding the catcher bearing when shocked, the higher stiffness is chosen. Reviewing the mode shapes at the chosen stiffness, Figure 7.71 reveals that there are sufficient relative displacements at the bearings for control of the first and second modes. The third and the fourth modes are lacking the displacement at one bearing. To help control the third mode, the displacements sensor is mounted at the outboard side of each AMB where the sensor sees more than only the center motion. The second design issue is to determine how many bending modes should be controlled. To keep the control electronics relatively simple, the frequency range with acceptable control response is limited by two factors, the inductance load and the filtering delay. It is imperative to have an adequately damped bending mode below the operating speed (the third mode in this case), because of the unbalance excitation during traversing the critical. The bending mode immediately above the operating speed (the fourth mode in this case) should be 15 to 20% away in frequency. However, it still can be excited by harmonics as the rotor is going up in speed, or by a shock load. But, less damping is required for controlling this mode. The higher bending modes normally are less likely to be excited. The rotor material damping is a source to resist the minor or occasional excitation. Oil-film bearings always provide positive damping but there is no guarantee of this for AMBs. The control current at the high critical frequency may lie behind the displacement measurement or the probe may be at the wrong side of the AMB. The AMB may become a small exciter for that mode. When it happens, a band-reject filter for the excitable mode can be inserted in series in the feedback loop to block the control at that modal frequency. For the example herein, the power amplifier and the sensor low-pass filter are assumed to have the cut-off frequencies at 500 and 5000 Hz, respectively. Applying the normalized stiffness and damping of Figure 7.69, the normalized frequency of 1.0 to 50 Hz, which is between the first and the second criticals can be read from the map. The values of ωB/K for the lower four modes range from 0.2 to 0.6, which is adequate for properly designed vibration modes. More damping can be achieved by increasing the value of Cv .
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design (a)
7-83
15 10 Coupling
Thrust disk
Inch
5
Wheel
AMB No. 1
AMB No. 2
0 –5
–10 –15 0
5
10
15
20 Inch
25
30
35
40
(b) 100,000
Inch
4th
Operating speed 15,000 rpm
10,000 3rd
2nd 1st 1,000 1,000
AMB stiffness √(K 2 + v2B2 10,000
100,000
1,000,000
Inch
FIGURE 7.70
Characteristics of the rotor model (a) Rotor model. (b) Rotor critical speed map.
7.6.3.4.2 Rigorous Dynamic Analysis After component sizing and cursory analysis, rigorous system analysis is needed to prove the expected rotor vibration behavior. Considering the fact that the AMB stiffness and damping are functions of excitation frequency, the state vector of the conventional rotor model is extended to include the state variables of the AMBs. The mathematical rotor model is the same as the conventional model including sections of shaft with specified ID, OD, length, and concentrated masses and inertias. The model for each bearing would be the bering station number, the measurement station number, and the key parameters of Figure 7.69. Using this electromechanical model, the lower four damping frequencies of the rotor running at 15,000 rpm were computed and are presented in Table 7.28. The first three modes are adequately damped because the associated log decrement values are all significantly above 0.4. The latter is a damping value generally accepted for a rotor system supported in oil-film bearings. The fourth mode has a log decrement value of 0.06 without considering the rotor material damping. It should be acceptable since it is much higher in frequency that the third mode and the operating speed. In Table 7.28 the cross-coupling stiffness produces a destabilizing seal effect at the wheel, but it only affects the first mode damping. The reason for this is that only the first mode shape has significant lateral displacement at the wheel. The rotor/AMB system can sustain a value of 4000 lb/in. (7 × 105 N/m) before becoming unstable.
© 2006 by Taylor & Francis Group, LLC
7-84
Handbook of Lubrication and Tribology AMB No. 1
1.0
AMB No. 2
0.8
Relative amplitude
0.6 0.4 0.2 –0.0 –0.2 –0.4 –0.6 –0.8 –1.0 0
FIGURE 7.71
5
10
15 20 25 Rotor length (in.)
30
35
40
Rotor critical mode shapes. Table 7.28
Damped Natural Frequencies of Forward Modes Kxy = −4000 lb/in.(−7 × 105 N/m)
Kxy = 0 Mode 1st 2nd 3rd 4th
Frequency (cpm)
Log decrement
Frequency (cpm)
Log decrement
2,318 5,263 11,154 34,156
1.05 2.20 0.89 0.06
2,374 5,623 11,144 34,156
−0.00 2.19 0.88 0.06
Note: Rotor speed = 15,000 rpm; Cross-coupling stiffness (Kxy ) at wheel.
Figure 7.72 presents an unbalance response at the third critical speed using the same electromechanical model. It indicates that the response peak is well damped and far away from the operating speed of 15,000 rpm. The peak dynamic current was calculated to be 0.45 (0-peak) at 11,500 rpm. It specifies that a current slew rate no less than 500 A/sec must be provided by the power amplifier design.
7.7 Design Considerations In practice a designer must obtain quantitative data to ascertain on the one hand whether the bearing will meet his operational requirements; and on the other hand find out what the power losses, flows, temperatures, etc. will be to properly plan the layout of the facility. In Sections 7.2 to 7.6 the graphs and tables offered values for the performance of various bearing designs. These, however, do not exhaust the information required for rational design. What is needed is some orientation as to how the various geometrical and operational parameters affect bearing operation and how to go about improving or even optimizing a given bearing design. Within their restricted space the following paragraphs should offer some guidance as to how to go about approaching this task.
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-85
1.40
Response (m, O-peak)
1.20
1.8 gm-cm at thrust disk 1.8 gm-cm at coupling –3.6 gm-cm at wheel
1.00
AMB No. 1
0.80 0.60 0.40 AMB No. 2 0.20 Wheel 0.00 7.00
8.00
9.00
10.00 11.00 12.00 Speed (1000 rpm)
13.00
14.00
15.00
FIGURE 7.72 Unbalance response at third critical speed. (Taken from Allaire, P.E., Li, D.F. and Choy, K.C., J. Lubr. Technol., Trans. ASME July 1980. With permission.)
7.7.1 Performance Parameters The expressions required for calculating the more important items of bearing performance are the following: Film thickness: For an aligned journal the film thickness is given by h = (h/C = 1 + cos(θ − φ))
(7.53)
The attitude angle φ is defined as the angle between the line centers — a line passing through the centers of bearing and journal — and the load vector. When the treatment is restricted to vertical loads, φ denotes the angle between location of hmin and the vertical and therefore the importance of φ lies in that it determines the location of hmin . Sommerfeld number (load parameter): The Sommerfeld number, given by µN S= P
2 R C
(7.54a)
has traditionally been the most important parameter. However, a more convenient quantity is the inverse of S, here called the Load Parameter, given by
W =
P µn
2 2 W C C = R LDµn R
(7.54b)
where P = (W /LD) is the unit loading. What this parameter says is that any combination of P, µ, N , C, and R such as to leave the value of W unchanged, would result in the same bearing eccentricity ratio, , and attitude angle, φ.
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Handbook of Lubrication and Tribology
0.1
Hydrodynamic lubrication regime
Boundary lubrication
Low loads
f
0.001
High loads
S = (mN/P )(R/C)2
FIGURE 7.73
Behavior of friction coefficient in fluid film bearings.
Minimum film thickness: is given by
This is the smallest distance between the journal and bearing surfaces and it
h min =
hmin = (1 − ) C
(7.55)
What is normally referred to as load capacity relates to the load, W , which this hmin can support. Friction coefficient: This is the ratio between the frictional force and bearing load. It is normally expressed in the form of R (R/C)FT f = C W
(7.56)
The general shape of f as a function s is given in Figure 7.73. The region of sudden rise in f denotes the limit of hydrodynamic lubrication, followed by a regime of “boundary lubrication” characterized by partial contact between the mating surfaces. Power loss: This, of course, can be obtained from the value of Fτ , namely H = FT · Rω = f · W · Rω H=
H Ho
=
(7.57)
H [π 3 µN 2 LD 3 /c]
The quantity by which H is normalized, represents the power loss in an unloaded concentric journal bearing, that is, one in which = 0. It is know as the Petroff Equation. Flow: An amount of lubricant, Q1 , enters the bearing at the leading edge; an amount, Qs leaks out the two sides of the bearing (one-half Qs at each side), and an amount Q2 leaves the trailing end of the pad. In most cases, since a journal bearing extends over 2π , Q2 is not discharged outside but reenters the next oil groove, so that the net amount of lubricant to be made up from an outside source is Qs . The latter is referred to as side leakage. Clearly we must always have Q1 = QS + Q2
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
7-87
All of these flows are given in dimensionless form as Q=
Q (π/2)NDLC
(7.58)
the denominator representing the flow in an unloaded, concentric bearing, that is, at = 0 (for which case QS = 0 and Q1 = Q2 ). The above flows, Q1 , QS , and Q2 are what may be called hydrodynamic flows induced by the shearing action and pressure gradients of the fluid film. QS is the minimum amount of oil to be delivered to the bearing so as to maintain a full fluid film with all its potentialities. In practice designers supply more than this required minimum, using a supply pressure ps < pa . The effect of the supply pressure, usually of the order of 10 to 30 psig, can be ignored as far as bearing hydrodynamics are concerned. Temperature rise: A bulk temperature rise can be estimated from the values of power loss and side leakage, namely T = (Tav − T1 ) = Dynamic coefficients:
H cp wQs
(7.59)
The dimensionless stiffness is given by K = (K /2µNL ) (C /R )2
while the damping coefficient reads B = (π B /µL )(C /R )3 from which the dimensional values of K and B can be obtained. The coefficients , f , H , Q 1 , Q 2 , K , and B which serve to evaluate bearing performance are obtained from solutions of the Reynolds Equation for the specific geometries and operating conditions of the various bearing designs.
7.7.2 Bearing Configuration The behavior of a bearing is naturally a function of its geometry. However, even for a given design there are a number of variables that will affect its performance. Among the more known parameters are the L /D and C /R ratios and the degree of preload. Of the less familiar ones one can cite load orientation, the geometry of the oil grooves, or the relative proportions of a bearing’s geometrical elements. 7.7.2.1 Journal Bearings Although one often hears about the use of full, that is 360◦ arc bearings, it is very rarely that such sleeves are employed in machinery. Most journal bearings consist of two or more pads separated by horizontal oil grooves making them in fact partial bearings, used either singly or in tandem. The number and distribution of these angular pads on bearing performance is one of the more important considerations in bearing design. 7.7.2.1.1 Partial Bearings Whenever a single pad of an angular extent β < 2π is used, it is called a partial bearing. When β is very small its load capacity is low, as illustrated in Figures 7.74 and Figure 7.75. However, soon a limit is reached at about β = 140◦ beyond which no further gains are registered. The reason for this asymptotic behavior is due to oil cavitation at the trailing end of the pad where the pressures decrease close to or even below ambient pressure. This, if a partial bearing is used there is no need to go beyond a 140◦ arc. The effect of temperature in partial bearings is a combination of two phenomena. The higher the arc the longer the
© 2006 by Taylor & Francis Group, LLC
7-88
Handbook of Lubrication and Tribology W
O Or
100
1 Dimensionless load capacity, W
10 1/2 1/4
1
1.0 1/2
1/4 0.10
e = 0.2 e = 0.8 Nos. refer to (L/D) 0.01
FIGURE 7.74
20
60
100 140 180 Bearing arc b°
200
360
Effect of bearing arc on load capacity.
dissipation path and the higher the temperatures; however, a longer arc produces thicker films and thus less heating. Consequently, as shown in Figure 7.76, a crossover point occurs; at high loads low values of β are preferred, if low T s are desired; at low loads a longer arc is preferred. 7.7.2.1.2 Grooved Bearings Partial bearings are not used extensively. The most common designs are grooved bearings which consist of a number of pads arranged in tandem by cutting axial oil grooves in the 360◦ circumference. There is a great variety of such designs, the most common being a 2-pad bearing with two grooves at the horizontal split. Others may have 3, 4, or 6 grooves forming the same number of individual pads. The more the number of grooves the lower the load capacity, as shown in Figure 7.77 and 7.78. Thus, if load capacity is the primary objective, a 2-groove bearing is best; however, those with a larger number of grooves are somewhat more stable. Related to the above is the fact that any hole or disruption in the bearing surface will reduce the load capacity. Figure 7.79 shows the effects on the pressure profile of cutting a slit or circular hole in the loaded part of a bearing. The larger the incursion, the more drastic the reduction in the hydrodynamic pressures which translates directly into reduced load capacity. 7.7.2.1.3 Tilting Pad Bearings The primary characteristic of this family of bearings is that the individual pads are not fixed but are pivot supported so that during operation not only does the journal move but so do the pads and each in a different fashion. A general picture of a tilting 3-pad bearing is shown in Figure 7.80. The structural and
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Principles of Gas Turbine Bearing Lubrication and Design
7-89
1.0 0.9 Minimum film thickness (hmin/C)
b = 360 0.8 b = 180
0.7 0.6 0.5 0.4 0.3
b = 120 0.2 b = 60
0.1 0 0.001
0.01
0.1 Load capacity, (mNP) (R/C)2
1.0
10.
FIGURE 7.75 Effect of bearing arc on value of hmin . (Taken from Pinkus, O., “Manual of Bearing Failure and Repair in Power Plant Rotating Equipment,” EPRI, July 1991. With permission.)
103 8 6
L/D = 1
4
Temperature rise (wc D T/P)
2 102 8 6 4
b = 60° b = 120°
2 10 8 6 4 2
b = 180° b = 360°
1
0.01 0.02
0.06 0.1
0.2
(m N/P)
0.6 1.0
2.0
6.0 10.
(R/C)2
FIGURE 7.76 Effect of bearing arc on temperature rise. (Taken from Pinkus, O., “Manual of Bearing Failure and Repair in Power Plant Rotating Equipment,” EPRI, July 1991. With permission.)
© 2006 by Taylor & Francis Group, LLC
7-90
Handbook of Lubrication and Tribology fL = 0 +fL –fL L/D = 1 e = 0.6
10 9
1
2
3
Load capacity, W
1 8 7 2
6 3
5 4 3 2 1 0 –60 –50 –40 –30 –20 –10 0
10 20 30 40 50 60 70 80 90 100 110 120 Load angle fL°
FIGURE 7.77 Load capacity of grooved bearings. (Taken from Pinkus, O., J. Lubr. Technol., Trans. ASME, Oct. 1975. With permission.) 1.0 0.8 0.7 0.6 0.5
drica
cylin
ial cy
e
4 Ax
Plain
0.3
5 = 0. L/D 5 brg, = 0. ove , L/D l gro brg Axia ical 1 r d L/D = ylin brg, in c Pla rical n li =1
0.4
l brg
0.2
, L/D
0.1 0.01
0.1
1.0 S
FIGURE 7.78
Comparisons of 2- and 4-axial bearings.
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10
Principles of Gas Turbine Bearing Lubrication and Design
7-91
1.0 L/D = 1/2; e = 0.5
L/D = 1/2; e = 0.5
No slot
No hole 0.05 L
5 L
L0
d 0.05
0.20 Extent of hole at C.L. Mean axial pressure, P
0.60
7.2°
0.40
Values (L/L0)
0.02
0.04
1.0
0.5 0.60
v
0.80 Values (d/L)
Load 54°
90°
Pmax 0.1 126°
162°
54°
199°
90°
126° Position from hmax
Position from hmax
FIGURE 7.79 Effect of a slot and a hole on hydrodynamic pressure.
Pivot Pivot clearance circle
Pad R-Cm e
v
R
Shaft
bp
FIGURE 7.80 A tilting 3-pad journal bearing.
© 2006 by Taylor & Francis Group, LLC
b
162°
199°
7-92
Handbook of Lubrication and Tribology Table 7.29 Relative Load Capacity and 3-Pad and 5-Pad Bearings Values of m On-pivot load
Load between pivots
W
3 Pads
5 Pads
3 Pads
5 Pads
20 40 60 80 100
0.72 0.79 0.80 0.82 0.83
0.82 0.87 0.90 0.92 0.93
1.42 1.58 1.60 1.61 1.62
0.97 1.05 1.09 1.10 1.11
Source: Pinkus, O., J. Tribol., Oct. 1986.
analytical complexities of these bearings are more than compensated by their great reliability and the fact that they have no rival in their stability characteristics. The number of possible design parameters and operating modes in a tilting pad bearing is very large. Some of them are discussed below. 1. Number of pads: Table 7.29 gives a comparison of a 3-pad vs. a 5-pad centrally pivoted bearing having zero preload. When the load is in line with the pivot the 3-pad design has a higher load capacity but the reverse is true when the load direction is between the pads. For loads of engineering interest the 5-pad design consumes less power. 2. Pivot location: In order to assure two-directional rotation and for ease of assembly, most tilting pad bearings are centrally pivoted. However, a 10 or 15% displacement of the pivot in either direction would not significantly alter the general performance, a slight preference being a downward shift. 3. Preload: From many standpoints a high preload is desirable. Its effect on preventing the scraping of the top pads has been discussed previously and from this standpoint a m of at least 0.5 is required. High preloads also yield higher stiffness and damping. However, the penalty is that the film thickness over the pivot and often also the absolute hmin is reduced. Likewise, the power losses and temperatures rise with an increase in preload. 4. Mode of loading: In general the shaft eccentricity will be lower when loaded over the pivot. It is characteristic of tilting padis bearings that regardless of whether the load vector is over the pivot or between the pads the locus of shaft center is along a vertical line, which has a direct beneficial effect on stability. Results for the two modes of loading on stiffness and damping are given in Figure 7.81 for a bearing of zero preload. As seen, both the spring and damping coefficients are lower for the between-pads mode of loading. 7.7.2.1.4 Oil-Ring Bearings As pointed out previously, oil-ring bearings operate under starved conditions. It is thus the main task of the designer to find ways to increase as much as possible the amount of oil delivered to the bearing surface. Some of the important parameters that play a role in accomplishing it are geometry shape of contact surface, weight, the material, and the size of the ring relative to the shaft. In an experimental study a series of rings portrayed in Table 7.30 was tested with the purpose of both increasing the flow of lubricant and of extending the regime of stable ring operation. The conclusions reached were as follows: 1. An optimum ring shape is one with a quasi-trapezoidal cross section and a series of straight teeth at the contact surface shown in Table 7.30 as Ring No. 2. 2. The best ring material is bronze with a weight of 23 N per meter of ring circumference. 3. For bearing diameters in excess of 6 in. dual rings are recommended. 4. An anchored spring leaf inserted between the ring and journal raises the amount of oil delivery and extends the ring’s region of stable operation. One such stabilizer is shown in Figure 7.82.
© 2006 by Taylor & Francis Group, LLC
103
10
1
102
1
Cv Bxx /W
10
CKxx /W
10–1
Cv Byy /W
CKyy /W
1
10–2
10–3
10–1 10–1
1 10 S = (m NDL/W) (R/C)2
102
CWMCR/(mLD(R/C)2)2
Critical mass
102 CK/W; Cv B/W
10
CK/W; Cv B/W
103
7-93
Critical mass Cv Bxx /W
10
CKxx /W
1
10–1
10–2
CWMCR/(mLD(R/C)2)2
Principles of Gas Turbine Bearing Lubrication and Design
10–3
10–1 10–1
1 10 S = (m NDL/W) (R/C)2
102
FIGURE 7.81 Effect of mode of loading on bearing stability in a 4-pad tilting pad bearing. (a) Load between pivots. (b) Load over pivots.
7.7.2.1.5 Load Angle Bearing loads are usually directed midway of a pad or between grooves. However, improved performance can be obtained by shifting the load vector toward the trailing edge of the bearing pad. A comprehensive mapping of the effects of shifting the load vector around the circumference of a 2-groove bearing is shown in Figure 7.83. Normally the load would be straight down, that is along φL = 0. However, as seen in the figure by moving the load toward the trailing edge, improved performance is obtained for the entire range of bearing operation. At low loads an optimum occurs at a load of φL = 10; at high loads the value is some 30. The lowest load capacity would occur at a load angle of 60◦ from the midway point. Supplementary data is given in Table 7.31 where it is seen that the worst angular position results in a load capacity reduction of 70.70%. Similar data for a 3-groove bearing has been given in Figure 7.77. Achievement of an optimum bearing position requires no special effort. It is sufficient to rotate the bearing in the housing the required 10 to 30◦ to obtain this. Attention should only be given to the oil delivery path since now the oil grooves would no longer be at the horizontal split. This can be taken care of by cutting a short oil-supply channel on the outside of the bearing shell. 7.7.2.1.6 Misalignment It was pointed out in an earlier section that an overhung impeller will cause bearing misalignment. A full treatment of misalignment would exceed the available space here but the qualitative consequences should be pointed out. As shown in Figure 7.84 in severe misalignment the journal at one end may find itself in the upper half of the bearing even though the load is downward. As a consequence, a fluid film and hydrodynamic pressures may develop in both the lower and upper portions of the bearing. Stretching from the end where the hydrodynamic film is at the bottom, this film will wrap itself in helical fashion around the entire bearing circumference. In all cases the load capacity, that is the value of h for the imposed load, will be drastically reduced. 7.7.2.2 Thrust Bearings Unit loads in thrust bearings are higher than in journal bearings and consequently their hmin will be smaller. But it should also be realized that, except for a bearing with a flat at the end, hmin in thrust bearings occurs not along a line as in journal bearings but at a point, namely the outer downstream edge
© 2006 by Taylor & Francis Group, LLC
7-94
Handbook of Lubrication and Tribology Table 7.30 Oil ring configuration
Ring No.
T Split T-Section Brass Ring Made from Rolled Stock Fastened Together
1
3.4
3.4 C
11.7
A
7.35
E
19.9 "
Split Trapezodial Section Machined from Bronze SAE660 "=30E,$b=0E
2
Unit Weighta (N/m)
Cross section (mm)
Description
3.2
1.3
14.3
D
23.6
C
2.4
E
B 25.4
Split Trapezodial Section Machined from Bronze SAE660"=30E,$b=90E
3
1.7 24.9 B $
Split Trapezodial Section Machined from Bronze SAE660 "=30E,$b=45Eor135E
4
3.0
1.7 24.3 B 3.0
$ Split and Relieved, Trapezodial Section Machined from Bronze SAE 660 "=30E,$b=0E
5
aUnit
1.5
18.0 B L
22.9
5.2
weight equalsring weight/circumferential length. $ is measured from the direction of ring rotation.
bAngle
Source: Heshmat, H. and Pinkus, O. “Experimental Study of Stable High-Speed Oil Rings,” J. Trib. Trans. ASME, 107, 1985, 14–22.
of the pad. This point is also where Tmax will occur and again it will be higher than in journal bearings. This is due to the smallness of hmin but also to the higher linear velocities of the runner at the outer radius of the pad. 7.7.2.2.1 Tapered Land Bearings A conventional tapered land bearing was shown in Figure 7.31. There are three parameters here; the taper (h1 − h2 ), the pad arc β, and the (L/R2 ) ratio. The angular extent also determines the number of pads in
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Principles of Gas Turbine Bearing Lubrication and Design (a)
7-95
Ring
NR
Oil film Journal N
(b)
Ring backward swing Ring stable position
A
Ring
Ring forward swing
N
View A–A
FIGURE 7.82 Configuration of oil-ring stabilizer. (a) Spring stabilizer. (b) Stabilizer and ring positions during oscillations. Table 7.31 Effect of Load Angle on Load Capacity in Conventional 2-Groove Bearing Worst Condition
W at Worst φL W at φL = 0
L /D
W at φL = 0
φL
W
0.5
0.6 0.95
3.1 83
50 55
1.05 14.5
0.32 0.175
1.0
0.6 0.95
8 115
52 60
2 20
0.25 0.17
a thrust bearing. Table 7.32 shows the results of an optimization study giving the values of (h1 − h2 ) and β for the entire range of (L /R2 ) ratios. From this an optimum set of design parameters can be obtained for a particular application. It is worth noting that in general the optimum configuration is that which yields nearly square bearing pads. An improved version of a plain tapered land bearing is one with a flat surface at the trailing end, as shown in Figure 7.32. The additional merit of this design is that upon starting and stopping the runner rides on a flat surface reducing wear. Here a new parameter is the ratio of the tapered to the flat portion. The plot in Figure 7.85 shows such a variation from 60 to 100% taper, the latter being the tapered land bearing discussed previously. The load capacity peaks at a taper value of about 80% of the pad arc, that is the tapered portion should be four times that of the flat. Interestingly the value of (power loss/load capacity) achieves a minimum at the same point. 7.7.2.2.2 Misalignment In properly operating thrust bearings the load carried by each pad is the same. When the shaft and consequently the runner is misaligned, this is no longer true and some of the pads are much more heavily
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7-96
Handbook of Lubrication and Tribology
1000
Location of uG1 500 groove
Central loading
Location of uG2 0.986 groove 0.98
200
0.95
100 0.9
Load capacity, W
50
0.8 20 10
0.6
L/D = 1
5
+fL –fL
e = 0.4 2
Load 1
v uG1
0.5
uG2
0.2 0.1 –80 –60 –40 –20
FIGURE 7.83
0
20 40 60 80 Load angle fL, degrees
100
120
140
160
Effect of load angle on load capacity in 2-groove bearing.
Table 7.32 Optimum Pad Arrangement |25| L /R2 1/3
1/2
2/3
h1 /δ0
β, deg
Number of pads
1 1/2 1/4 1/8 1 1/2 1/4 1/8 1 1/2 1/4 1/3
<30 <30 35 40 40 45 50 60 50 60 80 >80
>10 >10 9 8 8 7 6 5 6 5 4 4
Source: Pinkus, O., Trans. ASME, Ser.D., 81, 1959.
loaded than the other. A pictorial representation of this situation is given in Figure 7.86. As seen, the loads carried by the heavily loaded pads as well as their maximum temperatures can be ten times as high as the ones located opposite them where the runner is furthest from the pads. The values of h in the two sets of pads will be of the same ratio. The span of severity of bearing operation goes up with the number of pads used in the misaligned bearing. Thus, if misalignment is expected one should not use more than four to six pads.
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Principles of Gas Turbine Bearing Lubrication and Design
7-97
F2
O
O⬘2
O⬘1
F1 N F2 Viewed from 2
Hydrodynamic film
Hydrodynamic film
F1
FIGURE 7.84
Hydrodynamic forces and films under misalignment.
7.7.2.2.3 Hydrostatic Bearings In a conventional hydrostatic bearing portrayed in Figure 7.4 the load capacity is given by W =
π R22 (po − pa )[1 − (R1 /R2 )2 ] 2n (Re /R1 )
There are therefore two parameters that determine the level of W ; (po − pa ) and (R2 /R1 ). The variation of load with these quantities is shown in Figure 7.87. As seen no optimum for load capacity occurs; it rises with p and drops with a rise in (R1 /R2 ). A minimum occurs in the power loss but power loss in a hydrostatic bearing is not of great concern and, when it is, it is due not to bearing geometry but to the onset of turbulence in the fluid.
7.7.3 Qualitative Guidelines In selecting design parameters it must be kept in mind that the choice often depends on the size of the bearing. Small bearings, less than 2 in. in diameter, can tolerate relatively lower values of hmin , higher unit loads of P, and operate close to isothermal conditions, whereas larger bearings require larger values of hmin , lower values of P, and tend to run close to adiabatic conditions. On the other hand (C/R) ratios must be higher for small bearings. With this as an introduction Table 7.33 gives some typical design practices in the field of journal bearings. The performance characteristics of one’s design should fall somewhere within the range of values listed in the table.
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7-98
Handbook of Lubrication and Tribology 150 D(DT )max 140
130 W = 0.0286 120 DW 110 DH 100
90 D Qs
D H/D W 80
70 Tapered land thrust bearing 60
b = 32°; (L/R2) = 2 h11 = 3; d21 = 0;
50 Rei = 700; E = 3.3 60
FIGURE 7.85
70
80 b%
90
100
Effect of extent of taper or flat. Table 7.33
Typical Design Limits for Journal Bearings
Minimum film thickness Temperature rise Maximum temperature Loads L/D Ratio C/R Ratio Preload, m Bearing arcs Inlet oil temperature, T
0.001–0.01 in. (0.0025–0.25 mm) Up to 80◦ F (27◦ C) (on babbitt) Up to 300◦ F (150◦ C) (on babbitt) 500 psi (3.4 Mpa) 0.25–1.0 0.001–0.002 0.25–0.75 150–60◦ for fixed pad 80–30◦ for tilting pad 80–130◦ F (27–55◦ C)
Nearly all the bearing data given in the present write-up are for bearings operating under laminar conditions. Should turbulence set in, the operating characteristics will change. One may expect turbulence when the bearing Reynolds Number reaches a value between 750 and 1500. The higher the Reynolds Number the more intense will be the effect of turbulence. Table 7.34 shows what will be the impact of the turbulent regime on the major items of bearing operation. In a more comprehensive way Table 7.35 provides a guide in which direction design modifications should head in order to ameliorate unsatisfactory results in a chosen design. Finally, Table 7.36 offers a cursory look at the relative advantages and disadvantages in choosing journal bearings of different designs.
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Principles of Gas Turbine Bearing Lubrication and Design
7-99
Pressure protites in 12-pad bearing: ax = 0 .75, ay = Q, E = 12.7 9
Multipad thrust bearings E =1.27 ax = 0.75 ay = 0 n = 12 Pmax/P n=6
8
Multipad thrust bearings E =12.7 ay = 0 ax = 0.75
3.6
7 3.2 6
n = 12
2.8
DTmax
n=8
2.4
5
2.0 WT
4
n=6
1.8 3 1.2 2 Q2
0.8
hmin
0.4
1 M
0
0 0
p u Performance of individual pads in multipad thrust bearing
FIGURE 7.86 permission.)
2p
0
Bearing pad number
n
Pad loadings in multipad thrust bearings
Effects of misalignment in thrust bearings. (Taken from Pinkus, O. ASME, Series D, 83, 1961. With
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7-100
Handbook of Lubrication and Tribology
Q
Q H
W
W
H
0
1 (R1/R2) h
R1
R2
FIGURE 7.87
Performance of incompressible hydrostatic bearing.
Table 7.34 Effect of Turbulence on Bearing Performance Reynolds Number a Parameter Re = ρRw h /µ Regimes • Re < 750 Laminar • 750 < Re < 1,500 Transition • Re > 15,000 Turbulence Item
Effect
Load capacity Oil flow Power loss Temperatures Stiffness and damping
⊕ ⊕ ⊕ ⊕ or
⊕ — Increase; — Decease
7.8 Advanced Bearing and Seal Applications Driven by the goals of increased power density and higher performance, advanced turbomachinery is operating at ever higher speeds and temperatures. Some of these advanced designs will require equally advanced bearing systems that do not suffer from the temperature limitations of liquid-lubricated bearings or the limited life of REBs when operating at extreme conditions. One such oil-free bearing, which has been applied successfully in advanced, high-speed machinery, is the air-lubricated, CFB as described in Section 7.6.2 (see, e.g., [38–41]). A recent summary of progress in the state of the art for foil bearings is presented in Reference 42. These bearings have been applied to a wide range of high-speed machinery, with operating environments ranging from cryogenic to temperatures in excess of 1500◦ F and speeds in excess of 700,000 RPM.
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Principles of Gas Turbine Bearing Lubrication and Design Table 7.35
7-101
Effect of Design Parameters on Journal Bearings Performance
Code: (+) means increase magnitude of parameter to achieve effect in left-hand column. (-) means decrease magnitude of parameter to achieve effect in left-hand column. Example: To decrease temperature rise, one or more of the following can be done: decrease L/D ratio, increase C/R, decrease oil viscosity, etc. Objective
L/D
C/R
To decrease temperature rise (−) (+) To reduce power loss (−) (+) or (−) (+) (+) To reduce Tmax To increase oil flow (+) (+) To improve stability (−) (−) To increase load capacity (+) or (−) (−) To avoid turbulence (+) or (−) (−) Stability (+) (−)
Geometry
Viscosity
Preload, m
Arc, β
(−) (−) (−) (−) (+) (−) (+) (+)
(−) (−) (+) or (−) (+) (−) (−) (−) (+)
Elliptical Circular Elliptical Elliptical
(−) (−) (−) (−) a (+) or (−) Circular (+) 3-lobe (+) Tilting pad (+) or (−)
Supply Oil Pressure, ps (+)b (−) No effect (+) No effect No effect (−) No effect
a The stability of a journal bearing increases in the following order: circular, pressure, elliptical, 3-lobe, tilting pad. b Apparent effect only.
Table 7.36
Characteristics of Various Journal Bearings Journal Bearing Summary Table
Bearing type
Advantages
Disadvantages
Axial groove
1. Easy to make 2. Low cost
1. Subject to oil whirl
Elliptical
1. Easy to make
3- and 4-lobe (tapered land, etc)
2. Low cost 3. Good damping at critical speeds 1. Good suppression of whirl
1. Subject to oil whirl at high speeds 2. Load direction must be known
2. Subject to whirl at high speeds
Pressure dam (Single dam)
2. Overall good performance 3. Moderate cost 1. Good suppression of whirl 2. Low cost
2. Dam may be subject to wear or build up over time 3. Load direction must be known
Hydrostatic
Tilting pad
3. Good damping at critical speeds 4. Easy to make 1. Good suppression of oil whirl 2. Wide range of design parameters. 3. Moderate cost 1. Will not cause whirl (no cross coupling) 2. Wide range of design parameters
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1. Some types can be expensive to make properly
1. Goes unstable with little warning
1. Poor damping at critical speeds
Comments Round bearings are nearly always “crushed” to make elliptical or multi-lobe Probably most widely used bearing at low or moderate speeds
Currently used by some manufacturers as standard bearing design
Very popular with petro-chemical industry. Easy to convert elliptical over to pressure dam
Generally high stiffness properties used for high precision rotors
2. Requires careful design 3. Requires high pressure lubricant supply 1. High cost
2. Requires careful design 3. Poor damping at critical speeds 4. Hard to determine actual clearances 5. High horsepower loss
Widely used bearing to stabilize machines with subsynchronous nonbearing excitations
7-102
Handbook of Lubrication and Tribology
Khydro
Hydrodynamic component of stiffness
Kstruct Compliant component of stiffness
Bhydro
Bstruct
FIGURE 7.88
Hydrodynamic component of damping
Compliant component of damping
Foil bearing operating mechanisms.
As shown in Figure 7.88, the essential feature of such a bearing is its twofold mechanism of imparting stiffness and damping to the system. One is via the geometry and materials of the complaint support structure. The other is due to the hydrodynamic film between journal and top foil. The compliant support can consist of one or more corrugated bump foils, which offer great flexibility in obtaining dynamic characteristics geared to a specific system. Additional flexibility is afforded through construction as a single top foil, or a multipad bearing with multiple top foil segments. The complaint construction allows the foils to be forced radially outward as speed increases, forming a converging wedge. This converging wedge becomes more pronounced with increasing speed and load, thereby increasing load capacity. The compliant foil surface readily accommodates itself to rotor centrifugal and thermal growth, as well as thermal and mechanical deformations of the bearing housing. These bearings have also demonstrated good performance under shock loading [43]. From a historical perspective there has been a considerable advance in CFB technology over the last decade. The capabilities of advanced designs now meet the requirements for advanced applications, with unit loadings in excess of 100 psi (689 kPa) [40] for journal bearings, 85 psi (586 kPa) for thrust foil bearings, and adequate damping to allow successful operation above the first system bending critical speed [44–46]. However, the issue of the scalability of early designs, which were typically 25 to 50 mm, to the much larger sizes required by turbomachinery such as gas and steam turbines had not been widely addressed until the late 1990s. The largest bearing discussed in the open literature, which was located by the authors, is the 89 mm diameter bearing described in Reference 38. This bearing was presented as having an ability to support a steady-state operating load of 23.5 lb (109 N), and a dynamic load of 335 lb (1490 N). These loads were well below that required by even most small gas turbine engines. However, recent efforts by Heshmat et al. 40–46 have been completed that have substantially extended the applicability of foil bearings to bearing sizes and loads that are substantially higher than previously demonstrated. In particular, static loads of 950 lb (4200N) have been supported by a 100 mm diameter foil journal bearing operating at a speed of 22,000 rpm. Figure 7.89 shows the measured and predicted load-carrying capability of two modern fourth generation foil journal bearings as a function of speed
© 2006 by Taylor & Francis Group, LLC
Principles of Gas Turbine Bearing Lubrication and Design
2600
7-103
150 mm bearing
2400 2200 2000
Applied load (lb)
1800 1600 1400 1200 1000 800 100 mm bearing 600 400 200 0 0
FIGURE 7.89
5
10 15 20 Speed (krpm)
25
30
35
Load-carrying capability of 100 and 150 mm diameter foil journal bearings.
capable of supporting static loads in excess of 2500 lbs. Figure 7.90 shows that a new class of oil-free foil thrust bearings has also been developed that are capable of operating at surface velocities above 1200 ft/sec (speeds to 80,000 rpm for the specific design tested) and has demonstrated a load capacity of over 80 psi1 . This demonstrated performance is more than three times the capacity of available thrust foil bearings just a few years ago. The features that have made it possible to significantly enhance the load-carrying capability of our compliant foil thrust bearings are: 1. 2. 3. 4. 5.
A hydrodynamic pressure profile which is more uniform over the thrust pad surface. Spatially variable stiffness to accommodate axial motion and misalignment of the rotating part. Enhanced thermal management capabilities. Unique surface coatings, which can accommodate thermal demands. Light weight and low parts count for high frequency response and reliability.
These tremendous recent advances point to the very real viability of applying foil thrust bearings in gas turbine engines. Figure 7.91 and Figure 7.92 are examples of very large compliant foil journal and thrust bearings that are being fabricated today. Besides high-load capacity, additional developments have been made in the high-temperature operating of foil bearings due to improved high-temperature coatings. As seen in Figure 7.93, a 150 mm diameter foil journal bearing is being operated at temperatures above 1200◦ F and speeds above 20,000 rpm. At these operating conditions the shaft journal expansion exceeds 0.030 in. The inherent design features of the compliant surface of the foil bearing are such that it readily accommodates this substantial rotor growth. Figure 7.94 shows potential application locations 1 Where
load capacity is defined as the maximum load carried by the bearing divided by its projected area.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology G G G
80
Achievement 2004 70 Achievement 2002 60
Load capacity (psi)
50 Bearing OD 3.82 in. Coating: Korolon 800 40
30 Achievement 2001 20 Baseline 1994 10
0 0 Speed (krpm)/(ft/sec)
Uncoated baseline MiTi Proprietary and confidential 10 170
20 340
30 510
40 680
FIGURE 7.90
Foil thrust bearing developments.
FIGURE 7.91
Large diameter, high-load capacity CFBs.
© 2006 by Taylor & Francis Group, LLC
50 850
60 1020
70 1190
80 1360
90 1530
Principles of Gas Turbine Bearing Lubrication and Design
FIGURE 7.92
7-105
Large diameter, high-load capacity thrust foil bearing.
for oil-free bearings and seals in advanced gas turbines. In certain applications where rotor loads are very high, compliant foil or hybrid foil/magnetic bearings may be considered. Many other oil-free hybrid combinations are also under consideration. This foil bearing technology base was instrumental in the successful test of a WJ24-8 240 lb thrust turbojet engine with a foil bearing directly behind the turbine. The foil bearing, tail cone, and rotor used in this test are all shown in Figure 7.95. The engine ran flawlessly, vibrations were low, and engine fuel consumption was reduced. Posttest inspection of the engine and bearings revealed them to be in perfect condition. Besides the foil bearing employed in the turbojet testing, foil bearings ranging from 6 mm in. diameter to 150 mm in diameter (see Figure 7.96 through Figure 7.98) have been developed and tested. What is interesting to note from Figure 7.96, is that as bearing size has decreased to 15 mm in. diameter and smaller, the predicted power loss due to viscous shearing of the air has been reduced. This reduction in power loss is due primarily to the fact that the static loading on the bearing is approximately one order of magnitude smaller than the larger bearings. This low power loss, which is already substantially lower than ball bearings, will be instrumental in the overall engine system thermal management for mesoscopic and MEMS class systems. One set of the small foil bearings indicated in Figure 7.96, are shown in Figure 7.97 (a) and (b). The 15 mm journal bearings and 35 mm diameter thrust bearings shown in (a) have been successfully tested at speeds in excess of 258,000 rpm and been subjected to shock loads in excess of 90 g without failure.
7.8.1 Rotor-Bearing Dynamics and Engine Integration Due to its principal of operation, the behavior of a CFB is strongly coupled to the dynamics of the interfacing rotor. Modifications to the rotor often impact the behavior of the bearing and vice versa. The design of a CFB and oil-free rotor system is therefore an iterative process that cycles between component
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Largest foil bearing ever built in the world
150 mm foil bearing under 400 Ib radfial load 2,500
20,000
Speed (20,000 rpm)
Operated to 27,500 rpm (4.05 MDN) Avg. housing temp ~ 1500°F+
1,500
15,000
Foil brg temp ~ 1200°F 10,000
1,000
Speed rpm
Temperature °F
2,000
Maximum ambient temperature to date 1500°F 120 lb rotor weight
5,000 500
00 4:
00 4:
:0 11
00 :4
4: :2
10
00 10
:0
4:
00 10
:4
4:
00 09
4:
00 :2
4: 09
:0 09
08
:4
4:
00
0
Run time
FIGURE 7.93
150 mm diameter foil journal bearing tested at elevated temperature.
Advabced bearing and seal technologies for next generation gas turbine engines Hybird foilmagnetic bearing
High performance foil bearing High temperature compliant foil seal
FIGURE 7.94 Advanced oil-free bearings and seals for gas turbine engines.
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FIGURE 7.95 WJ24-8 rotor, tailcone and MiTi® foil bearing after engine test.
Power loss (Hp)
1.0
15 mm diameter bearings (predicted)
0.5 2.339E-009* × (1.76) 6 mm diameter bearings (predicted)
0.0 0
FIGURE 7.96
10,000
20,000 30,000 40,000 Surface velocity (ft/min)
50,000
60,000
70,000
Power loss vs. speed for foil bearings from 6 to 200 mm in diameter.
and system level performance assessment. A generalization of this process is illustrated in Figure 7.99. The initial steps involve the creation of a rotordynamic model using assumed rotor and bearing geometries. The model is then used to solve for static bearing loads, and to generate a critical speed map which is, essentially, a parametric study of the system critical speeds as a function of bearing stiffness. The functional requirements of the foil bearings are generated from this data. The effort then shifts to the component level wherein MiTi®’s specialized computer codes are used to predict the speed-dependent stiffness, damping, and power loss. The data yielded by this analysis constitutes a preliminary bearing design. Focus is then shifted back to system behavior by applying the CFB stiffness and damping coefficients to the existing rotordynamic model and investigating the response. The goal at this stage is, generally speaking, to obtain acceptable bearing loads and stable operation throughout the run range. If such results are achieved then the bearing design is complete, excepting optimization and enhancement. More often than not, however, additional design iterations are required, either on the rotor or bearing or both.
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(a)
(b)
FIGURE 7.97
Miniature foil bearings tested to over (a) 250,000 rpm and (v) 700,000 rpm.
FIGURE 7.98
MiTi® foil bearings from 15 mm to 150 mm diameter.
The prediction of CFB performance involves a blend of computational techniques and empirical data drawn from past experience. In the case of the former, a sophisticated coupled structural-hydrodynamic computer code is needed to analyze the hydrodynamics within the bearing at various speed, load, and temperature conditions. The information gained from the analyses includes, but is not limited to, the bearing’s power loss, eccentricity, and speed-dependent stiffness coefficients. MiTi®’s experience with CFBs has shown that the predicted values agree very well with experimental data. Damping, however, is more difficult to predict. Although analytical tools do exist to forecast damping, experience in this case has shown that the best predictions are made using past experimental data and engineering judgment. Although damping is often simplified to pure viscous damping in dynamic models, the available data clearly indicates that such assumptions are not valid in CFB rotor systems. The difficulty in obtaining analytical values stems from the fact that damping in CFBs is essentially a friction phenomenon and that the physics underlying friction are not well understood. 7.8.1.1 Foil Bearing Analysis Like the rotor-bearing system analysis the foil bearing analysis is also iterative in nature combining both structural and hydrodynamic compressible gas solutions. In certain critical applications, a finite element
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Principles of Gas Turbine Bearing Lubrication and Design
7-109
Assume Assumerotor rotorconfiguration configuration Available brg space Assume Assumebearing bearinggeometry geometry Brg geometry Create Createrotordynamic rotordynamicmodel model Finite element model
Static analysis
Critical Critical speed speed map map
Required stiffness
Brg loads
Hydrodynamic bearing analysis Dynamic brg coefficients (k, b) Rotordynamic Rotordynamicanalysis analysiswith with dynamicbearing bearingcoefficients coefficients dynamic Predicted rotor behavior
Viable design?
NO
YES DONE
FIGURE 7.99
Rotor-bearing system design process.
analysis (FEA) of the top foil to assess the expected stress and deformations may be required. With proper design, the stresses in the foils may be held well below the fatigue endurance limit levels of the foil material. For example MiTi strives to maintain stresses less than 2200 psi. Such an FEA is performed only if extremely high loading conditions are expected. Having already established the influence of the top foil on bearing performance through FEA, MiTi® does not routinely require the FEA to be completed unless extreme loads, speeds, and temperatures are expected, and where the contribution of the top foil may be crucial to the successful operation of the rotor-bearing system. However, most bearing systems may be successfully completed by conducting parametric design tradeoff studies of key foil bearing design parameters, such as those shown in Figure 7.100, including diameter, length, operating clearance, and structural stiffness, which is based upon the number of layers of bumps, thickness of the bump material, change in material properties with temperature, and so on. The significant number of parameters available to the designer in meeting the requirements for rotor centrifugal and thermal growth while yielding the desired stiffness characteristics offer considerable flexibility but also challenges due to the large number of parameters one has at their disposal. As such a well-defined design procedure is needed that includes the entire rotor- bearing system as noted in Figure 7.99. For example, once static rotor load, speed, and ambient temperature are known, the MiTi® procedure is to establish a nominal bearing geometry based on the length and diameter of the bearing that will maintain a low load on the foil and coating so that the
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Speed
Pressure
Static load
Dynamic load
Temperature
Thermal gradient
Bearing length
Bearing diameter
Clearance
Preload
Bump stiffness (thickness, No. of layers, No. of bumps)
Number of bearing pads (1, 3, 5, 7)
Material properties
Gas properties
Hydrodynamic pressure profile
(a)
Structural stiffness elements Gas film
Coulomb friction
Pressure (psi)
30 20
Lower
Upper
Interface coating
Bearing support/mount struture
10 (b) 0
Pressure
8
Ax
Low load condition Rotor load Original position Deflected position
100
6
ia
ld
ire
4
ct
FIGURE 7.100
io
n
2
25
Coulomb 75 frication tion c e 50 ir al d renti fe m Circu MiTi Proprietary information
High load condition
Foil bearing design parameters resulting in coupled hydrodynamic and structural stiffness.
rotor will lift off and become airborne as quickly as possible. Low-speed lift off is essential for long life of the high-temperature coating and to minimize starting sizing requirements. Once the bearing geometry has been established an initial structural stiffness is defined and the coupled structural and hydrodynamic analysis is run to determine hydrodynamic pressures and hence load-carrying capability, power loss, and stiffness coefficients. Due to the compliance of the bearing surface, an iterative solution is needed to converge on a pressure profile that matches with the deformed foil surface. 7.8.1.2 Hybrid Bearings As noted above, these recent developments are making CFB technology suitable for use as main rotor support for gas turbine engine rotors weighing as little as a few grams to more than a thousand pounds. CFBs with load capacities and damping significantly greater than previously demonstrated have been developed and combined with magnetic bearings to support rotor loads in excess of 1350 lb for a 100 mm diameter bearing. The hybrid foil/magnetic bearing combines two oil-free bearing technologies to take advantage of the strengths of each. Foil bearings have good load-carrying ability and transient shock response at high shaft speeds. Magnetic bearings provide nearly constant load carrying ability over the operating speed range, but are susceptible to overload during transient events, and have potentially catastrophic failure modes if power, sensor, or winding failures occur. Unlike more conventional magnetic bearing applications, the hybrid foil/magnetic bearing does not require a separate backup bearing. In the hybrid bearing, the foil bearing component provides the transient/failure protection, as well as significantly increases the load capacity for a given bearing system size and weight. For example, a hybrid foil/magnetic bearing system designed by MiTi® for these tests has been tested with an applied load of almost 1350 lbs combined with a dynamic load of approximately 100 lbs. This bearing system is shown in Figure 7.101. Measured performance of the 100 mm diameter bearing is shown in Figure 7.102. As seen the combination of the two bearings exceeds the capacity of either alone.
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Foil brg
Foil brg
FIGURE 7.101 Hybrid foil magnetic journal bearing.
To demonstrate the ability of the foil bearing component of the hybrid bearing to act as a backup bearing for the magnetic bearing, the system response to potential magnetic bearing failures and failure recoveries was investigated. For this investigation, 13 different simulated failures were investigated at speeds of 15,000 rpm and 25,000 rpm. The failure modes selected for this investigation were: 1. 2. 3. 4. 5. 6. 7. 8.
Vertical Sensor Failure Horizontal Sensor Failure Both Sensors Failure Loss of Vertical Displacement Signal Loss of Horizontal Displacement Signal Loss of Both Displacement Signals Loss of Left Horizontal Amplifier Output Loss of Both Horizontal Amplifier Outputs
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Handbook of Lubrication and Tribology 1400 Hybrid foil/mag bearing
Permissible load capacity (lb)
1200 1000 800
Foil bearing
600 400 200
Magnetic bearing
0 0
FIGURE 7.102
5
10
15 20 Speed (krpm)
25
30
35
Experimentally measured performance of hybrid foil magnetic bearing.
Displacement (Mil)
3 2 Recovery
Failure
1 0 –1 –2 0
2
4
6
8
10
12
14
16
18
Time (sec)
FIGURE 7.103
Rotor response due to vertical sensor open failure at 25,000 rpm.
Displacement (mil)
3 Recovery
Failure
2 1 0 –1 –2 0
5
10 Time (sec)
FIGURE 7.104
Loss of vertical displacement signal at 15,000 rpm.
© 2006 by Taylor & Francis Group, LLC
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Principles of Gas Turbine Bearing Lubrication and Design
9. 10. 11. 12. 13.
7-113
Loss of Top Vertical Amplifier Output Loss of Both Vertical Amplifier Outputs Loss of Bias Current Loss of All Control Amplifier Outputs Loss of All Amplifier Outputs and Bias Current
To simulate a worst case scenario, the bearing load was supported almost entirely by the magnetic bearing prior to the failure event. In addition, lightly damped magnetic bearing characteristics were selected for use during this testing. Based on previous experience, MiTi engineers examined both the initial failure transient as well as the system recovery In many cases, the recovery transient was actually more severe than the failure transient under these operating conditions. Figure 7.103 presents a typical shaft vertical response at the foil bearing displacement sensors for a failure that simulates sensor or cable damage which results on a loss of electrical continuity between the sensor and the sensor signal conditioning unit (-24 VDC control system input) and subsequent recovery of sensor signals. Horizontal shaft motion for this test was negligible. Figure 7.104 presents a typical shaft vertical response at the foil bearing displacement sensors for any system failure, which results in a loss of the vertical displacement measurement (0 VDC control system input). Again, the time-history trace of shaft displacement shows both failure and recovery. As was noted previously, horizontal shaft motion was negligible. In all 26 test cases, the transient during both failure and recovery was well controlled by the foil bearing. In addition, a foil bearing alone coast-down from full speed under shaft load was demonstrated in later testing. The combination of large static load and simulated failures demonstrate that the foil bearing component of a foil/magnetic hybrid bearing is an effective back up bearing for the magnetic bearing, allowing both continued operation following a failure, as well as damage-free equipment shut down and that the hybrid bearing system can support more load than either alone. 7.8.1.3 Compliant Foil Seals Due to the need to track rotor dynamic excursions without loss of seal performance either during the excursion or thereafter, advanced dynamic seals are needed. As with advances in bearings the improved low leakage film riding seals are needed to compliment the life and performance gains possible by the oil-free bearings. A Mohawk Innovative Technology, Inc. (MiTi®) novel, non-contacting, Compliant Foil gas Seal (CFS) has been designed and successfully tested at temperatures to 600◦ C and surface velocities of over 2100 ft/sec in a dynamic simulator representative of a small gas turbine engine hot section. Measured and analytical comparisons of leakage flow rates at varying differential pressures were made, showing that the CFS capability significantly exceeds the performance of both brush and labyrinth seals. The brush and CFS tests were performed with the rotor operating at speeds to 48,000 rpm. The labyrinth and CFS comparisons were made under nonrotating conditions, but with each seal mounted in the rig. Besides the performance benefits, the CFS offers improved life and durability benefits when applied to most rotating machinery, as noted later on. The performance of both the CFS, like the foil bearing, is based on the hydrodynamically generated high-pressure gas film that is built up between the journal and the seal top surface (top foil) due to the shaft rotation. This thin gas film separates the seal surface from the rotating shaft sealing surface, creating a film pressure (See Figure 7.105) resulting in noncontact, continuous operation. Strict accounting of all variables results in a seal with optimized performance attributes. While its primary purpose is the control of leakage, the similarity to a CFB results in its having a load-bearing capacity as a secondary function. This capacity is predictable and can be incorporated in the overall system design to provide additional support stiffness and damping for improved rotor system dynamics. Through compression system aerodynamic improvements, advanced materials, and weight reducing designs, improved engine efficiencies are being achieved. However this emphasis on efficiency, reduced emissions, and lowered noise levels is placing more stringent requirements on operating systems, including higher shaft surface speeds and increased environmental temperatures. This increased severity causes greater wear and mechanical distress for the seals due to their material and design limitations [47].
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Handbook of Lubrication and Tribology Compliant foil seal pressure profile including top foil
Pressure (Psi)
60 45 30 15 1
FIGURE 7.105
2 Ax 3 ial pos 4 5 itio n
100
6
50 75 on l directi ferentia Circum
25
CFS pressure profile. Compliant gas foil seal concept to application Shoulder
PH Top foil
Shaft
PL Spring bump Fabrication 1.4 in.
Concept
2.84 in.
FIGURE 7.106
5.950 in.
8.5 in.
CFS sizes.
An important conclusion from this discussion is that seals need improvements to accommodate the influence of operating speed, temperature, and pressure on seal performance and life. The higher operating speeds significantly affect seal heat generation and subsequent thermal distortions of mating components. Thus, as pressure–velocity product increases, seal durability, reliability, performance, and cost may all be adversely affected. Consequently, new materials, new designs, or some combination of both, are needed. Noncontacting labyrinth seals are insufficient due to the large gaps and hence leakage flows. Contact seals such as brush seals experience wear. Improved air seals will provide increased engine performance and improved life-cycle costs [47–49]. Advanced engine seals must have reduced losses and they must maintain these performance benefits over the service interval of the engine [50]. However, to achieve these benefits as surface speeds and ambient temperatures increase, new technologies that result in extremely lightweight design concepts, employ improved or novel sealing approaches with reduced leakage, and take advantage of new materials are needed. Damage-tolerant, high-performance designs are required
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6 in. seal rotor motion under 80 psig differential pressure 90 120
mi
60
5 4
150
30
3 2 e 1
180
a
4
3
2
1
d
b c
0 5
f g
0
1
2
0 3
4
1
5
mi
Speed (rpm)
Temperature (°F)
a = 1500 b = 3000 c = 4000 d = 5000 e = 6000 f = 7000 g = 8000
72 73 76 75 75 74
2 3
210
330
4 5 300
240 270
FIGURE 7.107 Demonstrated CFS rotor excursion capability.
to accommodate the high differential pressures, extreme temperatures, high surface speeds, and normal rotor excursions. Advanced seals must also accommodate large rotor excursions; whether due to axial thermal gradients, maneuvers in aerospace applications, bowed rotor starts, and passage through bending critical speeds. Attempts to meet these stringent requirements with fixed geometries (i.e., labyrinth seals) has generally resulted in design compromises with larger than desired steady-state operating clearances. Brush seals have been investigated with a fairly high degree of success in numerous military and commercial engines and seem to address some of these limitations. For example, brush seals have shown leakage rates from 2 12 to 10 times lower than labyrinth seals and have demonstrated performance improvements in engine applications [50–52]. Although brush seals have demonstrated significantly lower leakage than conventional designs, contacting seals are subject to wear, which degrades their effectiveness with time. Good brush seal performance requires a smooth, wear resistant runner interface and wear resistant bristles. Further, the ability of brush seals to continue to operate reliably and efficiently under large rotor excursions is very limited. Experience has shown that the bristles plastically deform under large excursions, thereby changing the seal clearance and increasing leakage. As reported at the 1997 Seal Workshop at NASA/GRC [52], there are a number of concerns related to brush seals. Specifically, brush seals are costly than existing seals; they are heavier; the suppliers are limited; there is considerable concern over the durability; and there is concern over damage during construction and installation. In particular, care must be taken during installation to prevent reverse rotation of the shaft, which can damage the seal. While significant efforts are being expended to address the limitations present in brush seals, it will always be a contacting seal and thus prone to wear. In essence then, the performance of the brush seal is a function of time, shaft excursions, environment, and the tribo-material system. Given this background it is evident that there is a need for a highly efficient, long-life seal that is capable of maintaining performance during and after large excursions. Recent developments [53–56] indicate that noncontact CFB technology has great application and development potential for use as aircraft gas turbine engine seals. Dr. Heshmat of Mohawk Innovative
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Technology, Inc., (MiTi) has developed CFB’s with load capacities and damping significantly greater than previously demonstrated. This bearing technology foundation has been instrumental in the development of the CFS as reported by Salehi and Heshmat [57–59]. They have demonstrated the successful operation of a 1.4 , a 2.84, and a 6.0 in. CFS, and have even manufactured an 8.5 in. CFS for NASA that was tested to 30,000 rpm (See Figure 7.106). The current seal test has shown us that the lift off process and film riding does take place to produce the desired noncontact operation. The seal has also shown large excursion capability in excess of 6-mi (diametral) for a 150 mm diameter seal with an 80 psig differential pressure and while operating at speeds to 14,000 rpm at MiTi® (see Figure 7.107).
References The selection of the following references was made with the intent of providing sources from which additional data could be culled for bearing design purposes. Della Corte, C. and Pinkus, O., “Tribological Limitations in Gas Turbine Engines: A Workshop to Identify the Challenges and Set Future Directions,” ASME Publication Allaire, P.E., Li, D.F., and Choy, K.C., “Transient Unbalance Response of Four Multilobe Journal Bearings,” J. Lubr. Technol., Trans. ASME, July 1980. Chen, H.M. “Active Magnetic Bearing Technology: A Conventional Rotordynamic Approach,” Proceedings of 15th Leeds-Lyon Symposium on Tribology, Sept., 1988. Chen, H.M., “Magnetic Bearings and Flexible Rotor Dynamics,” Proceedings of STLE Annual Meeting at Cleveland, Ohio, May 9–12, 1988. Chen, H.M. et al., “Stability Analysis for Rotors Supported by Active Magnetic Bearings,” Proceedings of 2nd International Symposium on Magnet Bearings, July 12–14, 1990, Tokyo, pp. 325–328. Chen, H.M., “Design and Analysis of a Sensorless Magnetic Damper,” presented at ASME Turbo Expo, June 5–8, 1995, Houston, Texas, 95GT180. Heshmat, H. and Chen, H.M., Compressor handbook, Chapter 19, “Principles of Bearing Design,” McFrawHill, 2001. Gross, W.A., “Gas Film Lubrication,” John Wiley, 1962. Heshmat, H., Walowit, J.A. and Pinkus, O., “Analysis of Gas-Lubricated Compliant Thrust Bearings,” ASME Paper 82-LUB-39, 1982. Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982. Heshmat, H., and Dill, J. “Fundamental Issue in Cryogenic Hydrodynamic Lubrication,” Proceedings of AFOSR/ML Fundamentals of Tribology Work Shop, (February 1987). Heshmat, H. “Analysis of Compliant Foil Bearings with Spatially Variable Stiffness,” presented at AIAA/SAE/ASME/ASEE 27th Joint Propulsion Conference, June 24–26, 1991, Sacramento, CA, Paper No. AIAA-91-2101. Heshmat, H. “A Feasibility Study on the Use of Foil Bearings in Cryogenic Turbopumps,” presented at AIAA/SAE/ASME/ASEE 27th Joint Propulsion Conference, June 24–26, 1991, Sacramento, CA, Paper No. AIAA-91-2103. Heshmat, H. and Hermel, P., “Compliant Foil Bearing Technology and Their Application to High Speed Turbomachinery,” Proceedings of 19th Leeds-Lyon Symposium on Thin Film in Tribology — From Micro Meters to Nano Meters, Leeds, Sep. 1993, D. Dowson et al. (Eds.), Elsevier Science Publishers B.V., 1993, pp. 559–575. Heshmat, H. and Pinkus, O. “Performance of Starved Journal Bearings with Oil Ring Lubrication,” J. Trib. Trans. ASME, 107, 1985, 23–32. Heshmat, H. and Pinkus, O. “Experimental Study of Stable High-Speed Oil Rings,” J. Trib. Trans. ASME, 107, 1985, 14–22.
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Heshmat, H. and Pinkus, O. “Performance of Oil Ring Bearing,” International Science Conference on Friction, Wear, Lubr., Tashkent, May 1985. Hustek, J.F. and Peer, O.J. “Design Considerations for Compressors with Magnetic Bearings,” Proceedings of 3rd International Symposium on Magnetic Bearings, July 93, Alexandria, VA. Jones, G.J. and Martin, F.A. “Geometry Effects in Tilting-Pad Journal Bearings,” ASLE Paper No. 78-AMatA-2, 1978. Ku, C.-P.R. and Heshmat, H. “Compliant Foil Bearing Structural Stiffness Analysis: Part I–Theoretical Model Including Strip and Variable Bump Foil Geometry,” J. Trib. Trans. ASME, 114, 1992, 394 –400. Pinckney, F.D. and Keesee, J.M., “Magnetic Bearing Design and Control Optimization for a Four-Stage Centrifugal Compressor,” Proceedings of Mag. ’92, pp. 218–227. Pinkus, O. and Sternlicht, B., “Theory of Hydrodynamic Lubrication,” McGraw-Hill, 1961. Pinkus, O. and Wilcock, D.F., “Low Power Loss Bearings for Electric Utilities: Volume II: Conceptual Design and Optimization of High Stability Journal Bearings; Volume III: Performance Tables and Design Guidelines for Thrust and Journal Bearings,” MTI Report Nos. 82TR42, 82TR43, April 1982. Pinkus, O., “Analysis of Elliptical Bearings,” Trans. ASME., Vol. 78, 1956, pp. 965–973. Pinkus, O., “Analysis and Characteristics of the Three-Lobe Bearing,” Trans. ASME, Series D., 81, 1959. Pinkus, O., “Solution of the Tapered-Land Sector Thrust Bearing,” Trans. ASME, 80, 1958. Pinkus, O., “Analysis of Non-circular Gas Journal Bearings,” J. Lubr. Technol., Trans. ASME, Oct. 1975. Pinkus, O. “Solution of Reynolds Equation for Arbitrarily Loaded Journal Bearings,” Trans. ASME, Series D, 83, 1961. Pinkus, O. “Misalignment in Thrust Bearings Including Temperature and Cavitation Effects,” J. Tribol., Oct. 1986. Pinkus, O. “Optimization of Tilting Pad Journal Bearings Including Turbulence and Thermal Effects,” Isr. J. Technol., 22, 1984/85. Pinkus, O., “Manual of Bearing Failure and Repair in Power Plant Rotating Equipment,” EPRI, July 1991. Raimondi, A.A. and Boyd, J., “A Solution for the Finite Journal Bearing and Its Application to Analysis and Design — III,” Trans. ASLE, 1, 1959. Reddickoff, J.M. and Vohr, J.H., “Hydrostatic Bearings for Cryogenic Rocket Engine Turbopumps,” J. Lubr. Technol., 1969. Schmied, J.L. and Predetto, J.C., “Rotor Dynamic Behaviour of a High-Speed Oil-Free Motor Compressor with a Rigid Coupling Supported on Four Radial Magnetic Bearings,”proceedings of 4th International Symposium on Magnetic Bearings, August 23–26, 1994, ETH Zurish, Switzerland, pp. 441–447. Vohr, J.H., “The Design of Hydrostatic Bearings,” Columbia University, NY. Walton, J.F. and Heshmat, H.,“Compliant Foil Bearings For Use in Cryogenic Turbopumps,” Proceedings of Advanced Earth-to-Orbit Propulsion Technology Conference Held at NASA/MSFC May 17–19, 1994, NASA CP3282, Vol. 1, Sept. 19, 1994, pp. 372–381. Wilcock, D.F. and Booser, E.R., “Bearing Design and Application,” McGraw Hill, 1957. Suriano, F.J., Dayton, R.D., and Woessner, F.G., “Test Experience with Turbine-End Foil Bearing Equipped Gas Turbine Engines.” ASME Paper No. 83-GT-73. Heshmat, H., Chen, H.M., Walton, J.F., 1998, “On the Performance of Hybrid Foil-Magnetic Bearings,” Proceedings of 43rd ASME Gas Turbine and Aeroengine Congress, Stockholm, Sweden, ASME Paper No. 98-GT-376. Heshmat, H., “Advancement in the Performance of Aerodynamic Foil Journal Bearings- High Speed and Load Capacity,” ASME J. Tribol., 116, 1994, 287–95. Heshmat, H. and Ku, C.-P., “Structural Damping of Self-acting Compliant Foil Journal Bearings.” ASME J. Tribol., 116, 1994, 76–82. Heshmat, C.A. and Heshmat, H., “An Analysis of Gas-Lubricated, Multileaf Foil Bearings with Backing Springs.” ASME J. Tribol., 117, 1995, 437–443. Heshmat, H. and Hermel, P., “Compliant Foil Bearing Technology and their Application to High-speed Turbomachinery,” Proceedings of 19th Leeds-Lyon Symposium on Thin Film in Tribology — From Micro Meters to Nano Meters, Leeds, D. Dowsen, et al., (Eds.), Elsiver, 1992, pp. 559–575.
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Walton, J.F. and Heshmat, H. “Compliant Foil Bearings for use in Cryogenic Turbopumps.” NASA CP3232 Vol. 1, 1994, pp. 372–381. Heshmat, H., “Operation of Foil Bearings Beyond the Bending Critical Mode.” ASME Paper 99-TRIB-48. Swanson, E.E., Walton, J.F., Heshmat, H., 1999, “A 35,000 RPM Test Rig for Magnetic, Hybrid, and Back-Up Bearings.” ASME Paper No. 99-GT-180, 1999. Swanson, E., Heshmat, H, and Shin, J., “The Role of High Performance Foil Bearings in an Advanced, Oil-Free, Integral Permanent Magnet Motor Driven, High-Speed Turbo-Compressor Operating above the First Bending Critical Speed,” Proceedings of Turboexpo 2002, ASME Turbo Expo 2002, June 3–6, Amsterdam, The Netherlands, 2002, GT-2002-30579. NASA. “Starting a Turbomachinery Revolution” NASA Glenn Research Center (216) 433-5573 Heshmat, H. “Advancements In The Performance of Aerodynamic Foil Journal Bearings: High Speed and Load Capability,” ASME Paper 93-Trib-32, STLE/ASME Tribology Conference, October 24–27, New Orleans, LA, 1993. Heshmat, H. “Advancements In The Performance of Aerodynamic Foil Journal Bearings: High Speed and Load Capability,” ASME Paper 93-Trib-32, STLE/ASME Tribology Conference, October 24–27, New Orleans, LA, 1993. Hendricks, R.C., Griffin, T.A., Kline, T.R., Csavina, K.R., Pancholi, A., and Sood, D., “Relative Performance Comparison Between Baseline Labyrinth and Dual-Brush Compressor Discharge Seals in a T700 Engine Test,” proceedings of 39th ASME International Gas Turbine and Aeroengine Conference, The Hague, Netherlands, June 13–16, 1994. Paper No. 94-GT-266 Flower R., “Brush Seal Development Systems, AIAA Paper 90-2143, 1990. Hendricks, R. and Steinitz, B. Editors, “Seals/Secondary Flows Workshop 1997, V.1, October 1998. Heshmat, H. and Hermel, P. “Compliant Foil Bearing Technology and Their Application to High Speed Turbomachinery,” proceedings of the 19th Leeds-Lyon Symposium on Thin Film in Tribology — From Micro Meters to Nano Meters, Leeds, Sep. 1992. Heshmat, H. “Analysis of Compliant Foil Bearings with Spatially Variable Stiffness,” proceedings of AIAA/SAE/ ASME/ASEE 27th Joint Propulsion Conference, June,1991, Sacramento, CA, Paper AIAA-91-2102. Heshmat, C.A. and Heshmat H. “An Analysis of Gas Lubricated, Multi-Leaf Foil Journal Bearings with Backing Springs,” ASME Paper 94-Trib-61. Heshmat, H. and Ku, C-P R., “Structural Damping of Self-Acting Compliant Foil Journal Bearings,” ASME Trans., J. Tribol., 116, 1994, pp. 76–82. Salehi, M., Heshmat, H., Walton, J., and Cruzen S., “The Application of Foil Seals to a Gas Turbine Engine,” AIAA paper 99-2821, proceedings of 35th AIAA/ASME/SAE/ASEE joint Propulsion Conference and Exhibit, June 20–24, Los Angeles, CA, 1990. Salehi, M, and Heshmat, H., 2000, “High Temperature Performance Evaluation of a Compliant Foil Seal,” presented at 36th AIAA/ASME/SAE/ASEE joint Propulsion Conference and Exhibit, July 17–19, Huntsville, AL, 2000. Salehi, M., and Heshmat, H., “On the flow and thermal analysis of the compliant gas foil seals and foil bearings,” STLE Trans., 43, 2000, pp. 318–324.
Additional References Salehi, M., Heshmat, H., and Walton, J.F., II “On the Frictional Damping Characterization of Compliant Bump Foils,” Presented at the International ASME/STLE Joint Tribology Conference, October 2002, Cancun, Mexico, ASME J. Tribol., 125, 2003, pp. 804–813. Salehi, M., Heshmat, H., and Walton, J.F., II “Advancements in the Structural Stiffness and Damping of a Large Compliant Foil Journal Bearing — An Experimental Study,” Paper GT2004-53860, Published in the ASME J. Eng. Gas Turbines Power, and presented at the International Gas Turbine Institute Conference, Vienna, Austria, June 2004.
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Salehi, M., Heshmat, H., Walton, J.F., II and Tomaszewski, M.J. “Operation of a Mesoscopic Gas Turbine Simulator at Speeds in Excess of 700,000 rpm of Foil Bearings.” Paper GT2004-53870, Published in the ASME J. Eng. Gas Turbines Power, and presented at the International Gas Turbine Institute Conference, Vienna, Austria, June 2004. Walton, J.F., II, Heshmat, H., and Tomaszewski, M.J. “Testing of a Small Turbocharger/Turbojet Sized Simulator Rotor Supported on Foil Bearings.” Paper GT2004-53647, Published in the ASME J. Eng. Gas Turbines Power, and presented at the International Gas Turbine Institute Conference, Vienna, Austria, June 2004. Hryniewicz, P., Locke, D.H. and Heshmat, H. “New-Generation Development Rigs for Testing High-Speed Air-Lubricated Thrust Bearings.” Tribol. Trans., 46, 2003, pp. 556–569. Heshmat, C.A., Xu, D.S., and Heshmat, H. ”Development of Advanced Thrust Foil Bearings.” Presented at the 54th STLE Annual Meeting, May 23–27, Las Vegas, NV, 1999. Submitted for publication in the Tribology Transactions. Heshmat, C.A., Xu, D., and Heshmat, H. “Analysis of Gas Lubricated Foil Thrust Bearings using Coupled Finite Element and Finite Difference Methods.” Presented at the ASME/STLE Joint Tribology Conference, October 10–13, Orlando, FL, 1999, published in the J. Trib. Trans. ASME, 122, 2000, pp. 199–204. Heshmat, C.A., Xu, D.S. and Heshmat, H. “Analysis of Gas Lubricated Foil Thrust Bearings Using Coupled Finite Element and Finite Difference Methods,” Trans. ASME, J. Tribol., 122, 2000, pp. 199–204.
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8 Steam Turbines 8.1 8.2
Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Features of Steam Turbines . . . . . . . . . . . . . . . . . . . . . . . . . . .
8-1 8-2
Classification
8.3
Turbine Design and Construction . . . . . . . . . . . . . . . . . . . .
8-3
Bearings • Bearing Housings and Bearing Housing End Seals • Steam Control Valves, Governors, and Control Systems • Turning Gear • Couplings • Additional Tribological Components and Issues • Driven Units
8.4
Lube Oil Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
8-17
Nonpressurized Oil Ring Lubrication • Pressurized Lubrication Systems
8.5
Turbine Oil . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
8-23
Physical Properties • Formulation
8.6
Performance Features of Turbine Oils. . . . . . . . . . . . . . . .
8-26
Viscosity • Oxidation Stability • Freedom from Sludge and Deposits • Corrosion Protection • Water Separability (Demulsibility) • Air Separability and Resistance to Foaming
8.7
Degradation of Turbine Oils in Service . . . . . . . . . . . . . .
8-28
Contamination • Additive Depletion • Thermal and Oxidative Degradation • Biological Deterioration • Turbine Oil Severity
8.8
Lubricant Maintenance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
8-30
New Oil Makeup • Lube Oil Purification • Refortification
8.9
B.C. Pettinato Elliott Company
Fire-Resistant Fluids . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
8-32
Properties • Degradation • Condition Monitoring • Maintenance
References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
8-34
8.1 Introduction Steam turbines are used extensively in the power generation industry as prime movers for generators. They are also used for mechanical drive application in petrochemical and other industries where they power centrifugal pumps, compressors, blowers, and other machines. In addition, they continue to be used for shipboard propulsion. Sizes range from as low as 0.75 kW for some mechanical drive applications to as high as 1,500 MW for electric generator drives in large nuclear power plants [1]. Steam turbines are particularly well suited for continuous operation, and in many cases are operated for years without shutting down.
8-1
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Handbook of Lubrication and Tribology
8.2 Features of Steam Turbines Steam turbines operate by taking high-pressure steam and converting it into useful mechanical work through expansion. The steam is fed into an inlet casing then throttled through a set of inlet valves, which control the rate of steam admission into the turbine. The steam is then allowed to expand and accelerate through stationary blades or nozzles, which directs the flow onto the rotating blades. The rotating blades convert the steam’s kinetic energy into torque, which results in rotation of the turbine shaft along with a loss of pressure and temperature in the steam. The rotating shaft is used to drive machinery coupled to the exhaust end of the turbine shaft. Absence of lubrication from the steam path is an important feature. Since the exhaust steam is not contaminated with oil vapor, this allows the steam to be condensed and returned directly to the boilers for reheat, or extracted and used for direct heating or other purposes. The lack of internal lubrication also results in a relatively low rate of lubricating oil consumption [2].
8.2.1 Classification Steam turbines have numerous configurations and means of classification. A steam turbine is generally classified as being either high-pressure or low-pressure, condensing or noncondensing, single-stage or multi-stage, single-valve or multi-valve, extraction or nonextraction, direct drive or gear drive, and for either electric generator, mechanical drive, or propulsion service [3]. In addition, steam turbines are classified in accordance with recognized engineering standards, which govern various aspects of turbine design and construction. Some of these classifications are discussed further. High-pressure designs refer to the internal pressure to be contained by the main shell and casing parts. High pressure generally refers to pressures in excess of 13,800 kPa (2,000 psig) where double shell construction is often used. The pressure and temperature of steam are interrelated. Higher inlet steam pressure is often accompanied by higher steam temperature. Temperatures can range from 200◦ C to over 600◦ C. High temperature generally refers to applications with inlet temperatures in excess of 540◦ C (1,000◦ F). Condensing turbines exhaust steam at less than atmospheric pressure, whereas noncondensing (back pressure) turbines exhaust steam at higher than atmospheric pressure. Condensing machines tend to be larger and more complex than noncondensing designs due to the increased volume expansion of the steam at the exhaust end as well as the additional hardware required to drop the exhaust end pressure below atmospheric. In direct drive arrangements, the turbine is directly coupled to the driven machine; whereas gear drive applications have either a speed increasing or speed reducing gear between the turbine and driven equipment. The use of a speed increasing or reducing gear creates added complexity, cost, and power losses along with additional requirements of the lube oil system. However, the use of gears greatly increases the application range whether the need is for high torque as in marine propulsion or high-speed requirements such as integrally geared compressors. Gear drives also enable the efficient use of small turbines, which can operate at higher speeds when a reduction gear is used. Generator drive turbines operate at single speeds to synchronize the generators with the electric grid. Typically, the synchronization speed is either 1,800 or 3,600 rpm in regions with 60 Hz power, or 1,500 or 3,000 rpm in regions with 50 Hz power. On the other hand, mechanical drive turbines are variable speed with shaft speeds as low as 1,000 rpm or as high as 20,000 rpm depending on the turbine and the application. A number of different engineering standards have been developed for the design and procurement of steam turbines as shown in Table 8.1. American Petroleum Institute (API) standards pertain to design, manufacture, and testing of mechanical drive turbines for petrochemical application [4,5]. National Electrical Manufacturers Association (NEMA) standards pertain to design and application of mechanical drive turbines and turbine generator sets for electric utility application [6,7]. Military standards generally apply to steam turbines for shipboard use [8–10]. Other international
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Steam Turbines TABLE 8.1
8-3 Steam Turbine Design and Procurement Standards
Standard designation API 611 API 612 / ISO 10437 IEC 60045-1 NEMA SM-23 NEMA SM-24 MIL-T-17286D MIL-T-17600D MIL-T-17523
Standard title General-purpose steam turbines for refinery service Special-purpose steam turbines for refinery service Steam turbines — part 1: specifications Steam turbines for mechanical drive service Land-based steam turbine generator sets 0 to 33,000 kW Turbines and gears, shipboard propulsion, and auxiliary steam; packaging of Turbines, steam, propulsion naval shipboard Turbine, steam, auxiliary (and reduction gear) mechanical drive
recognized standards such as IEC 60045-1 are also used to assist in steam turbine specification and procurement [11]. Figure 8.1 shows a general-purpose (API 611) turbine. These turbines are either horizontal or vertical units used to drive equipment that is usually spared, is relatively small in size (power), or is in noncritical service. General-purpose steam turbines for refinery service are intended for applications where the inlet gauge pressure does not exceed 4,800 kPa (700 psi), the inlet temperature does not exceed 400◦ C (750◦ F), and the speed does not exceed 6,000 rpm [4]. The turbine shown in Figure 8.1 has lubrication consisting of sumps at each journal bearing with oil ring-lubricated bearings. An isolated mechanical–hydraulic governor with oil sump is used to control speed. Figure 8.2 shows a special-purpose turbine for refinery application that meets API 612/ISO 10437 specifications. Such units are usually not spared and are used in uninterrupted continuous operation in critical service. They are not limited by steam conditions, power, or turbine speed. The equipment (including auxiliaries) covered by these standards are designed and constructed for a minimum service life of 20 yr and at least 5 yr of uninterrupted operation [5]. The turbine shown in Figure 8.2 has lubrication provided by a circulating oil system console (not shown) providing oil at high volumes to the bearings and to the servo valve actuator.
8.3 Turbine Design and Construction The parts of a steam turbine may be thought of as being in four groupings (1) the rotor, or spindle, (2) stationary parts, (3) the governing and trip systems and valves, and (4) auxiliary systems consisting of the lubrication system and other components such as the condition monitoring system. The rotor, depending on turbine type, may consist of wheels mounted on a shaft or may be machined from a solid forging or a forging made up of welded sections. In each case, the rotor carries securely fastened radial blades or buckets. Principle stationary parts consist of the steam-tight casing, nozzles, shaft seals, and bearings. Turbine governors control speed by controlling steam-admission valves through mechanical, pneumatic, or hydraulic actuators. Those parts of the turbine requiring lubrication consist of the bearings supporting the rotor, hydraulic actuators and governor components, and the trip system; and in some cases, a turning gear, geared couplings, and front standard. Lubricated parts reside external to the steam path, and when properly isolated will not contaminate the steam or become contaminated by the external environment. The lubrication system may be simple reservoirs in the pedestals of ring-oiled bearings, or elaborate circulation systems, having pumps, coolers, filters, and monitoring devices [12]. Figure 8.3 shows a typical unit of an oil-piping diagram for a turbine, gear, and generator string. Lubricating oil is supplied at two pressures by an oil console (not shown). Lube oil is supplied at low pressure of 100 to 125 kPa (15 to 18 psig) to the bearings. High-pressure oil of 1,000 kPa (150 psig) is supplied to the trip and throttle valve, to the valve actuator, and, if needed, to the governor mechanism. Bearing and coupling housings are part of the lube oil circuit and act to return oil to the reservoir.
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8-4
Sentinel valve
Rotor disk assembly
Casing cover
Shaft sleeve seal
Exhaust end sealing gland Carbon ring assembly
Steam end sealing gland
Rotor locating bearing Overspeed thip assembly Coupling (governor drive)
Carbon ring assembly
Governor Governor linkage
Oil rings
Oil rings
Rotor shaft
Steam end bearing housing
Shaft sleeve seal Steam chest Governor valve
Exhaust end bearing pedestal Exhaust end casing
FIGURE 8.1
Reversing blade assembly
Steam end casing Nozzle ring
Steam end journal bearing
Steam end support
General-purpose steam turbine. (From Installation, Operation, and Maintenance Instructions for YR Turbines, Elliott Company, Jeannette, PA, 2003. With permission.)
© 2006 by Taylor & Francis Group, LLC
Handbook of Lubrication and Tribology
Exhaust end journal bearing
Steam Turbines
Rocker arm bearing
Lubrican connection
Governor linkage assembly Valve stem and packing Valves, seat, and bar assembly Breather cap Bearing Journal housing bearing Bearing housing deflector Shaft end with coupling bolt pattern
Steam chest
Turbine case
Gland packing assembly case
Exhaust end packing gland assembly
Breather cap Interstage shaft seals
Journal bearing Thrust bearings
Steam end packing gland assembly
Rotor
Bearing housing end seal
Oil drain Steam exhaust
Bearing housing end seals
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Gland seal leak off
8-5
Steam turbine for special-purpose refinery service.
Bearing housing
Steam end flexible support Casing drain
FIGURE 8.2
Oil drain
8-6
Lube oil supply to unit conn’s
P1
Flexible hose
PSH PSLL
S P1
Breather Servo motor Driven equipment
IP
Coupling
Breather Gear
To reservoir
FIGURE 8.3
SG
SG
1/2” per foot minimum
Concentric reducer I
Inlet SY servo motor/ valves
Trip and throttle valve S
e
Slop
Thermowell Thermowell
P
Current/pneumatic convertor
PCV
Pressure control valve
P1
Pressure indicator
PSH
Pressure switch, high
PSLL
Pressure switch, trip
SG
Site glass
SY
Speed relay
T1
Temperature indicator
T1
T1
SG
SG
Unit oil piping diagram for turbine-gear-generator set.
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Orifice PSLL
SG
Handbook of Lubrication and Tribology
SG
T1
2 valve manifold instrument valve/bleed valve
High pressure control oil supply
Trip solenoid dump valve
Spare vent
Spare vent
Solenoid valve (2-way) Ball valve
Nitrogen precharge
Turbine
Coupling
SG
Governor control signal
N2
Control oil accumulator
Steam Turbines
FIGURE 8.4
8-7
Tilt pad journal bearing. (From Elliott Company. With Permission.)
8.3.1 Bearings Proper rotor position is maintained by journal and thrust bearings. Journal bearings are used in pairs for radial positioning of the turbine shaft supporting the gravitational load of the shaft. Thrust bearings are used for axial positioning and support thrust loads that arise from steam forces within the turbine case. Thrust bearings are located at the steam end of the turbine opposite the coupling, and are used in pairs to accept thrust loading in either direction along the axis of the rotor. Steam turbine bearings can be either hydrodynamic, rolling element, or magnetic. Hydrodynamic bearings are the most prevalent. 8.3.1.1 Hydrodynamic Bearings Hydrodynamic bearings are highly advantageous because they suffer little or no wear and have exceptionally long life thereby enabling long periods of continuous operation, often in excess of 5 yr. In addition, the bearings possess dynamic characteristics that allow for vibration control thereby enabling high-speed operation, and traverse of rotor critical speeds. For this reason, hydrodynamic bearings are the most common type of bearing applied to steam turbines. Journal bearings are most often of the plain cylindrical, elliptical, multilobed, pressure dam, or tilt pad design. Figure 8.4 shows a schematic of a tilt pad journal bearing. Tilt pad journal bearing designs consist of several pads arranged in a ring around the shaft with the pads free to tilt about their respective pivots. Tilt pad journal bearings may include several design variations such as self-aligning features to compensate for misalignment, and special oil feed and drain configurations for temperature and power loss control [13,14]. One particular advantage of tilt pad journal bearings is their dynamic characteristics and inherent resistance to rotordynamic instability, which allows for control of vibration even at high speeds. Thrust bearings are usually of the tapered land or tilt pad design. Figure 8.5 shows a six shoe selfequalizing tilt pad thrust bearing. Tilt pad thrust bearings may also have features to compensate for misalignment, as well as special oil feed and drain configurations for temperature and power loss control [15,16]. Hydrodynamic bearings are lubricated with turbine grade oil either by a low-pressure circulating supply system or by ring lubrication where appropriate. In low-pressure supply systems, the oil flow is metered to each bearing by an orifice or other flow-controlling device. The oil flows into the clearance spacing of the bearing where it forms a wedge separating the bearing and shaft surfaces. The oil exits axially out the sides of journal bearings; and exits radially and tangentially from thrust bearings. Observation of drain oil flow through sight boxes can be taken as an indication of at least partial flow through the bearing and is often used as a quick indication that the oil pump is running and that the oil supply is probably sufficient. Oil supplied to the bearings functions as both a lubricant and as a coolant to
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Handbook of Lubrication and Tribology
FIGURE 8.5 Self-equalizing tilt pad thrust bearing. (From A General Guide to the Principles, Operation and Troubleshooting of Hydronamic Bearings, Publication HB, Kingsbury, Inc., Philadelphia, PA, 1997. With Permission.)
counteract the heat generated by shearing of the oil during operation and conduction from the hot rotor. Hydrodynamic bearings are limited with respect to minimum film thickness, maximum bearing temperature, and peak oil film pressure. These restrictions are inherently related to the load, speed of operation, and design of the bearing [17]. The bearings may be boundary lubricated during startup and turning gear operation, developing a full film shortly after startup. Operational film thickness is typically 25 to 75 µm (0.001 to 0.003 in.). Bearing metal temperature at the instrumented location may range from less than 55◦ C (130◦ F) for an unloaded inactive thrust bearing up to 130◦ C (265◦ F) for a bearing operating near its design limits. Peak oil film pressure is typically 2.5 to 3 times the specific load defined as P=
W A
(8.1)
such that P is the specific load (N/mm2 ), W is the load (N), and A is the projected area (mm2 ) [17]. For journal bearings, the area is the product of the diameter and length. For thrust bearings, the area is the area of the loaded surface. Bearing surfaces consist of a soft metal bonded to a hard metal backing. For North American operation, the soft metal surface is most often an ASTM B23 grade 2 babbitt comprised of 89% tin alloyed with antimony, lead, copper, iron, and trace amounts of other metals. Equivalent specifications can be found in ISO 4381 as SnSb8Cu4 [18], and Federal Spec Q-T-390 Grade 2. In some cases, an ASTM B23 grade 3 babbitt is used. Babbitt bearing surfaces generally cause the least damage to steel shafts when operated with inadequate lubrication or with contaminants. Babbitts are good for embedding hard contaminant particles and for resistance to seizure and galling [19]. In addition, tin-based babbitt is highly resistant to corrosion from organic acids and can provide satisfactory operation in the presence of oxidized and contaminated oils. A disadvantage of babbitt bearing materials is their relatively low compressive, tensile, and fatigue strengths especially at high temperature. To provide additional strength, the babbitt surface is cast and bonded as a thin layer to a hard metal backing, which may be steel, bronze, or chromium copper. Steel is the most prevalent and least expensive backing material. Chromium copper is used for its superior thermal conductivity enabling reduced bearing metal temperature. A good babbitt bond is critical, and can be inspected by nondestructive ultrasonic testing as described in ISO 4386 Part 1 [20]. The journal or thrust collar/disk is usually polished steel with surface finishes not exceeding 0.8 µm (32 µin.) Ra . The rotating element is either an integral part of the turbine shaft or else attached mechanically to the shaft. Bearing surface materials are normally steel containing less than 2.5% Cr, to prevent
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Steam Turbines
8-9
a type of failure known as wire wool [20]. In those cases where 12-Cr shaft material is required due to erosion concerns, a sleeve may be used as the bearing journal or the rotor may be inlaid with an acceptable steel material. To supervise the satisfactory and safe operation of turbine bearings, one or more of the following quantities may be monitored: inlet and drain temperature of oil, bearing metal temperature, drain flow of oil, shaft position, vibration amplitude, oil film pressure, and lift oil pressure. Bearing metal temperatures are measured using temperature sensors embedded in the bearing backing metal near the babbitt bond line and bearing surface [21]. High bearing metal temperature can be indicative of potential bearing failure. Bearing metal temperature that rises in an upward trend without corresponding change to load or speed is also indicative of potential bearing failure. There is varying opinion with respect to metal temperature limitation. In general, the manufacturer’s recommendation should be followed especially for new equipment lacking in historical data. Drain temperatures are also useful for identifying problems. Drain temperature of oil is an indicator of bearing power loss if measured separately for each bearing. It is also an indirect indication of bearing health, but not as reliable as bearing metal temperature. For this reason, drain oil temperature is not relied upon as an indication of safe operation unless the bearings are not instrumented. Radial and axial shaft position and vibration are measured with noncontacting eddy-current probes. Drastic shaft movement is an indication of bearing distress that occurs with wiping of babbitted surfaces [22]. Excessively high radial vibration is another sign of potential bearing distress and needs to be monitored as it may cause babbitt surfaces to fatigue or internal rubs to occur. Depending on the bearings, hydrodynamic bearing maintenance can consist of inspection, repair, or replacement. After installation, a lift check should be performed on each journal bearing, and the thrust bearing endplay and rotor axial position should be checked and recorded. These same checks should also be performed prior to bearing removal especially if abnormal bearing conditions were observed prior to shutdown such as high metal temperature or vibration. During visual inspection, the bearings are examined for signs of wear or distress such as scoring, cracks, pivot fretting or brinelling, heat discoloration, electrostatic discharge machining, corrosion, flaking, signs of overheated or contaminated oil such as varnish deposits, and loss of babbitt bond. The rotor journal and thrust areas also need to be examined for signs of distress such as scoring. Causes of bearing distress and failure include overloading, insufficient oil flow, insufficient bearing clearance or endplay, excessive overspeed, excessive vibration, and too high inlet oil temperature. Corrosion failures for tin babbitt bearings are fairly uncommon, but can occur in certain cases. The formation of hard deposits of tin oxide on tin rich white metal has been a problem with bearings in steam turbines caused by electrolytic action in certain environments such as when the oil contains free water with salt in solution [20]. Oil contamination from process gases that originate from the seal oil systems of driven units such as compressors can be particularly corrosive and may attack the components found in babbitt. 8.3.1.1.1 Hydrostatic Jacking Hydrodynamic bearings may include additional features such as an externally pressurized hydrostatic jacking system. The purpose of hydrostatic jacking is first to reduce the required breakaway torque during either start-up or turning gear operation and second to reduce bearing wear during turning gear operation. Hydrostatic jacking is effective by simply reducing the loading on the bearing surface such that it is within acceptable ranges. One manufacturer recommends consideration of hydrostatic jacking when the specific load on startup exceeds 1,300 kPa (190 psi) for plain journal bearings, 1,200 kPa (175 psi) for tilt pad journal bearings, and 60% of the maximum load for thrust bearings [23–25]. The need for hydrostatic jacking depends on the frequency of start-ups, duration of any baring condition, and available starting torque. Hydrostatic jacking systems are typically designed to lift the rotor off the bearings; however, this is not always practical. Hydrostatic jacking is effective so long as the friction torque is acceptable, the loading on the babbitt surface is reduced, and associated wear is negligible. Bearings with hydrostatic lift features require a high-pressure oil system, which typically supplies oil
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
at 7,000 to 14,000 kPa (1,000 to 2,000 psig). The high-pressure oil is turned off shortly after start-up and turned on during coast down. Lift oil pressure may be indicated by reading the pressure supplied to the lift pocket. 8.3.1.2 Rolling Element Bearings Rolling element bearings (also known as antifriction bearings) are used where service is not critical or the steam turbine is spared. These bearings can be used as a complete set to accommodate both radial and axial loads or are used as a thrust bearing in conjunction with ring-lubricated bearings. Rolling element bearings are generally less reliable than pressure fed hydrodynamic bearings and are only applied when they meet specific criteria with respect to their speed and life, which are designated by dN and L10 parameters, respectively. The dN parameter is the product of d, the journal diameter (mm) and N , the rated speed in revolutions per minute. Operation of dN in excess of 300,000 generally requires oil lubrication. The L10 parameter describes the basic rating life expressed in number of operating hours, or millions of revolutions with 90% reliability. The latest revised and updated L10 equation considers the bearing design, dynamic load, reliability factor, and life adjustment factor that involves the complex interaction of lubrication conditions, contamination, bearing material properties and other factors [26]. Rolling element bearings are generally designed and retained in accordance with American Bearing Manufacturers Association (ABMA) standards. The Conrad type or deep groove ball bearing is a typical design. The bearings are lubricated either by grease with protection against overgreasing, or by oil supplied by bath, mist, or jet lubrication. Grease fittings are required to extend outside the machine to permit regreasing during operation. Venting is provided to prevent pressure buildup in the housing. One particular disadvantage of rolling element bearings is that they cannot be horizontally split without reducing their life and degrading their performance. As a result, most rolling element bearings cannot be replaced without removing the rotor and coupling. Presence of water in oil is particularly detrimental to the life of a rolling element bearing [27].
8.3.2 Bearing Housings and Bearing Housing End Seals Bearing housings support and position the bearings such that the rotor is centered in its respective packing bores. These housings are also used to mount vibration monitoring and other condition monitoring devices. The steam end bearing housing further encases the overspeed trip assembly; as well as the governor speed sensor, which may consist of a notched wheel and speed pickup, or it may consist of flyweights or other devices. At times, a turning gear is also present. Grounding brushes may be mounted to the outboard end of the bearing housing to prevent the buildup of high voltage between the shaft and the case, which can damage the bearings through electrostatic discharge. Bearing housings also function as a part of the lube oil circuit, keeping oil in while keeping contaminants, such as steam out. In the case of pressure-lubricated hydrodynamic bearings, the housings are arranged to minimize foaming through proper design of the drain and vent system to maintain oil and foam levels below shaft end seals. Proper sizing of drains is important to minimize foaming. Bearing housings are equipped with replaceable labyrinth end seals and deflectors where the shaft passes through the housing to minimize contamination and leakage. Bearing housings and gland seals are spaced to help prevent leaking oil from entering the glands and gland steam from entering the bearings. For ring-oiled bearings, the housings further act as oil sumps and may contain water jackets for cooling the oil. Bearings housing oil seals may suffer from oil carburization, contaminant leakage into the seal or oil leakage from the seal. Contaminant leakage into the bearing housing can be a problem when using a vapor extractor on the main oil tank, which creates a slight vacuum in the bearing housings through the oil drain lines. Pressurizing the annulus in the oil seal with a gas purge such as nitrogen or air can assist with
© 2006 by Taylor & Francis Group, LLC
Steam Turbines
8-11
seal leakage. This may also cool the oil seal to prevent carburization. Overheating of the oil seal may also be prevented by improvements to the heat guards [28].
8.3.3 Steam Control Valves, Governors, and Control Systems As shown in Figure 8.6, steam is directed into the turbine steam chest through either a trip or trip and throttle (T&T) valve. Trip valves are opened by fluid pressure and mechanically closed by spring force. The trip system is controlled by an overspeed governor. A trip due to overspeed or other unsafe operating condition causes the solenoid valve to open thereby causing system depressurization and immediate closure of the trip valve, which shuts off the steam thereby bringing the turbine to an eventual stop. After passing through the trip valve, the steam is directed through the steam chest, and then through control valves (also called governor valves). The control valves throttle the steam into a nozzle ring matching the turbine power to the load thereby controlling speed. The control valves may be operated by mechanical linkage, by bar-lift arrangement (Figure 8.6), by cams, or by individual hydraulic cylinders. Mechanical and pneumatic actuators can be found on the smallest turbines whereas hydraulic actuators are required on most other units. Extraction turbines have additional valves located at an intermediate stage in the turbine. Extraction valves may be of poppet or spool type for higher pressure, or of grid type for controlling large volumes of steam at lower pressure. In each case, the valve actuators are controlled by the main governor. The main governor operates independently from the overspeed governor. The main governor can be either a relatively simple system that acts directly upon a steam-admission valve; or a complex system that may control speed, extracted steam, and devices separate from the turbine such as a compressor or the boiler. Figure 8.7 shows a mechanical–hydraulic governor with hydraulic actuator. In this case, hydraulic accumulators are used to supply the high volume of fluid that is required for rapid control action during sudden changes in load. To achieve the high force levels required in multivalve applications, the governor typically controls a servo (prepilot or slave) to a master pilot that controls the flow of high pressure oil to a large piston as shown in Figure 8.7. The assembly of servo, pilot valve, and piston is called a servomotor. In such a control system, a few ounces in governor force can be multiplied through a hydraulic mechanical advantage to generate the thousands of pounds of force that may be required to operate the turbine governor valves [29]. Required hydraulic oil pressures typically range from 350 kPa (50 psi) on small turbines to 18,000 kPa (2,600 psi) on very large turbines [30,31]. Turbine oils are typically employed at pressures below 2,000 kPa (290 psi) whereas fire-resistant fluids are often used at pressures exceeding 2,000 kPa and in installations where steam pipe temperatures exceed the auto-ignition temperature of turbine oil, particularly in power plant applications [30]. The governor and actuator control system may be supplied from the same lube system as the bearings or may be fed independently from a separate system. In small turbines, hydraulic and mechanical–hydraulic governors are often self-contained units featuring a shaft driven oil pump, and an oil sump with sight glass for determining the oil level. In medium sized turbines, typical of process industry applications, the control system is normally fed off the same lube system as the bearings. In large power plant turbines, two separate circulating systems are usually employed: one for the bearings using turbine oil and one for the control system using a fire-resistant phosphate ester fluid. Control systems have long had high visibility due to reliability and maintenance shortcomings. Large quantities of mechanical components such as pins, links, levers, rod end bearings, hydraulic relays, springs, gearing, and flyball governor assemblies are present and subject to wear. The use of electronic speed sensors and electronic governor controls has enabled the elimination of some wearing mechanical parts, and has improved control and flexibility through use of noncontacting pick-ups and nonmechanical feedback circuits. Actuators have remained primarily hydraulic due to the large forces and quick response time required. Governor maintenance depends considerably on the type of governor in use, and the manufacturer’s recommendations should be followed. The proper oil must be selected, and it must be kept clean, dry, and at
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High-pressure oil supply from oil console Variable-pressure control oil Drain oil to oil console reservoir
Steam to turbine
Trip and throttle valve
Inlet steam valves
Trip pin
Spring-loaded handle Solenoid valve
Trip lever Knife-edge
Bearing housing Servo motor
Orifice
High-pressure oil from oil console
FIGURE 8.6
To oil console drain
Trip system. (From Elliott Multivalve Turbines, Bulletin H-37B, Elliott Company, Jeannette, PA, 1981. With permission.)
© 2006 by Taylor & Francis Group, LLC
Handbook of Lubrication and Tribology
Electrical leads
Steam chest
Variable-pressure control oil
Steam Turbines
KEY High-pressure oil supply from oil console
Drain oil to oil console reservoir Governor internal high-pressure oil supply
Steam inlet
Governor intermediate pressure oil
Woodward PG governor
Trapped governor oil Pilot valve
Inlet steam valves
Drain oil to governor internal reservoir
Flyweights Pre-pilot valve
Accumulators
Oil pump Governor drain Governor drain
Lube system drain
High pressure oil from lube system
Mechanical–hydraulic governor system. (From Elliott Multivalve Turbines, Bulletin H-37B, Elliott Company, Jeannette, PA, 1981. With permission.)
© 2006 by Taylor & Francis Group, LLC
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FIGURE 8.7
Worm and wheel governor drive Inlet servo-motor
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Handbook of Lubrication and Tribology
the proper level and temperature. Water contamination, even in trace amounts, contributes significantly to early failure as well as forming oxides that also contribute to failures. As noted by one governor manufacturer, dirty oil causes most governor/actuator troubles [32]. Oil contamination and degradation is particularly problematic on self-contained sump units that do not have oil conditioners. Other parts of the governing system must also be maintained. Governor valves and linkages must be free from binding or sticking. Valve sticking may be related to steam impurity, which may lead to deposits on the valve stem [33], or it may be related to oil contamination causing deposits or corrosion in tight clearance hydraulic components [34], which may have clearance as small as 5 µm. Hydraulic actuators may operate on the valves through a linkage. Loose or worn linkage components can cause unacceptable governor control. Linkage bearings are usually hand-oiled or greased. Some, however, are made from low-friction materials, which may require little or no lubrication.
8.3.4 Turning Gear High-temperature steam turbines are sometimes equipped with a turning gear to prevent bowing of the rotor when at rest, especially after shutdown. The need for a turning gear depends upon the probability of rotor bow, which is related to the steam temperature, shaft diameter, and bearing span. Turning gears are primarily found on large turbines with long bearing spans, though they are sometimes needed for small turbines as well to allow for oil circulation through the bearings during cool down. The turning gear is operated prior to turbine run-up and immediately after shutdown. Turning gears are electric motor driven with a means for disengagement such as a clutch or retractable gear. The turning gear motor is typically grease lubricated whereas the actual turning gear and bearings are lubricated with oil supplied from the main circulation system. A separate, relatively small, motor-driven oil pump is generally provided to supply oil to the bearings of the turning gear system. The auxiliary oil pump, which backs up the main oil pump, may also be used for this service. During turning gear operation, oil inlet temperature may be kept cool to increase oil viscosity thereby maintaining a thick oil film in the turbine bearings during low-speed operation.
8.3.5 Couplings Couplings are used to connect the steam turbine to the driven equipment. They are made from corrosionresistant or coated materials. Couplings can be either rigid or flexible. Rigid couplings are essentially two flanges bolted together. Such couplings require no lubrication, but do not readily accommodate changes to machine position, which can be caused by thermal expansion of the equipment, foundation settling, and strain due to loading. Flexible couplings accommodate some misalignment; however, their use does not preclude the need for proper machine alignment of both the turbine and driven equipment [6]. Flexible couplings are described by a number of standards such as API 671, ISO 10441, and MIL-C-23233A. For turbine applications, special attention may be required with respect to machinery alignment due to thermal expansion. Quill shafts, membrane couplings, and contoured disc couplings run dry and without lubrication and are often preferred for their low maintenance. Gear couplings must be lubricated. 8.3.5.1 Gear Couplings Gear couplings can be advantageous because of their light weight and minimal required overhang, and because they allow for maximum axial movement between turbine and driven equipment shaft ends as caused by expansion of various parts under hot conditions [35]. In general, however, the need for lubrication and maintenance means that geared couplings are seldom used in new turbine applications though there is still a considerable population of geared couplings that must be maintained. The life of a geared coupling is primarily dependent on alignment and lubrication. The majority of geared tooth coupling failures are due to improper or insufficient lubrication [36]. Gear coupling lubrication is complicated by the centrifugal effect that a spinning coupling has on lubricants. Packed lubrication with grease can only be applied at relatively low speeds since the thickener tends to separate out of the grease under high
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Steam Turbines
8-15
centrifugal force [37]. Packed lubrication with either grease or oil may require lubricant replenishment at 6 to 12 month intervals [38]. Typical grease used for high-speed geared coupling is NLGI #1 or #2 grade with R&O inhibitors. Greases that are specifically formulated for high-speed coupling application use thickeners, which have a density closer to that of oil [39]. These formulations resist separation due to centrifugal effects. Special grease formulations can also extend replenishment intervals beyond the 6 to 12 months typically cited. A test method for evaluating grease separation is ASTM D4425. In this test, the grease is subjected to 36,000 G centrifugal acceleration at 50◦ C for a period of at least 6 h. Results from the test are presented as K 36 = V /H
(8.2)
where V is the oil separation in volume percent, and H is the accumulated time of testing in hours. High speeds and low maintenance in gear couplings require the use of continuous lubricant feed at each hub that is provided by either oil spray or jet using filtered oil piped from the system bearing oil supply. Coupling teeth for such applications are often hardened usually by nitriding [39]. Lubricants must be carefully selected with additives that resist separation from centrifugal force. Oil additives, in particular silicone antifoam compounds, can separate out of the lubricating oil and form sludge [40]. The coupling lubricant must also resist reaction with metal particles that may exist in the coupling due to wear [41]. Such measures are unnecessary when using dry couplings. The coupling housings provide safe enclosure of the coupling. Coupling housings may also act as part of the lube oil system. The housings are oil tight and include provision for coupling lube oil supply if needed and drainage back to the reservoir. The drains also handle any oil that may be carried over from the coupled equipment and are consequently featured on housings for both dry and lubricated couplings. A filter breather is attached to the coupling housing to allow proper drainage or the housing is connected to the bearing oil vent system of the equipment train. Regardless of the type of coupling used, proper design of coupling housings is important due to windage losses, heat generation, and potential for oil leakage from the joined equipment [42].
8.3.6 Additional Tribological Components and Issues Several components outside of the lubricating oil circuit require batch lubrication and special material consideration to limit wear and corrosion. Among these components is the turbine casing and steam patch components. Gland seals may also be subject to wear and have considerable effect upon water contamination of the lubrication system. 8.3.6.1 Casings Steam casings expand and contract due to changes in casing temperature caused by the use of high temperature steam. Thermal movement is typically accommodated at the steam end by either a flexible support or sliding pedestals. Sliding pedestals are most common on large turbines and rarely used on small and medium sized units. Sliding pedestals may operate dry, or they may be lubricated by either grease or oil depending on the load, temperature, and expansion. The use of lubrication reduces friction thereby allowing casing thermal expansion without binding. Binding of the casing can cause distortion, misalignment, and vibration. Grease lubricated casing supports are often used for large central station steam turbines. Grease may be supplied by either a common system or grease gun. The type and application should follow the manufacturer’s recommendation. NLGI number 1 or 2 grease of sodium, lithium, or sodium–calcium soap base have been used for lubricating sliding pedestals; in addition a mixture of graphite and cylinder or turbine oil mixed to a paste consistency has also been used [43]. Problems associated with grease separation have been noted on high temperature, heavy turbines used in central station applications due to the high temperature and heavy loads associated with these applications [44].
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Handbook of Lubrication and Tribology
8.3.6.2 Steam Path Parts that are located in the steam path consist of the steam chest, rotating blades, stationary nozzles, diaphragms’, seals, valves, and valve guides. These parts have tribological issues that must be solved without lubrication, which would contaminate the steam. Stationary nozzles and rotating blades are damaged by erosion and corrosion mechanisms. Turbine blades in particular suffer the most damage in turbines resulting in loss of power, efficiency, or operation. There exist many causes for wear through the steam path; however, three mechanisms are prevalent. These are (1) moisture impingement erosion, (2) erosion–corrosion, and (3) solid particle erosion. Moisture impingement erosion is caused by the presence of water droplets in the steam. When turbine blades operate in wet steam, the moist steam may cause blade erosion. Erosion is dependent on speed of the rotating blade, wetness of the steam, and blade design. Blade velocities can exceed 250 m/sec (825 ft/sec) at the tip. Moisture impingement erosion has been noted to be particularly problematic in the final stage of long multi-stage low-pressure turbines due to condensation. Allowable wetness is related to steam conditions, blade velocities, and design. In some cases, there is no need for a moisture limit. In other cases, 8% [45], 12% [46], or other moisture limit is used depending on the application. Moisture erosion also effects seals and can lead to degradation of performance and changes to the thrust loading [47]. Erosion–corrosion problems are caused by reactive steam chemistry. Steam of insufficient purity may cause deposits on the casing, nozzles, blades, seals, and sealing surfaces. These deposits may contain corrosive agents such as chlorine, which can attack the material used on these components. This results in eventual pitting and stress corrosion cracking [48]. Geothermal steam applications are known to have particularly corrosive steam with constituents of silica, sodium, ammonia, calcium, and sulfate. The acidity of geothermal waters can be very high with pH as low as 1.8 [49]. Solid particle erosion (SPE) is caused by entrainment of erosive materials in the steam. Solid particle erosion is traceable to exfoliated material coming from the boiler tubes, and in some cases, the steam leads. This type of erosion appears to be related to both the size of the unit and the pressure being employed. The solid particle erosive mechanism is most prevalent on large central station utility turbines; and it is rarely observed on small turbines operating under 540◦ C (1000◦ F) [50]. An important factor in each of these erosive wear mechanisms is the condition of the steam. For this reason, steam conditioning may be used for ensuring reliable operation. Monitors have been developed to quantify the particle loadings from the boiler [46]. Various separators and moisture removal devices may be employed upstream and inside of the turbine. Strainers are used to remove the largest particles and trap foreign objects. Some recommendations for steam purity are specified by NEMA for lowpressure turbines relating to the amount of dissolved solids, alkalinity, conductivity, and content of silicon oxide, iron, copper, sodium, and potassium [6]. Original equipment manufacturers also provide steam purity recommendations. Design methods for combating erosive wear include the use of either hardened materials or hard coatings such as Stellite on turbine blades. Stainless steels such as 12% chromium steel are also used. For turbine nozzles and internals, chromium steel cladding may be used [50]. Corrosion-resistant coatings have also been developed for this service. Other forms of wear and degradation internal to the steam path also exist. Turbines that are not in operation can experience a form of corrosion known as stand-by corrosion [45,49]. The corrosion is due to steam leaking into the turbine past a valve, which is not tight. Once the steam has leaked into the unit, it can condense and corrode the unit. This type of corrosion may cause severe pitting on stainless steel buckets. Brown specs (known as tubercles or scabs) form on carbon steel parts, such as discs and diaphragms [46]. It is therefore important that idle turbines have the inlet valve tightly seated and that all the casing drains be open [49]. Additional measures such as an additional drain between the turbine and steam inlet valve, and blanketing turbine internals with a positive flow of dry gas along with running the lubrication system and rotating the journals have also been performed [43]. 8.3.6.3 Seals and Gland System In order to maintain efficiency and performance, seals are required to limit steam leakage from the turbine case and between each stage. Casing end seals, also referred to as packing seals, are provided where the
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Steam Turbines
8-17
shaft ends pass through the casings. They are used to seal against the leakage of steam to atmosphere, and to seal against air suction into the low-pressure condensing case of condensing turbines. The seals may be carbon ring, labyrinth, or noncontacting mechanical design. Properly functioning gland seals are important for maintaining turbine performance. Improperly functioning gland seals can cause excessive steam leakage, which may ingress into the bearing housing leading to oil contamination. Likewise, a loss of vacuum on condensing turbines can also cause excessive steam leakage past the gland seals. Gland leakage does not have to be visible to cause a problem.
8.3.7 Driven Units Driven units such as gears, compressors, and generators create additional complexity to a lube oil system. Many of these units will have similar requirements to the turbine with respect to journal and thrust bearing lubrication. The oil selected for a common lube oil system must be suitable to all the pieces of equipment to be supplied. Low-viscosity rust and oxidation inhibited (R&O) oils, commonly called turbine oils, are used in many high-speed gear units where the gear tooth loads are relatively low [51] and the high entraining velocity of the gear develops thick elasto-hydrodynamic (EHD) oil films. Slower speed gears, as used for propulsion, tend to be more heavily loaded. These gears generally require higher viscosity lubricants with antiscuff additives [51]. High-pressure oil seals, as used in compressors; and hydrogen seals, as used in generators, can cause contamination of the seal oil by gas such that natural or vacuum degassing is required [31]. In some cases, a separate, isolated, lube oil system is used to provide seal oil due to potential contamination of the lubricating oil [52].
8.4 Lube Oil Systems Lube oil systems may be classified as either nonpressurized or pressurized systems. Nonpressurized lube systems consist of ring lubrication and are common on very small steam turbines. Larger turbines use pressurized lubrication.
8.4.1 Nonpressurized Oil Ring Lubrication Ring-lubricated hydrodynamic bearings are used where service is not critical or the steam turbine is spared. These bearings have the advantage of not requiring an external lube oil system thereby enabling steam turbine application where initial cost is a primary concern. Figure 8.8 shows an oil ring-lubricated journal bearing. The oil ring lubrication system employs metal rings to deliver oil to the turbine bearings. The rings are rotated by the journals carrying oil from a sump below the bearings to the top half bearing liners where it is fed into the clearance between the bearing liners and the shaft journals. Oil is drained from the ends of each bearing liner and returned to the bearing housing reservoirs to be cooled. Some of the supplied oil may be used to feed a rolling element bearing that is normally required in conjunction with ring-lubricated journal bearings for thrust positioning. The use of ring-lubricated bearings is limited with respect to load capacity, journal rotational speed, and by the need for cooling. Bearing housings may be double walled to allow water circulation to remove heat from the oil bath. Under conditions of high inlet steam temperature, the bearings can be damaged after shutdown because there is no longer oil circulation to carry heat away from the shaft, and a turning gear is sometimes used to continue the rotation of the shaft and subsequent oil ring lubrication. Ring-lubricated bearing housings are equipped with constant-level sight-feed oilers that maintain a constant reservoir oil level. A permanent indication of the proper oil level is clearly marked on the outside of the bearing housing. Low oil level in the housing will cause inadequate bearing lubrication. Excessively high oil levels can also be detrimental as it may restrict oil ring rotation also causing inadequate bearing lubrication. Housings for ring-lubricated bearings are provided with plugged ports positioned to allow visual inspection of the oil rings while the turbine is running.
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Handbook of Lubrication and Tribology Inspection plug Rotor shaft
Oil ring
Journal bearing
Oiler
Oil reservoir Cooling water Cooling chamber
Lubricating oil Cooling water
FIGURE 8.8 Ring-lubricated journal bearings. (From Installation, Operation, and Maintenance Instructions for YR Turbines, Elliott Company, Jeannette, PA, 2003. With permission.) TABLE 8.2
Lube System Design and Procurement Standards
Standard designation API 614 ASTM D4248 ASTM D6439
Standard title Lubrication, shaft-sealing, and control-oil systems and auxiliaries for petroleum, chemical, and gas industry services Design of steam turbine generator oil systems Standard guide for cleaning, flushing, and purification of steam, gas, and hydroelectric turbine lubrication systems
8.4.1.1 Ring Lubrication System Maintenance Proper oil level should be maintained at all times. Since oil ring lubrication systems have no means of filtering solids from the oil or removing water, periodic sampling and frequent oil changes are necessary to ensure a clean oil supply. The range of cooling water temperature must also be controlled to ensure good heat transfer without promoting condensation in the oil sump. To avoid condensation, the minimum inlet water temperature to the bearing housings should preferably be above the ambient air temperature.
8.4.2 Pressurized Lubrication Systems The pressurized lubrication system is essentially a closed loop system designed to provide an uninterrupted supply of cooled and filtered oil at the proper pressure to the bearings, control-oil system, shutdown system, and other components such as continuously lubricated couplings, as well as gears and seals on adjoining equipment. Oil consoles vary widely depending on the make, size, type, and purpose of the turbine and its adjoining equipment. Lubrication systems are designed according to the application, which may require the use of design standards such as those shown in Table 8.2. Proper lube system design is vital to machine reliability. A typical system is shown in Figure 8.9 and Figure 8.10. The oil is taken from the reservoir and passed through a cooler then filtered. The flow is split into two legs. One leg delivers high-pressure oil to a
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Steam Turbines
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FIGURE 8.9 API 614 lube oil console. (From Elliott Company, Jeannette, PA. With permission.)
common header for the governing and control mechanisms. A second leg delivers reduced pressure oil to a common header for the bearings. Bearing oil pressure is typically 100 to 125 kPa (15 to 18 psi), but may range from as low as 55 kPa (8 psi) in some systems to 345 kPa (50 psi) in others. Oil from the bearings and governor mechanisms will drain back to the reservoir. Figure 8.10 shows the following major components: • • • • • •
Oil reservoir Pumps and drivers Filters Coolers Control valves Piping
Additional accessories may include relief valves, transfer valves, accumulators, and instrumentation as shown in Figure 8.10. Not shown is the oil conditioning hardware. Each of the major components is described briefly along with instrumentation, commissioning, and system maintenance. 8.4.2.1 Oil Reservoir The reservoir is usually of rectangular shape, carbon steel construction with an interior coating of rust proofing paint. Solid stainless steel or stainless steel clad construction is also used. Normally the reservoir will have a sloping bottom to drain, clean out manways, gasketed openings, fill opening with strainer, oil level sight gage, and vent with weatherproof breather. The various oil levels as defined in the reservoir are shown in Figure 8.10. Depending on design requirements, the reservoir is sized to contain an amount of oil for anywhere from 3 to 5 min working capacity as measured from the minimum operating level. Large reservoir capacity enables disengagement of entrained air or gas and the settling of water and solid contaminants. A high and low oil-level indicator and alarm are usually provided. A free oil surface in the reservoir of at least 0.37 m2 /lps (0.25 ft2 /gpm) of oil is required to enhance air disengagement from the oil [53]. In addition, the oil reservoir is designed with a sloping bottom (1 unit in 24) such that supplementary water and dirt can accumulate at the area of the low point drain, and thus be drawn off during operation.
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Lube oil supply
8-20
High pressure control oil supply To oil reservoir To oil reservoir
Cooling water out
Oil return from units
A
A
PCV
Fill line
Cooling water in
PCV
TE
TI
Oil coolers VENT
Cooling water out
Vent
PCV
Main oil filters
TI PI
e
Purge
RV
RV
Minimum oil level and alarm level
PI
PI
TI
LS
LI Reservoir
Heater
Transfer valve 6-way
Relief/safety valve
Check valve
Pump suction strainer
Globe valve
Primary oil pump and driver
Change capacity (oil required for initial system fill) 800 U.S. gallons
Secondary oil pump and driver
Electric motor driver AC or DC power
Cooling fan Oil filter
Pump suction level
Inside bottom of reservoir
LIT
Tank drain TE
(3 min. retention)
Level indicator
PSH
Pressure switch, high
LIT
Level indicator transmit
PSL
Pressure switch, low
LS
Level switch
RV
Relief valve
Pressure control valve
TE
Temperature sensor
Pressure indicator
TI
Temperature indicator
LI
Orifice Gate valve Ball valve
FIGURE 8.10
Lube oil console P and I diagram.
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PCV Double pass, shell and tube heat exchanger
Concentric reducer PI
Handbook of Lubrication and Tribology
1/2” per foot minimum Oil to clairifier
Top of oil reservoir Maximum oil level Working capacity 360 gallons
Purge Fill conn w/strainer Oil from clairifier
PSL
Start secondary pump o setpoint (PSIG) falling press
Dirty oil drain
PDIT
Exhauster vent
Slop
PI
Drain
Cooling water in
Retention volume 600 gallons
Coolers and filter vents
Clean oil drain
B
B
Steam Turbines
8-21
Most pressurized lubrication systems are constructed with some provision for ventilation although some systems enjoy satisfactory operation with no such provision. Effective ventilation of the lubrication system enables the reduction of moist air that affects the service life of the turbine oil. The provision of adequate ventilation is also helpful in reducing foaming where trouble from this source is encountered [54]. The following methods of ventilation are commonly used: Natural ventilation, vacuum ventilation, or dehumidifier system [54]. Figure 8.10 shows a system equipped with a vapor extractor. The extractor pulls a slight negative pressure that should result in no more than −0.5 kPa (−0.07 psi) in the bearing housing to keep oil vapors from escaping, but without pulling in atmospheric contaminants. Reservoirs normally have a connection for an oil conditioning system. Such oil conditioners can provide further purification by removing water, acids, and other contaminants not removed by the filters. These are discussed in more detail under oil maintenance. 8.4.2.2 Pumps and Drivers Two or more oil pumps are normally supplied with the lube oil system. One pump is considered the main oil pump and the other, the auxiliary. The pumps are sized with additional flow capacity to provide a positive flow of oil under all normal operating conditions and most abnormal conditions to the turbine and the driven equipment. Additional smaller pumps may be used to supply oil for special purposes such as turning gear operation; hydrostatic lift oil for highly loaded bearings; seal oil for hydrogen-cooled generators; or oil transfer through filters [28]. Positive displacement pumps have relief valves located at each pump discharge line to protect the pumps and system against excessive pressure. The main oil pump can be driven off the main turbine shaft, by an electric motor, or by a small steam turbine. Most often, the auxiliary oil pump is driven from a different source of power than the main oil pump. The auxiliary pump driver is selected to reflect availability of power or steam under emergency conditions. Should the main pump fail, the auxiliary pump will automatically start. If the pressure continues to drop, the turbine and driven equipment will shut down. Emergency situations where both pumps fail are handled by either an emergency oil pump sized to provide last-resort lubrication for coastdown or a rundown tank that provides lube oil by gravity flow. Rundown tanks are common in marine applications. On lube systems where the auxiliary oil pump is driven by a small steam turbine, an accumulator is incorporated into the system. The accumulator will maintain the required oil flow while the turbine (auxiliary pump driver) is accelerating to speed preventing a system shutdown in case of main pump failure. 8.4.2.3 Filters Twin filters with multiple cartridge filtering elements are normally used in the lubricating oil system. Filters are operated in the full-flow mode such that all oil being circulated to the turbine passes through the filter. Using two filters permits filtering element changes while the equipment is in operation. Filtration ratings should be a minimum of 25 µm, and filtration of 10 µm is typically required. Filters are sized for a maximum pressure drop of 35 kPa (5 psid) when clean and passing oil at the design temperature. Filtering elements are typically replaced when pressure drop reaches approximately 100 kPa (15 psid) above original clean value [55]. The effect of water on the filter must be considered. Water and corrosion-resistant filter cartridge materials are preferred. Such water-resistant filter cartridges should not deteriorate even if water contamination reaches 5% by volume and an operating temperature as high as 70◦ C (160◦ F). Depth type elements (e.g., cotton and nylon) can suffer from a phenomenon termed “cartridge erosion,” where oil velocity enlarges or erodes the filter passages over time, which effectively invalidates the filtration rating of the element [55]. Cartridge erosion problems are eliminated by conservatively replacing filter elements every 6 months. Filters are also replaced if the pressure drop from clean increases by 100 kPa (15 psid). Recommendations of the filter element manufacturer should be considered.
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8.4.2.4 Coolers Oil coolers are usually conventional shell and tube heat exchangers with removable tube bundles. Normally, water flows through the tube bundle thus allowing for easy waterside cleaning. Oil flows through the shell side in a single pass. Coolers are usually operated so that the oil is at a higher pressure than the water, thus, reducing the severity of water contamination caused by tube failure. Typical cooling requirement is to cool oil to 120◦ F. Where conditions do not lend themselves to water-cooled heat exchangers, such as desert or subzero installations, air blast oil coolers must be considered. Coolers may be used for heating during initial oil system installation and cleansing and it is important that the system be designed for such use if desired [56]. Problems with maintaining oil temperature could be caused by improper venting, malfunctioning temperature regulators, incorrect water pressure, or badly fouled coolers [28]. Tube failure may be caused by fatigue and erosion. Excessive water flow can cause flow-induced vibration of the cooling tubes, but maintaining proper flow will reduce fatigue related problems. Water treatments and sacrificial anodes are used to retard corrosion failure of the cooler [56]. Cooler failures are responsible for the worst water contamination of the turbine oil. 8.4.2.5 Control Valves The backpressure regulator is designed to maintain a constant header pressure for all operating conditions. Normally, it is a self-operated valve, but where wide control ranges are required, pneumatic regulators complete with valve positioners are used. In addition to the backpressure regulator, pressure-reducing valves are required for all pressure levels below main header pressure. These valves are normally self-operated reducing valves but where wide control range is required, pneumatic operators complete with valve positioners may be furnished. 8.4.2.6 Piping Lubrication system components are joined together by the necessary piping to make the system functional. This includes provisions for the mounting of control instrumentation, such as pressure gauges, temperature gauges, switches, and monitoring and safety devices. Piping may be either carbon or stainless steel. Stainless steel is preferred due to superior corrosion resistance and is used extensively in refinery applications. The header piping connects the lube oil console to the various components being lubricated, such as bearings and seals. Used oil is returned to the reservoir through drain piping. Oil drains are sized to run no more than half full when flowing at a velocity of 0.3 m/sec (1 ft/sec) and are arranged to ensure good drainage. Horizontal runs slope continuously, at least 40 mm/m (1/2 in./ft), toward the reservoir [57]. 8.4.2.7 Safety and Monitoring Devices Lubrication system instrumentation is located throughout the system as shown on the schematic oil flow diagram. Monitoring devices, such as pressure gauges, safety devices, and alarm and trip switches are generally mounted on header piping close to the components being lubricated. A low-pressure start-up switch signals the auxiliary pump to start if pressure is too low. Temperature indicators are provided at bearing and seal outlets and at the inlet and outlet of coolers. Pressure indicators are generally provided at each pressure level. A sight-flow indicator is provided at the outlet of each turbine shaft bearing and each turbine thrust bearing. 8.4.2.8 Cleaning and Flushing All reasonable effort must be made to limit the introduction of contaminants into the lube oil system during construction. Proper cleaning and preservation of lube system components must be performed prior to system shipment. Different preservatives are used depending on the environment and expected storage time [58]. All units employing forced-feed oiling systems should have the entire lubrication system thoroughly flushed before operation. The importance of this step cannot be overemphasized. All dirt, rust scale, weld
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slag, or other contaminants that have been introduced into the oil system during storage, transportation, and fabrication at the jobsite must be removed by a continuous flushing operation or in extreme cases that do not involve stainless steel pipe, by pickling and cleaning. In addition to flushing during the commissioning, the system should also be flushed if left idle for a long time. Most turbine manufacturers provide special instructions for the oil flush. In the absence of such instructions, industry recommendations should be consulted such as those detailed in ASTM D6439 [59] or API RP 686 [58]. Flushing the system may require the use of external pump and such preparations should be made in advance. The bearings and bearing area should be bypassed until the system is proven to be clean. The flushing should continue until the required cleanliness is achieved based on inspection of the flushing filters or strainers, patch test, particle counters, or ISO 4406 cleanliness level. Flush oils, operating oils, and preservative oils must be compatible to preclude foaming, formation of emulsions, or breakdown of oil additives. Compatibilities and limitations may generally be obtained from the oil supplier. A system that is to use phosphate ester fluids must be flushed with phosphate ester fluid since such fluid is incompatible with mineral oil. The same may apply to other synthetic oils. 8.4.2.9 Lube System Maintenance Lube systems must be periodically inspected and maintained to ensure their proper operation. As a minimum, the following regular checks should be performed: • Check filter pressure drop and replace elements as recommended. • Check the oil reservoir level and add oil as required. • Periodically check operation of auxiliary oil pump by operating pump and returning to auxiliary duty. In addition, turnaround maintenance of the lube oil system should be performed at 1 to 3 yr intervals, as normal plant maintenance permits. Care must be taken to keep contaminants out of the lube oil circuit during bearing changes, filter changes, top up, and other maintenance activities.
8.5 Turbine Oil Equipment vendors often have turbine oil standards detailing the minimum characteristics required for successful turbine operation. In the absence of such standards, an internationally recognized turbines oil specification such as shown in Table 8.3 should be used.
8.5.1 Physical Properties Turbine oil performs four functions (1) Lubricate bearings and gears; (2) cool lubricated parts, carrying heat away from hot surfaces such as bearings and shafts; (3) act as a hydraulic fluid for governor, control valves, and safety devices; and (4) act as a sealant for gas seals such as hydrogen shaft seals in generators or gas seals on compressors. Each of these functions require an oil that is suitable with respect to several physical, chemical, and performance properties. Some physical properties frequently used to characterize turbine oils with corresponding American Society for Testing and Materials (ASTM) test methods TABLE 8.3
Standards for Turbine Oils and Hydraulic Fluids
Standard designation ASTM D4293 ASTM D4304 ISO 8068 MIL-PRF-17672D MIL-PRF-17331H
Standard title Standard specification for phosphate ester-based fluids for turbine lubrication Standard specification for mineral lubricating oil used in steam or gas turbines Petroleum products and lubricants — petroleum lubricating oils for turbines (categories ISO-L-TSA and ISO-L-TGA) Performance specification: hydraulic fluid, petroleum, inhibited Performance specification: lubricating oil, steam turbine and gear, moderate service
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8-24 TABLE 8.4
Handbook of Lubrication and Tribology Standardized Lubricating Oil Analytical Techniques
Physical or chemical property
ASTM designation
ISO designation
ISO viscosity grade Kinematic viscosity at 40◦ C, 100◦ C Viscosity index
D2422 D445 D2270
ISO 3448 ISO 3104 ISO 2909
Pour point Flash point Total acid number (TAN)
D97 D92 D974
ISO 3016 ISO 2592 ISO 6618
Foaming characteristics Air release
D892 D3427
ISO 6247 DIN 51 381
Water separability (Demulsibility) Rust prevention
D1401 D665
ISO 6614 ISO 7120
Corrosiveness to copper
D130
ISO 2160
Oxidation stability (TOST) Rotating pressure vessel oxidation test (RPVOT) Acid number Karl Fischer titration Color ISO cleanliness code
D943 D2272
ISO 4263
D664 D1744 D1500
ISO 6619 ISO 6296 ISO 2049 ISO 4406
Purpose Overall viscosity classification Relates to viscosity at normal operating conditions Empirical comparison of viscosity and temperature characteristics Measures low temperature flow properties Low value indicates volatile components Determination of acidity of new and used oils by titration with KOH Foaming characteristics of lubricating oils The oil’s capacity to separate entrained air over a period of time Emulsion characteristics of oil Ability of oil to prevent rusting of steel surfaces in presence of water Indicates tendency of oil to corrode copper and copper alloys Oxidation stability of mineral oils Tests remaining oxidation life of in-service oils Indicates acid level Measures the water content of oil Measures color Measures oil cleanliness
are summarized in Table 8.4. Detailed descriptions of the ASTM methods are available in the ASTM Handbook [59]. The most important physical property is viscosity. Table 8.5 gives the viscosity ranges for typical mineral lubricating oils used in steam turbines. Typical viscosity grade numbers are ISO-VG-32, VG-46, VG-68, VG-78, and VG-100 such that the viscosity grade numbers indicate the average oil viscosity in centiStoke units at 40◦ C (104◦ F). In order to reduce the power losses at the bearings and improve the responsiveness of hydraulic components, the lowest acceptable lubricant viscosity is normally selected. As a result, the usual lubricant employed in a common oil system is ISO VG-32 turbine oil cooled to a supply temperature of 120◦ F after the cooler. Other viscosity grades are also used. ISO VG-46 turbine oil cooled to a supply temperature of 140◦ F after the cooler is commonly used in desert, arid, and offshore applications where air blast coolers are utilized, or where the ambient temperature is quite high [56]. Oils used for ringoiled turbine bearings tend to be higher viscosity such as ISO VG-68 or VG-100. Oils used for shipboard propulsion may be ISO VG-68 to VG-100 and may have mild antiscuff additives. It is important to note that the lube oil system and the turbine rotordynamics are designed considering a specific oil viscosity. Turbine lube systems must be maintained with lubricants of the recommended viscosity, and the viscosity specification should not be changed without proper engineering review.
8.5.2 Formulation To achieve the desired physical, chemical, and performance properties, turbine oil is formulated with a base fluid and additive package consisting of rust and oxidation (R&O) inhibitors. Steam turbine oils are essentially special grades of R&O oils, formulated to give better oxidation resistance and longer life in a steam turbine [60]. While industry standard lube oil bench tests can provide great insight into the performance and life expectancy of turbine oils, both turbine original equipment manufacturers (OEMs) and oil suppliers generally agree that past successful performance of a particular oil under similar conditions is the best overall representation of quality and performance [61].
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Steam Turbines TABLE 8.5
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Physical Requirements for Turbine Oils
ISO viscosity grade Kinematic viscosity, mm2 /sec at 40◦ C, min. at 40◦ C, max. Military specification Military symbol ISO viscosity grade Kinematic viscosity, mm2 /sec at 40◦ C, min. at 40◦ C, max. At 100◦ C Pour point, ◦ C, max. Flash point, ◦ C, min. Viscosity index, min. Total acid number (TAN), mg KOH/g, max. Corrosiveness to copper, max. Rust prevention Water, percent Valve sticking characteristics Foaming characteristics Sequence 1, mL max. Sequence 2, mL max. Sequence 3, mL max. Air release Water separability Oxidation stability, min.
Light turbine oil
Medium turbine oil
Medium-heavy turbine oil
Heavy turbine oil
Test method
32
46
68
100
D2422 D445
61.2 74.8
90 110 MIL-L-17331J 2190 TEP
28.8 35.2 2075 T-H 32
41.4 50.6 MIL-L-17672D 2110 T-H 46
2135 T-H 68
D2422 D445
28.8 35.2 Report −29 157 94 0.20
41.4 50.6 Report −23 163 94 0.20
61.2 74.8 Report −18 171 94 0.20
74 97 8.0 −6 204
1 Shall pass None Shall pass
1 Shall pass None Shall pass
1 Shall pass None Shall pass
1 None None Shall pass
65/0 65/0 65/0
65/0 65/0 65/0
65/0 65/0 65/0
40/40/3 1000 h
40/40/3 1000 h
40/40/3 1000 h
65/0 65/0 65/0 20 40/—/3 1000 h
0.3
D97 D92 D2270 D974 D130 D665 D95 D892
D3427 D1401 D943
8.5.2.1 Base Oil The base oil stock of a turbine oil comprises more than 98% of the formulation. The base oil is categorized as either conventional solvent refined mineral-based (API Group I), or hydroprocessed mineral-based (API Group II) oil. Group II base oils contain fewer heteroatoms (sulfur, nitrogen, oxygen), and have less aromatic content than Group I base oils. When properly formulated, Group II turbine oils will have longer oxidation life, less deposit forming tendencies, improved water shedding ability, and overall higher performance than do Group I turbine oils [60]. One advantage of the conventional mineral-based (Group I) turbine oils is better innate solvency than the hydroprocessed (Group II) oils. The better solvency of the Group I turbine oils provides better additive package retention and increased ability to dissolve oxidation products that could otherwise potentially lead to varnish and sludge. While Group I and Group II base stocks are compatible with each other, the additive packages used to formulate the respective turbine oils may be incompatible with the overall mixture. Mixing oils can therefore cause sludge formation and additive dropout [62]. For this reason, compatibility between products is an important consideration when mixing two oils. 8.5.2.2 Additives Additives are used to improve the performance of the oil. Although additives are to some extent consumed in performing their functions, they can be replenished through normal lubricant make-up thereby enabling suitable performance for longer periods. Note that newer machine designs offer less oil loss and therefore do not benefit as much from this effect as did older machines exhibiting greater oil loss. The main types of additives include oxidation inhibitors, rust inhibitors, foam inhibitors, and demulsifiers.
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8.5.2.2.1 Oxidation Inhibitors Antioxidants are the additives, which have the strongest influence on the useful life of turbine oils. They generally function either by free radical inhibition, by hydroperoxide decomposition, or by deactivation of metal catalysts. The two major types of antioxidants used in turbine oils are arylamines and hindered phenols [63], and they work as free radical inhibitors. A mixed phenol-amine has certain advantages over the use of a single antioxidant system. Other additives and combinations of additives are also used to suppress oxidation. In particular, metal deactivators are used to suppress oxidation by reacting with metal ions and surfaces to inhibit their catalytic activity [64]. 8.5.2.2.2 Corrosion Inhibitors Highly refined oils lose their metal-wetting ability and are easily displaced by water. For this reason, corrosion inhibitors are necessary to prevent corrosion. New turbine oils contain a rust-inhibitor additive and must meet ASTM Test Method D 665. These corrosion inhibitors typically work based on the physical adsorption principle. In action, the corrosion inhibitor “plates out” on surfaces, forming a film that resists displacement by water and, therefore, protects the surfaces from contact with water [65]. Corrosion inhibitors used in turbine oils are polar and thus susceptible to water washout. Alkenyl succinic acids are therefore widely used due to their resistance to water washout [66]. 8.5.2.2.3 Foam Inhibitors Foam additives must be carefully selected in order to prevent excessive foam formation, but still retain short air release times [67]. Highly refined hydrotreated base oils have lower foaming tendencies than conventionally refined base oils. Foam inhibitors work by decreasing the gas-lubricant interfacial tension. Liquid silicones are an effective antifoamant, but also act as an air-emulsion stabilizer, negatively influencing the air release properties of the turbine oil as it resides in the stilling portion of the equipment. 8.5.2.2.4 Demulsifiers Demulsifiers destabilize oil–water emulsions by changing the interfacial tension of oil and water thereby allowing their separation [64]. Conventional mineral-based (Group I) turbine oils usually contain demulsifying additives whereas hydroprocessed (Group II) turbine oils have good demulsibility without an additional additive.
8.6 Performance Features of Turbine Oils The following oil performance features must be retained to ensure safe and continuous operation of the turbine (1) viscosity; (2) oxidation stability; (3) freedom from sludge; (4) anticorrosion protection; (5) water separability [68]; and (6) air separability and resistance to foaming.
8.6.1 Viscosity Viscosity is measured by ASTM Test Method D 445. Viscosity is the most important characteristic of turbine oil, as the oil film thickness under hydrodynamic lubrication conditions is critically dependent on the oil’s viscosity characteristics. Viscosity also affects journal bearing stiffness and damping properties, which determine the vibration characteristics of the turbine. Viscosity of most new oils may vary by ±10%. A change in viscosity up to 10% is not in itself likely to cause trouble; however, a change in viscosity of 5% from its original value should be investigated for the cause. A change in viscosity is usually caused by contamination or top off with the wrong lubricant rather than by degradation of the oil. Drop in oil viscosity is a particular concern where turbine driven compressors are used in the compression of hydrocarbon gases because the viscosity change may be caused by contamination of the oil from the lighter hydrocarbons [56].
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8.6.2 Oxidation Stability During steam turbine operation, the lubricating oil is subjected to relatively severe oxidizing conditions. These are due primarily to the influence of heat, the presence of water and entrained air, and the catalytic action of substances in contact with the oil, particularly copper and ferrous metals [69]. Under these influences, the antioxidants are gradually used up and the oxidation stability decreases. ASTM D943 Turbine Oil Stability Test (TOST) is used to evaluate the oxidation characteristics of new inhibited steam turbine oils. This is an accelerated oxidation test; actual service should be much longer than test report hours [61]. Since TOST testing can take longer than a year, it is impractical as an in-service oil test.
8.6.3 Freedom from Sludge and Deposits Deposits are generally formed due to oxidation of the turbine oil, soap formation, microbiological growth, contamination by water containing salts, and solid particulate contamination [66]. Process gases can also react with the oil and its additives to form deposits. One such example is a turbine driven ammonia compressor in which the oil became contaminated with ammonia. The acidic rust inhibitor used in the turbine oil reacted with ammonia to form an insoluble resinous product [70]. Filtration and centrifugation can remove sludge and other products from oil as they are formed, but if oil deterioration is allowed to proceed too far, sludge will deposit in parts of the equipment and system flushing and an oil change may be required [71].
8.6.4 Corrosion Protection Protection against rusting is very important due to the common presence of water in turbine oils and water vapor in the ambient air. Rusting may occur below the oil surface, at the oil surface, or in the vapor spaces above the oil surface. Rusting requires oxygen, water, a corrodible surface, and time. Effective corrosion protection requires the elimination of any one of these items. Oil acts to protect against corrosion by coating structural surfaces with corrosion inhibitor thereby denying access of water to corrodible surfaces. ASTM D665 is used to evaluate the rust-preventing characteristics of steam-turbine oil in the presence of water. Procedure A is used for land turbines where condensed steam or humidity from air is the water source. Procedure B is used for marine-service ocean-going vessels where salt water can be a water source. Present additive technology has been found to be highly effective at preventing rusting problems below the oil surface in full flow conditions. When rusting occurs below the oil surface, it is frequently caused by galvanic corrosion, and it is noticed in areas where there is little oil movement and where free water collects, such as the bottom of the oil reservoir. Galvanic corrosion is caused by contaminant particles settling out of the oil and the presence of water. Particulate matter can create galvanic cells and act as nuclei for air bubbles [34]. Factors that influence galvanic corrosion are impurity concentrations, the pH of the water, and temperature [66]. Galvanic corrosion shows up as black rust. Rusting at the oil surface is typically caused by liquid water standing on the surface [72]. Most rust problems occur above the oil in what is known as the vapor space. Vapor spaces are present in steam turbine bearing pedestals, oil return lines, sumps, and gear cases. The air in these vapor spaces will contain water vapor from the relative humidity of the air drawn into the system and from the evaporation of water entrained in the oil. In addition, salt particles that can act as corrosion-sponsoring nuclei also may be present [73]. Water vapor tends to condense on the cooler parts of the circulation system, such as the underside of the reservoir top, inside return-oil piping above oil level, in bearing pedestals, and around governor parts [74]. Corrosion in the vapor space results in formation of scaly red rust.
8.6.5 Water Separability (Demulsibility) A lubricant’s ability to separate readily from water is one of the most important requirements of a turbine oil. Water must readily separate from oil in the drain tank so that it is dry when pumped to the system.
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Demulsibility is influenced by oxidation and contamination from dirt or metallic particles. Resistance to oxidation helps preserve the demulsibility characteristics of the oil. Normally, if the oil is in good condition, water will settle to the bottom of the storage tank, where it should be drained off as a routine operating procedure [71]. Water may also be removed by purification systems. If turbine oil develops poor demulsibility, significant amounts of water will stay in the system and create problems such as increased oxidation, additive depletion, and corrosion. ASTM D1401 is used to test the demulsibility characteristics of oil.
8.6.6 Air Separability and Resistance to Foaming All oils will foam in some degree. Foaming of the present day turbine oils should not, however, occur unless the oil is contaminated or subjected to abnormal aeration. Antifoam additives suppress foam, but in doing so may also slow down air release leading to air entrainment. Air entrainment in the oil has been known to cause pressure surge in oil systems, interruptions in oil supply, excessive formation of foam [75], and reduced hydraulic control. Care must be taken such that improving the antifoam characteristics of turbine oil does not lead to unacceptable air separability characteristics. Turbine circulation systems have been constructed to eliminate conditions that have been found to cause foaming such as leaky pump suctions, excessive splashing of oil returning to the reservoir, oil-return lines of insufficient size or capacity, and insufficient venting. Wide differences in temperature between the fresh oil (as added) and the oil in the system may contribute to foaming [76]. Serious cases of excessive foaming may be due either to mechanical faults of the type listed or to oil contamination [77]. Problems with excessive foaming may also be due to mixing of incompatible lubricants [63] or the use of excessive antifoam inhibitor. Air entrainment issues are also affected by system design. In particular, the stilling period of the lube oil system can affect the air entrainment characteristics of oil. Machines that provide short stilling periods for the oil have displayed air entrainment/release characteristics that seem to counter those displayed during standard air release testing (ASTM D3427). Such machines with very short stilling periods have displayed increased air entrainment when nonsilicone antifoamants have been used and it is suspected that the silicone antifoams discourage the initial air entrainment during the agitation period [78].
8.7 Degradation of Turbine Oils in Service Factors responsible for oil degradation in service include contamination, additive depletion, oxidation, and bacteriological deterioration.
8.7.1 Contamination Contaminants will unavoidably find their way into the lubricating oil. The following types are most common: water, oil soluble contaminants, and solid particles. 8.7.1.1 Water Water is always present in oil in solution and may also be present in free or emulsified form. The solubility of water in oil is temperature dependent. Water in solution has no adverse effect on lubricating properties and will not cause corrosion; however, when hot oil subsequently cools, some water may come out of solution as very fine droplets dispersed throughout the oil [28]. This water is very likely to cause corrosion of steel parts and may also cause other problems, (e.g., foaming, sludge formation, and change of viscosity). In addition, water can also lead to oxidation, additive removal, bacteriological contamination, as well as reducing filter element life. Water enters the oil system from the condensation of humid air by system temperature fluctuation; from steam through the turbine gland seals; or from leaking oil coolers [73]. Leaking gland seals is the
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most common source of water. Leaking oil coolers is the most detrimental particularly since cooling water leaks will have moderate to high concentrations of dissolved solids. In extreme cases, a rupture of the heat exchanger can cause massive amounts of water to enter the machine compartment [79]. Free water generally exists above a saturation level of around 120 to 150 ppm, and oil becomes cloudy in the range of 200 to 500 ppm [80]. A centrifuge is effective in removing free water down to about 30 ppm above the saturation level. Different methods for the testing of water exist. The simplest is visual inspection followed by “crackle” or hot-plate test, which can indicate the presence of water in oil due to boil off. Another test is the Fourier Transform Infra-Red (FTIR) spectrometry. The Karl Fischer Titration, ASTM Test Method D 1744, is the most accurate method for water testing. Differing limitations for water are noted by different manufacturers. In general, the water content should never be allowed to exceed 2500 ppm (0.25%). ASTM D4378 cites 1000 ppm (0.1%) as a warning limit. Depending on the design and application, some manufacturers will require a limit of 500 ppm (0.05%). 8.7.1.2 Soluble Contaminants Oil soluble contaminants may include gases, solvents, other lubricants, flushing oils, preservatives, and sealants. Gases and some light solvents can be removed by vacuum dehydration methods. Other contaminants cannot be removed. The presence of such contaminants requires the consultation of the oil supplier and the turbine manufacturer. A common source of dissolved gases is the oil seals used in some generators and compressors. 8.7.1.3 Solid Particles Abrasive contaminants can damage bearings, journals, and control mechanisms. Improved practices such as better preservation of the turbine and its components when not in operation, high velocity system flushing during commissioning, and use of full flow filtration during operation have led to a significant reduction in failures due to abrasive contaminants [81]. Cleanliness of the system oil can be determined by gravimetric means by ASTM F 311 or F 312 or by particle counting. Allowable contamination level is dependent upon the individual turbine application and components in the system. ISO 4406 cleanliness levels ranging from 18/16/13 to 16/14/11 are commonly applied to steam turbine service. The three digits of the ISO 4406 code refer to the number of particles per milliliter greater than 4-, 6-, and 14-µm respectively. It should be noted that further reductions in contamination beyond manufacturer recommendations might lead to improvements in reliability that can be cost justified [82].
8.7.2 Additive Depletion Additives are used up in the performance of their function. In other cases, the additives are removed due to reaction with contaminants or drop out due to problems of compatibility. Oil suppliers are often able to replenish additives by sweetening the oil.
8.7.3 Thermal and Oxidative Degradation The oil acts as a heat transfer fluid with the overall system design determining the heat load on the oil. Factors such as smaller oil reservoirs, higher shaft surface velocity, and higher shaft and bearing temperatures all contribute to environmental conditions that degrade the oil by thermal stress leading to oxidation. Oxidation occurs by chemical reaction of the oil with oxygen. The first step in the oxidation reaction is the formation of hydroperoxides. Subsequently, a chain reaction is started and other compounds such as acid, resins, varnishes, sludge, and carbonaceous deposits are formed [71]. Oxidation products may further lead to rust and corrosion, and promotion of foaming and poor demulsibility. The oxidation rate is influenced by the presence of water, contaminants, entrapped air, and temperature. The oxidation rate of a fully inhibited mineral oil is quite low at temperatures less than 60◦ C and will double for every 10◦ C rise in temperature [83].
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For in-service oil testing, the oxidation stability reserve is best determined by the rotating pressure vessel oxidation test (RPVOT), ASTM test Method D 2272, and by total acid number (TAN), ASTM Test Method D 974.
8.7.4 Biological Deterioration Lubricating oils are susceptible to biological deterioration if the proper growing conditions are present. Procedures for preventing and coping with biological contamination include cleaning and sterilizing, addition of biocides, frequent draining of moisture from the system, and avoidance of dead-legs in pipes [71]. Sustained high water content can lead to bacterial and fungal growth in the system. This can cause filter blocking and formation of deposits. The most effective antimicrobial measure is the establishment of preventative procedures such as frequent draining of free water from the oil reservoir. Biocides are used to prevent microorganism growth. Sterilization by heat is also effective.
8.7.5 Turbine Oil Severity The expected service life of a turbine lubricant depends considerably on the severity of the application. Many low severity steam turbines have a history of requiring a full lubricant changeout only every 10 to 20 yr or longer, with periodic top-up with fresh oil [67]. Certain environmental conditions, however, can result in or accelerate lubricant degradation and reduce life. As noted, factors responsible for oil degradation in service include contamination, additive depletion, and thermal, oxidative, or physical breakdown. Other important factors affecting service life are (1) type and design of lubrication system, (2) condition of the system after construction, and (3) oil makeup rate. These factors vary from unit to unit so that service life is difficult to predict solely on original oil properties [84]. One method for determining the service conditions for each operating unit is to use a property called the turbine severity level (B), which is defined as the percent of fresh oil oxidation resistance or oxidation inhibitor lost per year due to oil reactions [85]. The equation for turbine severity is B = M · (1 − X /100)/(1 − e−M ·t /100 )
(8.3)
where B is turbine severity, % of fresh oil oxidation resistance lost per year due to oil reactions in the turbine, M is fresh oil makeup, % per year, t is years of oil use, and X is used oil oxidation resistance by ASTM D2272, % of fresh oil. A lubrication system with a high severity level requires frequent makeup or completely new charges, whereas one with a low severity level may have no problems with routine makeup [86]. The method requires periodic testing of the lubricating oil. Large steam turbines should have their turbine severity determined. The severity constant is different for each turbine, and varies widely between 5 and 30 for large turbines [87]. Figure 8.11 shows the importance of makeup rate for maintaining oil quality in a high-severity turbine where B = 25%/yr [85].
8.8 Lubricant Maintenance Small turbines with ring-lubricated bearings, and governors with sumps require periodic changes in lubricant. The quantities of oil are small, and it is often more economical to change the oil rather than to maintain it. Change periods of 1 yr are not uncommon and are set by regular change intervals, by monitoring the acid number or by more sophisticated monitoring. In larger turbine systems that employ circulating oil systems using more than 200 l (roughly 50 gal) of oil that require long periods of continuous operation, oil analysis generally proves more profitable than a routine time/dump program [88]. A life of up to 30 yr is desirable because of the outage and oil change
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100 90 80 70 60
40 30 M =3
0%
20 M 5%
=2
0%
10 9 8 7
M=2
M = 10%
M = 0%
Percent of original oil life remaining, X
50
6 5 0
4
8
12
16
20
Years of use
FIGURE 8.11 Effect of makeup rate on oil degradation for turbine severity, B = 25%/yr. (From DenHerder, M.J., Vienna, P.C., Lubrication Engineering, 37, 67, 1981. With permission.)
TABLE 8.6
Standards for Turbine Oil Maintenance
Standard designation ASTM D4378 ASTM D4057 ASTM D6224 ASTM D6439 IEC 60962 IEC 60978
Standard title Standard practice for in-service monitoring of mineral turbine oils for steam and gas turbines Standard practice for manual sampling of petroleum and petroleum products Standard practice for in-service monitoring of lubricating oil for auxiliary power plant equipment Standard guide for cleaning, flushing, and purification of steam, gas, and hydroelectric turbine lubrication systems Maintenance and use guide for petroleum lubricating oils for steam turbines Maintenance and use guide for petroleum lubricating oils for triaryl phosphate ester turbine control fluids
costs involved. In such systems, regular sampling and testing can indicate the need for oil conditioning. Many oil suppliers offer programs to meet specific lubrication maintenance requirements. Standards for turbine oil maintenance are listed in Table 8.6. Such standards offer a guideline for oil-monitoring and maintenance. Other methods may be applied depending on the application. One such standard, ASTM D4378, Standard Practice for In-Service Monitoring of Mineral Turbine Oils for Steam and Gas Turbines is used in the power generation industry [89]. As with any oil monitoring program, proper sampling is important. In-service oil should be tested at sufficient intervals to detect contamination, oxidation, and additive depletion. Key tests include appearance and color, water content, viscosity, total acid number, rust test, cleanliness, and RPVOT [89]. Systems that are exposed to volatile gases or liquids
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may also benefit from flash-point testing. Maintaining the lubricant may require new oil makeup, lube oil conditioning, and refortification.
8.8.1 New Oil Makeup Oil is lost due to leakage and to system maintenance such as draining off impurities, and filter changes. There is considerable variation with respect to the amount of makeup oil required for a steam turbine. Makeup rates can range from less than 5%/yr to more than 30%/yr in extreme cases. The average makeup rate in the United States is 7 to 10%/yr [89]. The compatibility of the system oil and the makeup oil are of critical importance. Compatibility is described as a lubricant’s ability to be mixed with another lubricant without detriment to the properties and the characteristics of either lubricant. The introduction of Group II oils has caused some concerns with respect to compatibility with Group I oils. In particular, the different additives and the solubility of those additives is a concern when mixing different oils especially those involving different base stocks. Problems involving excessive air entrainment, varnish particle build up, development of sludge, sticking of governor proportional valves, and plugging of governor filters has been noted on hydroturbines operating on turbine lubricants [90]. In some cases, a complete system flush may be required to introduce a new oil. The use of makeup oil that is the same oil as is already in the system is preferred for the elimination of compatibility issues.
8.8.2 Lube Oil Purification All circulating lube oil systems use filters to remove particle contaminants and purify the oil. Devices for removing liquid contaminants such as water will also improve system reliability. The most common devices for removing liquid contamination are settling tank, centrifuge, coalescing filter, and vacuum dehydrator. The settling tank works best on a batch basis. The centrifuge, coalescing filter, and vacuum dehydrator are applied continuously with 10 to 20% of the volume of oil in the turbine system every hour. Systems of this type tend to remove impurities as fast as they enter the oil, thereby avoiding accumulations. Settling tank — Oil contaminants that are heavier than oil can be separated by gravity alone. Such settling is best accomplished in a settling tank that is separate from the main oil tank. Settling times can be very long and the results are often less adequate than the onstream methods. Centrifuge — In a centrifugal purifier, or centrifuge, centrifugal force is used to accomplish the separation of contaminants heavier than oil. A separating force several thousand times that of gravity is produced by rotating the liquid at 7,000 to 15,000 rpm. The centrifuge is particularly effective in removing water and larger, heavier particles of solid impurities. The extent to which extremely fine solid particles are removed depends on the rate of throughput and other factors. Centrifuges are capable of removing free water and solids. Coalescing filter — A coalescing filter system uses special cartridges to combine small, dispersed water droplets into larger ones. The larger water drops are retained within a separator screen and fall to the bottom of the filter while the dry oil passes through the screen. Coalescers are capable of removing free water and solids. Vacuum dehydrator — A vacuum dehydration system removes water from oil through the application of heat and vacuum. The contaminated oil is exposed to a vacuum and heat. The water is removed as vapor. The vacuum dehydrator removes not only the free water, but also the dissolved and suspended water well below the solubility point (down to 10 ppm). In addition, vacuum dehydrators also deaerate and degasify the oil [91].
8.8.3 Refortification Refortification refers to the act of adding a predetermined amount of additive to a clean, dry, used lubricant to replenish some of the depleted additives [92]. In most cases, refortification and purification are used together.
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8.9 Fire-Resistant Fluids Fire-resistant fluids are used in the hydraulic actuators of large steam turbines operating at very high temperatures in excess of the auto-ignition temperature of turbine oil. Since the early 1970s, phosphate esters have been the only fire-resistant fluids approved by the turbine builders for use as a turbine control fluid although small amounts of more flammable carboxylate or synthetic esters have been used in refurbished systems [93].
8.9.1 Properties The main advantage of phosphate esters is their fire-resistance. Phosphate esters tend to have higher flash and fire points, higher auto-ignition temperatures and perform better in spray flammability and wick-type fire propagation tests [94]. Auto-ignition temperatures are in the region of 550–590◦ C. The triaryl version of the phosphate ester possesses inherent self-extinguishing properties because the fluid does not create enough energy to support its own combustion. Triaryl phosphates, in addition to their fireresistant properties, have good thermal stability, excellent boundary lubrication properties, low volatility, fair hydrolytic stability [94], adequate air release, and low-foaming properties. The density of phosphate esters is roughly 30% higher than mineral oils necessitating some additional consideration with respect to lube oil system design. Phosphate ester-based fluids are described in ASTM D4293. Viscosity grades are either ISO VG-32 or ISO VG-46. Phosphate ester fluids can be incompatible with some seal and insulative materials as well as certain paints thus making the pressurization system design and maintenance critical.
8.9.2 Degradation In service, phosphate esters are subject to deterioration as a result of hydrolysis, oxidation, and contamination. In the case of triaryl phosphate ester hydraulic fluids, contamination may be by water, particulates, mineral oil, and chlorine or chlorinated materials [95]. The principle degradation pathway for phosphate esters in steam turbine-generator lubrication systems is hydrolysis. While water is inevitably present in the fluids, its continued high concentration can be tolerated if fluid acidity is controlled [96]. As the solubility of water in phosphates is very much higher than in oil (reaching about 2500 ppm at 25◦ C), free water is not usually a problem and the level of fluid acidity will normally determine the suitability of the fluid for continued use. Many of the problems with the use of phosphate esters in turbine applications are associated with the development of acidity due to hydrolysis or oxidation. Since acidity development can cause corrosion, further accelerate the rate of hydrolysis, and is probably an early stage in the process of deposit formation, the maintenance of acidity levels of less than 0.5 mg KOH/g and preferably less than 0.2 mg KOH/g is strongly recommended [97]. Contamination by mineral oils can impair fire resistance, as well as being incompatible with various seals. High chlorine content can cause servo valve electrokinetic wear [95].
8.9.3 Condition Monitoring The following properties are considered necessary for the in-service testing of phosphate esters; appearance, chlorine content, color, mineral oil content, total acid number or neutralization number, fluid cleanliness, particle size, resistivity, viscosity, water content, and air release. The parameters that are of most concern are the increase in acidity, water content, and particulate contamination level. When triaryl phosphates degrade the most common result is an increase in acidity with little effect on viscosity change. Triaryl phosphate ester fluids are condemned if the acid number exceeds 0.2 over the original value (typically 0.03) [98]. Water should be kept below 2000 ppm. Alternate guidelines for maintenance and use of triaryl phosphate ester turbine control fluids can be found in IEC 60978.
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8.9.4 Maintenance The key to the cost-effective use of phosphate esters is the use of conditioning media to remove acid degradation products. Fuller’s earth and activated alumina have provided years of acceptable service; however, new adsorbents based on ion-exchange resins, may allow the fluid to be kept in the system for many years. Vacuum dehydration is required to remove the displaced water [93]. Phosphate ester hydraulic fluids require additional consideration of the lube oil system. Their use in high-pressure (1000 psi) systems requires fine filtration (0.5 to 5 µm) to protect more closely fitted pumps and control valves [99]. In addition, adsorbent filtration of phosphate ester hydraulic fluids using fullers’ earth, activated alumina, or ion exchange resin is needed to control fluid acidity. Adsorbent filters remove dissolved contaminants, such as acids, that are not removed economically or at all by other processes [100]. There is a tendency for these types of filters to remove additive materials. For this reason, adsorbent clay filters are typically not used on turbine oils, but are used for purifying fire-resistant phosphate ester hydraulic fluids as used in turbine control systems. The filters are most often used in a continuous bypass mode with 1–3% treatment ratio or they are used intermittently in accordance with changes in the acid number [97]. A fine particulate filter must be placed in series and downstream of the fuller’s earth filter to control particulates.
References [1] Steam, Its Generation and Use, 40th ed., The Babcock & Wilcox Company, New York, chap. 57. [2] Church, E.F., Steam Turbines, 3rd ed., McGraw-Hill Book Company, Inc. New York, 1950, chap. 1. [3] Electric Power Plant Design, Technical Manual TM 5-811-6, Department of the Army, Washington, DC, 1984. [4] API Standard 611, General-Purpose Steam Turbines for Petroleum, Chemical, and Gas Industry Services, 4th ed., American Petroleum Institute, Washington, DC, 1997. [5] API Standard 612, Special-Purpose Steam Turbines for Petroleum, Chemical, and Gas Industry Services, 4th ed., American Petroleum Institute, Washington, DC, 1997. [6] NEMA Standard SM-23, Steam Turbines for Mechanical Drive Service, National Electrical Manufacturers Association, Washington, DC, 2002. [7] NEMA Standard SM-24, Land-Based Steam Turbine Generator Sets 0 to 33,000 kW, National Electrical Manufacturers Association, Washington, DC, 2002. [8] Military Specification MIL-T-17286D, Turbines and Gears, Shipboard Propulsion and Auxiliary Steam; Packaging of, Department of the Navy, Washington, DC, 1989. [9] Military Specification MIL-T-17523, Turbine, Steam, Auxiliary (And Reduction Gear) Mechanical Drive, Department of the Navy, Washington, DC, 1982. [10] Military Specification MIL-T-17600D, Turbines, Steam, Propulsion Naval Shipboard, Department of the Navy, Washington, DC, 1982. [11] IEC Specification 60045-1, Steam Turbines — Part 1: Specifications, IEC, 1991. [12] Wills, J.G., Lubrication Fundamentals, Marcel Decker, Inc., New York, 1980, chap. 9. [13] Zeidan, F.Y. and Herbage, B.S., Fluid film bearing fundamentals and failure analysis, Proceedings of the 23rd Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1999, 161. [14] Nicholas, J.C., Tilting pad bearing design, Proceedings of the 23rd Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1999, 81. [15] New, N.H., Experimental comparison of flooded, directed, and inlet orifice type of lubrication for a tilting pad thrust bearing, Journal of Lubrication Technology, 96, 22, 1974. [16] Mikula, A.M., The leading edge groove tilting pad thrust bearing: recent developments, Trans. ASME, Journal of Tribology, 107, 423, 1985. [17] Gardner, W.W., Hydrodynamic Oil Film Bearings: Fundamentals, Limits and Applications, Waukesha Bearings Corporation, Waukesha, WI, 1998.
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[18] Kingsburg, G.R., Friction and wear of sliding bearing materials, in Friction, Lubrication, and Wear Technology, ASM Handbook, Vol. 18, ASM International, 1992, 741. [19] Ludema, K.C., Failures of sliding bearings, in Friction, Lubrication, and Wear Technology, ASM Handbook, Vol. 18, ASM International, 1992, 483. [20] Summers-Smith, J.D., A Tribology Casebook, Mechanical Engineering Publishing Ltd., Bury St. Edmunds, 1997, pp. 5–7, 91–92, 110–113. [21] API Standard 670, Machinery Protection Systems, 4th ed., American Petroleum Institute, Washington, DC, 2000. [22] Jackson, C., How to prevent turbomachinery thrust failures, in Compressor Handbook for the Hydrocarbon Processing Industry, Gulf Publishing Company, Houston, TX, 1975, 152. [23] Glacier Rotating Plant Bearings, Glacier Designers’ Handbook No. 5B: Tilting Pad Thrust Bearings, The Glacier Metal Co. Ltd., 1989. [24] Glacier Rotating Plant Bearings, Glacier Designers’ Handbook No. 10: Standard Tilting Pad Journal Bearings, The Glacier Metal Co. Ltd., 1989. [25] Glacier Rotating Plant Bearings, Glacier Designers’ Handbook No. 14: Medium and Thick Wall Journal Bearings, The Glacier Metal Co. Ltd., 1989. [26] Snyder, D.R., Selecting rolling element bearings for modern applications: how performance criteria, operating environments, materials, lubrication requirements and other factors shape the bearingselection process, Tribology and Lubrication Technology, 60, 28, 2004. [27] Cantley, R.E., The effect of water in lubricating oil on bearing fatigure life, ASLE Transactions, 20, 244, 1977. [28] Kure-Jensen, J., Large steam turbine-generators, in CRC Handbook of Lubrication (Theory and Practice of Tribology) Volume I: Practice, Booser, E.R., Ed., CRC Press, Inc., Boca Raton, FL, 1983, 91. [29] Bloch, H.P., A Practical Guide to Steam Turbine Technology, McGraw-Hill, New York, 1996, 136. [30] Ogle, A.W. and Render, M., Impact of market changes upon power plant control, in IMechE Seminar Publication: Steam Turbine Governing and Overspeed Protection, Professional Engineering Publishing Ltd., London, 1998, 5. [31] Hayler, M.G. and Wilson, A.C.M., Lubrication of water and steam turbines, in Lubrication in Practice, 2nd ed., Robertson, W.L., Ed., Marcel Dekker, Inc., New York, 1984, chap. 4. [32] Governing Fundamentals, Manual 25195, Woodward Governing Company, 1999, chap. 7. [33] Clark, E.E., Protecting against rotating equipment loss, Chemical Engineering, 104, 106, 1997. [34] Wilson, A.C.M., Problems encountered with turbine lubricants and associated systems, Lubrication Engineering, 32, 59, 1976. [35] Neale, M.J., Drives And Seals, A Tribology Handbook, Society of Automotive Engineers, Inc., Butterworth-Heinemann, Ltd., Oxford, 1994, 29. [36] Crease, A.B., Design principles and lubrication of gear couplings, International Conference on Flexible Couplings for High Powers and Speeds, Michael Neale and Associates Ltd., 1977, Paper B1. [37] Calistrat, M.M., Extend gear coupling life — Part I, in Compressor Handbook for the Hydrocarbon Processing Industry, Gulf Publishing Company, Houston, TX, 1975, 161. [38] Wright, J., Principle engineering features required by high performance gear couplings, International Conference on Flexible Couplings for High Powers and Speeds, Michael Neale and Associates Ltd., 1977, Paper B4. [39] Calistrat, M.M., Shaft-coupling lubrication, Lubrication Engineering, 37, 9, 1980. [40] Calistrat, M.M., Mechanical shaft couplings, in CRC Handbook of Lubrication (Theory and Practice of Tribology) Volume II: Theory and Design, Booser, E.R., Ed., CRC Press, Inc., Boca Raton, FL, 565. [41] Calistrat, M.M., Extend gear coupling life – Part II, in Compressor Handbook for the Hydrocarbon Processing Industry, Gulf Publishing Company, Houston, TX, 1975, 166. [42] Carter, D., Garvey, M., and Corcoran, J.P., The baffling and temperature prediction of coupling enclosures, Proceedings of the 23rd Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1999, 115.
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[43] Errichello, R., Friction, lubrication, and wear of gears, in Friction, Lubrication, and Wear Technology, ASM Handbook, Vol. 18, ASM International, 1992, 541. [44] Bloch, H.P., A Practical Guide to Compressor Technology, McGraw-Hill, New York, 1996, 372. [45] Transamerica Delaval Engineering Handbook, 4th ed., Welch, H.J., Ed., McGraw-Hill Book Company, New York, 1983, chap. 5. [46] Low, M.B.J., The sliding of high temperature components in steam turbines, Proceedings of IMechE, 194, 171, 1980. [47] Hackel, R.A. and Keyes, H.M., Steam turbines in process industries, in Sawyer’s Turbomachinery Maintenance Handbook, Volume 2: Steam Turbines/Power Recovery Turbines, Sawyer, J., Ed., Turbomachinery Intl. Pubns., Norwalk, 1980, chap. 4. [48] Steam turbines and auxiliaries, Power, 133, June, 1989. [49] Pilicy, F.X. and Dundas, R.E., Insurance for steam turbines, in Sawyer’s Turbomachinery Maintenance Handbook, Volume 2: Steam Turbines/Power Recovery Turbines, Sawyer, J., Ed., Turbomachinery Intl. Pubns., Norwalk, 1980, chap. 10. [50] Westhofen, B., Enhancing the availability of industrial turbines, Brown Boveri Review, 73, 31, 1986. [51] Cameron, J.A., Materials for use in geothermal steam turbines, in Geothermal ’77, Carrier Corporation, 1977, 39. [52] Ortolano, R.J., Steam turbine blading maintenance: part IV, Turbomachinery International, 24, 56, 1983. [53] D’Innocenzio, M., Oil systems – design for reliability, Proceedings of the First Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1970, 163. [54] Ventilation of steam turbines, Gargoyle Lubricants Technical Bulletin, 1945. [55] Sassos, M.J., Babyak, M.R., and Zerbe, J.P., Lubrication and seal oil systems — common problems and practical solutions, Proceedings of the 22nd Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1993, 51. [56] Salisbury, R.J., Stack, R., and Sassos, M.J., Lubrication and seal oil systems, Proceedings of the 13th Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1984, 151. [57] API Standard 614, Lubrication, Shaft-Sealing, and Control-Oil Systems and Auxiliaries for Petroleum, Chemical and Gas Industry Services, 4th ed., American Petroleum Institute, Washington, DC, 1999. [58] API Recommended Practice 686, Recommended Practice for Machinery Installation and Installation Design, 1st ed., American Petroleum Institute, Washington, DC, 1996. [59] Petroleum Products, Lubricants, and Fossil Fuels, Annual Book of ASTM Standards, Section 5, ASTM International, West Conshochocken, PA, 2003. [60] Schwager, B.P., Hardy, B.J., and Aguilar, G.A., Improved response of turbine oils based on Group II hydrocracked base oils compared with those based on solvent refined base oils, in Turbine Lubrication in the 21st Century, ASTM STP 1407, Herguth, W.R. and Warne, T.M., Eds, American Society for Testing and Materials, West Conshohocken, PA, 2001, 71. [61] Hannon, J.B., How to select and service a turbine oil, Machinery Lubrication, July, 2001. [62] Potential performance problems caused by mixing different types of turbine oils, Engineering and Construction Bulletin 2003-17, US Army Corps of Engineers, 2003. [63] Dolby, G.W. and Kofke, W.A., The role of base oils and additives in modern-day turbine oils, in Symposium on Turbine Oils, ASTM STP 321, American Society for Testing and Materials, Philadelphia, PA, 1962, 1. [64] O’Brien, J.A., Lubricating oil additives, in CRC Handbook of Lubrication (Theory and Practice of Tribology), Volume II: Theory and Design, Booser, E.R., Ed., CRC Press, Inc., Boca Raton, FL, 1984, 301. [65] von Fuchs, G.H., Rust inhibitors — their evaluation and performance, in Symposium on Turbine Oils, ASTM STP 321, American Society for Testing and Materials, Philadelphia, PA, 1962, 28. [66] vd Merwe, D.G., Morgan, P.M., Roets, P.N.J., and Botha, J.J., Additives in turbine and dynamic compressor oils, The South African Mechanical Engineer, 47, 13, 1997.
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[67] Swift, S.T., Butler, K.D., and Dewald, W., Turbine oil quality and field application requirements, in Turbine Lubrication in the 21st Century, Herguth, W.R. and Warne, T.M., Eds, ASTM STP1407, American Society for Testing and Materials, West Conshochocken, PA, 2001, 25. [68] Roberton, R.S., Background and development of ASTM D 4378: practice for in-service monitoring of mineral turbine oils for steam and gas turbines, in Turbine Oil Monitoring, ASTM STP 1021, Young, W.D. and Roberton, R.S., Eds, American Society for Testing and Materials, Philadelphia, PA, 1989, 3. [69] The Lubrication of Steam Turbines, Shell, Alabaster Passmore Sons, Ltd., London, 1958. [70] Summers-Smith, D., The unacceptable face of lubricating oil additives, Tribology, 11, 318, 1978. [71] Engineering and Design — Lubricants and Hydraulic Fluids, Engineer Manual EM 1110-2-1424, Department of the Army, Washington, DC, 1999, chap. 12. [72] Furby, N.W., Hanly, F.J., and Vincent, J.A., Rusting in turbine oil systems, in Symposium on Steam Turbine Oils, ASTM STP 211, American Society for Testing and Materials, Philadelphia, PA, 1956, 40. [73] Layne, R.P., Vapor space corrosion inhibition of steam turbine lubricating and cleaning oils, in Turbine Lubricating Problems, ASTM STP 437, American Society for Testing and Materials, Philadelphia, PA, 1968, 73. [74] Steam Turbines and their Lubrication, Mobil Oil Corporation, New York, NY, 1981. [75] Enz, W.E. and Hausermann, A., Particular problems of steam turbine lubrication, Proceedings of the Seventh Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1981, 125. [76] Steam Turbine Lubrication, 2nd ed., The Texas Company, 1947. [77] Fowle, T.I., Problems in the lubrication systems of turbomachinery, Proceedings of the Instrumental Mechanical Engineers, 186, 705, 1972. [78] Bice, C.D., Air entrainment issues in equipment using new generation turbine oils, Presented at STLE, Pittsburgh Section, November, 2004. [79] Fitch, J.C. and Jaggernauth, S., Moisture — the second most destructive lubricant contaminate, and its effects on bearing life, P/PM Magazine, December, 1994. [80] Coleman, W.L., Water contamination of steam turbine lube oils — how to avoid it, Proceedings of the Seventeenth Turbomachinery Symposium, 51. [81] Missana, A. and Steenburgh, J.H., Ensuring clean lube oil for large steam turbines, Power Engineering, 88, 46, June, 1984. [82] Bissett, W., Cost effective condition monitoring of large steam turbine/generator oil systems, Transactions of Mechanical Engineering, IEAust., ME20, 61–68, 1995. [83] Abner, Jr., E., Lubricant deterioration in service, in CRC Handbook of Lubrication (Theory and Practice of Tribology) Volume I: Practice, Booser, E.R., Ed., CRC Press, Inc., Boca Raton, FL, 1983, 517. [84] Lamping, G.A., Cuellar, Jr., J.P., and Silvus, H.S., Summary of maintenance practices for large steam turbine-generator lubrication systems, ASME/IEEE Power Generation Conference, ASME Paper 86-JPGC-Pwr-14, 1986. [85] DenHerder, M.J. and Vienna, P.C., Control of turbine oil degradation during use, Lubrication Engineering, 37, 67, 1981. [86] McCloskey, T.H., Troubleshooting bearing and lube oil system problems, Proceedings of the 24th Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1998, 147. [87] Ohgake, Ryoji, Sunami, M., Yoshida, T., and Watanabe, J., The reliable control of oil quality in Japanese turbine units, in Turbine Oil Monitoring, Young, W.C. and Roberton, R.S., Eds, ASTM STP 1021, American Society for Testing and Materials, Philadelphia, PA, 1989, 35. [88] Bloch, H.P., Criteria for water removal from mechanical drive steam turbine lube oils, Lubrication Engineering, 36, 699, 1980.
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[89] Roberton, R.S., ASTM in-service monitoring program for steam and gas turbine oils, Lubrication Engineering, 42, 466, 1986. [90] Micetic, J., New Generation of turbine oils, US Army Corps of Engineers, Hydroelectric Design Center, Portland District. [91] Adams, M.A. and Bloch, H.P., Vacuum distillation methods for lube oils increase turbomachinery reliability, Proceedings of the 17th Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1991, 41. [92] Stein, W.H. and Bowden, R.W., Shell Global Solutions, Inc., Turbine oil reclamation and refortification, Machinery Lubrication Magazine, July, 2004. [93] Phillips, W.D., The use of a fire-resistant turbine lubricant: Europe looks to the future, in Turbine Lubrication in the 21st Century, Herguth, W.R. and Warne, T.M., Eds, ASTM STP1407, 2001, 1. [94] Brown, K., Utility Service Associates, Condition-monitoring of phosphate ester hydraulic fluids, Machinery Lubrication Magazine, November, 2002. [95] Brown, K.J. and Staniewski, J.W., Condition monitoring and maintenance of steam turbinegenerator fire resistant triaryl phosphate control fluids, in Condition Monitoring and Preventive Proceedings, STLE SP-27, Society of Tribologists and Lubrication Engineers, 1989, 91. [96] Wolfe, G.F., Experience with phosphate ester fluids as industrial steam turbine-generator lubricants, Lubrication Engineering, 8, 413, 1978. [97] Phillips, W.D., The conditioning of phosphate-ester fluids in turbine applications, Lubrication Engineering, 12, 766, 1983. [98] Thibault, R., Converting to condition-based oil changes — part II, Practicing Oil Analysis Magazine, March, 2001. [99] Steele, F.M., Filtration and reclamation of turbine oils, Lubrication Engineering, 34, 252, 1978. [100] Steele, F.M., Contamination control in turbomachinery oil systems, Lubrication Engineering, 40, 487, 1984.
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9 Compressors and Vacuum Pumps 9.1 9.2
Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Part I: Basic Principles of Gas Compression . . . . . . . . .
9.3
Part II: Compressor Design and Lubrication . . . . . . . .
9-1 9-2
Gas Laws • Gas Compression Cycle
T. Kazama Muroran Institute of Technology Department of Mechanical Systems Engineering
G.E. Totten Portland State University Department of Mechanical and Materials Engineering
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Compressor Classification • Compressor Lubrication • Reciprocating Positive Displacement Compressors • Rotary Positive Displacement Compressors • Dynamic Compressors • Lubricants for Compressors
9.4
Part III: Vacuum Pump Design and Operation . . . . . .
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Introduction to Vacuum Pumps • Fundamental Vacuum System Relationships • Vacuum Pump Selection Criteria • Vacuum Pumps and Applications • Seals for Vacuum Pumps • Vacuum Measurement • Lubricants for Vacuum Pumps
References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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9.1 Introduction Compressors and vacuum pumps are vitally important industrial equipment. Gas compressors are mechanical devices that are used to pressurize and circulate gas through a process, facilitate chemical reactions, provide inert gas for safety or control systems, recover and recompress process gas, and maintain correct pressure levels by adding and removing either gas or vapors from a process system. Compressor and hydraulic pumps are similar devices with the main difference being that the fluid delivered by compressors is usually a gas which is compressed and under pressure at the time it is delivered, even if there is no load on the system. Gas compressors are used in nearly every industry sector including: automotive, steel, mining, food, gas and petroleum production, and storage and energy conversion. Global demand for compressor pumps is expected to grow and additional market opportunities exist for compressor manufacturers in the global energy production sector due to increasing consumption. Similar equipment market opportunities are expected in the natural gas extraction and transportation industries. Compressors are used not only for gas compression but also for cooling air conditioning and refrigeration. Refrigeration is that process used to remove heat from a material or gas so that its temperature is less than that of its surroundings. A refrigerant is the working fluid used to transmit heat in a refrigeration compressor, which is used to remove heat from the heat-laden refrigerant vapor in the evaporator
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and which compresses the refrigerant gas to a pressure that will liquefy in the condenser. Refrigeration compressors are either electrically or mechanically driven and they are used in stationary applications such as home and business air-conditioning and in mobile applications which include automotive and truck air conditioning and refrigeration. Rapid and significant market growth is expected for most types of refrigeration compressors. Compressors are loosely related to vacuum pumps. A compressor is used to reduce the volume of a gas, whereas a vacuum pump is essentially a compressor which operates with an intake pressure below atmospheric pressure. Vacuum pumps are vitally important equipment for nearly every industrial sector including the electronic semiconductor industry, steel making, automotive manufacturing, forestry, aerospace, and many others. Compressor manufacture is not only an enormous and growing industry, nearly all compressors require a lubricant to assure adequate cooling, sealing, and lubrication of internal components. The choice of a compressor lubricant depends on the type and construction of the compressor, the gas being compressed, the degree of compression, and the final outlet temperature. Therefore, it is important to understand the operation of various types of compressors and their lubrication mechanisms in order to select a proper lubricant for the compressor. This chapter contains three parts. Part I presents basic equations relating to gas compression. Part II provides an overview of industrial gas and refrigeration compressor design and lubrication requirements. Part III provides an overview of vacuum pump design and operation.
9.2 Part I: Basic Principles of Gas Compression 9.2.1 Gas Laws For the purpose of this discussion, it will be assumed that gases being compressed by a compressor follow the well-known Ideal Gas Laws. For example, Boyles Law states that pressure times volume of a gas is a constant if the temperature is constant: p1 V1 = p2 V2 The reference value used to determine pressure must be indicated. If the reference value is a vacuum, then the pressure is absolute pa . However, if the reference value is atmospheric pressure (patm ) then it is called gauge (pg ) pressure. They are related by: pa = pg + patm Absolute pressure must be used for ideal gas law relationships. When the pressure is constant, Charles Law states that the volume of a gas increases proportionately with temperature: V1 /V2 = T1 /T2 If the temperature of a gas increases as the pressure increases and the volume is kept constant, then Amonton’s Law is followed: p1 /p2 = T1 /T2 For these calculations, all temperatures are in reference to absolute zero. If the temperature is in ◦ C, then degrees Kelvin is calculated from: Kelvin = ◦ C + 273 Similarly, if the temperature is given in ◦ F, then degrees Rankine is calculated from: Rankine = ◦ F + 460
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The well-known Ideal Gas Law is obtained by combining the expressions for Boyle’s and Charle’s Laws: p1 V1 /T1 = p2 V2 /T2 Avogadro’s Law states that equal volumes of gas at the same temperature and pressure will contain the same number of molecules: pV = nRT where n is the number of moles, R is the gas constant that is selected to be consistent with the units of temperature, pressure, and volume used in the calculation. Dalton’s Law states that the total pressure (pt ) of a mixture is equal to the sum of partial pressures of each component (a, b, c, etc.): pt = pa + pb + pc + · · · . Similarly, the total volume is equal to the sum of the partial volumes of the constituent gases according to Amagat’s Law: V T = Va + V b + V c + · · · . If the temperature of a gas decreases or if the pressure increases sufficiently, the gas will undergo a change of state to a liquid. Further decreases in temperature or increases in pressure will convert the liquid into a solid. If the temperature increases sufficiently, a point will be achieved where the gas can no longer be liquefied by increasing pressure. The highest temperature at which a gas can be liquefied by increasing pressure is called the critical temperature of the gas. The pressure required to liquefy the gas at the critical temperature is called the critical pressure. When the pressure, temperature, and volume variation of a gas follow the Ideal Gas Law, the gas is referred to as an ideal gas. However, as the pressure increases, the behavior of a gas deviates from that predicted by the Ideal Gas Law. This is due to the compressibility of the gas and is accounted for in the Ideal Gas Law calculation by using the compressibility factor (Z ): p1 V1 /T1 Z1 = p2 V2 /T2 Z2 Values for compressibility factors for gases are obtained from charts called “general compressibility charts,” which are typically published in reference books. A series of thermophysical constants for commonly encountered compressed gases is provided in Table 9.1.
9.2.2 Gas Compression Cycle Figure 9.1 shows the p–V curve, namely, the relation of volume V , pressure p, and work [1]. Work is equal to force × distance where pressure corresponds to the force on the cylinder and volume corresponds to the distance the compressor piston moves. The area under the curve p × V is equal to the work performed during the gas compression cycle. There are two ways that a positive displacement compressor can be operated: either isothermally or adiabatically. For isothermal operation, temperature is held constant during compression by removal of the heat of compression and the work corresponds to: p1 V1 = p2 V2 Adiabatic compression (or expansion) occurs when there is no heat added or removed during the process. In adiabatic processes, pressure will vary with an exponential value of volume: p1 V1k = p2 V2k
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TABLE 9.1
Thermophysical Constants of Selected Gases
Gas
a
NH3 CO2 H2 N2 O2 H2 S CH4 C 2 H6 C 2 H4 C 3 H8 C 3 H6 C4 H10 C4 H10
Molecular weight (g/mol)
Critical temperature (◦ C)
Critical pressure (bar)
28.95 17.03 44.01 2.016 28.0134 31.9988 34.08 16.043 30.069 28.054 44.096 42.08 58.123 58.123
−140.5 132.4 31 −240 −147 −118.6 100 −82.7 32.2 9.5 96.6 91 152 134.9
37.71 112.8 73.825 12.98 33.999 50.43 89.37 45.96 48.839 50.76 42.5 46.1 37.96 36.48
1.202 0.73 1.87 0.085 1.185 1.354 1.45 0.68 1.282 1.178 1.91 1.81 2.52 2.51
Specific heat at CP CV (1.013 bar (1.013 bar and 21◦ C and 21◦ C [70◦ F]) [70◦ F]) kJ/(mol K) kJ/(mol K) 0.029 0.037 0.037 0.029 0.029 0.029 0.034 0.035 0.053 0.042 0.075 0.062 0.097 0.095
0.02 0.028 0.028 0.021 0.02 0.021 0.026 0.027 0.044 0.034 0.066 0.054 0.088 0.086
CP /CV Gamma-value
Specific volume (1.013 bar and 21◦ C [70◦ F]) m3 /kg
Compressibility (Z -factor) (1.013 bar and 15◦ C [59◦ F])
1.4028 1.309623 1.293759 1.384259 1.403846 1.393365 1.326 1.305454 1.193258 1.242623 1.134441 1.156832 1.093586 1.095845
0.833 1.411 0.547 11.986 0.862 0.755 0.699 1.48 0.799 0.862 0.543 0.587 0.4 0.406
0.9992 0.9929 0.9942 1.001 0.9997 0.9994 0.9915 0.998 0.9912 0.9935 1.0193 0.984 0.9625 0.9675
a The composition of air is: N = 78.09% by vol.; O = 20.94% by vol.; Ar = 0.93% by vol.; CO = 330 ppm by vol.; Ne = 18 ppm by vol.; He = 5.2 ppm by vol.; 2 2 2 Kr = 1.1 ppm by vol.; H2 = 500 ppb by vol.; Xe = 86 ppb by vol.; and Rn = 6 × 10−11 ppb by vol. b The data was obtained from Air Liquide from their Web site: http://www.airliquide.com/en/business/products/gases/gasdata/list.asp. c Except where otherwise noted, the data was obtained from Air Liquide website address provided in Note b.
© 2006 by Taylor & Francis Group, LLC
Handbook of Lubrication and Tribology
Airb Ammoniab Carbon dioxideb Hydrogenb Nitrogenb Oxygenb Hydrogen sulfidec Methaneb Ethaneb Ethyleneb n-Propaneb Propyleneb n-Butaneb i-Butaneb
Chemical formula
Density at (1.013 bar and 15◦ C [59◦ F]) kg/m3
Compressors and Vacuum Pumps
9-5
FIGURE 9.1 p–V curve. (From Figure 4 in Garg, D., Totten, G.E., and Webster, G.M., Compressor Lubricants, in Fuels and Lubricants Handbook: Technology, Properties, Performance, and Testing, ed. G.E. Totten, ASTM International, USA, 2003, chap. 14. With permission.)
where k is the ratio of specific heats. When the adiabatic process is reversible, it is referred to as an isotropic process (k = 1). When the operation is assumed to be adiabatic, the discharge temperature T2 can be easily estimated. For example, if T1 = 293 K, p2 /p1 = 11 and n = 1.4 (the ratio of specific heats, Cp /CV = k, of air), T2 is determined to be 308◦ C. This example indicates that the greater the amount of gas compression, the higher the final temperature. Although less work is required for an isothermal process, it is impossible to achieve it even though compressors are designed for as much heat removal as possible. Similarly, adiabatic processes are also impossible to achieve since some heat is always added or emitted in use. Therefore, what is typically achieved is a polytropic cycle: p1 V1n = p2 V2n where n (the polytropic index) is experimentally determined for each type of compressor and is usually not equal to k. The index n is unity for isothermal operation and n is the specific heat of the gas for adiabatic operation. (Thermodynamically, polytropic compression is defined as that process where: pV n = constant. A polytropic process differs from an adiabatic process in that the change does not occur at constant entropy since heat is either added to, or removed from, the gas. When heat is extracted by cooling, the n-value will be less than the adiabatic k-value.) The value of n may also be calculated from: T2 /T1 = (p2 /p1 )(n−1)/n where p is the absolute pressure, T is the absolute temperature, and the subscripts 1 and 2 are corresponding to the suction and discharge period respectively. Thermodynamically, adiabatic and isothermal processes are reversible but polytropic processes are irreversible, steady-state processes. A typical gas compression cycle where a gas is compressed in a piston cylinder from the inlet pressure ps to the discharge pressure pd along the line 1-2 is shown in Figure 9.2 [1]. Since it is impossible to discharge all of the gas due to the volume of the space not covered by the piston stroke, there will be a residual
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Handbook of Lubrication and Tribology
FIGURE 9.2 p–V cycle and piston stroke. (From Figure 3 in Garg, D., Totten, G.E., and Webster, G.M., Compressor Lubricants, in Fuels and Lubricants Handbook: Technology, Properties, Performance, and Testing, ed. Totten, G.E., ASTM International, USA, 2003, chap. 14. With permission.)
amount of gas referred to as the clearance volume. This is typically the area between the cylinder and the head of the piston illustrated in Figure 9.2(a). Clearance volumes typically range from 4 to 20%. Figure 9.2(b) illustrates the completion of the compression stroke along path 2-3. When the piston reaches the point 3, the discharge valve closes and the piston undergoes the expansion stroke 3-4 (shown in Figure 9.2[c]) until the pressure drops below the inlet pressure at the point 4. At the point 4, the inlet valve opens and the gas fills the cylinder as shown in Figure 9.2(d) and the process is repeated. Figure 9.2 illustrates the relation of the p–V curve and the piston stroke [1]. The schematic representations illustrated by Figures 9.2(a), (b), (c), and (d) correspond to the compression, discharge, expansion, and suction strokes respectively. There is a residual value referred to as clearance (dead) volume, since all of the gas is unable to discharge. Typical values of dead volumes range from 4 to 20%. The piston stroke is also illustrated in Figure 9.2. For Figures 9.2(c) and (d), the actual stroke is less than that represented by the line 4-1. The ratio of the actual capacity to the total displacement is referred to as volumetric efficiency. The volumetric efficiency is always less than the theoretical value because: • • • •
The reexpansion of the gas trapped in the cylinder clearances Entrance losses due to the pressure drop at the inlet Piston valve and ring leakage Slight increase in gas volume due to the heat rise from the warm cylinder
Theoretical volumetric efficiency ηv for polytropic compression for a diaphragm compressor is calculated from: ηv1 = 100 − C(R 1/k − 1) = 100 + C − CR 1/n
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Compressors and Vacuum Pumps
9-7
where R is the compression ratio that is defined as absolute discharge pressure divided by the absolute inlet pressure of the compressor and C is the cylinder clearance (%). For a piston pump, the equation for volumetric efficiency is: ηv2 = 100 − C(R 1/k − 1) = 100 + C − CR 1/k where the symbol is a leakage factor for piston compressors. It is determined experimentally for compression ratios of 6 to 10:1. The above equations refer to ideal gases. To be more generally applicable, these equations must be modified to account for the influence of the compressibility of the gas. This is done by using the compressibility factors for the gas at the inlet (Z1 ) and outlet (Z2 ) resulting in the following corrected equations for a diaphragm: ηvZ1 = (100 − C(R 1/k − 1) = 100 + C − CR 1/n )(Z1 /Z2 ) ηvZ2 = (100 − C(R 1/k − 1) = 100 + C − CR 1/k )(Z1 /Z2 ) For an intermediate pressure air compressor using a petroleum lubricant, the correction of the volumetric efficiency may be approximately 5%. It is necessary to determine the amount of power required to drive the system in order to properly size a compressor. This is done by determining the amount of brake horsepower required to compress a given volume of gas from the incoming inlet pressure to the desired discharge pressure [1]. Brake horsepower is defined as the ideal isoentropic (theoretical) horsepower plus any fluid (valve, flow, and other leakage) or mechanical friction losses. Theoretical horsepower (hptheory ) may be calculated from [1]: hptheory =
144 33,000
k (r − 1)k−1/k ps Vs k −1
where the units of the value 144 is in.2 /ft2 , 33,000 foot-pounds/min = 1 hp, k = the k-value for the gas, ps is the inlet pressure in PSIA, Vs is the inlet volume in ft3 /min, and R is the compression ratio. Compressors may also be characterized by specific power consumption pspec [1]: pspec = power consumption (kW)/volume flow (m3 /min) As indicated by the gas laws described earlier, gas compression is accompanied by a temperature rise. The greater the compression, the greater the temperature rise. If high discharge pressures are required, the compression process must be accompanied by two or more cooling stages, which improves efficiency and reduces power consumption. The temperatures actually encountered in gas compression must be considered in lubricant selection since the viscosity of the lubricant is temperature dependent and increased temperatures increase the potential for oxidation and deposit formation.
9.3 Part II: Compressor Design and Lubrication There are numerous compressor designs available. In this section, the design and operation of only the more commonly encountered compressor types will be discussed.
9.3.1 Compressor Classification Compressors, blowers, and fans increase the pressure of a gas or mixture of gases such as air, nitrogen, or a refrigerant and move it to where it can be applied to move actuators or refrigerating cycles [2,3]. Compressors may be classified in terms of their discharge pressure: high (>2000 kPa, gauge), intermediate
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Handbook of Lubrication and Tribology
(800–2000 kPa, gauge), and low (100–800 kPa, gauge). Compressors should be contrasted with blowers that are used to handle large volume of gases at gauge pressures ranging from 10 to 100 kPa and fans that are used at gauge pressures, which are typically less than 10 kPa [4]. Only compressors and their lubrication will be addressed here. Typical ISO 6743 compressor categories are provided in Table 9.2. There are two functional types of compressors: positive displacement and dynamic (or turbo). Displacement compressors increase the pressure of the gas by reducing its volume [5–7]. This is accomplished by taking in successive volumes of air that is confined within a closed space (such as a piston in a cylinder) and elevating the entrapped gas to a higher pressure. Displacement compressors are preferred for high pressures and relatively small volumes of gas. Examples of displacement compressors include: rotary and reciprocating. Rotary pumps can be one-rotor such as: sliding vane, liquid piston (liquid ring), rolling piston, scroll, and single screw types or they may be two-rotor types such as lobe or screw compressors. Reciprocating compressors include: crosshead, trunk, labyrinth (reciprocating piston), diaphragm compressors, and rocking piston compressors. Reciprocating compressors are available either as aircooled or water-cooled in lubricated and nonlubricated configurations, may be packaged, and provide a wide range of pressure and capacity selections. Dynamic compressors are turbo-compressors where impellers transfer rotational energy to a gas to be compressed. Examples of dynamic (or turbo) compressors include axial, radial, and centrifugal compressors. Centrifugal compressors, for example, produce high-pressure discharge by converting angular momentum imparted by the rotating impeller (dynamic displacement). To do this efficiently, centrifugal compressors rotate at speeds that are higher than those of other types of compressors. Dynamic compressors are also designed for higher capacity because flow through the compressor is continuous. Dynamic compressors may be further subclassified by: the number of compression stages, cooling method (air, water, oil), drive method (motor, engine, steam, other), lubrication (oil, oil-free), packaged or custombuilt. However, with the exception of lubrication, these additional classifications will not be discussed in detail here.
TABLE 9.2 Category of Compressors Compressors Positive displacement Reciprocating Piston Single acting Double acting Axial piston Swash plate Rotating Straight lobe Roots Multi-lobe Rotary Sliding vane Rolling piston Swing Screw Single Twin Scroll Liquid piston (Liquid ring) Dynamic (Turbo) Centrifugal (Radial) Axial Mixed
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Compressors and Vacuum Pumps TABLE 9.3
9-9
Duty Classification for Reciprocating Oil-Lubricated Air Compressors Operating conditions
Duty
Symbol
Normald
DAA
Severee
DAB
a b c d e
Intermittent or continuous operation Intermittent or continuous operation
Discharge temperaturea (◦ C)
Differential pressureb (bar)
Discharge pressurec (bar)
<180
<25
<70
≥180
≥25
≥70
Maximum temperature at discharge flange of any cylinder. Maximum differential between suction and discharge flanges of any stage. Maximum pressure at discharge flange of any cylinder. All criteria must be met. Applicable when any or all criteria are met. Source: Table 2A in Scales, W., Air Compressor Lubrication, in Tribology Data Handbook, ed. Booser, E.R., CRC Press, New York, 1997, chap. 19.
TABLE 9.4 Duty
Duty Classification for Rotary Oil-Flooded Air Compressors Operating conditionsa
Symbol
Normal
DAG
Severe
DAH
Intermittent or continuous operation Intermittent or continuous operation
Maximum air/oil temperature at discharge flange of air end, any stage, <100◦ C Maximum air/oil temperature at discharge flange of air end, any stage ≥100◦ C
a In some countries, temperatures higher than 90◦ C are not allowed because of legal restrictions.
Note: High air humidity and low circulation volume may be such as to recommend an oil for heavier duty. Source: Table 2B in Scales, W., Air Compressor Lubrication, in Tribology Data Handbook, ed. Booser, E.R., CRC Press, New York, 1997, chap. 19.
It should be noted that these operational categories have been given specific codes as described in ISO 6743 — Part 3A and shown in Tables 9.3 and 9.4 [8,9]. Choosing a compressor that best meets design requirements depends on the application.
9.3.2 Compressor Lubrication Compressor lubrication is essential and appropriate lubrication is essential to provide adequate heat removal, minimize friction and wear of bearings and other moving surfaces in contact, reduce internal leakage paths, protect against rust and corrosion, and prevent deposit formation on hot discharge surfaces [8,9]. The manner in which a compressor is lubricated will vary with the design of the compressor. Lubrication requirements [10] vary fundamentally with both major compressor types. Positive displacement compressors typically possess lubricated counter surfaces with rolling or sliding motion as well as bearing and sealing parts in the compression chamber. Dynamic compressors, however, utilize a hydrodynamic journal and thrust bearings or rolling element bearings for supporting the main shaft and are typically isolated from the compression chamber. Lubrication requirements may be further influenced by the operating conditions as working pressures and temperatures as well as by the characteristics of the gases. Lubrication problems are often dependent on the operation temperature of the compressor. For example, the temperature of rotary screw compressors typically varies from 80 to 115◦ C, which is sufficient to cause such operational problems as filter blocking deposits and varnish on bearings. Vane compressor temperatures typically vary from 80 to 150◦ C, which also lead to deposit formation that result in not only filter plugging but increased vane wear as well. Operational temperatures of reciprocating compressors are 270◦ C for single stage and 160 to 210◦ C for multistage compressor types. These temperatures are
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Handbook of Lubrication and Tribology
FIGURE 9.3 Effect of gas solubility on viscosity. (From Figure 2 Effects of Gas Solubility on Viscosity in George E. Totten, G.E. Totten and Associates, Inc. and Roland J. Bishop, Jr., Dow Chemical Company, Lubrication Requirements, Properties and Maintenance for Natural Gas Compressors. Machinery Lubrication Magazine. September 2002. With permission.)
sufficient to cause varnish and carbon deposit formation on exhaust and inlet valves and increased piston ring wear, which results in increased leakage [9]. Oil lubrication systems may be classified as a wet sump or a dry sump system. In a wet sump system the lubricating oil is contained beneath the crankshaft in an oil pan. In a wet sump, the oil pump sucks oil from the bottom of the oil pan through a tube into the compressor. A wet sump system should be contrasted to a dry sump system where the compressor oil is stored in a tank outside the compressor rather than in the oil pan. There are at least two oil pumps in a dry sump — one pulls oil from the sump and sends it to the tank, and the other takes oil from the tank and sends it to lubricate the compressor. The minimum possible amount of oil remains in the compressor. In “wet sump,” such as may be used for reciprocating and rotary screw compressors, the compressed gas and the lubricant come into contact with each other [12]. When this occurs, lubricant absorption into the gas being compressed will result in a lower viscosity as illustrated in Figure 9.3 [11] leading to a depletion of the lubricant film in the cylinder area leading to metal–metal contact and higher wear. (This occurs more often at pressure >3500 psi.) Some compressors, such as reciprocating compressors, do not recirculate the lubricating oil back into a reservoir for reuse. For nonrecirculating or once-through lubrication system, gas solubility in the lubricant is a particularly important issue because the physical properties, especially, viscosity and viscosity–pressure index are affected by the base oil and degree of gas solubility. In addition, phase separation during the cycle can result in serious problems such as accumulation of oil on the inside of the tubes, reducing heat transfer, oil starvation, and thus potential breakdown [12]. Specific lubrication requirements and systems for several compressors are discussed in subsequent sections. The solubility of natural gas and other hydrocarbons is much greater in petroleum oils and polyalphaolefin (PAO) synthetics compared to other commonly used synthetic base stocks such as diesters and polyalkylene glycols (PAG) because both hydrocarbon gas and petroleum-based oils exhibit similar polarity (consistent with the classic understanding that “like dissolves like”) unlike diesters and PAGs, which are more polar than hydrocarbons. Relative methane solubility in a petroleum oil, PAO, and a PAG fluid is illustrated in Figure 9.4 [11]. Hydrocarbon gases are considerably more soluble in mineral oil and PAO-based compressor lubricants, whereas the solubility of hydrocarbon gases increases with increasing pressure at a constant temperature in a less compatible fluid such as a PAG lubricant illustrated in Figure 9.5 [11,13]. Conversely, increasing the
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Compressors and Vacuum Pumps
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FIGURE 9.4 Solubility of methane in different lubricants. (From Figure 3. Solubility of Methane in Different Lubricants in George E. Totten, G.E. Totten & Associates, Inc. and Roland J. Bishop, Jr., Dow Chemical Company, Lubricant Requirements, Properties and Maintenance for Natural Gas Compressors, Machinery Lubrication Magazine. September 2002. With permission.)
FIGURE 9.5 Lubricant viscosity dilution chart for methane. (From Figure 4. Lubricant Viscosity Dilution Chart for Methane in George E. Totten, G.E. Totten & Associates, Inc. and Roland J. Bishop, Jr., Dow Chemical Company, Lubricant Requirements, Properties and Maintenance for Natural Gas Compressors. Machinery Lubrication Magazine, September 2002. With permission.)
temperature at a constant pressure will result in lower gas solubility as illustrated in Figure 9.6 [11]. Because increasing gas solubility decreases viscosity, at some point the viscosity reduction of the compressor lubricant may be sufficient to lead to lubrication failure due to the loss of hydrodynamic lubrication [14]. For refrigeration compressors, sealing as well as lubrication are important and the lubrication mechanisms of stationary and mobile compressors are different. The major differentiation between the compressors of room air conditioners, which are classified as stationary, and automotive air conditioners, which are classified as mobile, is the sealing mechanism between the drive shaft and the prime mover. Some room air conditioner compressors possess a hermitically sealed electric motor. In contrast, automotive
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Handbook of Lubrication and Tribology
FIGURE 9.6 Solubility of hydrocarbon gases in a PAG. (From Figure 1. Solubility of Hydrocarbon Gases in a PAG in George E. Totten, G.E. Totten & Associates, Inc. and Roland J. Bishop, Jr., Dow Chemical Company, Lubricant Requirements, Properties and Maintenance for Natural Gas Compressors. Machinery Lubrication Magazine, September 2002. With permission.)
air conditioner compressors are driven by the engine through the drive shaft. Therefore, the shaft seal is semihermetically sealed. Other differentiation regarding compressor types and potential problems may be caused by gases such as ammonia, carbon dioxide, natural gas, methane, and propane. The issues are given briefly in the section on lubricants.
9.3.3 Reciprocating Positive Displacement Compressors Reciprocating compressors are the most common compressors on the market. They are basically piston compressors, also known as “articulated” (jointed) piston compressors. These designs utilize the same operating principle as a conventional car piston engine. Sizes vary from <1 to >5000 hp. Reciprocating compressors (recips) are typically intended to be operated intermittently; loaded approximately 60% of the time (no more than 80%) and unloaded (approximately 40%) the remainder of the time. Reciprocating compressors are preferred in applications requiring higher discharge pressures or great pressure differences and are normally selected for applications where the inlet flow rate is no greater than 6800 m3 /h. These compressors also provide the most flexibility, including multiple staging and the ability to add or remove staging as needed. Reciprocating compressors are available in hermetic, semihermetic, and externally driven versions. In the hermetic-type reciprocating compressor, the motor and compressor are contained in a common single housing. Failure results in replacement of the entire unit. In the semihermetic type compressor, the motor is connected to but not sealed in the same housing as the compressor and therefore can be serviced. In externally driven-type units, the motor is connected to the compressor by a flexible coupling; however, this type of compressor is not commonly used today. 9.3.3.1 Reciprocating Piston Compressors Figure 9.7 [15] is a schematic illustration of a conventional piston compressor. The essential components of this assembly are: crankshaft, connecting rod and piston, cylinder, and valve head at the top of the cylinder that contain the inlet and discharge valves. A crankshaft is a shaft with one or more cranks, or “throws,” that are coupled by connecting rods to the pistons. The articulated (jointed) piston rod assembly
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Compressors and Vacuum Pumps
P
9-13
a
d
b
c
V
FIGURE 9.7 Schematic of conventional piston compressor and p–V cycle (D.R. Garg, G.E. Totten and G.M. Webster, Compressor Lubricants, in, Fuels and Lubricants Handbook: Technology, Properties, Performance and Testing G.E. Totten, S.R. Westbrook, R.J. Shah eds, 2003, ASTM International, West Conshocken, PA, pp. 383–412. With permission.)
transfers rotational energy from a motor into a reciprocating motion used to compress gas in a cylinder. The crankshaft may be powered by either an electric motor or gasoline engine. Both valves, as illustrated in Figure 9.7 [15], are essentially thin metal flaps, one of which is mounted on the top and the other underneath a valve plate. As the piston moves down, a vacuum is created above it, which allows the gas to push open the inlet valve and fill the area above the piston. As the piston moves up, the gas is compressed and holds the inlet valve shot and the compressed gas pushes open the discharge valve and the gas moves to the compressed gas storage tank. With each stroke of the piston, more gas enters the storage tank, which increases the tank pressure. A pressure switch stops the filling process when the desired storage tank pressure is attained. For safety, the tank has a safety valve that opens if the pressure switch malfunctions. A reciprocating compressor may be single- or double-acting. It is single-acting when the compression is accomplished using only one side of the piston. A compressor using both sides of the piston is considered double-acting. As the piston moves in one direction to compress a gas, a vacuum is created on the other side. In the return direction, the opposite occurs. Double-acting compressors lubricate the frame or crankcase by constant circulation. A stuffing box (a packing box) as shown in Figure 9.8 [16] or piston rod seal must be added with a crosshead and piston rod connection to the crankshaft. Larger double-acting compressors are usually water-cooled. The crosshead is attached to the connecting rod using a wrist pin and the surfaces between the crosshead and the crosshead slide are lubricated by the same system as the main bearings and wrist pin. The piston rod is connected to the crosshead by one of two methods: direct threading or flanging. These connections are illustrated schematically in Figures 9.9(a) and (b) [17]. Most typical compressors are available in one- or two-cylinder models, although a larger number of cylinders are possible, typically up to four. Most two-cylinder models operate like one (single)-cylinder versions except that there are two strokes per revolution instead of one. Some two-cylinder compressors are two-stage compressors where one piston pumps gas into a second cylinder where the heat of compression is removed using an “intercooler,” then the gas is compressed to a final pressure in a second compression stage. Single-stage compressors are generally used for outlet pressures in the range of 70–100 psig. Twostage compressors are generally used for higher pressures in the range of 100–250 psig. In applications where very high pressures are required, a compressor may have as many as four stages. Pressures of up to 6500 psig are possible. For practical purposes most compressors over 100 hp are built as multistage units
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Handbook of Lubrication and Tribology Entrained air in pump discharge
Stuffing box
Air
Suction under vacuum
Air
FIGURE 9.8 Stuffing box. (From Figure in 1. Stuffing Box Components in Process & Industrial Training Technologies, Mechanical Seals & Packing Video Based Educational Packages, http://www.iglou.com/pitt/ volume3.htm. No permission.)
in which two or more steps of compression are grouped in series. The gas is normally cooled between the stages to reduce the temperature and volume entering the following stage. There are three distinct types of pistons used for reciprocating piston compressors: trunk-type, crosshead-type, and labyrinth-type. These three piston compressor types are schematically illustrated in Figure 9.10 [18]. In a single-acting trunk-type piston compressor, the piston is elongated, hollow, and open at one end in which the end of the connecting rod is pivoted. The piston rod, crosshead, and stuffing box are thus eliminated. The trunk-type pistons are of one-piece construction and have grooves and lands to hold the piston rings in position, the number of which will vary with the size and type of piston. The skirt of the piston absorbs the side thrust from the rotation of the crankshaft and keeps the piston properly aligned in the cylinder. The crosshead piston has a separate crown and skirt, which are held together by a piston pin. The downward load of the crown is absorbed by a bushing and since it is separate from the skirt, it encounters much less side thrust from the movement of the crankshaft. The separate skirt, which guides the piston in the cylinder, absorbs side thrust and transports lubricating oil to the scraper rings, exhibits less thermal distortion than the crown, and because it is separated from the crown it does not encounter any downward load. The term “crosshead” is a traditional term that refers to crosshead connection of the piston to the crankshaft described earlier. An example of a labyrinth piston compressor and a labyrinth piston design is illustrated in Figure 9.11 [19]. A labyrinth piston assembly is characterized by a piston where there are many “labyrinths” or helical concentric grooves in a direction opposite to that of the crankshaft to provide lubrication by a noncontacting, airtight, Labyrinth Effect. The piston is designed to provide minimum gas between cylinder and piston. A labyrinth piston type compressor is designed to be of airtight construction with labyrinth effect and the gas with normal specifications can be procured at completely dried condition. This type of compressor is most suited for handling gases such as N2 as well as O2 and other potentially combustible gases that may be ignited by frictional heat as complete dried gases (N2 etc.), combustible gases liable to be ignited by heat created by friction (O2 etc.). Incursion of mist and dust is minimized by the use of a labyrinth seal. Labyrinth compressors may be used at pressures up to approximately 4500 psig. When considering the lubrication process for reciprocating piston compressors, it is convenient to divide the parts that need to be lubricated into two categories: cylinder parts and running parts. The cylinder parts include pistons, piston rings, cylinder liners, cylinder packing, and valves. All parts associated with
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Compressors and Vacuum Pumps
9-15
(a)
(b)
FIGURE 9.9 (a) Threaded piston rod attachment and (b) flange piston rod attachment. (From Figure 16a Thread Piston Rod Attachment, Technical Bulletin of Burton Corblin, Basics of Gas Compression Bulletin BCTB-101, http://www.burtoncorblin.com/BCTB101.pdf. From Figure 16b Flange Piston Rod Attachment, Technical Bulletin of Burton Corblin, Basics of Gas Compression Bulletin BCTB-101, http://www.burtoncorblin.com/BCTB101.pdf. With permission.)
Trunk-type
Crosshead-type
FIGURE 9.10 Three distinct types of reciprocating piston compressors [18].
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Labyrinth-type
9-16
Handbook of Lubrication and Tribology
Piston sleeve surface
FIGURE 9.11 Labyrinth piston compressor and a labyrinth piston design. (From JSW Japan Steel Works Product. Reciprocating compressors what’s labyrinth piston. http://www.jsw.co.jp/en/product/machinery/compressor/ compressor_01_en.html. With permission.)
the driving end, crosshead guides, main bearing and wristpin, crankpin and crosshead pin bearings are running parts [20]. Running parts are usually lubricated by crankcase “splash” lubrication where the oil does not come into contact with the gas. Crankcase lubrication is typically achieved using a SAE 30 (ISO 100) or SAE 40 (ISO 150) grade oil which may be an R&O (rust and oxidation inhibited) or a zinc-containing antiwear oil. (Note: The equipment manufacturer specifies the required viscosity grade [20].) There are various methodologies for determining the lubricant feed rate to the cylinders, packings, and the like. One method is the use of a nomogram such as that shown in Figure 9.12 [21] or the guidelines provided in Table 9.5 [22]. Alternatively, one may use an equation to determine the feed rate such as [23]: Q = B × S × N × π × 1440/2,000,000 where Q is the amount of lubricant necessary to provide an oil film sufficient to cover the cylinder-ring contact area, B = bore diameter (in.), S = stroke (in.), and N = rotational speed (revolutions per minute, rpm). The value of B is cylinder bore diameter but for low pressures, it is the rod external diameter. This equation is applied by calculating the feed rates for all cylinders and rods. To calculate the oil feed rate for the rod packings, use the above equation except instead of cylinder bore diameter, use rod diameter. For break-in periods, it is generally recommended that 1.5 to 3 times the normal feed rate be used. When selecting an equation such as the example provided above, it is important to understand the intended purpose of the equation being used. For example, some equations are used for lubrication during the break-in period only and some assume a specific degree of water saturation of the gas being compressed, which will affect the lubrication requirements. Single-acting machines, which are usually open to the crankcase, utilize splash lubrication for cylinder lubrication. In single-acting trunk compressors, the same oil usually lubricates both cylinders and bearings. The cylinders often are splash-lubricated whereas the bearings may be lubricated by splash or pressurized feed. In some machines, both cylinders and crankcase components are force-fed for operation at higher speeds [2]. Crankcase lubrication is isolated from the lubrication system used for the compressor cylinder and rod-packings. Oil in the cylinder must also assist the sealing of piston rings [24], valves, and packing while
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Compressors and Vacuum Pumps
9-17
Stroke (in.) 4
6
8
10
12
32
Oil feed rate for compressor cylinders having:
30 14
28
16
26
Multiply by
24 22
18 16 14
1.0
(B) Bronze rings
1.7
(C) Wet field gases
3.3
(D) Wet CO2 and bronze rings
2.5
(E) Transmission on gases and 0.4 PTFE trim (F) Mimi-lube transmission gases and PTFE trim
0.2 0.01
8
(G) Micro-lube transmission gases, PTFE trim, and special synthetic oil
6
Oil atomizer in inlet port
12 10 20” × 15” @ 400 RPM and 500 PSIG
Bore - in
20
(A) PTFE rings, packing, and riders
4.0 X
4 Oil free rate (pints per day) 2
Speed
1000
500
300
150
50
1600
1400
1200
1000
800
600
400
3000
200
5000
7500
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18
Discharge pressure (psig)
FIGURE 9.12 Oil feed rate. (From Figure 18.4, Majors, G., Compressor Lubrication, in Compressor Handbook, ed. Hanlon, P.C., McGraw-Hill, New York, USA, 2001, 18.8.)
© 2006 by Taylor & Francis Group, LLC
9-18 TABLE 9.5
Handbook of Lubrication and Tribology General Limits of Compressor Lubrication Oil Feed, drops/min (total for cylinder and packing)a Discharge pressure of cylinder, psig (bars)
Cylinder diameter (in.) 6–8 8–10 10–12 12–14 14–16 16–18 18–20 20–22 22–24 24–26 26–28 28–30 30–32 32–34 34–36 36–38 38–40
25–75 (1.72–5.17)
75–150 (5.17–10.34)
150–250 (10.34–17.24)
300–600 (20.68–41.37)
600–1500 (41.37–103.4)
1500–3000 (103.4–206.9)
3–6 4–7 5–9 6–10 7–12 8–13 9–15 10–17 11–18 12–20 13–21 14–23 16–24 17–26 18–27 19–28 20–30
5–8 6–9 7–11 9–12 10–15 12–17 13–19 15–21 17–23 18–25 20–26 — — — — — —
6–9 8–11 9–13 11–16 13–18 15–20 17–22 19–24 21–26 23–28 25–30 — — — — — —
7–12 10–15 12–18 14–21 — — — — — — — — — — — — —
9–16 12–20 — — — — — — — — — — — — — — —
12–18 — — — — — — — — — — — — — — — —
a The quantities given are based on an average of 8000 drops per pint of oil at 75 ◦ F (24◦ C). Note: The figures given are for gravity- and vacuum-type sight feed lubricators. For glycerin sight-feed lubricators, divide the figures by 3. Feeds to cylinder bores should never be less than one drop per feed per minute under any conditions. Source: From Table 2.8 Compressor Lubrication, General Limits in Compressed Air and Gas Handbook, 5th edn, Rollins, J.R., ed. Prentice Hall, 1988, 136.
resisting the effects of elevated temperatures and pressures. For these compressor components, lubrication is achieved by injecting the oil directly into the cylinders, packings, and in some situations, directly into the incoming gas upstream of the suction valves [15]. Oil injection systems may be point-to-point systems, which utilize a number of positive displacement pumps that pump oil from a small reservoir to the desired lubrication point on the compressor. Alternatively a distribution block system utilizes one or two larger pumps to force oil through sequenced valves to the desired point of lubrication [15]. Both cases are once-through systems where the oil is not recirculated or recovered. Since the cylinders are lubricated on a once-through basis, the oil must provide the lubrication and sealing function at the minimum possible rate of feed to avoid overconsumption and the possibility of deposit buildup on valves and cylinder heads. Compressor valves are lubricated from the atomized gas-lubricant in the system [20]. Large compressors used in heavy industrial process and pipeline service are usually of the double-acting crosshead type as shown in Figure 9.13 [2]. In these units, cylinders are separated from the crankcase by distance pieces that pass through a cylinder stuffing box. (Stuffing boxes are described in Section 9.3.3.2.) A distance piece is a long compartment with access doors to separate the cylinder from the crankcase and may be pressurized internally — if required. The crankcase side contains an oil seal with a scraper ring and a sealing ring. Although distance pieces are often a single design, a double compartment design with a wiper packing may be used for containment of flammable, hazardous, or toxic gases. As the crankshaft turns, the connecting rod converts the rotation to a reciprocating linear motion via the crosshead. The crosshead transfers this motion to the piston via the piston rod. The gas contained in the cylinders is compressed through the suction valve and discharged through the discharge valve. The double-acting compressor compresses the gas in both directions of piston movement. The same oil or different oils may be used to lubricate cylinders and crankcase components.
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Compressors and Vacuum Pumps
9-19
It is critically important that it be understood that reciprocating compressors used for air compression are fundamentally different than those used for compression of process gases such as methane, hydrogen, carbon dioxide, and nitrogen. Improper use of air compressors for a process gas application may lead to disastrous results including excessive leakage, fire, and even explosion. For example, air compressors are sealed from the crankcase by piston rings and the crankcase may be pressurized. Process gas compressors are designed with a distance piece to prevent leakage of gas into the atmosphere and crankcase. Due to its small diameter, the piston rod of a process gas compressor can be sealed within a packing case. Figure 9.14 [3] shows one of the reciprocating compressors, which is classified into trunk or automotive piston type that has a single-acting cylinder and is limited to refrigeration service and to smaller air compressors. The piston is connected to a connecting rod, which is in turn connected directly to the crankshaft [9]. Figure 9.15 [9] is an illustration of a typical double-acting reciprocating compressor, which consists of a crankshaft, connecting rods, crossheads, piston rods, pistons, cylinders and liners, and piston rings, which
Oil wiper ring Unloader cylinder Suction Inlet valve
Piston rod Connecting rod Flywheel
Crosshead
Crankshaft
Piston
Stuffing box piston rod seal Oil gauge
Foundation
Outlet Cooling water jackets
Discharge valve
FIGURE 9.13 Double acting, single stage compressor. (From Figure 2 in Burke, L.J., Compressors and Vacuum Pumps, in Handbook of Lubrication: Theory and Practice of Tribology, Volume 1, Application and Maintenance, ed. Booser, E.R., CRC Press, USA, 1994, 136. With permission.)
FIGURE 9.14 Cutaway of trunk-piston type two-stage compressor. (From Figure 3-2 in Brown, R.N., Compressors: Selection and Sizing, 2nd edn, Gulf Professional Publishing, USA, 1997, 50. With permission.)
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9-20
Handbook of Lubrication and Tribology
Cylinder liner
Crosshead
Crosshead guide
Piston rod Water jacketed packing cylinder
FE650 Piston rod Valve
Frame
Crankshaft Connecting rod
Counter weight
Variable volume clearance pocket
FIGURE 9.15 Cutaway of double-acting reciprocating compressor. (From Figure 7 in Garg, D., Totten, G.E., and Webster, G.M., Compressor Lubricants, in Fuels and Lubricants Handbook: Technology, Properties, Performance, and Testing, ed. Totten, G.E., ASTM International, USA, 2003, 387. With permission.)
are all mounted onto a suitable frame. As the crankshaft turns, the connecting rod converts the rotation to a reciprocating linear motion via the crosshead. The crosshead transfers this motion to the piston via the piston rod. The gas contained in the cylinders is compressed and discharged through the discharge valve. The piston rod packing, also known as cylinder packing, seals the high-pressure gas from the low-pressure crankcase [9]. In the lubrication of reciprocating compressors, since the physical properties of lubricants are dependent on the amount of the dissolved gas, that is, the viscosity and viscosity–pressure coefficient decrease with increasing the refrigerant mixing fraction, a distinction is made between the requirements of the parts that come in contact with the gas (cylinders, valves), and those that do not (crankcase bearings, connecting rods [25]). In both categories the lubricant must minimize friction and wear as well as provide protection against rusting from condensed moisture. Because gas temperature increases with increasing pressure, if heat is not removed, the lubricant will be exposed to high temperatures and undergo severe decomposition. Therefore, compressor cylinders are equipped with cooling jackets. One of the most important roles of the compressor cylinder lubricant is as a coolant. The coolant is usually water or a water–glycol refrigerant. Although the same lubricant can be used to cool both the cylinder and the running parts, there are many cases where different lubricants are used because the cylinder lubricant is exposed to compressed gas at high temperatures. Therefore, the lubricant should also exhibit thermal and oxidative stability [20]. 9.3.3.2 Stuffing Box For proper operation and to prevent outward leakage, it is necessary to provide a seal for the compressor shaft. One of the simplest types of shaft sealing devices is called a stuffing box. The stuffing box is a cylindrical space in the pump casing surrounding the shaft. Rings of packing material are placed in this space to control the rate of leakage along the shaft. The packing rings are held in place by a gland. The gland is, in turn, held in place by studs with adjusting nuts. By tightening the adjusting nuts, an axial force is applied, which compresses the packing to radially forming a tight seal between the rotating shaft and the inside wall of the stuffing box. High-speed rotation of the shaft generates a significant amount of heat as it rubs against the packing rings. If no lubrication and cooling are provided to the packing, the temperature of the packing increases to the point where damage occurs to the packing, the pump shaft, and possibly nearby pump bearings. Stuffing boxes are designed to allow a small amount of controlled leakage along the shaft, which provides lubrication and cooling to the packing. The leakage rate can be adjusted by tightening and loosening the packing gland.
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Compressors and Vacuum Pumps
9-21
FIGURE 9.16 Schematic of dry-running piston rod sealing systems. (From Hydrocarbon Processing, September 2004, Courtesy of Dr N. Feistel, Burckhardt Compression A.G. With permission.)
It is not always possible to use a standard stuffing box to seal the shaft of a centrifugal pump. These conditions require a modification to the standard stuffing box. In such cases, a lantern ring, which is a perforated hollow ring located near the center of the stuffing box that receives relatively cool, clean liquid from either the discharge of the pump or from an external source distributes the liquid uniformly around the shaft to provide lubrication and cooling. The fluid entering the lantern ring can cool the shaft and packing. Examples are provided in Figure 9.16 and Figure 9.17. Figure 9.16 shows the behavior of dry-running piston rod sealing systems in crosshead compressors and Figure 9.17 illustrates different concepts of heat dissipation from piston-rod sealing systems [26,27]. To obtain maximum packing and service life, it is desirable to seal the gas escaping the compressor at minimum pressure. In positive displacement compressors, such as the reciprocating piston compressors, suction pressure is reduced by use of internal and external porting from discharge to suction. In many pump designs, recirculation is required to cool and lubricate tight bearings and bushings. 9.3.3.3 Swashplate, Wobble Plate, and Bent Axis Piston Compressors Swashplate type axial piston compressors consist of an inclined plate called a swashplate, pistons, shoes (slippers), and driving shaft (Figure 9.18 [28]). The rotating displacement pistons (in the cylinder barrel) are supported by the static swashplate, which is rigidly attached to the rotating shaft. The angle of the swashplate determines the piston stroke and therefore the amount of displacement of the compressor. Thus, the amount of displacement can be varied by varying the angle of the swashplate. As the swashplate approaches the vertical position, the amount of displacement decreases. The unidirectional rotation of the shaft is transformed to simple reciprocal motion of the piston through the swashplate/shoe contact [28]. Relatively high loads are applied to the pistons through the slippers to the swashplate, using springs, thus allowing the slippers to maintain contact and slide on the swashplate at high rotational speeds. An alternative to the swashplate piston compressor is the so-called wobble plate piston compressor. Wobble plate compressors utilize a stationary piston block containing a number of parallel pistons arranged radially around the center of the block. The end of each piston is forced against a rotating wobble plate by springs. The wobble plate is designed to have varying thickness and is connected to a rotating drive shaft. As it rotates, the pistons reciprocate at a fixed stroke. The pistons intake gas (fluid, such as refrigeration fluid) from the cavity during half of a revolution and compress the fluid out at the rear of the compressor during the remaining half of a revolution. Gas (fluid) flow is controlled using nonreturn valves for each piston.
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Handbook of Lubrication and Tribology
FIGURE 9.17 Different concepts of heat dissipation from piston-rod sealing systems, in the 2005 January issue Compressor Tech Two. (Courtesy of Dr N. Feistel, Burckhardt Compression AG, Switzerland.)
Shoe
52100 steel shoe
Piston
390 A1 plate
FIGURE 9.18 Schematic configuration of swashplate compressor. (From Figure 1 in Yoon, H., Sheiretov, T., and Cusano, C., Scuffing Behavior of 390 Aluminum Against Steel under Starved Lubrication Conditions, Wear, 237(2), 163, 2000. With permission.)
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Compressors and Vacuum Pumps
9-23
Another reciprocating piston design is that of the Bent-Axis Axial Piston pump. In an angle- or a bentaxis-type piston design, the pistons are at an angle to the drive shaft and thrust plate. The piston block shaft is connected to the drive shaft by a universal joint. The drive shaft, piston block shaft, pistons, and thrust plate all revolve. As this unit revolves, half of the pistons suck in fluid as they pass over the intake port and the remaining pistons discharge their fluid through the outlet port. As in the swashplate design, the offset angle determines the displacement of the compressor. Lubrication of swashplate, wobble plate, and bent-axis compressors is due to slippage. Slippage is defined as the fluid leakage from a higher pressure outlet to a lower pressure area or back to the inlet. A drain passage is designed within the compressor to permit the leaking oil to return to an inlet or a reservoir. A certain amount of slippage is designed into these compressors to provide lubrication. Slippage will increase with system pressure and wear. Oil flow through a given orifice size depends on the pressure drop and an internal leakage path is the same as an orifice. Therefore, if pressure increases, more flow will occur through a leakage path and less from an outlet port. Any increase in slippage is a loss of efficiency. One of the tribological problems with these compressor designs is scuffing. The transitions from mild to severe wear and severe wear to scuffing are related to the destabilization of the protective layers caused by subsurface flow (see Section 9.3.3.6 on Materials of Reciprocating Compressors). 9.3.3.4 Reciprocating Diaphragm Compressors Diaphragm compressors are used in many industries including medical, environmental, printing industry, and chemical processing in addition to their use as vacuum pumps because of their advantages when compared with other compressor designs. Some of these advantages include: oil-free, maintenancefree, and uncontaminated operation. Because the gas being compressed is completely isolated from the lubricant, it is as pure as the gas in the incoming stream. Only clean, dry metallic surfaces are contacted. However, diaphragm compressors can be constructed for use with corrosive gases that will prevent contamination of the lubricant and the displacement element, piston, crankcase, and the like [29]. Depending on the design specifications, diaphragm compressors may be used in applications ranging from vacuum to pressures of up to 60,000 psig. However, most commonly these compressors are 1 hp or less in size. Piston and diaphragm compressors possess many of the same components: crankcase, crankshaft, piston, and connecting rods. The primary difference between the two compressor designs lies in how the gas is compressed. In a piston compressor, the piston is the primary gas displacing element. However, diaphragm compressors are reciprocating compressors where gas compression is achieved by the flexing of a thin metal or fabricated disk called a diaphragm, which is caused by the hydraulic system operated by the motion of a reciprocating piston in a cylinder under the diaphragm. The diaphragm completely isolates the gas from the displacement element (piston) during the work cycle. A hydraulic fluid transmits the motion of the piston to the diaphragm. A schematic illustration of a representative diaphragm compressor is provided in Figure 9.19 [30]. As Figure 9.19 indicates, the flexing of the diaphragm causes the gas chamber (between the diaphragm and the cover) to increase or decrease in volume. At the beginning of the work cycle, gas is sucked into the gas chamber through the suction tube via the suction valve. When the gas space is reduced, gas compression results and the compressed gas exits through the discharge valve into the discharge tube. The diaphragm completes one cycle as the crankshaft completes one revolution (and piston completes one stroke). A schematic of the suction and discharge valves of a plate-ring design is provided in Figure 9.20 [30]. In this design, the valves are sealed by metal sealing rings, which are held in place by thrust pieces. The diaphragm assembly is composed of three separate plates that are clamped gas-tight at the periphery between the cover and perforated plate. A seal is achieved with a metal O-ring. In this arrangement, only the following components come into contact with the gas: diaphragm cover, gas side-plate of the diaphragm assembly, compressor valves with thrust pieces and sealing rings, and the metal O-ring used to seal the diaphragm assembly. Figure 9.21 provides a schematic of a diaphragm pump and valve assembly [17].
© 2006 by Taylor & Francis Group, LLC
9-24
Handbook of Lubrication and Tribology Diaphragm
Perforated plate Piston
Gas space
Cylinder Crank drive Gas valves
Cover
Oil space Oil flange
FIGURE 9.19 Schematic of a diaphragm compressor. (From Figure 3: Scheme of a 1 stage diaphragm compressor in Technical essays, 1. Essay, Diaphragm Compressors by Manfred Dehnen, http://www.andreashofer.de/english/htm/info_essay_diaphragm_compressor.htm.), Courtesy of Andreas Hofer Hochdrucktechnisk GmbH. Discharge valve
Suction valve
Thrust place Cover
Calcher
Seat
Spring Plate
Plate
Spring
Seat
Calcher Diaphragm
FIGURE 9.20 Schematic of the suction and discharge valves of a plate-ring design. (From Figure 5: Compressor valves in plate/ring design in Technical essays, 1. Essay, Diaphragm Compressors by Manfred Dehnen, http://www.andreashofer.de/english/htm/ info_essay_diaphragm_compressor.htm.), Courtesy of Andreas Hofer Hochdrucktechnisk GmbH.
Lubrication in the upper end of the diaphragm compressor is not required since there are no moving surfaces in contact such as piston rings against a cylinder. Because diaphragm compressors operate at such high speeds, the diaphragm needs to be replaced at 3 to 4 month intervals. Bearing service life is limited owing to their very high operational speeds [31]. 9.3.3.5 Reciprocating Rocking Piston Compressors Another compressor design that is related to both piston and diaphragm compressors is the rocking piston compressor. A schematic illustration of a rocking piston compressor is provided in Figure 9.22 [15]. Instead of a piston or diaphragm, the rocking piston compressor has a flexible cup mounted, without a wrist pin, on a connecting rod with an eccentric bearing such as that used in a diaphragm pump, which is driven by a crankshaft. The flexible cup, which is constructed from an elastomeric material such as Teflon, provides
© 2006 by Taylor & Francis Group, LLC
Compressors and Vacuum Pumps
9-25 Discharge check valve Suction check valve Gas path Diaphragm group Head Piston ring Piston sleeve Piston Frame
FIGURE 9.21 Schematic of the suction and discharge valves of a plate-ring design. (From Figure 14, Typical Diaphragm Compressor Head Assembly in Technical Bulletin of Burton Corblin, Basics of Gas Compression, Bulletin BCTB-101, http://www.burtoncorblin.com/BCTB101.pdf. With permission.)
FIGURE 9.22 Rocking piston compressor. (From Figure 16, The rocking piston principle can be viewed as a combination of the reciprocating piston and diaphragm Ideas in Vacuum Pressure Handbook, http://www.gastmfg.com/pdf/ vacpresshdbk.pdf.)
both a seal (similar to rings on a piston) and also acts as a guide for the connecting rod. The seal is maintained as the elastomeric cup expands when the assembly is in compression. This compensates for the rocking motion. Rocking piston compressors may be used for air compression, such as for pond and fountain aeration or as vacuum pumps. They are capable of producing air pressures of 1-psig. Single-stage rocking piston
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
vacuum pumps are capable of ≥29 in. Hg and double-stage units are capable of even greater vacuum levels. Advantage of the rocking piston designs is that they are quiet and oil-less. However, a limitation of rocking piston compressors is that they cannot generate very high levels of air flow, typically <10 cfm. 9.3.3.6 Materials of Construction Sliding, bearing, and sealing parts can be abrasive and worn when lubrication is inappropriate. In addition, the parts can be corroded, particularly when used with corrosive gases or refrigerants. Material selection and surface treatment, control of surface roughness, and lubricant selection are very important. For example, nitride or tungsten carbide coated parts are the most common materials used with nonlubricated compressor designs. For small reciprocating refrigeration compressors, cylinder and crank case materials are cast aluminum or high grade cast iron. Crankshafts are constructed with forged steel or spheroidal graphite cast iron, whose surfaces are hardened by high-frequency induction hardening or tufftriding. The pistons are constructed from aluminum alloy. At the end of the connecting rod, bearing steel or white metal is inserted. For large compressors, the crankcase and crankshaft are constructed of high grade case iron and forged steel or spheroidal graphite cast iron respectively [6]. These different materials may be selected on the basis of material strength, manufacturing process, tribological aspects, and cost. 9.3.3.7 Lubricated vs. Nonlubricated Compressor Designs Both lubricated and nonlubricated (oil-less) compressors are available. Some of the advantages of lubricated compressors are: increased piston ring life, use of metallic rings, and the ability to use air-cooling or noncooling systems. Disadvantages include: oil contamination of discharge gas stream, oil deposits in pressure vessels reducing capacity to store gas, and an increase in maintenance of requirements on lubrication systems. On the other hand, the advantages of nonlubricated compressors are: low to no oil contamination of discharge gas, a reduction in lubrication requirements, and reduced filtration requirements. Disadvantages include: an increase in cooling requirements, lower maximum discharge temperatures, and a reduction in piston ring and rod packing life [32]. Nonlubricated compressors are typically used in food, textile, paper, pharmaceutical, and chemical industries. The use of nonlubricated compressors eliminates the potential for air/oil contamination that may lead to stains and increased degradation or present a safety hazard.
9.3.4 Rotary Positive Displacement Compressors Reciprocating compressors are widely used, efficient, rugged, and their operation is generally well understood. Reciprocating compressors operate at approximately 70% volumetric efficiency. However, reciprocating compressors do exhibit a number of potential disadvantages. For example, reciprocating compressors are often noisy and they are designed to be operated intermittently, not continuously (typically approximately 80% duty cycle). Approximately 20% idle time is required to allow for heat dissipation. In contrast, rotary positive displacement compressors usually run more quietly than reciprocating compressors. As a class, they are designed to be operated up to 100% of the time. They typically have fewer moving parts and therefore encounter less downtime for repair and part replacement. Rotary positive displacement compressors include strait lobe (roots and multilobe), sliding vane, rolling piston, screw, scroll, and liquid piston (or liquid ring) compressors [2,3,6,7]. 9.3.4.1 Rotary Compressor Lubricants The choice of a rotary compressor lubricant is also dependent on compressor types and construction as well as the operating conditions. These compressor types and their recommended lubrication requirements are classified in ISO 6743 — Part 3A.
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Compressors and Vacuum Pumps
9-27 Tolerances between helical screw and casing
Inlet area
Roller bearings
Seal
Balls bearing Helical screw
Positive lubrication
FIGURE 9.23 Tribological parts in screw compressor. (From Figure 1 in Carnero, M.C. et al., Control of Wear Applied of Compressors: Trends in Lubricant Analysis, Wear, 225–229, Pt. 2, 905, 1999. With permission.)
P PV k = Const.
0
V
FIGURE 9.24 Theoretical p–V curve (D.R. Garg, G.E. Totten and G.M. Webster, Compressor Lubricants, in, Fuels and Lubricants Handbook: Technology, Properties, Performance and Testing G.E. Totten, S.R. Westbrook, R.J. Shah eds, 2003, ASTM International, West Conshocken, PA, pp. 383–412.)
9.3.4.2 Rotary Screw Compressor (Lysholm Compressors) A helical lobe, or a screw compressor as it is commonly known, is constructed so that two (male and female) intermeshing rotors with helical or spiral contours, also known as helical lobes or screws (twin screw compressor), rotate on parallel axes within a close-fitting casing as illustrated in Figure 9.23 [33]. The rotors are mounted on bearings, which hold their position with design tolerances with respect to intersecting cylinder bores. The screw-shape rotors possess a specified numbers of male and female rotors, depending on the desired speed and displacement. In some compressors, there are four or five male lobed rotors driving six or seven female lobes through an oil film. The driving device, an electric motor or engine, is usually connected to the male rotor. However, the driving device may be connected to the female rotors to obtain greater rotor speeds and displacement [34]. To understand the compression process, consider the simplified schematic of a screw compressor shown in Figure 9.24 [35]. The concept of lobe numbers with respect to the helical screw construction can be visualized in Figure 9.25 [35]. Initially, the gas enters the compressor through the suction port at the top of the compressor drawing shown as the lobes begin to unmesh during rotation. Gas fills the so-called flute spaces until the trailing lobe crosses the suction port.
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Handbook of Lubrication and Tribology (a)
(b)
(c)
Discharge port
Discharge port
Secondary rotor
Secondary rotor
Discharge port Secondary rotor
Main rotor
Main rotor
Inlet Inlet
Main rotor Inlet
FIGURE 9.25 Compression cycles with rotor terminology for screw compressor. (From Figure 5 in Burke, L.J., Compressors and Vacuum Pumps, in Handbook of Lubrication: Theory and Practice of Tribology, Volume 1, Application and Maintenance, ed. Booser, E.R., CRC Press, USA, 1994, 136. With permission.)
At the bottom of the compressor, the lobes begin to mesh. As the lobe meshes in the flute space, the total gas volume is reduced and the pressure increases. Volume between the advancing rotor helix and the compressor endplate decreases with corresponding pressure increase as long as there is gas in the flute space. The gas will be discharged from the compressor as the leading edge of the lobe crosses the discharge port. The total volume of gas at this point is referred to as the discharge volume [35]. There is a finite volume of gas remaining after the leading lobe crosses the discharge port. Although further rotation forces this remaining gas into the discharge port, there is no additional pressure increase since the flute space is no longer sealed. A rotary screw compressor has no clearance volume at the end of the discharge process. Therefore, it is theoretically possible to have 100% efficiency. However, the seal between the rotors and the housing is not gas-tight leading to leakage between flutes, which reduces volumetric efficiency. Rotary screw compressors are characterized by their volume ratio (Vi ), which is defined as: Vi = Inlet volume, acf/Discharge volume, acf where acf is the actual volume of gas trapped (cubic feet). The volume ratio is related to the internal pressure ratio (Pi ) that is developed in the flute spaces in the pump by: Pi = Vik where k is the specific heat ratio of the gas being compressed. A compressor with a fixed volume ratio will develop the same internal pressure ratio, which is independent of the line pressure. (Volume ratio can be varied for some types of compressors by varying the location or size of the discharge port.) Ideally, compressor should be sized so that the volume ratio of the rotary screw compressor is equal to the system compression ratio to achieve optimal compressor efficiency [35]. The increase in compression ratio and volume ratio with discharge pressure is shown in Figure 9.26 [36]. The increase in volume ratio with decreasing temperature is shown in Figure 9.27 [36]. There are two basic types of rotary screw compressors, dry or wet-flooded types [37–39]. In the dryscrew type, the rotors are encased inside of a stator without a lubricant. The heat of compression is removed from the gas after exiting the compressor. In dry-screw compressors, timing gears position the rotors so that they intermesh closely but do not touch and therefore, no internal lubrication is required. Contact between the rotors is prevented by timing gears that mesh outside the working chamber and are lubricated externally. In addition, to prevent the lubricant entering the working chamber, internal seals are used on each shaft between the working chamber and the bearings. In the case of process gas compressors,
© 2006 by Taylor & Francis Group, LLC
Compressors and Vacuum Pumps
9-29
6.5
Volume or compression ratio
6.0 5.5
on
si res
o rati
mp
5.0
Co
4.5 4.0 e Volum
3.5
ratio
3.0 k = 1.37
2.5
Saturated suction temperature = 08F 2.0 120
140
160
180
200
Discharge pressure (psig)
FIGURE 9.26 Volume of compression ratio vs. discharge pressure. (From Reindl, D.T. and Jekel, T.B., Selection of Screw Compressors for Energy Efficient Operation, IRC Tech Note, Industrial Refrigeration Consortium, Madison, WI, USA, 2002. With permission.)
Screw compressor volume ratio
6.0 5.5 5.0
Tsuction,sat = –208F
4.5 4.0 3.5 Tsuction,sat = 08F
3.0 2.5
Tsuction,sat = 208F
2.0 1.0 120
140
160
180
200
Discharge pressure (psig)
FIGURE 9.27 Screw compressor volume ratio vs. discharge pressure. (From Reindl, D.T. and Jekel, T.B., Selection of Screw Compressors for Energy Efficient Operation, IRC Tech Note, Industrial Refrigeration Consortium, Madison, WI, USA, 2002. With permission.)
double mechanical seals are used. Dry-screw compressors deliver oil-free gas, which is useful in process and vacuum service. Operating speeds are relatively high at 8000 rpm or more. In the wet-flooded rotary screw compressor, a fluid, usually a compressor oil, is injected directly into the compression chamber to lubricate moving parts, seal the gaps, and reduce the temperature rise during compression. These flood-lubricated machines, often called “wet”-screw compressors, may be constructed without timing gear, in which case they will run at lower speeds than dry screw compressors. Although the compressors require no internal seals and are simple in mechanical design, the flood-lubrication system requires peripheral equipment, such as an oil separator, to separate oil from the discharge gas. The lubricated parts of screw compressors always include the shaft and thrust bearings and timing gear if any. High-quality engine oils and automatic transmission fluids are among the oils that have been used successfully in the lubrication of oil-flooded screw compressors. These oils have been particularly useful
© 2006 by Taylor & Francis Group, LLC
9-30
Handbook of Lubrication and Tribology
FIGURE 9.28 Stright lobe rogary compressor. (From Figure 4 in Burke, L.J., Compressors and Vacuum Pumps, in Handbook of Lubrication: Theory and Practice of Tribology, Volume 1, Application and Maintenance, Ed. Booser,E.R., CRC Press, USA, 1994, 140.)
in mobile air compressors in construction service, where their detergent/dispersant properties may well be of benefit considering the dust-laden environment. Detergent/dispersant additives can have an adverse effect, however, on the ability of the oil to separate from the water (condensed moisture) that accumulates under certain operating conditions. This, in turn, could lead to oil–water emulsions, which may clog filters and oil passages, accelerate oxidative deterioration of the oil, or promote corrosion. Consequently, a growing number of manufacturers now advise against the use of engine oils, detergent-containing oils, or stipulate that only oils with good water separating properties be used in their flooded-screw compressors. Typical recommendations include R&O turbine and hydraulic oils in ISO viscosity grades from 32 to 46. Diester-based synthetic lubricants are also being used in some of these applications where the higher initial cost can be offset by a performance benefit, reduced maintenance, or longer oil drain interval. The dry screw compressors require lubrication only for bearings and gears. The same oils as used in flooded screw machines may be employed although requirements are less severe as there is no contact between the lubricant and the compressed gas. Inhibited circulating oils are most often used, at relatively low viscosity like VG 32 to suit the high rotational speeds. The screw compressors are reviewed in detail in References 34, 35, and 40. 9.3.4.3 Straight Lobe Compressors In its simplest form, the straight-lobe compressor consists of a casing containing a matched set of symmetrical rotors (lobes) usually having a cycloidal or figure-eight profile (Figure 9.28) [3]. If the number of lobes is two (a figure-eight profile), it is called a twinlobe compressor. If the total number of lobes on the shaft is three (a cycloidal profile), it is called a trilobe compressor. If there are more than three lobes, the compressor is called a multilobe compressor. These compressors are commonly known as Roots designs or Roots compressors [41]. The term straight lobe means that the lobe runs the length of the shaft. The operation of the Roots compressor is similar to the twin screw compressor described earlier. The lobes are mounted on parallel shafts and gas flow is at right angles to the shaft axes. A pair of matched
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Compressors and Vacuum Pumps
9-31
lobes on the shafts rotates in opposite directions to trap incoming air and compress it against the casing. Since the rotors do not actually touch, they are driven by timing gears as they intermesh. Although the operation of the twin-screw and Roots compressor is similar, there are significant advantages for the twin-screw compressor including: lower power consumption, cooler discharge temperature, and quicker boost delivery. The greater temperature rise of the discharge gas for the Roots-type compressor is due to a longer path and greater turbulence during the gas compression process relative to the twin-screw compressor. Although both compressors are widely used and perform well, it is important to note that there are subtle but important differences in their operation. The Roots compressor may utilize either twin- or trilobe design. The trilobe design is often used at higher pressures because it produces more stable performance at any inlet pressure and smoother discharge performance since it produces six (6) pushes per revolution as opposed to the four (4) pushes per revolution produced by the twinlobe design. However, the total displacement and total leakage path of both designs is approximately the same. Roots compressors are also available as a “multistage” design which means that there is more than one “compression” stage. The gas is compressed sequentially in stages that may be accomplished through internal passages within a single pump, or more simply by stacking compressors (blowers) together, starting with a large blower and going down to smaller blowers. This arrangement provides better vacuum if used as a vacuum pump of higher gas pressure if it is used as a gas compressor. Typically, a single stage roots compressor exhibits a compression ratio of approximately about 10 : 1. With additional compression stages, it is possible to achieve a multiple of this compression ratio. Twinlobe and trilobe machines are sometimes called Roots blowers and are used primarily where large volumes of oil-free low-pressure air is needed, for example, in aeration, drying, and pneumatic conveying, or in engine exhaust scavenging. Very large displacement capacities, >20,000 m3 /h can be generated but only against low-pressure differentials of 10 to 10−3 mbar [46]. The Roots compressor is primarily a displacement machine (blower) and not a true compressor since it is not well suited to discharge near atmospheric pressures. It is, however, possible to obtain somewhat higher pressures if a relief value is used to assure that a maximum pressure differential is not exceeded. Roots compressors are available in power ranges of 7 to 3000 hp and may be designed to deliver pressures up to 250 psig. For greatest efficiency, these compressors should be operated at maximum rotational speed. Since there is no contact between the rotors or between rotors and casing, no internal lubrication is required. Lubrication is required, however, for shaft bearings and for the precision timing gears, which maintain the rotors in their proper position. Turbine quality R&O oils are most frequently specified, in a viscosity suitable for the gears and the operating temperature. ISO viscosity grades 100 and 150 are recommended for normal and higher temperature operation whereas 68 or lower may be required in lowtemperature applications. Some designs with splash lubricated gears have grease lubricated roller bearings for which a lithium soap grease of National Lubricating Grease Institute (NLGI) consistency number two is typically used. 9.3.4.4 Rotary Sliding Vane (Rotary Vane) Compressors Sliding vane compressor is a rotary compressor [42] that comprises a slotted rotor assembly with a series of sliding, flat vanes (typically eight or more), mounted eccentrically within a cylindrical casing, which is also known as the stator (Figure 9.29) [2]. The eccentric positioning of the rotor causes the volume between the vanes and the body to vary during each half turn. The rotor is supported by water-cooled bearings housed in the end plates of the compressor body. The discharge port is located at the position where the rotor is closest. The inlet port is located just past the discharge port. Because of its design, no suction valve is needed. As the rotor turns, the vanes slide radially outward and are held against the inner surface of the casing by centrifugal force, thus forming a series of sealed compartments of varying volume. As the rotor turns one revolution, the volume is greatest at the intake ports and least (and the pressure is greatest) at the discharge port. Gas is sucked into a centrifugal pump because of the position of the inlet port so that the volume behind the last vane to pass increases, allowing gas to expand into it until the next vane passes.
© 2006 by Taylor & Francis Group, LLC
Handbook of Lubrication and Tribology
Co
m
pr es sio n
9-32
Oil in
ke
Inta
Discharge
FIGURE 9.29 Side view of sliding vane compressor. (From Figure 7 in Burke, L.J., Compressors and Vacuum Pumps, in Handbook of Lubrication: Theory and Practice of Tribology, Volume 1, Application and Maintenance, ed. Booser, E.R., CRC Press, USA, 1994, 136. With permission.)
The cylinder walls, sliding vanes, rotor support bearings, and shaft seals all require oil lubrication. Therefore, oil is injected into the inlet and along the stator walls to cool the gas and lubricate the bearings as a “once-thru” lubrication system. The same oil also lubricates the vanes and rotor and the provides a seal for the vane contact with the stator wall to prevent leakage and helps to protect internal surfaces against rust and corrosion. The oils used for rotating vane compressor lubrication are typically higher in viscosity than oils typically used in reciprocating compressors to reduce friction and to improve the seal. After compression, the gas must be separated from the oil. However, oil-free rotary vane compressors are also available. Rotary sliding vane compressors are designed to be used in harsh environments. They have low operating cost, no pulsation, are free of vibration, and require no special foundations. Since rotary vane compressors do not operate at close tolerances, their life expectancy is greater than other compressors that depend on close tolerances. Compressor capacities typically range from 2 to 300 l/sec. They are available in power ranges of 10 to 500 hp and at pressures ranging from 500 to 1000 kPa. 9.3.4.5 Rolling Piston Rotary Compressors Another type of rotary compressor is the rolling piston type rotary compressor, which operates on an eccentric shaft that rotates within the cylinder (Figure 9.30) [43]. Gas or refrigerant enters the compression chamber consisting of the rolling piston, the cylinder, and the stationary vane that remains in contact with the roller through the suction port. The eccentrically moving roller produces a reciprocating motion of the vane. Gas is compressed as the shaft revolves by the eccentricity of the roller (piston). When the piston is in contact with the top of the cylindrical housing, the hot gas exits through the discharge port. The displacement (Vd , m3 /sec or CFM) for a rolling piston compressor can be calculated from [44]: Vd = π H (A 2 − B 2 )/4 where H is the cylinder block height (m or in.), A is the cylinder diameter (m or in.), and B is the roller diameter (m or in.) for a given rolling piston rotary compressor. The contacting conditions at the counterfaces between the rolling piston and the vane [45] as well as the piston and the cylinder are lubricated with the mixture of the gas or refrigerant and oil [46]. High reaction force from the compressed gas is applied to the shaft, which is supported by hydrodynamic journal bearings [47]. To improve the tribological performance, especially for use of refrigerant, a rotary compressor with the rotor combined rigidly with the blade (vane) in the chamber creates a so-called swing compressor [48]. The swing compressor takes the advantage to avoid the concentrated nominal line contact between the vane tip and the cylinder.
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Compressors and Vacuum Pumps
9-33
Although its performance does not equal that of other rotary compressors, it does match or exceed that of reciprocating compressors. The rolling piston rotary compressor possess a number of advantages relative to other compressor designs, including fewest number of parts, low cost, less vibration, less suction and discharge losses, and reduced torque fluctuations. 9.3.4.6 Scroll Compressors Scroll compressors are positive displacement machines consisting of two identical intermeshing scrolls, namely, a fixed scroll and an orbiting scroll (Figure 9.31) [49]. A scroll is a fundamental compressing
Crank angle, u
Rolling piston
Suction port Suction volume
Stationary vane Cylinder Compression volume Shaft Discharge valve
Cylinder head
Discharge port
Eccentric
FIGURE 9.30 Mechanical components of rolling piston compressor. (From Figure 1 in Prater, G. Jr. and Hnat, W.P., Optical Measurement of Discharge Valve Modal Parameters for A Rolling Piston Refrigeration Compressor, Measurement, 33(1), 75, 2003. With permission.) Discharge Drive-shaft Motor Bearing Frame body Cross frame Orbiting scroll A
A Fixed scroll
Section A–A
FIGURE 9.31 Structure of refrigeration scroll compressor. (From Figure 1 in Jiang, Z., Harrison, D.K., and Cheng, K., Computer-Aided Design and Manufacturing of Scroll Compressors, Journal of Materials Processing Technology, 138(1–3), 145, 2003. With permission.)
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9-34
Handbook of Lubrication and Tribology
element in a scroll compressor. Section A–A of Figure 9.31 is a schematic illustration of the scroll compressor design, which is called an involute spiral. The shape of the involute spiral allows opposing machine elements to mesh so that the opposing faces roll and not slide against each other, which reduces friction and wear and produces a constant angular velocity during meshing. During the meshing process, a series of crescent-shaped gas pockets (chambers) are formed. The orbiting scroll is driven directly by the drive shaft connected with the motor rotor. The orbiting path is conventionally described by scroll compressor manufacturers as compliance. A radially compliant compressor allows the orbiting scroll to follow a flexible path that is defined by its contact with the stationary scroll. Axial compliance refers to the ability of the orbiting and stationary scrolls to separate axially. When there is both radial and axial compliance, scroll compressors wear in not wear out. A con-compliant compressor is one where the orbiting scroll follows a fixed path where the orbiting and stationary scroll never touches [50]. As the orbiting scroll rotates around its own axis, gas is drawn into the outer pocket created by the two scrolls, sealing off the open passage. (Actually, the scroll does not “rotate”; instead the orbiting scroll basically“wobbles”on a cam inside the fixed scroll.) The trapped gas is forced toward the center of the scroll as the pocket continuously becomes smaller in volume, creating higher gas pressures as the spiral motion continues. When the compressed gas reaches the center of the fixed scroll member, it is discharged [51]. In the scroll compressor, several pockets of gas are compressed simultaneously. This produces a smooth, nearly continuous compression cycle. Scroll compressors have a fixed compression ratio. In addition, they control output by turning on and off. Variable capacity can be achieved by using several compressors, using relatively expensive electronic controls and opening ports that bleed some gas from the initial stages of compression. The volumetric efficiency of a scroll compressor is defined as: Volumetric efficiency = (suction volume flow at the pump inlet)/(compressor displacement) Scroll compressors are used in various industries but their main application is in air-conditioning and refrigeration such as: supermarket racks, bulk milk cooling, truck transport, and marine containers. In air conditioning, smaller compressors (1 to 6 tons) are used in residential systems such as heat pump systems used to heat and cool homes or businesses. Larger compressors (7 to 25 hp) are used in commercial applications including process chillers and in a variety of condensing unit systems. Scroll compressor technology has also been successfully applied in cryogenics and natural gas. 9.3.4.7 Liquid Piston (Liquid Ring) Compressors Liquid piston-type rotary compressors that are often referred to as liquid ring compressors are generally used for lower-pressure service. The compressor employs a shrouded rotor with forward-curved blades (impeller), which upon rotation, cause a centrifugal force that leads to the formation of a gas capturing ring of liquid within an elliptical (eccentric) casing onside the pump body creating a vacuum. There is no metal-to-metal contact between the rotor and the casing. Although the liquid is most often water, other liquids, usually oil, can be used as well. The liquid ring [52] forms an annulus, which maintains a seal between blades and housing as illustrated in Figure 9.32 [2]. A fixed port cylinder inside the rotor directs the gas into the suction ports. The gas is trapped as it is forced radially inward toward the central port cylinder and is compressed at the point of maximum eccentricity between the rotor blades as the liquid seal recedes from the port cylinder. Stationary inlet and discharge ports are located in cones that extend inward from the cylinder heads. The gas pockets that are formed between the blades and annulus increase and decrease in volume forming a liquid piston during rotation of the impeller. The type of load for which the liquid ring compressor will be used affects its capacity. For example, if the compressor is being used for dry air, the dry air will saturate itself by evaporating water from the liquid ring, which will take up space in the gas pocket formed in the impeller, which reduces the amount of incoming air load. A saturated air mixture will also enter the impeller gas pocket through the inlet but if the temperature of the incoming air is greater than the liquid ring seal, condensation of the moisture
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Compressors and Vacuum Pumps
9-35 Discharge port
Inlet port Discharge
Elliptical casing
Inlet Discharge port
Inlet port Rotor (Rotates clockwise)
FIGURE 9.32 Liquid piston compressor. (From Figure 8 in Burke, L.J., Compressors and Vacuum Pumps, in Handbook of Lubrication: Theory and Practice of Tribology, Volume 1, Application and Maintenance, ed. Booser, E.R., CRC Press, USA, 1994, 136. With permission.)
in the incoming air will occur, increasing the capacity of the compressor. Therefore, when handling saturated vapor–gas mixtures the liquid ring acts as a condenser, greatly increasing the effective capacity of the pump. Liquid ring compressors that are designed for compressing gases in the overpressure range may utilize the same relationships that apply for liquid ring vacuum pumps except that the influence of service fluid temperature Ti must be considered at temperatures different from the published conditions used for comparative rating of liquid ring pumps and compressors. Typical published data are for 59◦ F water as the sealing liquid for the liquid ring compressor. As shown earlier in this chapter, Boyle’s Law states that the volume and pressure of a gas will change in inverse proportion to one another according to: p1 V1 = p2 V2 . When using the Ideal Gas Laws such as Boyle’s Law, the values used for calculation must be in absolute terms ( Hg absolute or torr). It is important to understand the difference between SCFM and ACFM. SCFM is measured at standard conditions (68◦ F or 20◦ C and 29.92 Hg or 14.7 psia). ACFM is measured at actual inlet conditions. Interconversion between SCFM to ACFM is performed using Boyle’s Law. One may also use a nomogram for an expansion such as that shown in Figure 9.33 [53] to calculate pressure and volumes required in either compressors (or vacuum pumps). This will be illustrated by the following example [53]. Assume that the volume of a tank including connecting piping is 750 ft3 . The initial atmospheric pressure is 760 mm HgA (torr). The required vacuum level is 24 HgV (150 torr). The amount of gas to be removed from the vessel (Q, ft3 ) is calculated from: Q = V × Ln(p1 /p2 ) where V is the volume of reservoir plus connecting piping, p1 is the initial pressure, and p2 is the required pressure. Using Figure 9.33 or by calculation, Ln(p1 /p2 ) = Ln(760/150) = Ln 5.067 = 1.62. When using Figure 9.33, locate the required vacuum level on the chart (150 torr) and read expansion factor (F ) off the scale (1.62). The total amount of air to be removed to reduce the pressure inside the vessel from atmospheric pressure to a vacuum level of 24 HgV: 750 × 1.62 = 1215 ft3 . If an evacuation time of 3 min is required, the average pump capacity from 760 to 150 torr should be: 1215/3 = 405 ACFM, which is the capacity of the pump to be selected.
© 2006 by Taylor & Francis Group, LLC
9-36
Handbook of Lubrication and Tribology Vacuum in ”Hg 29.7 29.6 29.5 29.4 29.1 29 5
28 27
26
24 22 20 1816 12 8 4 0
Expansion factor F
4
3
F=
Ln
(P P ) 1
2
2
1
0 5
10
20
30 40 50
100
200 300
500 760
Absolute pressure in mm HgA (torr)
FIGURE 9.33 Gas expansion factor. (From the figure entitled: Expansion Factor, http://www.groundeffects.org/ Tech_Data_2.htm.)
Performance curves for liquid ring compressors and vacuum pumps are always derived relative to atmospheric pressure at sea level and it is important to account for the barometric pressure difference affecting compressor performance. As a general rule, it is often assumed that for each increase of 1000 feet of elevation, the barometric pressure will decrease by 1 Hg [53]. Alternatively, one can use a nomogram [53]. The vapor pressure of the service liquid exhibits a direct influence on the compressor capacity. When the vapor pressure of the service liquid is less than that of water at 59◦ F, the pump capacity will increase and when the vapor pressure of the service liquid is higher, the pump capacity will decrease. The diagrams in Figure 9.34 permit the selection of the correct compressor by utilizing inlet temperature correction factors provided in Figure 9.34 [54] or values provided in Reference 29. The following example illustrates the use of these nomograms. Assume that a high seal-water temperature (∼860◦ F) can exhibit a significant effect on compressor capacity. For a single-stage compressor, the capacity correction factor when operating at 75 torr (27 HgV), is 0.76. This means that if the published capacity of the compressor is 300 ACFM at 75 torr, it will possess corrected capacity of 228 ACFM with the higher seal-water temperature [53]. Figure 9.35 [54] provides average condensing factors for vacuum compressors and pumps in using saturated air. When handling air–water vapor mixtures, the pump capacity will increase depending on the saturated air temperature, sealing liquid temperature, and the sealing liquid temperature entering the compressor. These condensing factors are multiplied by the volume of dry air (cfm) to obtain the actual volumetric flow rate (cfm) value. During the gas compression cycle, the heat of compression is absorbed by the liquid ring seal. This additional heat must be removed by cooling to maintain the proper vapor pressure. This is done by the addition of sufficient cool liquid to equal the amount of sealing fluid lost through the discharge port in addition to the volume of the compressed gas mixture. The gas mixture is subsequently separated from the sealing liquid. During the pumping cycle the gas is in intimate contact with the sealing liquid and compression is nearly isothermal.
© 2006 by Taylor & Francis Group, LLC
Compressors and Vacuum Pumps
9-37
Single-stage liquiding vacuum pumps 1.3 1.2
5 08F
s e r vi
1.1
t e r te m
perature
598F
1.0
8F 63
0.9 0.8 0.7
0.3 20
25
30
wa te
r te
40
50
io n
pre s
8F s
er v
ic e
u ct
0.4
122
ab le s
1138 F
a ll ow
104 8F
um
0.5
95 8F
xim
86 8F
Ma
0.6
re tu
mp
er a
68 8
F 77 8F
Temperature factor
ce wa
s ure
70 100 150 200 250300 Suction pressure in mm HgA (torr)
400 500 700 760
Two-stage liquiding vacuum pumps 1.3 50 53
1.1
5 58F 5 78F
8F
r vi ce
1.0
5 98F 6 18F
0.9
64
wa
te r
te m p
eratur
8F
0.3 20
25
30
40
te m p
re
vic e
wa ter
0.4
50
u c ti
on p
122 8F
38F
ab le s
11
a ll ow
18F
m
u
0.5
atu er
ser
8F
xim Ma
0.6
10
0.7
95
0.8
FIGURE 9.34
e
6 7 88 7 5 28F F 7 8F 8 2 98F 8 6 8F 8F
Temperature factor
8F
se
1.2
re s s u r e
70 100 150 200 250300 Suction pressure in mm HgA (torr)
400 500 700 760
Temperature factors. (From http://www.groundeffects.org/Tech_Data_2.htm.)
© 2006 by Taylor & Francis Group, LLC
9-38
Handbook of Lubrication and Tribology Using 778F service water
Using 598F service water
2.5
2.5 1228F 1228F
1228F
1228F
2.0
Inlet air/vapor mixture temperature
1228F
Condensing factor
Condensing factor
1228F
1228F 1.5
1228F
1.0 20 25
30
45
50
60
80 100
1228F 1228F 1228F 1228F 1228F 1228F 1228F
2.0 1228F
1.5
1.0 20 25
150 200 250 300 400 500 700 760
30
45
50
Pump inlet pressure in mm HgA (torr)
60
Using 688F service water
Using 778F service water
Inlet air/vapor mixture temperature
1228F 1228F 1228F 1228F Condensing factor
Condensing factor
150 200 250 300 400 500 700 760
2.5 1228F 1228F 1228F 1228F 1228F 1228F
2.0 1228F
1228F
1.0 20 25
30
45
50
60
1228F 1.5
1228F
30
45
50
60
80 100
150 200 250 300 400 500 700 760
Pump inlet pressure in mm HgA (torr)
Using 868F service water
Using 958F service water
2.5
2.5 Inlet air/vapor mixture temperature
1228F 1228F 1228F Condensing factor
1228F 1228F 1228F 1228F
2.0
Inlet air/vapor mixture temperature
2.0
1.0 20 25
150 200 250 300 400 500 700 760
80 100
1228F
1228F
Pump inlet pressure in mm HgA (torr)
Condensing factor
80 100
Pump inlet pressure in mm HgA (torr)
2.5
1.5
Inlet air/vapor mixture temperature
1228F
1228F 1.5 1228F
Inlet air/vapor mixture temperature
2.0 1228F
1228F
1.5
1228F 1228F
1228F
1228F 1.0 20 25
30
45
50
60
80 100
150 200 250 300 400 500 700 760
1.0 20 25
30
Pump inlet pressure in mm HgA (torr)
45
50
60
80 100
150 200 250 300 400 500 700 760
Pump inlet pressure in mm HgA (torr) Using 778F service water
2.5 Inlet air/vapor mixture temperature
Condensing factor
1228F 1228F
2.0 1228F
1228F 1.5 1228F 1228F 1228F 1228F 1.0 20 25
30
45
50
60
80 100
150 200 250 300 400 500 700 760
Pump inlet pressure in mm HgA (torr)
FIGURE 9.35 Gas condensing factors. (From http://www.groundeffects.org/Tech_Data_2.htm.)
© 2006 by Taylor & Francis Group, LLC
the
figure
entitled:
Gas
condensing
factors,
Compressors and Vacuum Pumps
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Use of oil as the liquid media will potentially increase the capacity of the compressor because of its low vapor pressure (<12 mm HgA). When a liquid seal other than water is used, it is necessary to account for the specific gravity, specific heat, viscosity, molecular weight, and vapor pressure differences of the liquid being used relative to water. The dry gas capacity of the compressor is increased if the vapor pressure is less than that of water because the seal liquid will flash under the vacuum conditions of the compressor and take up space between the impeller blades. These pumps are ideally suited for wet vacuum applications or where condensable vapors are present. However, because either bronze or stainless steel pumps can undergo corrosion under these conditions, the selection of compatible materials of construction can be vitally important. The rolling element shaft bearings in one of these machines are the only parts requiring lubrication. They are located in external housings and are designed for either oil or grease lubrication. Premium quality rust and oxidation-inhibited oils such as ISO grade 32 turbine oils are generally supplied by the bath method. In the case of grease-lubricated equipment, an NLGI No. 2 grade premium ball and roller bearing grease is supplied by cup or pressure fitting [2]. Because the design and operation of these pumps offer various advantages over other types of rotary pumps including: long life and low maintenance costs, vibration free and noiseless operation, pulsation free gas flow, environmentally clean (when no oil is used), and low starting torque, they are being more important in plant production processes, refinery gas desulfurizing plants, ozone and hydrogen peroxide compression, and in flare gas recovery applications [54]. Liquid ring compressors are available in sizes ranging from 5 to 7500 l/sec with maximum pressures from single-stage units of approximately 240 kPa and two-stage units of 900 kPa [55] or approximately 0.2 and 11 bar.
9.3.5 Dynamic Compressors 9.3.5.1 Centrifugal (Turbo) and Axial Compressors Dynamic compressors are a class of compressors that use dynamics for gas compression by spinning the gas at high speed and then using that kinetic energy to compress the gas. A dynamic compressor does not compress (squeeze) a gas as do reciprocating compressors discussed earlier. There are two general types of dynamic compressors. The most common type is a centrifugal-flow (turbo) compressor where an impeller accelerates the gas as it enters the compressor and a diffuser is used to convert the kinetic energy of the rapidly moving gas into a pressure increase. The impeller of a centrifugal impeller is characterized by a large frontal area. Centrifugal compressors typically consist of three impellers (three stages), each in a housing called a scroll. These compressors are intended to operate 100% of the time. The second type of dynamic compressor is the axial-flow compressor, which is characterized by alternating rows of rotor blades that accelerate the incoming gas (typically air since these are used as aircraft engines), and a ring of stationary stator vanes, which direct flow into each successive row of rotor blades and act as a diffuser that converts kinetic energy of the rapidly moving air into a pressure increase. Each set of rotor blades and stator vanes is called a compressor stage. Axial-flow compressors are typically used for very high flow and relatively low-pressure applications. The primary disadvantage of an axial-flow compressor is that it requires many more stages to achieve the same pressure increase. A sectional schematic illustration of a dynamic compressor is provided for the centrifugal compressor shown in Figure 9.36 [56]. In this figure, part A indicates the outer compressor casing, B the diaphragm bundle, C the shaft, D the impeller, E the balance drum, F the thrust collar, G the hub, H the journal bearing, I the thrust bearing, L the labyrinth seal, and M the oil film end seal. The illustration of the labyrinth seal is illustrated in Figure 9.37 [10]. In a dynamic compressor, the gas is sucked into the inlet and spun at very high rotational speeds. After entering the inlet port at the center of the impeller, the vanes on the rotating impeller spin the gas between outward vanes providing centrifugal acceleration. Because the vanes are curved, the gas exits tangentially in a radial direction by centrifugal force. Pressure is increased as the kinetic energy of the rapidly moving
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F I H
L M
G
H
A B
C
D
E
FIGURE 9.36 Sectional view of centrifugal compressor schematic: (A) outer casing; (B) diaphragm bundle; (C) shaft; (D) impeller; (E) balance drum; (F) thrust collar; (G) hub; (H) journal bearing; (I) thrust bearing; (L) labyrinth seal; and (M) oil film end seal. (From Figure 3.1 in Bendinelli, P. et al., Compressor Performance-Dynamic, in Compressor Handbook, ed. Hanlon, P.C., 2001, McGraw-Hill, 3.3. With permission.)
Inject inert gas
Process gas Atmosphere
FIGURE 9.37 Labyrinth seal. (From Figure 16.1 in Netzel, J., Rotary Compressor Seals, in Compressor Handbook, ed. Hanlon, P.C., 2001, McGraw-Hill, 16.4. With permission.)
gas is reduced in a stationary diffuser (volute). The diffuser is generally a curved funnel-shaped channel, which may or may not have vanes on the inner surface, in the compressor casing that increases in area to the discharge port and as the area increases, the speed of gas flow decreases as the pressure of the gas increases. Multistage compression is obtained by sequentially directing the pressured gas from one stage into the center of the impeller in the subsequent stage. This process is repeated until the desired gas pressure is achieved. Centrifugal compressors typically use 2–6 stages, deliver up to 18,000 cfm at speeds of up to 60,000 rpm.
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Mechanistically, gas compression in a dynamic compressor follows Bernoulli’s Law, which is defined as the behavior of a fluid under varying conditions of flow and height where: P + ρv 2 /2 + ρgh = [constant] where P is the static pressure (in N/m2 ), ρ is the fluid density (kg/m3 ), v is the velocity of fluid flow (m/sec), and h is the height above a reference surface. The second term in this equation is known as the dynamic pressure. The effect described by this law is called the Bernoulli Effect, which means that for horizontal fluid flow, an increase in the flow velocity will result in a corresponding decrease in the static pressure. The forces exerted on the air by the compressor are centripetal, which are defined as any force that pulls an object in a circular or roughly circular path causing circular acceleration. The frictional force that the gas exerts on its surroundings are centrifugal where centrifugal forces are defined as one of the “frictional” forces that arises in a rotating system, which means that a body will travel in a straight line if no force is applied [57]. The head (H ) in feet is approximately equal to velocity (v) energy in ft/sec at the periphery of the impeller: H = v 2 /2g where g is acceleration due to gravity which is 32.2 ft/sec2 and the velocity (v) at the periphery of the impeller can be determined from: v = (N × D)/229 where N is the impeller rotational speed in rpm and D is the impeller diameter in inches. Dynamic compressors generally follow the performance characteristics of “the three fan laws” [57]: 1. Flow is proportional to impeller speed (CFM 1/CFM 2 = RPM 1/RPM 2) 2. Differential pressure across an impeller is proportional to the square of the impeller speed (SP 1/SP 2) = (RPM 1/RPM 2)2 3. Power absorbed by the impeller varies with the cube of the impeller speed (HP 1/HP 2) = (RPM 1/RPM 2)3 Centrifugal compressors are variable capacity, constant head devices. If the inlet capacity under one set of conditions is known, then the fan laws can be used to determine the compressor capacity at other rotational speeds. The performance of a centrifugal compressor is analyzed as a polytropic process (PV n = constant) described previously in this chapter. The polytropic head for a centrifugal compressor can be calculated from [58]: (n−1)/2 n p2 hp = Zave RT1 −1 n−1 p1 where hp is polytropic head (ft), Zave is the average compressibility, R is the Gas Constant (1545/molecular weight), T1 is the inlet temperature of the gas (R), P1 is the inlet pressure (psia), and P2 is the outlet pressure (psia). Theoretically, the polytropic head for a particular dynamic compressor remains constant, which is independent of the gas or its molecular weight, inlet temperature, or whether it is cooled during compression. Dynamic compressors do not require lubrication within the compression chamber and so can deliver oil-free air or gas. Therefore, they are frequently selected for process applications. For the bearings and seals [59,60] as well as for drive couplings, lubrication is required. One of the representative shaft seals, a labyrinth seal installation, is given in Figure 9.37 [10]. Small machines may use rolling-element bearings that are grease lubricated from a pressure system or from grease cups. The lubricant recommended for these bearings is a premium ball and roller bearing grease with good chemical and mechanical stability and antirust properties. An NLGI No. 2 grade consistency lubricant is typically used. Some such units have a “filled-for-life” provision and are not relubricated.
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Other machines equipped with either rolling-element or plain bearings receive oil lubrication from a pump in the housing or from a rotating disk or oil ring. On the other hand, larger high-performance compressors employ hydrodynamic bearings that are lubricated from a pressurized circulation system [2]. The lubricants typically recommended for these units are turbine quality R&O inhibited oils, usually in the range of ISO viscosity grade 32–68. Lower viscosity oils are generally required for high-speed compressors to minimize bearing power losses. A higher viscosity lubricant may be necessary for speedincreasing gears [2].
9.3.6 Lubricants for Compressors 9.3.6.1 Influence of Gases on Lubricants and Lubrication In many cases air would be selected as the medium of compressors, but in some cases a specific gas may be considered, which is classified by an inert or reducing gas, hydrocarbon gas, chemically reactive gas, or refrigerant [2,61,62]. In Table 9.1, the thermophysical constants of selected gases are provided. Petroleum oils and synthetic lubricants can be used as compressor lubricants. To increase the resistance and longevity, additives are often desirable for the petroleum oils. The particular contingencies are: phenolics and dithiophosphate as oxidation inhibitor, calcium petroleum sulfonate as detergent, organic acids and sodium petroleum sulfonate as rust inhibitor, trieresyl phosphate as wear preventive, chlorinated naphthalene and sulfurized hydrocarbon as boundary lubrication improver, polyisobutylene as viscosity index improver, polymethyacrylate as pour-point depressant and silicone oil as defoaming agent [63]. 9.3.6.2 Inert or Reducing Gases Nitrogen, hydrogen, helium, carbon monoxide, carbon dioxide, and ammonia have no chemical effect on oil. Straight mineral oils can therefore be used in the same viscosity grades that apply with air. Carbon dioxide should be kept dry to avoid formation of carbonic acid, which is slightly corrosive [2]. 9.3.6.3 Hydrocarbon Gases Methane, ethane, propane, butane, propylene, butylene, and the like, and natural gas also have no chemical effect on oil. Straight mineral oils may be used. However, hydrocarbon gases under pressure may dissolve in the oil thus reducing its viscosity. Cylinders handling these gases should receive an oil of one or two viscosity grades higher than for air at comparable pressures. If the gas contains moisture or is “wet” with condensed hydrocarbons, the oil should be compounded with 3 to 5% fatty oil [2]. 9.3.6.4 Chemically Reactive Gases Oxygen, either pure or in concentrations higher than in air, will react chemically with hydrocarbon oils with the release of heat. Because of the explosion hazard, oxygen is usually handled in nonlubricated compressors fitted with tetrafluorethylene or graphite rings. Halogen gas and hydrogen chloride compressors should not be lubricated with mineral oils. Although concentrated sulfuric acid is sometimes used as a lubricant in conventional compressors handling such reactive gases, it is best to use nonlubricated compressors. Hydrogen sulfide and nitrogen oxide compressors should be lubricated with a dry compounded oil as used for moist air. Hydrogen sulfide is corrosive in the presence of moisture and must be kept dry. Addition of rust inhibitor is beneficial. Nitrous oxide may react with engine oil additives, so such oils should be avoided. Sulfur dioxide may form sludges with mineral oils by solvent extraction of certain hydrocarbons from the oil. Highly hydrogenated or deeply solvent-refined white oils, or synthetics such as polybutenes should be used with this gas [2]. 9.3.6.5 Classification by Lubrication System Lubrication requirements [10] depend on the type of compressor: positive displacement or dynamic. Positive displacement compressors often possess lubricated counter surfaces with rolling or sliding motion as well as bearing and sealing parts. Dynamic compressors often contain hydrodynamic journal and thrust bearings or rolling element bearings for supporting the main shaft, although there are essentially no bearing/sealing parts in the compression chamber. These requirements may be further influenced by
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the operating conditions such as working pressures and temperatures and by the characteristics of the gases. There may exist lubrication problems due to temperature, for example, the temperature of rotary screw compressors range from 80 to 115◦ C, which yields deposits that block filters, produce varnish on bearings, and so on. The temperature of vane compressors typically varies from 80 to 150◦ C, which leads to deposits that block filters, varnish formation, and cause excessive vane wear. Operating temperatures of reciprocating compressors may be 270◦ C for single stage and 160 to 210◦ C for multistage types. Varnish and carbon deposits on the exhaust and inlet valves and piston ring wear increase leakage and deposits [9]. Appropriate selection of the lubricant is essential because it may minimize friction and wear, reduce internal leakage paths, protect against rust and corrosion, and leave virtually no deposits on hot discharge surfaces [8,9]. Some compressors, such as reciprocating compressors, do not circulate their lubricating oil back into a reservoir for reuse. For nonrecirculating or once-through lubrication system, gas solubility in the lubricant is an important issue because the physical properties, especially, viscosity and viscosity–pressure index decrease. In addition, phase separation during the cycle can result in serious problems such as accumulation of oil on the inside of the tubes, reducing heat transfer, oil starvation, and thus potential breakdown [12]. Each lubrication’s requirements and systems for several compressors is discussed in the following sections. The solubility of natural gas and other hydrocarbons is much higher in petroleum oils and PAO synthetics compared to other commonly used synthetic base stocks such as diesters and PAGs. That is expected because both hydrocarbon gas and petroleum-based oils are similar molecules consisting primarily of C–H bonds unlike diesters and PAGs, which are relatively polar. In fact, in a typical PAG molecule, every third atom in the polymer backbone is an oxygen atom, which makes it quite polar. Therefore, hydrocarbons are less soluble in PAGs. In wet sump reciprocating and rotary screw compressors, the compressed gas and the lubricant come into contact with each other [11]. Hydrocarbon gases are infinitely soluble in mineral oil and PAO-based compressor lubricants, whereas the solubility of hydrocarbon gases increases with increasing pressure at a constant temperature in a less compatible fluid such as an ISO 220 polypropylene glycol [13]. Conversely, increasing the temperature at a constant pressure will result in lower gas solubility. Because increasing gas solubility decreases viscosity, at some point the viscosity reduction of the compressor lubricant may be too much, and lubrication failure may result because of loss of hydrodynamic lubrication [14]. For refrigeration compressors, sealing as well as lubrication become more sensitive. The mechanisms of stationary and mobile compressors are different. For example, the major differentiation between the stationary compressors for room air-conditioning and mobile compressors used for automotive air-conditioning for mobile is the seal of driving shaft connecting to a prime mover. Some room airconditioning compressors have hermetically sealed, built-in electric motors. In contrast, automotive air-conditioning compressors are driven by the engine through the shaft, so that the shaft sealing is needed and the system is semihermetic. Other differentiation regarding compressor types and potential problems may be caused by the process gases such as ammonia, carbonic acid gas, natural gas, methane, and propane. The issues are given briefly in the section on lubricants. 9.3.6.6 Refrigerant Gases The refrigerant that was initially used was ammonia and this and other refrigerants such as sulfur dioxide, methyl chloride, and carbon dioxide were in common use. The introduction of CFC-compounds provided refrigerants that were at that time considered to be harmless to humans, environmentally safe, and incombustible [64]. However, Molina and Rowland presented a paper in 1974 [65] where they showed that CFC-compounds have a significant destructive effect on the earth’s ozone layer that provides protection from UV-radiation from the sun. CFCs are very stable and can disperse in the stratosphere where they decompose and form free chlorine. The free chlorine then acts as a catalyst that reacts with the ozone.
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Handbook of Lubrication and Tribology TABLE 9.6 Refrigerants and Alternative Refrigerants (ASHRAE 34 Designation and Safety Classification of Refrigerants, ODP: Ozone Depleting Potential, S.C.: Safety Class) Alternative Refrigerant refrigerant ODP
S.C.
Component
CFC-11
R-123
0.02
B1
HCFC-123
CFC-12
R-134a R-401A R-401B R-409A R-412A R-413A
0 0.037 0.04 0.048 0.055 0
A1 A1/A1 A1/A1 A1/A1 A1/A2 A1/A2
HFC-134a HCFC-22/HFC-152a/HCFC-124 HCFC-22/HFC-152a/HCFC-124 HCFC-22/HCFC-124/HCFC-142b HCFC-22/FC-218/HCFC-142b FC-218/HFC-134a/isobutane
Composition (%) — — 53/13/34 61/11/28 60/25/15 70/5/25 9/88/3
CFC-13
R-23
0
A1
HFC-23
R-500
R-407D R-412A
0 0.055
A1/A1 A1/A2
HFC-32/HFC-125/HFC-134a HCFC-22/FC-218/HCFC-142b
15/15/70 70/5/25
—
R-502
R-404A R-507A R-407A R-407B R-22 R-509A R-402A R-402B R-403B R-408A
0 0 0 0 0.055 0.024 0.021 0.033 0.031 0.026
A1/A1 A1 A1/A1 A1/A1 A1 A1 A1/A1 A1/A1 A1/A1 A1/A1
HFC-125/HFC-143a/HFC-134a HFC-125/HFC-143a HFC-32/HFC-125/HFC-134a HFC-32/HFC-125/HFC-134a HCFC-22 HCFC-22/FC-218 HFC-125/propane/HCFC-22 HFC-125/propane/HCFC-22 propane/HCFC-22/FC-218 HFC-125/HFC-143a/HCFC-22
44/52/4 50/50 20/40/40 10/70/20 — 44/56 60/2/38 38/2/60 5/56/39 7/46/47
R-503
R-508A R-508B R-23
0 0 0
A1 A1/A1 A1
HFC-23/FC-116 HFC-23/FC-116 HFC-23
HCFC-22
R-407C R-407E R-410A
0 0 0
A1/A1 A1/A1 A1/A1
HFC-32/HFC-125/HFC-134a HFC-32/HFC-125/HFC-134a HFC-32/HFC-125
39/61 46/54 — 23/25/52 25/15/60 50/50
Source: (ASHRAE Publications, © American Society of Healing, Refrigerating and AirConditioning Engineers, Inc., With permission.)
Since then, there has been an intensive and ongoing debate concerning the use of CFC-compounds; especially how their use may be reduced and what alternative chemical compounds can be used. A replacement has been sought for R11, trichlorofluorinemethane, R12, dichlorofluoromethane, and R502 (R115+R22) in refrigeration application. Some of the alternative refrigerant including the ozone depleting potential (ODP), component, and the composition are given in Table 9.6 [66]. Physical properties such as viscosity [67], viscosity–pressure coefficient [68], and rheological properties [69] of the refrigerant/lubricant mixture are quite different from the pure lubricating oils. The lubricant mixtures fundamentally influence the tribological characteristics, especially, elastohydrodynamic lubrication (EHL) [70] at the highly loaded and concentrated contact surfaces as well as wear [71,72]. Figure 9.38 provides the relation between central film thickness in elastohydrodynamic point contact and refrigerant pressure using HFC-134a/POE mixture [73]. The rolling speed is 1.0 m/sec. The film thickness was decreased with increasing refrigerant pressure at each temperature condition. The higher the fluid temperature, however, the smaller the decrease. Figure 9.39 shows the comparison of the experimental data and the predicted curve [73]. The estimation of the pressure–viscosity coefficient αmix is given by: αmix = (α2 − α1 )(ms2 )/[s2 (m − 1) + 1] + α1 where the subscripts 1 and 2 refer to the components, that is, the oil and the refrigerant respectively, s2 is the weight fraction of component 2, and m is the molecular weight ratio of the components defined by
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208C
408C
608C
808C
Viscosity–pressure coeff. (GPa–1)
35 (a)
HFC134a / POE (m = 1.75)
30 25
208C
20 408C 15 10 608C 808C
5 0 0
20
40
60
80
100
Weight fraction of refrigerant (%)
FIGURE 9.38 Elastohydrodynamic film thickness as a function of refrigerant pressure and temperature for HFC134a/POE at rolling speed 1.0 m/sec. (From Figure 14(a) in Akei, M. and Mizuhara, K., The Elatohydrodynamic Properties of Lubricants in Refrigerant Environments, Tribology Transactions, 40 (1), 1, 1997. With permission.)
208C
408C
608C
808C
Central film thickness (mm)
3 HFC134a/POE Rolling speed: 1.0 m/sec 1
0.1
0 0
0.2
0.4
0.6
0.8
1
1.2
Refrigerant pressure (MPa)
FIGURE 9.39 Experimental data and prediction curve of viscosity–pressure coefficient for HFC-134a/POE. (From Figure 6 in Akei, M. and Mizuhara, K., The Elatohydrodynamic Properties of Lubricants in Refrigerant Environments, Tribology Transactions, 40(1), 1, 1997. With permission.)
M1 /M2 where M is the molecular component. As the weight fraction of refrigerant increases, the viscosity– pressure coefficient decreases. This indicates that the viscosity–pressure coefficients of the refrigerant/oil solutions can be predicted from viscosity data of mixtures if the effective molecular weight ratio of the oil and refrigerant based on the concept of segmental flow is introduced.
9.4 Part III: Vacuum Pump Design and Operation Vacuum processing is a vitally important technology and is used in nearly every industrial sector such as: materials handling, metallurgical processing, electronics and lumber industry, automotive, railway, refrigeration, vacuum distillation, freeze drying, medical applications, and numerous others [74]. There are fundamentally two ways to produce a vacuum: with a venturi or a vacuum pump. A venturi “pump”
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Compressed air
FIGURE 9.40 Venturi vacuum generators. (From Figure 2, Venturi Vacuum Generators — What They Are and How to Design Them Into Your System, Technical Brochure [http://www.teknocraft.com/pdfs/manual-vacuum.pdf], Teknocraft Inc.) TABLE 9.7 Classification of the Vacuum Ranges Low vacuum Medium vacuum High vacuum Ultra high vacuum
102 –105 10−2 –102 10−6 –10−2 10−10 –10−6
Source: Tompkins, H.G. and Gessert, T.A., Pumps Used in Vacuum Technology, 2nd edn, AVS Science & Technology Society, USA, 2001.
is illustrated in Figure 9.40 [75] where compressed gas flows through a nozzle producing a venturi effect or vacuum flow. The air that was entrained in the system and the air used to create the supply air exhausts into the atmosphere. A similar vacuum producing source is the steam jet ejector where steam is used instead of air [76]. Although venturis are fast cycling, operate over a wide range of temperatures, produce low noise, nearly maintenance free, of low cost, and can be used in corrosive environments, they are often an inappropriate vacuum source when no compressed air is available and there is constant vacuum flow or large vacuum flow at high vacuum [77]. In these situations, a vacuum pump is required. The remainder of this discussion will focus on vacuum pump technology.
9.4.1 Introduction to Vacuum Pumps A vacuum is defined as any pressure that is lower than atmospheric pressure (760 torr and 1 torr = 1 mm Hg). Although vacuum is sometimes referred to as negative pressure, this is actually incorrect since pressure is always positive. A negative pressure can only exist relative to a higher pressure. The classification of the vacuum ranges is given in Table 9.7 [74]. Another classification that is encountered is: low (105 –102 Pa), medium (102 –10−1 Pa), high (10−2 –10−5 Pa), ultrahigh (10−5 –10−8 Pa), and extreme high (10−8 Pa) [78]. The representative unit indicating pressure includes torr or mmHg, mbar, psi, inches Hg and atm. Atmospheric pressure at sea level is 760 torr, which is equal to 760 mm Hg = 1013 mbar = 1013 hPa = 14.7 psi = 29.92 in. Hg = 1.0 atm. The percentage of vacuum is 0.0%. If the vacuum pressure is 100 torr, each value is given by 133.3 mbar, 1.93 psi, 3.94 in. Hg, and 0.132 atm, corresponding to 87% vacuum. Furthermore, for 10 torr, the vacuum percentage becomes 98.7%. Namely, the pressure reduction of 2 torr corresponds to 0.232% vacuum [74,78]. A vacuum pump is simply a compressor that is run with the inlet and outlet ports reversed so that the inlet is attached to a vacuum system and the outlet is open to exhaust. Another difference between
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a vacuum pump and a compressor is in the drive power required. Depending on the pressure rating, an air compressor may require from 150 to 400% greater power than a vacuum pump with the same open-capacity rating. Compressors are typically part of a pneumatic power system and they are used to push more gas, most typically air, into the system thus increasing the pressure above that of atmospheric air. In vacuum systems however, vacuum pumps are used to produce pressure differentials that are used to pull air out of the system thus decreasing the pressure below atmospheric pressure. A vacuum system has three primary components: working chamber, pumping system, and accessories required by the process. In a vacuum system, the working chamber determines the system design. The gas load of the system is determined by the working chamber and its contents. The gas load (throughput) — Q — depends on: Q = Ki Ai where K and A represent the outgasing constant and area of the internal surfaces respectively. Optimal system design will be achieved when this sum is as small as possible, which is achieved when the materials selected possess minimal outgasing coefficients and the internal surface area is minimized (such as by polishing). Three important terms in vacuum technology are: compression ratio, pumping speed, and throughput. Pumping speed may be generally specified in liters per second (l/sec) and is the basic parameter for selecting a vacuum pump for a given application. Pumping speed is dependent on rotor and the stator assembly geometry, rotational speed of the rotor blades, and the inlet flange diameter. Pumping speed decreases at higher pressures because of the normal mass-flow (throughput) limit. Throughput is the quantity of gas pumped per unit time, which is determined by multiplying the intake pressure and the pumping speed (assuming constant temperature). A typical unit of measure for throughput is torr·liters per second. The compression ratio is the ratio of its partial pressure of the gas being pumped, which is measured at the outlet of the pump to partial pressure of the gas at the inlet of the pump. The compression ratio is dependent on the number of stages, pump design, and rotational speed. The compression ratio for a given vacuum pump is also dependent on the molecular weight of the gas being pumped. The compression ratio (ratio of absolute discharge pressure/absolute inlet pressure) is also a critically important aspect of vacuum system design. Although the pressure differential across a vacuum pump is typically very low, the compression ratio may significantly exceed 500 : 1. Vacuum pumps may be modified to optimize performance in one application relative to another. Vacuum pumps fall roughly into one of three categories: positive displacement pumps, momentum exchange pumps, and capture pumps [74]. There are a variety of vacuum pumps available for application in the process vacuum range. The most common may include: oil-sealed rotary, dry mechanical, sorption, and ejector pumps for the ranges of low to medium vacuum pressures, molecular drag and diffusion pumps for the range of high vacuum, and turbo-molecular, getter, and cryo pumps for the range of ultrahigh vacuum [79]. ISO 3529-2 or JIS Z8126 [80], categorize vacuum pumps according to the scheme provided in Table 9.8 [81]. Some of these vacuum pumps are classified by the ultimate pressure shown in Figure 9.41 [74].
9.4.2 Fundamental Vacuum System Relationships The overall numerical relationship that describes all vacuum systems and processes is [82]: Q = SP where Q is the gas load, S is the pumping speed, and P is the pressure. This is also known as the “Q = SP Equation” and its very basic relationship can be used to determine the impact of changing one of the terms
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Handbook of Lubrication and Tribology TABLE 9.8 Category of Vacuum Pumps Positive displacement (vacuum) pump Piston vacuum pump Liquid ring vacuum pump Oil-sealed rotary vacuum pump Sliding vane rotary vacuum pump Rotary piston vacuum pump Rotary plunger vacuum pump Roots vacuum pump Kinetic vacuum pump Turbine vacuum pump Molecular drag pump Turbo-molecular pump Ejector vacuum pump Liquid jet vacuum pump Gas jet vacuum pump Vapor jet vacuum pump Diffusion pump Self-purifying diffusion pump Fractionating diffusion pump Diffusion-ejector pump Ion transfer pump Entrapment (capture) vacuum pump Adsorption pump Getter pump Sublimation (evaporation) pump Getter ion pump Sublimation (evaporation) ion pump Sputter ion pump Cryopump Cryosorption pump
[Pa] 1.E-09
1.E-07
1.E-05
1.E-03
1.E-01
1.E+01
1.E+03
1.E+05
Rotary vane Roots Cryosorption Turbomolec Diffusion Cryogenic Sputter-ion
FIGURE 9.41 Classification in terms of ultimate pressures (absolute pressure). (From Tompkins, H.G. and Gessert, T.A., Pumps Used in Vacuum Technology, 2nd edn, AVS Science & Technology Society, USA, 2001. With permission.)
of the equation on magnitude of the other two terms. Although in practice, all three terms may change simultaneously, this is still a useful measure of relative effects on a vacuum system. Vacuum pump capacity can be specified in different ways. However, the fundamental value that needs to be determined is ACFM (actual cubic feet per minute) inlet capacity at a specific vacuum level. ACFM is also defined as displacement corrected for volumetric efficiency. Vacuum pump capacity may also be specified as CFM (cubic feet per minute) or SCFM (standard cubic feet per minute) do not include pump efficiency at a specific vacuum level. SCFM is measured at standard conditions of 68◦ F, 29.92 in. Hg (14.7 psia). ACFM is measured at the actual inlet conditions of the vacuum pump. The ACFM value can
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be calculated from SCFM by [83]: ACFM = SCFM × volumetric efficiency where the SCFM and the volumetric efficiency values are those values reported for the specific vacuum level (and temperature) desired. The conversion calculation is the same for CFM values such as those reported for piston pumps. The data required for these calculations can be obtained from the pump manufacturer or from pump performance curves. Actual inlet volume (ACFM) can be calculated from SCFM reported at different pressure and temperatures using Boyle’s Law, which was described earlier in this chapter. The required inlet volumetric flow rate is determined from the system evaluation time (tE ), which is also called“pump down time,” using the following equation for dry gases and under isothermal conditions [84]: tE = 60 ·
V pB · ln S pE
where tE is the evacuation time (min), V is the enclosure volume (m3 ), S is the inlet volumetric flow rate of the pump (m3 /h), pB is the pressure at the start of the evacuation (mbar), and pE is the pressure at the end of the evacuation (mbar). It is often not possible to accurately calculate the leakage rate (qL ) of a vacuum system. However, the leakage rate may be calculated using the relationship for leakage mass flow rate (mL ) [84]: qL =
1 mL R 360
T M
where qL is the leakage rate (mbar l/sec), mL is the leakage mass flow rate (kg/h), T is absolute temperature (K), R is the gas constant (J/kmol K), and M is the molecular weight of the gas (kg/kmol). For a given suction pressure (pA ), the volumetric flow rate at the compressor inlet (S), which is required for pumping the leakage air qL is [84]: qL S = 3.6 pA where S is in m3 /h, qL is in mbar l/sec, and pA is in mbar. The leakage rate is determined by performing a pressure increase (p) measurement by comparing the pressure of an evacuated system with that obtained for the pump when isolated from the system. The leakage rate is then calculated from [84]: pV qL = tM where qL is the leakage rate (mbar l/sec), p is the measured pressure increase (mbar), V is the system volume (l), and tM is the time interval for the measurement (sec). In some cases, it is possible to estimate leakage rates from various tabular summaries such as those provided in Table 9.9 using seal length and in Table 9.10, which is based on total system volume [84].
9.4.3 Vacuum Pump Selection Criteria Fundamentally, there are three critical criteria that must be considered in vacuum pump selection (1) vacuum pump rating, (2) air removal rate, and (3) the power requirement. Methods of determination of these three criteria will briefly be described in this section. The first of the three critical criteria is vacuum pump rating, which relates to the amount of vacuum produced, whether continuous or intermittent. Vacuum pump manufacturers provide this value as a maximum vacuum rating expressed as either absolute pressure in mm Hg or as vacuum in inches of Hg. In both cases, these values are determined for standard atmospheric pressure (29.92 in. Hg). If the vacuum
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Handbook of Lubrication and Tribology TABLE 9.9 Approximate Leakage Rates for Joint Flange Gasket (Seal Length) Quality of joint
mL /lseal (kg/h m)a
Very good Good Normal
0.03 0.1 0.2
a l seal is length of the seal.
TABLE 9.10 Volume
Leakage Rates as a Function of System
Total system volume (V ) (m3 ) 0.1 1 3 5 10 25 50 100 200
Guide values for leakage air to be drawn off from the system mL (kg/h)
QL (mbar l/sec)
0.1–0.5 0.5–1.0 1–2 2–4 3–6 4–8 5–10 8–20 10–30
20–110 110–230 230–460 460–930 700–1400 900–1800 1100–2300 1800–4600 2300–7000
is determined at some pressure other than standard atmospheric pressure, it must be adjusted. In such cases, the adjusted vacuum rating — Va — (in. Hg) is calculated from: Va = (VO pa )/29.92 where VO is the original vacuum rating at standard conditions (in. Hg) and pa is the anticipated pressure (in. Hg) at the site where the vacuum pump will be used. The second criterion is the flow rate of a vacuum pump, which is rated under conditions where there is no pressure differential across the pump (open capacity) and no air is exhausted. Open capacity curves are published for a given vacuum pump at a rated speed for different vacuum levels ranging from 0 in Hg to maximum vacuum rating. Some manufacturers also provide additional capacity curves at different speeds for a given vacuum level. The third criterion, power requirement, is determined from pressure-flow curves, which typically also show input power and speed requirements for a particular vacuum pump. The overall efficiency, including mechanical and volumetric efficiency, is determined by dividing the free air capacity of the pump at a given vacuum level by the power required for that condition.
9.4.4 Vacuum Pumps and Applications 9.4.4.1 Positive Displacement Vacuum Pumps Positive displacement vacuum pumps are mechanical pumps. The vacuum pumps include reciprocating, rotary, and dynamic units, which are virtually identical to corresponding-type compressors [2]. In general, the lubrication requirements of the mechanical vacuum pumps are thus almost similar to those of their compressor counterparts. Reciprocating vacuum pumps operating on relatively clean, dry air or gas may be lubricated with a good quality straight mineral oil or R&O oil. Rotary machines such as the lobe, screw, or sliding-vane
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Intake
Contact point Outlet valve
Vanes
Rotor
Stator
FIGURE 9.42 Rotary-vane oil-sealed mechanical pump. (From Figure 7 in Tompkins, H.G. and Gessert, T.A., Pumps Used in Vacuum Technology, 2nd edn, AVS Science & Technology Society, USA, 2001. With permission.)
type and dynamic vacuum pumps are lubricated most often with turbine oils or other high-quality R&O circulating oils. The chemical stability and low-vapor pressure of diesters and phosphate esters make them a good choice in the above applications in cases where petroleum-based oils have proven to be unsuitable [2]. Liquid ring pumps, sometimes called “wet vacuum pumps,” consist of a round multifinned rotor spinning in an elliptical casing and have only one moving part. The construction is simple. The pumps do not yield metal-to-metal contact in the compression chamber and no internal lubrication is required. Further, there are a number of sealant options available, in other words, virtually any liquid, for instance one that is compatible with the process gas, can be used because the pump is independent of the liquid for lubrication. However, ultimate pressure is limited by the vapor pressure of the liquid. Also, condensing can lead to environmental concerns, as it may be easier to deal with waste in the vapor state than in the liquid state. Other considerations in applying liquid ring pumps include their potentially high cost on deep vacuum applications and high power consumption in situations where there is no condensable vapor in the gas stream [85]. A rotary-vane oil-sealed pump shown in Figure 9.42 is one of the mechanical pumps. This is a twobladed pump in which the rotor is eccentric to the stator, forming a crescent-shaped volume swept by the blades through the outlet valve. The two vanes are forced outward by springs and centrifugal forces. An oil film exists between the tip of the vanes and the stator, which provides lubrication, sealing, and cooling. The schematic of the operation is illustrated in Figure 9.43. As the rotor turns in a clockwise direction the pump inhale (A), compress (B,C), and discharge (D) the gas [74]. The rotary piston pump is another type of the same principle as the rotary-vane oil-sealed pumps and the schematic is given in Figure 9.44 [74]. The rotor consists of an inner drum, rotating eccentrically about the axis. The piston (or plunger) moves up and down alternately, opening and closing the inlet. This type of pump will tolerate more abuse and gas contamination. The technology of the mechanical vacuum pumps would be mature since the design methods of most equipment are almost established and proven. However, an exception to this is dry vacuum pumps. Dry vacuum pumps include rotary vane, roots, claw, screw, and scroll vacuum pumps. Dry vacuum pumps take the advances of the ability to operate with no liquid at all in the compression chamber. Most dry pumps are designed to operate hot to prevent condensing of vapors during the compression cycle. Accordingly, they can handle most corrosive vapors even though materials of construction are typically limited to iron, or in some cases, iron with stainless steel rotating elements. However, as these
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Handbook of Lubrication and Tribology A
B
C
D
FIGURE 9.43 The principle of operation of an oil-sealed rotary-vane pump. (From Figure 8 in Tompkins, H.G. and Gessert, T.A., Pumps Used in Vacuum Technology, 2nd edn, AVS Science & Technology Society, USA, 2001. With permission.)
Inlet Outlet
Plunger or piston Shell
Inner drum (rotor) Contact point
Stator
FIGURE 9.44 Rotary piston pump. (From Figure 11 in Tompkins, H.G. and Gessert, T.A., Pumps Used in Vacuum Technology, 2nd edn, AVS Science & Technology Society, USA, 2001. With permission.)
pumps typically run hot to prevent condensing, the user should be cautioned against the possibility of auto-ignition when flammable vapors are present [85]. Rotary lobe vacuum pumps are often called “roots” vacuum pumps. With timing gears, the pumps are operated without contact between the rotating lobes. Water, oil, or other liquids are not added into the chamber, so that it is categorized into dry pumps. In dry rotary sliding vane pumps, the vanes are made of carbon impregnated composite material that causes such a small amount of friction that oil is not needed [86]. The design of dry vacuum pumps looks like any type of typical compressors, but the design such as materials, rotor geometry, and sealing has been modified for vacuum service.
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9.4.4.2 Kinetic Vacuum Pumps Kinetic vacuum pumps include turbine vacuum pumps, molecular drag pumps, ejector vacuum pumps, diffusion pumps, diffusion-ejector pumps, and ion transfer pumps. In the turbo molecular pump, a high rotational speed rotor imparts momentum to the gas molecules, moving them along the small clearance between the rotor and stator. The pumps equip magnetic bearings or rolling element bearings to support the main shaft. Lubrication is required for the bearings of the rolling element bearing type pumps. In the ejector pumps, high-pressure steam or other gas is discharged through a nozzle, which directs a high velocity jet across a suction chamber into a venturi-shaped diffuser case. Molecules of the gas being pumped are picked up and entrained in the jet stream and discharged with it. As the device has no moving parts, no lubrication is required [2]. Diffusion pumps utilize a high velocity vapor stream to entrain and remove molecules of the gas being extracted from the system. Basic elements of diffusion pumps are a water-cooled jacket, a boiler, and a nozzle assembly. The pumps have essentially no moving parts, and thus there are no lubrication requirements. The oils used in this service must have very low-vapor pressure, good stability, and be resistant to the gases being pumped. Narrow-cut, white mineral oils, silicone fluids, and synthetic esters are used as working fluids [2]. 9.4.4.3 Entrapment Vacuum Pumps Entrapment vacuum pumps or capture vacuum pumps include adsorption pumps, getter pumps, sublimation (or evaporation) pumps, getter ion pumps, and cryo pumps. Sputter ion pump is one of the getter ion pumps. The pumps use a cathode material such as titanium that is vaporized or sputtered by bombardment with high velocity ions. The active gases are pumped by chemical combination with the sputtered titanium, the inert gases by ionization and burial in the cathode, and the light gases by diffusion into the cathode. The pumps have no moving parts [87].
9.4.5 Seals for Vacuum Pumps Elastomeric seals like O-rings are often used for vacuum pumps and vacuum chambers. The requirements for the seals are: lower level of transmission gas, lower level of emission gas, lower level of compressive permanent strain, appropriate elasticity and mechanical strength, heat resisting and antiwear. The seals are made of fluoro-, butadiene acrylonitrile, silicone, and chloroprene rubber. The fluoro-rubber is suited for the vacuum pressure ranging from high to ultrahigh and the relatively high temperature below 200◦ C; the butadiene acrylonitrile rubber is for the pressure from low to high and temperature below 80◦ C [88].
9.4.6 Vacuum Measurement There are a wide range of measurement devices that may be used to quantify pressure in a vacuum system but all are related to a number of fundamental principles relating to pressure measurement. Before providing an overview of pressure measurement it is important to understand the basic principles related to pressure exerted by a gas, particularly: particle number density, mean free path, particle flux density, and monolayer time. These fundamental principles will be briefly defined here and their interrelationship with respect to pressure is illustrated in Table 9.11 [89]. Pressure — The pressure that a gas exerts is proportional to the number of collisions of the gas molecules on the walls of the container and is defined as force/unit area and is typically calculated from the Ideal Gas Law: P = nRT /V Particle number density — According to the kinetic gas theory, the particle number density can be calculated from: n = P/kT where k is the Boltzman constant k = 1.38 × 10−23 J/K and T is the absolute temperature.
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Handbook of Lubrication and Tribology TABLE 9.11 Molecular Explanation of Pressure/Vacuum (Assuming Room Temperature) Pressure (torr)
Number density
Mean free path (cm)
Particle flux (cm−2 sec−1 )
Time for one monolayer (sec)
0.76 7.6 × 10−2 7.6 × 10−3 7.6 × 10−4 7.6 × 10−5 7.6 × 10−6
2.5 × 1016 2.5 × 1015 2.5 × 1014 2.5 × 1013 2.5 × 1012 2.5 × 1011
0.0065 0.065 0.65 6.5 65 650
2.9 × 1020 2.9 × 1019 2.9 × 1018 2.9 × 1017 2.9 × 1016 2.9 × 1015
3 × 10−6 3 × 10−5 3 × 10−4 3 × 10−3 3 × 10−2 3 × 10−1
Pressure (atm) 1/1,000 1/10,000 1/100,000 1/1,000,000 1/10,000,000 1/100,000,000
Mean free path — The mean free path (l) is related to the average distance that a gas molecule travels between collisions and can be calculated from the kinetic theory of gases: l = kT /Pσ where σ is the collision cross section with a typical value of approximately 1.3 × 10−19 m2 . Particle flux density — The particle flux density ( ) is the number of molecules passing through unit area of a plane in unit time; or the number of particles striking a surface per unit area molecules per square meter (or cm2 ) seconds, which can be estimated from: =n
kT 2π M
1/2
where M is the molecular mass of the gas molecule and n is gas density as number of molecules/unit volume. Monolayer time — The “monolayer time” is the time it takes to contaminate a surface with a single adsorbed molecular layer and it can be estimated from: t = 3.2 × 10−6 /P where t is in seconds, and P is in millibar. A molecular “sticking coefficient” = 1.0 is assumed. Another fundamental concept that is important to consider with respect to the pressure exerted by a vacuum system is conductance. The speed at which gas molecules flow, which is also defined as “throughput” — Q — will depend on both the pressure difference between the two locations, as well as on the geometry of the chamber in between. Q = d(PV )/dt The units of gas throughput are Pascal meters-cubed per second (Pa-m3 /sec). Gas throughput will be restricted and smaller if the hole through which the molecules must flow to reach the low-pressure region from the high-pressure region is very small. Throughput will become larger as the opening between the low-pressure and high-pressure regions is widened. The factor that accounts for this geometry difference is called “conductance” — C, which is always positive, and is defined as: C = Q/(P) = Q/(P2 − P1 ) Therefore, consideration must be given to gas flow and pressure differentials when designing vacuum pumping systems. However, detailed discussion of this topic is beyond the scope of this chapter and the reader is referred to an advanced discussion of vacuum system design.
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U-tube manometers were among the first pressure indicators. Typically, these tubes are made of glass, and scales for measurement, such as millimeter of Hg or inches of water are added as needed. These manometers are large, awkward, and poorly suited for use in automatic control loops. Depending on the reference pressure used, they could indicate absolute, gauge, and differential pressure [90]. This form of vacuum gauge was invented in 1874 by H.G. McLeod to measure pressures of the order of 10−2 to 10−7 torr and is known as a McLeod Gauge. The readings of the McLeod gauge are absolute, and may be used to calibrate other gauges, such as ionization gauges. Bourdon Tube, Diaphragm, and Bellows Element Pressure Gauges are examples of mechanical motionbalance-based pressure sensors. These gauges use flexible elements as sensors. As pressure changes, a flexible element moves causing motion of a pointer on a dial. In these mechanical pressure sensors, a Bourdon tube, a diaphragm, or a bellows element detects the process pressure, which causes a corresponding movement of the element. A Bourdon tube is C-shaped with an oval cross section with one end of the tube connected to the system to detect the pressure. The other end is sealed and connected to the pointer or transmitter mechanism. This group of flexible pressure sensor elements includes bellows and diaphragm gauges. Diaphragm designs typically require less space and the motion they produce is sufficient for operating electronic transducers. These mechanical gauges evolved so that instead of a gauge pointer, the movement of the sensing element was used to convert a process pressure into a transmitted electrical or pneumatic signal [90]. A piezoresistive diaphragm pressure sensor is based on a piezoelectric effect that occurs when a piezoresistive material, such as silicon, is mechanically stressed and there is a corresponding change in the resistivity of the material. This produces a change in electrical resistance, which can be detected when a constant current is applied and the change in voltage is measured. A piezoresistant diaphragm pressure sensor consists of a micro-machined silicon diaphragm with piezoresistive strain gauges diffused into it, fused to a silicon or glass backplate. The resistors have a value of approximately 3.5 k . Pressure-induced strain increases the value of the radial resistors. The resistors are connected as a Wheatstone Bridge, the output of which is directly proportional to the pressure. The resistance change can be as high as 30%. Capacitance pressure transducers, such as the capacitance diaphragm type, are based on the movement of a diaphragm element, which is usually a metal or metal-coated quartz. Stainless steel is the most common diaphragm material used, but for corrosive environments, high-nickel steel alloys are used. Tantalum may be used for highly corrosive, high-temperature applications. Silver diaphragms can be used to measure the pressure of chlorine, fluorine, and other halogen gases. To determine system pressure, the diaphragm is exposed to the process pressure on one side and to the reference pressure on the other side. Depending on the type of pressure, the capacitive transducer can be either an absolute, gauge, or differential pressure transducer. In a capacitance-type pressure sensor, a high-frequency, high-voltage oscillator is used to charge the sensing electrode elements. In a two-plate capacitor sensor design, the movement of the diaphragm between the plates is detected as an indication of the changes in process pressure. Thermocouple (T/C) gauges are suitable for pressure ranges between 10 and 10−3 torr. The T/C gauge measures the voltage of a thermocouple, which is spot-welded to a filament that is exposed to gas in a vacuum system. A constant current is fed to the filament, which reaches a temperature that is dependent on thermal losses (cooling) due to gas molecule collisions. At higher pressure, more molecules hit the filament causing greater cooling resulting in a change in the measured thermocouple voltage. These gauges are used extensively to provide the signal to automatically switch the main chamber from backing (roughing) and high-vacuum pumps at the crossover pressure. The Pirani pressure gauge sensor is based on the measurement of the electrical resistance of a heated wire, which is placed in a vacuum system. The resistance of the wire is proportional to its temperature. At atmospheric pressure, gas molecules collide with the wire that has the effect of cooling the wire. As the vacuum is applied and gas molecules are removed there are fewer molecules and therefore less collisions causing an increase in the temperature of the wire and a corresponding increase in its electrical resistance. A simple circuit utilizing a heated wire and detection in the change in resistance can be calibrated to provide a correlation between wire resistance and pressure. There are two types of Pirani gauges; constant
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current, and constant resistance. This Pirani effect only works in the pressure region from atmospheric to approximately 10−3 mbar. Therefore, a different gauge must be used to measure pressures lower than this. Penning ionization vacuum gauges (cold-cathode) utilize a cold discharge process in a measuring tube. These systems are analogous to the electrode system of a sputter ion pump. Cold-cathode ionization vacuum gauges utilize two unheated electrodes, a cathode and an anode, between which a so-called cold discharge is initiated and maintained by means of a d.c. voltage (of around 2 kV) so that the discharge continues at very low pressures, which is achieved by using a magnetic field so that the paths of the electrons are sufficiently long so that the rate of their collision with gas molecules will form the number of charge carriers required to maintain the discharge. The magnetic field is arranged so that the magnetic field lines of force cross the electric field lines, which causes the electrons to be confined to a spiral path. The positive and negative charge carriers that are produced by collisions migrate to the corresponding electrodes and form the pressure-dependent discharge current, which is indicated on the meter. The pressure reading in mbar is gas-dependent. The Penning discharge changes to a glow discharge with intense light output in which the current (at constant voltage) depends only to a small extent on the pressure at several mbar and is therefore not suitable for measurement purposes. Therefore, the measurement range is rather limited. Schulz-Phelps and Bayard-Alpert ion gauges are two related “tabulated” hot-filament ion gauges that differ only in the physical size and spacing of their electrodes. Both types have heated filaments biased to give thermionic electrons of 70 eV, which is sufficiently energetic to ionize any residual gas molecules during collisions. The positive ions formed drift to an ion collector held at about 150 V. The current measures gas number density, which is a direct measure of pressure. With a suitable controller, the commonly available Bayard-Alpert ion gauges will measure pressures between 10−4 and 10−9 torr. The electrode spacing of an Schulz-Phelps gauge can increase the upper pressure measurement limit to 1 torr. Vacuum systems are typically fitted with a total pressure measurement device such as those discussed earlier. However, these gauges indicate only how much gas remains in the vacuum system and do not provide any information on which gases are present. This is usually not a problem unless the required level of vacuum reduction cannot be achieved. In such cases, a Residual Gas Analyzer (RGA), which is based on a quadrupole mass spectrometer can quickly identify the source of the problem, which may be due to an air leak, hydrocarbon contamination, excess water vapor, or maybe the presence of some other volatile component such as a cleaning solvent. The principle of a quadrapole RGA is to ionize the gas. The ionized gas is then directed to a mass separator for ion identification. The RGA will permit unambiguous identification of the contaminant and assist in the elimination of the problem thus preventing unnecessary delays. Table 9.12 provides a summary of these different vacuum gauges and the pressures for which they may be used [91].
TABLE 9.12 Pressure Ranges for Various Types of Pressure Gauges Pressure gauge Bourdon Thermocouple (T/C) Pirani Capacitance McLeod Penning Schulz-Phelps ion gauge Bayard-Alpert ion gauge Inverse magnetron Residual gas analyzer (RGA)
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Pressure range (torr) 100 –>102 10−3 –100 10−4 –100 10−5 –>102 10−4 –100 10−5 –10−2 10−5 –100 10−9 –10−4 10−12 –10−2 10−13 –10−4
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9.4.7 Lubricants for Vacuum Pumps Vacuum pump fluids fall into two categories for use: representatively for the oil-sealed vacuum pumps and the diffusion pumps. The fluids are selected in consideration of the vacuum pump types, the process gases, the vapor contamination, the ultimate pressure, and the like [92]. Vacuum pump fluids for mechanical pumps need to be good lubricity. Mineral oils, esters, and fluorocarbons are typical mechanical and turbomolecular pump fluids [93]. Mineral oils are often acceptable for most applications. Under the specific conditions, synthetic type hydrocarbon oils (Alkyldiphenylethers, especially, dialkyl-diphenylethers) are suited and fluoro oils (perfluoro-polyether) are selected.
References [1] Garg, D.R., Totten, G.E., and Webster, G.M., Compressor lubricants, in Fuels and Lubricants Handbook: Technology, Properties, Performance and Testing, ed. Totten, G.E., Westbrook, S.R., and Shah, R.J., 2003, ASTM International, West Conshocken, PA, pp. 383–412. [2] Burke, L.J., Compressors and vacuum pumps, in Handbook of Lubrication: Theory and Practice of Tribology, Volume 1, Application and Maintenance, ed. Booser, E.R., CRC Press, Boca Raton, FL, 1994, 136. [3] Brown, R.N., Compressors: Selection and Sizing, 2nd edn, Gulf Professional Publishing, Woburn, MH, USA, 1997. [4] Yamaguchi, A. and Tanaka, H., Oil-Hydraulics and Pneumatics (in Japanese), Corona Publishing Co., Ltd., Tokyo, Japan, 1986. [5] Hanlon, P.C. (ed.), Compressor Handbook, McGraw-Hill, New York, 2001. [6] Japanese Association of Refrigeration (ed.), JAR Handbook — Equipments, 5th edn, JAR, Tokyo, Japan, 1993, chap. 1. [7] Ohashi, H. et al. (eds), Handbook of Fluid Machinery, Asakura-Shoten, Tokyo, Japan, 1998. [8] Scales, W., Air compressor lubrication, in Tribology Data Handbook, ed. Booser, E.R., CRC Press, New York, 1997, chap. 19. [9] Garg, D., Totten, G.E., and Webster, G.M., Compressor lubricants, in Fuels and Lubricants Handbook: Technology, Properties, Performance, and Testing, ed. Totten, G.E., ASTM International, USA, 2003, chap. 14. [10] Netzel, J., Rotary compressor seals, in Compressor Handbook, ed. Hanlon, P.C., 2001, McGraw-Hill, New York, 2001, chap. 16. [11] Totten, G.E. and Bishop, R.J. Jr., Lubricant requirements, properties and maintenance for natural gas compressors, Machinery Lubrication Magazine, September, 2002. [12] Marsh, K.N. and Kendil, M.E., Review of thermodynamic properties of refrigerants + lubricant oils, Fluid Phase Equilibria, Vol. 199, Nox. 1–2, 30 June, 2002, pp. 319–313. [13] Tolfa, J., Synthetic lubricants suitable for use in process and hydrocarbon gas compressors, Lubrication Engineering, 47, 289, 1990, from Totten, G.E. and Bishop, R.J. Jr., Lubricant requirements, properties and maintenance for natural gas compressors, Machinery Lubrication Magazine, September, 2002. [14] Patzau, S. and Szchawnicka, E., Oils for air and technical gas compressors, Trybologia, 20, 18, 1989, from Totten, G.E. and Bishop, R.J. Jr., Lubricant requirements, properties and maintenance for natural gas compressors, Machinery Lubrication Magazine, September, 2002. [15] Vacuum Pressure Handbook, Revised Electronic Edition, Gast Manufacturing, Inc., http:// www.gastmfg.com/pdf/vacpresshdbk.pdf, Citing: 2004, http://www.gastmfg.com/ Copyright: 1997–2003. [16] Process & Industrial Training Technologies, Mechanical Seals & Packing Video Based Educational Packages, http://www.iglou.com/pitt/volume3.htm, Citing: 2004, Pumps & Filtration On-Line, Copyright: 1996, Process & Industrial Training Technologies, Inc.
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[17] Technical Bulletin of Burton Corblin, Basics of Gas Compression, Bulletin BCTB-101, http://www.burtoncorblin.com/BCTB101.pdf, 1990, Citing: 2004, http://www.burtoncorblin.com/, Copyright: 2003, 2004, Burton Corblin S.A. [18] http://www.domnickh.co.uk/tech_Centre.asp?chapter=1§ion=3_Overview_1.htm& getIndex= false, Citing: 2004. [19] JSW Japan Steel Works Product Reciprocating Compressors, What’s Labyrinth Piston, http://www.jsw.co.jp/en/product/machinery/compressor/compressor_01_en.html, Citing: 2004, http://www.jsw.co.jp/en/index.html, Copyright: 1995–2003, The Japan Steel Works, Ltd. [20] Totten, G.E. and Bishop, R.J. Jr., Natural gas compressors and their lubrication, Machinery Lubrication, May–June, 58–63, 2002. [21] Majors, G., Compressor lubrication in Compressor Handbook, ed. Hanlon, P.C., McGraw-Hill, New York, 2001, chap. 18. [22] Rollins, J.R. (ed.), Compressed Air and Gas Handbook, 5th edn, Prentice Hall, New York, 1988.x [23] Scott, R., Noria Corporation, Reciprocating Natural Gas Compressors, Machinery Lubrication Magazine, November 2003. http://www. machinerylubrication.com/article_detail.asp?articleid= 552&relatedbookgroup=Lubrication [24] Nakai, H., Ino, N., and Hashimoto, H., Piston-ring lubrication problems for refrigeration compressor, Proceedings of International Tribology Conference, Yokohama, vol. 1471, 1995. [25] Vetter, G. and Feistel, N., Behavior of dry-running piston rod sealing systems in crosshead compressors, Hydrocarbon Processing, 01-SEP-04, 2004. [26] Feistel, N., Different concepts of heat dissipation from piston-rod sealing systems, January 2005, CompressorTechTwo, http://www.compressortech2.com/ [27] Different concepts of heat dissipation from piston-rod sealing systems, Compressor TechTwo, January, 2005. [28] Yoon, H., Sheiretov, T., and Cusano, C., Scuffing behavior of 390 aluminum against steel under starved lubrication conditions, Wear, 237, 163, 2000. [29] Livingston, E.H., Build your working knowledge of process compressors, Chemical Engineering Progress, February, 1993, 27–36. [30] Technical Essays, 1. Essay, Diaphragm Compressors by Manfred Dehnen, http://www.andreashofer.de/english/htm/info_essay_diaphragm_compressor.htm, Citing: 2004. [31] DeRose, D., A brief primer on air compressors, Fluid Power Journal, July/August, 2004, 30–34. [32] Epp, M., CNG compressors, in Compressor Handbook, ed. Hanlon, P.C., 2001, McGraw-Hill, New York, 2001, 8.7. [33] Carnero, M.C. et al., Control of wear applied to compressors: trends in lubricant analysis, Wear, 905, 225–229, 1999. [34] Pillis, J.W., Basics of operation, application and troubleshooting of screw compressors, York Refrigeration North America, 100 CV Avenue, Waynesboro, PA 17268, http:// www.petroassist.com/Tech%20Know/ScrewCompressors.pdf. [35] LaPlante, J.R., Rotary screw compression process, Presented at Gas Machinery Research Council Conference, October 7, 2003, Salt Lake City, Utah. Basics of Operation, http:// www.gmrc.org/gmrc/pdf/gmc03/33-Oil-FloodedRotaryScrewComp.pdf. [36] Reindl, D.T. and Jekel, T.B., Screw Compressors for Energy Efficient Operation — Part I, from The Cold Front, 2002, Vol. 2, IRC Tech Note, Industrial Refrigeration Consortium, Madison, WI, USA. Web site: http://www.irc.wisc.edu/file.php?id=69. [37] Fujiwara, M. and Osada, Y., Performance analysis of oil-injected screw compressors and its applications, in Proceedings of 1990 International Compressor Engineering Conference Purdue — Volume I, ed. Soedel, W., Purdue University, Indiana, 1990, 51. [38] Summers-Smith, J.D. and Taylor, G., Screw compressor bearing failures on process plant: a case study and some general lessons, Journal of Engineering Tribology, Proceedings of IMechE, 209, 77, 1995.
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[39] Stosic, N., Smith, I.K., and Kovacevic, A., Optimisation of screw compressors, Applied Thermal Engineering, 23, 1177, 2003. [40] Fleming, J.S., Tang, Y., and Cook, G., The twin helical screw compressor, part 1: Development, applications and competitive position, Journal of Mechanical Engineering Science, Proceedings of IMechE, 212, 335, 1998. [41] The Technology of Dry Vacuum Pumps, Publication No. 12-A526-20-895-R 5M0797, Edwards High-Vacuum International, One Edwards Park, 30 1 Ballardvale Street, Wilmington, MA 01887. [42] Fukuta, M. et al., Mathematical model of vane compressors for computer simulation of automotive air conditioning cycle, JSME International Journal, 38B, 199, 1995. [43] Prater, G. Jr. and Hnat, W.P., Optical measurement of discharge valve modal parameters for a rolling piston refrigeration compressor, Measurement, 33, 75, 2003. [44] Hui Jin, Parameter estimation based on models of water source heat pumps, Ph.D. Thesis, Oklahoma State University, December 2002. [45] Lee, Y.-Z. and Oh, S.-D., Friction and wear of the rotary compressor vane–roller surfaces for several sliding conditions, Wear, 255, 1168, 2003. [46] Yoshimura, T. et al., Analysis of lubricating characteristics of rotary compressors for domestic refrigerators, Journal of Tribology, Transactions of ASME, 121, 510, 1999. [47] Ahn, H.-J., Han, D.-C., and Hwang, I.-S., A built-in bearing sensor to measure the shaft motion of a small rotary compressor for air conditioning, Tribology International, 36, 561, 2003. [48] Tanaka, S. et al., Lubrication characteristics between bush and blade of swing compressor, in Proceedings of 2002 International Compressor Engineering Conference Purdue, ed. Soedel, W., Purdue University, Indiana, 2002, 263. [49] Jiang, Z., Harrison, D.K., and Cheng, K., Computer-aided design and manufacturing of scroll compressors, Journal of Materials Processing Technology, 138, 145, 2003. [50] Scroll Compressors — High Efficiency Compression for Commercial and Industrial Applications, Brochure 811-20065, Carrier Corporation, P.O. 4808, Carrier Parkway Syracuse, NY 13221-4808, October 2004. [51] Tsubono, I. et al., New back-pressure control system improving the annual performance of scroll compressors, ASHRAE Transactions, 41, 1433, 1998. [52] Hy-Vac Liquid Ring Vacuum Pumps, online Technical Brochure, HyVac Products, Inc., P.O. Box 389, Phoenixville, PA, 19460-0389 http://www.hyvac.com/products/O_pumps/LiqRing/ HydroVac%5820Liq%5920Ring.htm. [53] Conversion Formulas, online Brochure available from Ground Effects Environmental Services, Inc., E-Mail: Contact: [email protected], http://www.groundeffects.org/ Tech_Data_2.htm#Formulas. [54] Introduction to Vacuum, online Brochure available from Ground Effects Environmental Services, Inc., E-Mail: Contact: [email protected], http://www.groundeffects.org/Tech_Data_2.htm. [55] National Appliance and Equipment Energy Efficiency Program: Analysis of Potential for Minimum Energy Performance Standards for Packaged Air Compressors, Report prepared for Australian Greenhouse Office, March 1, 2001 by by Mark Ellis & Associates, 44 Albert Street, Wagstaffe, NSW 2257, Australia. [56] Bendinelli, P. et al., Compressor performance — dynamic, in Compressor Handbook, ed. Hanlon, P.C., McGraw-Hill, New York, 2001, 3.3. [57] History of Turbo Compressors, online Brochure, Cashflo Limited, Britons Lane, Beeston Regis, Norfolk, NR26 8TR, England, UK, Website: http://www.cashflo.co.uk/Turbo.html. [58] 2.0. Centrifugal Compressor Theory, Web site sponsored by: 2003 Asteroid Mining Technology, http://www.inprosys.bizland.com/website/theory.htm. [59] San Andres, L. and De Santiago, O., Imbalance response of a rotor supported on flexure pivot tilting pad journal bearings in series with integral squeeze film dampers, Journal of Engineering for Gas Turbines Power, Transactions of ASME, 125, 1026, 2003.
© 2006 by Taylor & Francis Group, LLC
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[60] Japan Society of Lubrication Engineers (ed.), Lubrication Handbook (in Japanese), 1987, Yokendo. [61] Lilje, K.C., Rajewski, T.E., and Burton, E.E., Refrigeration and air conditioning lubricants, in Tribology Data Handbook, ed. Booser, E.R., CRC Press, New York, 1997, chap. 34. [62] Michels, H.H. and Sienel, T.H., Refrigeration lubricants — properties and applications, in Fuels and Lubricants Handbook: Technology, Properties, Performance, and Testing, ed. Totten, G.E., ASTM International, USA, 2003, chap. 15. [63] Heshmat, H. and Ming Chen, H., Principles of bearing design, in Compressor Handbook, ed. Hanlon, P.C., McGraw-Hill, Two Penn Plaza, New York, USA, 2001, 19.126. [64] Jonsson, U., Lubricant rheology in refrigeration systems and magnetic storage devices, http:// epubl.luth.se/avslutade/0280-8242/95-04/licintro.html, Citing: 2004. [65] Molina, M.J. and Rowland, F.S., Stratospheric sink for chlorofluoromethanes: chlorine atomcatalysed destruction of ozone, Nature, 249, 810, 1974. [66] Obata, Y., Trend of usage and development of alternative refrigerant (in Japanese), 43, 176, 1998. [67] Geller, V.Z. et al., Viscosities of HFC-32 and HFC-32/lubricant mixtures, International Journal of Thermophysics, 17, 75, 1996. [68] Mizuhara, K. et al., Evaluation of physical properties of lubricants with alternative refrigerants, Proceedings of International Tribology Conference, Yokohama, 1995, 881. [69] Muraki, M., Sano, T., and Dong, D., Rheological properties of polyolester under an EHD contact in some refrigerant environments, Transactions of ASME, Journal of Tribology, 123, 54, 2001. [70] Jonsson, U.J., Lubrication of rolling element bearings with HFC–polyolester mixtures, Wear, 232, 185, 1999. [71] Ciantar, C. et al., The influence of lubricant viscosity on the wear of hermetic compressor components in HFC-134a environments, Wear, 236, 1, 1999. [72] Sung, H.C., Tribological characteristics of various surface coatings for rotary compressor vane, Wear, 221, 77, 1998. [73] Akei, M. and Mizuhara, K., The elastohydrodynamic properties of lubricants in refrigerant environments, Tribology Transactions, 40, 1, 1997. [74] Tompkins, H.G. and Gessert, T.A., Pumps Used in Vacuum Technology, 2nd edn, AVS Science & Technology Society, USA, 2001. [75] Venturi Vacuum Generators — What They Are and How to Design Them Into Your System, Technical Brochure, Teknocraft Inc., 425 West Drive, Melbourne, FL, 32904, Web site: http://www.teknocraft.com. [76] Martin, G.R., Lines, J.R., and Golden, S.W., Understand vacuum system fundamentals, Hydrocarbon Processing, October, 1994, pp. 1–7. [77] The Jet Eductor System: What is It, and How Does it Work? Pumps of Oklahoma, Inc., Web site: http://www.pumpsofoklahoma.com/pdf/eductor4.pdf. [78] Vacuum Society of Japan (ed.), Vacuum Technology (in Japanese), 2nd edn, Nikkan Kogyo Shinbun, Ltd., Tokyo, Japan, 1998. [79] Sawada, M., Classification of vacuum pumps, in Handbook of Fluid Machinery (in Japanese), eds. Ohashi, H. et al., Asakura-Shoten, Tokyo, Japan, 1998, 389. [80] JIS Z8126, Vacuum Technology — Vocabulary — Part 2: Vacuum Pumps and Related Terms, Japanese Standards Association, 1999. [81] Vacuum and Pressure Systems Handbook, Available from Gast Manufacturing Inc., A Unit of IDEX Corporation, P.O. Box 97, 2300 M-139, Benton Harbor, MI 49023-0097, Web site: http://www.gastmfg.com/pdf/vacpresshdbk.pdf. [82] Danielson, P., How to use the Q = SP fundamental vacuum relationship, R&D Magazine, March, 2001. [83] 2. Technical Data — Pump Capacity Ratings, Brochure No. 0300/3, Dekker Vacuum Technologies, Inc., 935 S. Woodland Avenue, Michigan City, IN 46360-5672, Web site: http://www.dekkervacuum.com.
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Compressors and Vacuum Pumps
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[84] Technical Information — Basic Principles for the Design of Liquid Ring Vacuum Pumps and Compressors, Brochure No. 120.70004.52.09E, Sterling Fluid Systems, SIHI Pumps, Inc., 303 Industrial Boulevard, P.O. Box 460, Grand Island, New York 14072-0460 http://www.sterlingfluidsystems.com. [85] Pastore, M.J., Process Vacuum Pump Technology, 2004 Press Release from: Access Intelligence, 110 William St., New York, 10038. Available at Web site: http://www.che.com/shop/sihi/press4.php. [86] Mc Nally, B., A Look at Vacuum Pumps, Paper 14-9, The Mc Nally Institute, 16231 Kalli Way, Dade City, Florida 33523, Web site: http://www.mcnallyinstitute.com/14-html/14-09.htm. [87] A Short History of Vacuum, Terminology and Technology, McAllister Technical Services, West 280 Prairie Avenue, Coeur d’Alene, ID 83815, http://www.mcallister.com/vacuum2.html. [88] Nishida, T. and Maruo, K., High performance and oil-free technology of vacuum pumps and seals (in Japanese), The Tribology, 91, 31, 1995. [89] Lecture by Jyhpyng Wang, Institute of Electro-Optic Engineering, Joint Laboratory of Optical Sciences and Ultrafast Technology, National Taiwan University, P.O. Box 23-166, Taipei, Taiwan (e-mail: [email protected]) Web site: http://ltl.iams.sinica.edu.tw/document/ training_lectures/2004/vacuum_04.pdf. [90] Transactions — Pressure Gauges and Switches, Handbook available from: OMEGA Engineering, INC., One Omega Drive, Stamford, Connecticut 06907-0047, Web site: http://www.omega.com/ literature/transactions/volume3/pressure.html. [91] Ranke, W., Lecture Notes: T- and p-Measurement, Fritz-Haber-Institute of the Max Planck Gesellschaft, Dept. of Inorganic Chemistry, Surface Analysis Group, Faradayweg 4-6, 14195, Berlin, Germany, Web site: http://w3.rz-berlin.mpg.de/∼jentoft/lehre/ranke_t_p_measurement_ 051104.pdf. [92] O’Hanlon, J.F., Vacuum pump fluids, Journal of Vacuum Science and Technology A: Vacuum, Surfaces, and Films, 2, 174, 1984. [93] O’Hanlon, J.F., A User’s Guide to Vacuum Technology, Wiley-Interscience, New York, 1980.
© 2006 by Taylor & Francis Group, LLC
10 Basic Hydraulic Pump and Circuit Design Richard K. Tessmann FES, Inc.
Hans M. Melief The Rexroth Corporation, Industrial Hydraulics Division
Roland J. Bishop Union Carbide Corporation
10.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10.2 Discussion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
10-1 10-2
Hydraulic Principles • Hydraulic Pumps and Motors • Hydraulic System Components • Basic Hydraulic System Design • Hydraulic Fluid Considerations
References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
10-51
10.1 Introduction Hydraulics, as defined by Webster, is “operated, moved, or effected by means of water.” In the 17th century it was discovered that a fluid under pressure could be used to transmit power [1]. Blaise Pascal (1623–1662) observed that if a fluid in a closed container was subjected to a compressive force, the resulting pressure was transmitted throughout the system undiminished and equal in all directions [1]. Hydraulics is by far the simplest method to transmit energy to do work. It is considerably more precise in controlling energy and exhibits a broader adjustability range than either electrical or mechanical methods. To design and apply hydraulics efficiently, a clear understanding of energy, work, and power is necessary. In this chapter, fundamentals of hydraulic pump operation and circuit design are provided. This will include: • • • •
Hydraulic principles Hydraulic system components Hydraulic pumps and motors System design considerations
10-1
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Handbook of Lubrication and Tribology
10.2 Discussion 10.2.1 Hydraulic Principles Work is done when something is moved. Work is directly proportional to the amount of force applied over a given distance according to Equation (10.1): WORK (ft.-lbs) = DISTANCE (ft.) × FORCE (lbs)
(10.1)
Power is defined as the rate of doing work and has the units of foot-pounds/second. A more common unit of measure is “horsepower (Hp).” Horsepower is defined as the amount of weight in pounds that a horse could lift one foot in one second (Figure 10.1) [1]. By experiment it was found that the average horse could lift 550 lbs one foot in one second, consequently Equation (10.2): 1 hp =
550 ft.-lbs second
(10.2)
Energy is the ability to do work. It may appear in various forms, such as mechanical, electrical, chemical, nuclear, acoustic, radiant, and thermal. In physics, the “Law of Conservation of Mass and Energy” states that neither mass nor energy can be created or destroyed, but only converted from one to the other. In a hydraulic system, energy input is called a “prime mover.” Electric motors and internal combustion engines are examples of prime movers. Prime movers and hydraulic pumps do not create energy; they simply convert it to a form that can be utilized by a hydraulic system. The pump is the heart of the hydraulic system. When the system performs improperly, the pump is usually the first component to be investigated. Many times the pump is described in terms of its pressure limitations. However, the hydraulic pump is a flow generator, moving a volume of fluid from a lowpressure region to a higher-pressure region in a specific amount of time depending upon the rotation speed. Therefore, the pump is properly described in terms of its displacement or the output flow rate that is expected from it. All pumps used in a hydraulic system are of the positive displacement type. This means that there is an intentional flow path from the inlet to the outlet. Therefore, the pump will move fluid from the inlet or suction port to the outlet port at any pressure. However, if that pressure is beyond the pressure capability of the pump, failure will occur. The pressure, which exists at the outlet port of the pump, is a function of the load on the system. Therefore, hydraulic system designers will always place a pressure-limiting component (i.e., a relief valve) at the outlet port of a pump to prevent catastrophic failures due to overpressurization. Most pumps in hydraulic systems fall into one of three categories, vane pumps, gear pumps, or piston pumps (Section 10.2.2). The action of the hydraulic pump consists of moving or transferring fluid from the reservoir where it is maintained at a low pressure and consequently a low energy state [1]. From the reservoir the pump moves the fluid to the hydraulic system where the pressure is much higher, and the fluid is at a much higher energy state due to the work that must be done by the hydraulic system. The amount
1 sec
550 lbs 1ft
FIGURE 10.1
Illustration of horsepower concept.
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Basic Hydraulic Pump and Circuit Design
10-3
of energy or work imparted to the hydraulic system through the pump is a function of the amount of volume moved and the pressure at the discharge port of the pump. Work ∝ Pressure × Flow
(10.3)
From an engineering standpoint it is common to relate energy to force times distance. Work ∝ Force × Distance
(10.4)
However, hydraulic pressure is force divided by area and volume is area times distance. Pressure =
Force Area
Volume = Area × Distance
(10.5) (10.6)
From these relationships, Equation (10.7) shows that “pressure” times “volume” is equivalent to “force” times “distance.” P (lbs/in.2 ) × V (in.3 ) = F (lbs) × D (in.)
(10.7)
The hydraulic pump is actually a “three-connection” component. One connection is at the discharge (outlet) port, the second is at the suction (inlet) port, and the third connection is to a motor or engine (Figure 10.2) [1]. From this standpoint, the pump is a transformer. It takes the fluid in the reservoir, using the energy from the motor or engine, transforms the fluid from a low-pressure level to a higher pressure. In fact, the hydraulic fluid is actually a main component of the hydraulic system and as we will see, throughout this book, has a major influence in the operation of the system. Hydraulic pumps are commonly driven at speeds from 1200 to 3600 rpm or higher while maximum pressures may vary from values less than 1000 psi to greater than 6000 psi. Tables 10.1 and 10.2 [1] show typical pressure and speed limitations for various kinds of pumps and motors. Besides pressure, there is also a temperature limitation imposed by the hydraulic fluid. This is caused by the decrease in viscosity as the fluid temperature increases. Generally, pump manufacturers set upper and lower viscosity limits on the hydraulic fluids used in their pumps. The upper viscosity limit determines
FIGURE 10.2 Illustration of pump connections.
© 2006 by Taylor & Francis Group, LLC
10-4
Handbook of Lubrication and Tribology TABLE 10.1 Hydraulic pump type
Typical Pump Performance Parameters for One Manufacturer Flow
Maximum pressure (psi)
Maximum speed (rpm)
Total efficiency (%)
External gear Internal gear
Fixed Fixed
3,600 3,000
500–5,000 900–1,800
85–90 90
Vane
Fixed Variable
2,500 1,000–2,300
900–3,000 750–2,000
86 85
Radial piston
Fixed
1,000–3,400
90
Axial piston (bent axis)
Fixed Variable
5,100–6,500 5,800
950–3,200 500–4,100
92 92
Axial piston (swash-plate)
Variable
4,600–6,500
500–4,300
91
TABLE 10.2
10,000
Typical Motor Performance Parameters for One Manufacturer
Hydraulic motor type
Flow
Maximum pressure (psi)
Maximum speed (rpm)
Total efficiency (%)
Gear
Fixed
3600
Radial piston
Fixed Variable
6100–6500 6100
1–500 1–500
Axial piston (bent axis)
Fixed Variable
5800–6500 6500
50–6000 50–8000
92 92
Axial piston (swash-plate)
Variable
4600–5800
6–4900
91
500–3000
85 91–92 92
TABLE 10.3 A Pump Manufacturer’s Viscosity Index and Temperature Guidelines Temperature (◦ F) Minimum Optimum Maximum
Viscosity index (50)
Viscosity index (95)
18 85–130 155
5 80–135 160
Viscosity index (150) 0 70–140 175
the minimum temperature for pump start-up to prevent cavitation. Cavitation occurs when there is insufficient fluid flow into the pump inlet. During cavitation the fluid will out-gas any dissolved gases or volatile liquids. The gaseous bubbles produced will then travel into the high-pressure region of the pump where they will collapse under high pressure and may cause severe damage to the pump. Moreover, overheating of the pump bearings could result due to insufficient cooling as a result of inadequate flow. The lower viscosity limit will establish the upper temperature limit of the fluid. If the upper temperature limit is exceeded, then there may be insufficient viscosity to bear the high operating loads in the pump and will thus result in lubrication failure in shortened pump life or catastrophic pump failure. Table 10.3 shows the effect of viscosity index on the pump operating temperatures for one major pump manufacturer [1]. The viscosity index is a measure of the viscosity change or resistance to flow of a liquid as the temperature is changed. A fluid having a higher viscosity index gives a smaller viscosity change with temperature than a fluid having a lower viscosity index (see Chapter 2).
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Basic Hydraulic Pump and Circuit Design
10-5
10.2.1.1 Torque and Pressure The input parameters to the hydraulic pump from the prime mover are speed and torque. Input torque to the pump is proportional to the pressure differential between the inlet port and the discharge port. Torqueinput ∝ (Pressureoutlet − Pressureinlet ) × Displacement
(10.8)
The torque required to drive a positive displacement pump at constant pressure [3–7], is Ta = Tt + Tv + Tf + Tc
(10.9)
where Ta is actual torque required, Tt is theoretical torque due to pressure differential and physical dimensions of the pump, Tv is torque due to viscous shearing of the fluid, Tf is torque due to internal friction, and Tc is constant friction torque that is independent of both pressure and speed. Substituting the operational and dimensional parameters (using appropriate units) into Equation (10.9) produces the following expression: Ta = (P1 − P2 )Dp + Cv Dp µN + Cf (P1 − P2 )Dp + Tc
(10.10)
where P1 is pressure at the discharge port, P2 is pressure at the inlet port, Dp is pump displacement, Cv is viscous shear coefficient, µ is fluid viscosity, N is rotation speed of the pump shaft, Cf is mechanical friction coefficient, and Tc is breakaway torque. Actual values of the parameters such as viscous shear coefficient, mechanical friction coefficient and breakaway torque is determined experimentally. From Equation (10.10), the required torque to drive a pump is primarily a function of the pressure drop across the pump (P1 − P2 ) and the displacement (Dp ) of the pump. 10.2.1.2 Rotational Speed and Flow The output flow of a hydraulic pump is described by: Qa = Qt − Q l − Qr
(10.11)
where Qa is delivery or actual flow rate of the pump, Qt is theoretical flow, Q l is leakage flow, and Qr is losses due to cavitation and aeration (usually neglected). Substituting the operational (speed, flow, and pressure) and dimensional (pump displacement) parameters into Equation (10.11) yields: Qa = ND p −
[Cs Dp (P1 − P2 )] µ
(10.12)
where Qa is delivery or actual flow rate of the pump and Cs is leakage flow (slip) coefficient. From Equation (10.12) it is apparent that pump delivery (Qa ) is primarily a function of the rotational speed (N ) of the pump, while the losses are a function of the hydraulic load pressure (P1 − P2 ). The leakage flow or slip that takes place in a positive displacement pump is caused by the flow through the small clearance spaces between the various internal parts of the pump, which are in relative motion. These small leakage paths are often referred to as capillary passages. Two flat parallel plates characterize most of these passages with leakage flow occurring through the clearance space between the flat plates (Figure 10.3) [1]. Therefore, the fundamental relationships for flow between flat plates are normally applied to the leakage flow in hydraulic pumps. However, to apply Equation (10.12), it is necessary to obtain the value of the slip flow coefficient through experimental results [2,3].
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10-6
Handbook of Lubrication and Tribology C ~0.25 mm under steady-state load Piston
C~(m /p)–3 where: m = viscosity p = pressure
Cylinder block Swash plate
FIGURE 10.3
Valve plate
Illustration of the leakage path between two flat plates in a piston pump.
10.2.1.3 Horsepower (Mechanical and Hydraulic) Of interest to the designers and users of hydraulic systems is the power (Hpinput ) required to drive the pump at the pressure developed by the load and the output power (Hpoutput ) which will be generated by the pump. The expression (using appropriate units) that describes the “mechanical” input horsepower required to drive the pump is: Hpinput =
Ta N 5252
(10.13)
where Hpinput is required input power (Hp), Ta is actual torque required (lb-ft), and N is rotational speed of the pump shaft (rpm). The “hydraulic” output horsepower of a hydraulic pump is described by the expression written as follows: Hpoutput =
PQ 1714
(10.14)
where Hpoutput is power output of the pump (Hp), P is pressure at the pump discharge port (psi), and Q is delivery or actual flow output (gpm). However, in the real world, Hpinput > Hpoutput . This is always true because no pump or motor is 100% efficient. As will be shown in the next section, this is mainly due to internal mechanical friction and fluid leakage within the pump. 10.2.1.4 Pumping Efficiency Hydraulic pumping efficiency or total efficiency (Et ) is a combination of two kinds of efficiencies, volumetric (Ev ) and mechanical (Em ). Volumetric efficiency is given by the following expression: Ev =
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Actual Flow Output Theoretical Flow Output
(10.15)
Basic Hydraulic Pump and Circuit Design
10-7
The second kind of efficiency is called the torque efficiency or the mechanical efficiency (Em ). The mechanical efficiency is described by the following expression: Em =
Theoretical Torque Input Actual Torque Input
(10.16)
The overall or total efficiency (Et ) is defined by the following relationship: Et = Ev × Em
(10.17)
Total efficiency is also related to power consumption by the expression: Et =
Power Output Power Input
(10.18)
Substituting Equations (10.13) and (10.14) into (10.18) produces the following expression for the overall efficiency of a hydraulic pump: Et =
0.326Ta N PQ
(10.19)
Theoretical pump delivery is also determined from the dimensions of the pump. However, if the dimensions of the pump are not known, they are determined by pump testing. Overall efficiency of a pump may also be measured by testing. However, it is difficult to measure mechanical efficiency. This is because internal friction plays a major role and there is no easy way of measuring this parameter within a pump. Rearrangement of Equation (10.17) gives the mechanical efficiency as the overall efficiency divided by the volumetric efficiency: Em =
Et Ev
(10.20)
10.2.1.5 Hydraulic System Design When designing a hydraulic system, the designer must first consider the load to be moved or controlled. Then the size of the actuator is determined. The “actuator” is a component of a hydraulic system, which causes work to be done, such as a hydraulic cylinder or motor. The actuator must be large enough to handle the load at a pressure within its design capability. Once the actuator size is determined, the speed at which the load must move will establish the flow rate of the system. For hydraulic cylinders: A×V 231
(10.21)
D × RPM 231
(10.22)
GPM = where A is area (in.3 ) and V is velocity (in./min). For hydraulic motors: GPM = where D is displacement (in.3 /rev).
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Handbook of Lubrication and Tribology
For example: calculate the hydraulic cylinder bore diameter (D) and flow (Q) required to lift a 10,000 lb load at a velocity of 120 in./min with a hydraulic load pressure not exceeding 3,000 psi. Using the fundamental hydraulic expression: Force = Pressure × Area
(10.23)
Force is equal to 10,000 lb and pressure is 3,000 psi. The area of the head-end of the cylinder is calculated by rearranging Equation (10.23) and solving for the area: Force 10,000 = Pressure 3,000
Area =
Area = 3.333 in.2 π D2 Area = 4
(10.24)
D = 2.06 in. Therefore, a double-acting single-rod cylinder with a cylinder bore of 2.25 or 2.5 in. may be used, depending on the requirements of the head side of the cylinder (Figure 10.4) [1]. However, on the rod side, the area of the rod reduces the area of the piston. Since the effective area on the rod side of the cylinder is less than that on the head side, the cylinder would not be able to lift as large a load when retracting due to pressure intensification at the rod-end causing the system relief valve to open. For this reason, all single-rod cylinders exert greater force at the rod-end when extending than retracting. On the other hand, a double-rod cylinder of equal rod diameters would exert an equal force in both directions (extending and retracting). Once the size of the cylinder is selected, the designer must consider the speed requirements of the system. The speed at which the load must be moved is dependent on the pump flow rate. Of course, the velocity of the cylinder rod and the speed of the load must be the same. By using the following equation: Q=
V ×A 231
(10.25)
where Q is flow rate into the cylinder (gpm), V is velocity of the cylinder rod (in./min), and A is cylinder area (in.2 ).
Smaller needle valve
Check valves Larger needle valve
FIGURE 10.4
Illustration of a single-rod double-acting cylinder.
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Basic Hydraulic Pump and Circuit Design
10-9
The flow rate (Q) needed to produce a specific velocity (V ) can now be calculated by combining Equations (10.24) and (10.25) and using a 2.5
diameter cylinder as an example: π D2V 924 (3.14)(2.5)2 (120) = 924
Q=
(10.26)
= 2.6 gpm Once the pressure needed to support the load and flow to produce the specified load velocity is determined, the pump selection process and system design can begin. There are several factors in pump selection and hydraulic system design that have not been addressed. For example, the service life, which is required by the system, must be decided along with the contamination level that must not be exceeded in the system. The piping sizes and the inlet conditions to the pump must be considered. The fluid that will be used in the hydraulic system is an obvious consideration. These and other factors will be addressed in later sections of this chapter.
10.2.2 Hydraulic Pumps and Motors There are three types of pumps that are used predominately in hydraulic systems, vane pumps, gear pumps, and piston pumps (Figure 10.5) [1]. While there are many design parameters, which differ between a hydraulic pump and a hydraulic motor, the general description is fundamentally the same but their uses are quite different. A pump is used to convert mechanical energy into hydraulic energy. The mechanical input is accomplished by way of an electric motor, gasoline, or diesel engine. The hydraulic flow from the output of the pump is used to power a hydraulic circuit. On the other hand, a hydraulic motor is used to convert hydraulic energy back into mechanical energy. This is accomplished by connecting the output shaft of the hydraulic motor to a mechanical actuator, such as a gearbox, pulley, or flywheel. 10.2.2.1 Vane Pumps A typical design for the vane pump is shown in Figure 10.6 [4]. The vane pump relies upon sliding vanes riding on a cam ring to increase and decrease the volume of the pumping chambers within the pump (Figure 10.7). The sides of the vanes and rotor are sealed by side bushings. While there are high-pressure vane pumps (>2500 psi), this type of pump is usually thought of as a low-pressure pump (<2500 psi), see Table 10.1. Positive displacement hydraulic pumps
Fixed displacement
Balanced vane
Gear
External
Internal
Crescent seal
Piston
FIGURE 10.5 The family of hydraulic pumps.
Vane
Bent axis
Radial
“Gerotor” seal
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Variable displacement
Axial
Piston
Radial piston Direct operated
Pilot operated
Bent axis
Axial piston
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Handbook of Lubrication and Tribology
Port plates Ring
Vane
FIGURE 10.6
Rotor
Illustration of a typical vane pump.
Inlet region
Outlet region
Cam ring Rotor Vane tip detail
Vane
Outlet region
FIGURE 10.7
Inlet region
Illustration of cam ring and vanes in a vane pump.
There are two vane pump designs. One is a balanced design while the other is unbalanced. In the balanced design, there are opposing pairs of internal inlet and outlet ports that distribute the thrust force evenly around the shaft (Figure 10.8) [4]. All modern vane pumps are of the balanced design. The vane pump is considered one of the simplest of all positive displacement pumps and can be designed to produce variable displacement, that is the output flow can be changed to suit the needs of the hydraulic system (Figure 10.9) [48]. Fluid leakage in vane pumps occurs between the high- and low-pressure side of the vanes and across the side bushings, which results in decreased volumetric efficiency and hence reduced flow output. The unbalanced design suffers from shortened bearing life due to the unbalanced thrust force within the pump. 10.2.2.2 Gear Pumps It is generally agreed that the gear pump is the most robust and rugged type of fluid power pump. While there are many gear type pumps, three are used predominately for hydraulic service. One is the external gear pump (Figure 10.10) [4] and the other two are internal gear pumps of the “crescent” seal and “gerotor” seal type, Figure 10.11 and Figure 10.12, respectively [4]. The external gear pump is the most prevalent (Figure 10.10) [4]. Notice that there are two gears, a drive and a driven gear. The number of teeth, the pitch circle diameter, and the width of the gears are the dominant parameters that control the displacement. The housing and a side plate enclose the gears. Fluid leakage in this type of pump occurs between the tips of the gears and across the side plate (Figure 10.13) [4].
© 2006 by Taylor & Francis Group, LLC
Basic Hydraulic Pump and Circuit Design
10-11 Cam ring Rotor
Inlet Inlet port
Outlet port
Vane
FIGURE 10.8 Illustration of a balanced design vane pump.
Maximum volume control Cam ring
A Thrust screw Rotor
Pressure area
Spring Pressure adjustment
Suction area
Case drain A
FIGURE 10.9 Illustration of a variable displacement vane pump.
Housing Driven gear Hydraulic loading Gear tip Drive gear
FIGURE 10.10 Illustration of an external gear pump.
© 2006 by Taylor & Francis Group, LLC
10-12
Handbook of Lubrication and Tribology Crescent seal Ring gear
Drive gear Inlet
FIGURE 10.11
Outlet
Illustration of an internal (crescent) gear pump. Rotor
Outlet
Inlet
Outside gerotor element
FIGURE 10.12
Illustration of an internal (gerotor) gear pump. Regions of wear
Discharge
Suction
Side-plate
FIGURE 10.13
Illustration of the wear and leakage areas of the external gear pump.
The crescent seal internal gear pump consists of a small internal gear and a larger ring gear (Figure 10.11) [4]. The small internal gear is driven by the prime mover. The internal gear meshes with the ring gear and turns it in the same direction. The sealing of the high-pressure chamber from the pump’s inlet is achieved by a crescent seal between the upper teeth of the internal small gear and the upper teeth of the ring gear.
© 2006 by Taylor & Francis Group, LLC
Basic Hydraulic Pump and Circuit Design
10-13
Flange Piston rod
Cylinder barrel
Piston
Housing
Drive shaft
Valve plate
Center support pin Outlet
FIGURE 10.14
Illustration of a bent-axis fixed displacement piston pump.
Rotary group Differential area
Valve plate Positioning cylinder
FIGURE 10.15
Control oil drain
Illustration of a bent-axis variable displacement piston pump.
In the gerotor gear pump, the inner gerotor has one tooth less than the outer element (Figure 10.12) [4]. The internal gear is driven by the prime mover and in turn drives the outer element in the same direction. There is no satisfactory gear pump design in which the displacement can be varied. 10.2.2.3 Piston Pumps The piston pump is operated at the highest pressure of all of the pumps normally found in hydraulic applications. The piston pump is manufactured in the axial, bent-axis, and radial configurations. In addition, there are both fixed and variable displacement bent-axis configurations, Figure 10.14 and Figure 10.15, respectively [4]. The axial design configuration is predominate in hydraulic systems and will be the basis of the discussions here. A typical example of an axial fixed displacement piston pump is shown in Figure 10.16 [5], while a typical example of an axial variable displacement piston pump is illustrated in Figure 10.17 [5]. Variable displacement piston pumps lend themselves to the incorporation of various valve mechanisms that will alter the performance of this pump. For example, it can be a pressure compensated pump in one configuration, where the valve mechanism will alter the displacement of the pump to limit the outlet pressure to some preselected value. A pressure-compensated piston pump is illustrated in Figure 10.17 [5].
© 2006 by Taylor & Francis Group, LLC
10-14
Handbook of Lubrication and Tribology Spring force transmitting pin (3 spaced evenly)
Shoe retainer plate (retracts pistons) Swash-plate
Cylinder block spring
Ports are rotated 90° to show all flow
Valve plate
Spherical washer Cylinder block
FIGURE 10.16
Piston
Shoe
Illustration of an axial fixed displacement piston pump. Adjustment spring
Compensator spool Yoke Yoke return spring
FIGURE 10.17
Yoke actuating piston
Illustration of an axial variable displacement piston pump.
As seen in Figure 10.16 and Figure 10.17, the major components of a piston pump are the pistons, the piston or cylinder block, valve plate, piston shoes, swash-plate, and the drive shaft. In operation, the shaft drives the piston block, which in turn rotates the pistons. Springs and a retainer plate hold the pistons against the swash-plate. For the piston pump to produce a flow the swash-plate must be at some angle relative to the centerline of the shaft, which is also the axial center of rotation of the pistons. The pistons ride on the surface of the swash-plate and the angle will force the pistons to move in and out of the piston or cylinder block. The greater the swash-plate angle the larger the piston stroke and the greater the displacement of the pump. The dependence of pump displacement (V ) on the swash-plate angle “α” is shown by the following equation and Figure 10.18 [6]: V =
© 2006 by Taylor & Francis Group, LLC
π dk2 × Dk (tan α) 4
(10.27)
Basic Hydraulic Pump and Circuit Design
10-15 h
dk
Dk
a
FIGURE 10.18
Sketch of an axial piston pump “swash-plate” design showing displacement angle “α.”
a dk rh
FIGURE 10.19
Sketch of an axial piston pump “bent-axis” design showing displacement angle “α.”
where V is pump displacement (in.3 ), dk is piston bore diameter (in.), Dk is piston bore circle diameter (in.), and α is swash-plate angle (◦ ). For an axial piston pump “bent-axis” design, the pump displacement (V ) is described by the bent-axis angle “α” according to the following equation and Figure 10.19 [6]: V =
π dk2 × 2rh z (sin α) 4
(10.28)
where rh is piston bore circle radius and z is number of pistons. The third kind of piston pump is the radial design (Figure 10.20) [4]. In general, the radial piston pump has the highest continuous pressure capability than any other type of pump (Table 10.1). Figure 10.21 [4] shows the basic configuration of a three-piston pump. The pistons are positioned radially to an eccentric drive shaft. Each hollow piston consists of an inlet check valve, a spring, a piston barrel, a pumping chamber, an outlet check ball, and a support bearing.
© 2006 by Taylor & Francis Group, LLC
10-16
FIGURE 10.20
Handbook of Lubrication and Tribology
Cross-section view of a radial piston pump. Outlet
Support bearing Outlet check ball Pumping chamber Piston barrel Spring Inlet check
Eccentric Housing
Inlet
FIGURE 10.21
Illustration of the interior of a radial piston pump.
As the drive shaft is rotated, the spring holds the base of the piston in contact with the eccentric camshaft. The downward motion of the piston causes the volume to increase in the pumping chamber. This creates a reduced pressure, which allows the inlet check valve to open, allowing oil to enter the pumping chamber. The oil enters the chamber by way of a groove machined into the camshaft circumference. Further rotation of the camshaft causes the piston to move back into the cylinder barrel. The rapid rise in chamber pressure closes the inlet check valve. When the rising pressure equals the system pressure, the outlet check valve opens allowing flow to exit the piston into the pressure port of the pump. The resulting flow is the sum of all the piston displacements. The number of pistons that a radial pump can have is only limited by the special restrictions imposed by the size of the pistons, housing, and camshaft. For a radial piston pump, the pump displacement (V ) is defined by the piston bore (dk ) and the cam thrust (e) according to the following equation and Figure 10.22 [6]: V = where e = Cam thrust (in.).
© 2006 by Taylor & Francis Group, LLC
dk2 × π × 2e × z 4
(10.29)
Basic Hydraulic Pump and Circuit Design
10-17
dk
e
FIGURE 10.22
Sketch of a radial piston pump showing the parameters used to calculate displacement.
Typical applications for radial pumps include cylinder jacking, crimping, and holding pressure on hydraulic presses. However, it should be noted that for extremely high-pressure applications, the displacements of radial pumps are usually not larger than 0.5 in.3 /rev. For example, at 1800 rpm, a 0.5 in.3 /rev displacement pump will only deliver ∼3.9 gpm. Assuming an efficiency of 93% at a load pressure of 10,000 psi, the pump would require a 24-hp electric motor (Equations [10.14] and [10.18]). 10.2.2.4 Sizing and Selection of Pumps As was mentioned earlier, the sizing of the system pump actually begins at the load. The specification for a hydraulic system only deals with the movement of a load and therefore the size of the pump must be calculated from this information. The pump is a flow generator that develops flow, pressure is a result of the pressure losses within the system and the pressure necessary to maintain the motion of the load. Hence, the first parameter in sizing a pump is to determine the required flow. Then, as was shown earlier, the pressure capability of the pump must be considered. The pressure necessary to deal with the load was determined in sizing the cylinder. Then it is only necessary to add the system losses to the pressure to arrive at the pressure capability of the pump. Once the size and pressure capability of the pump is known, the type of pump and the manufacturer must be selected. There are basically three types of pumps that are normally used. The main criterion, which will play heavily in the selection of the pump, is the type of control that is used. In an open center system, which is not extremely high pressure, any of the three pump types can be used. In this case, price and personal opinions of the designer will prevail. However, if the system is to be closed center with pressure or load compensation, the usual selection is the variable displacement piston pump. 10.2.2.4.1 Open Loop Circuit The open loop circuit is by far the most popular design. An example of an “open loop” circuit is shown in Figure 10.23 [4]. In this figure, an electric motor powers a variable displacement pump that draws hydraulic fluid from the reservoir and pushes the fluid through a directional control valve. The fluid from the control valve can be directed to either side of a reversible hydraulic motor and then is sent back to the reservoir. Pumps used in open loop applications can only pump fluid in one direction. In contrast to the reversible hydraulic motor, the pump’s ports are not the same size, the inlet port is always larger than the outlet port. The advantage of an open loop design is that, if necessary, a single pump can be used to operate several different actuator functions simultaneously. The main disadvantage is its large reservoir size. Generally, the reservoir is sized to hold at least three times the volume of fluid that can be supplied by the pump in one minute.
© 2006 by Taylor & Francis Group, LLC
10-18
Handbook of Lubrication and Tribology
Motor Directional control Pump Relief valve
FIGURE 10.23
Reservoir
Illustration of an open loop circuit.
Crossport relief valves
Bi-directional motor Leakage oil return
Supercharge relief
Heat exchanger
Main pump
Filter Supercharge pump
FIGURE 10.24
Illustration of a closed loop circuit.
10.2.2.4.2 Closed Loop Circuit In contrast to the open loop design, the closed circuit eliminates the need for a large storage reservoir. Figure 10.24 [4] shows an illustration of a “closed loop” circuit. In this design, a reversible pump is used to drive a reversible hydraulic motor. The closed loop design is always used in conjunction with a smaller “supercharge” circuit. The supercharge circuit consists of a small fixed displacement pump (usually about 15% of the displacement of the main pump), a small fluid reservoir, filters, and a heat exchanger. The supercharge circuit always works on the low-pressure side of the main loop. Its function is to pump freshly filtered fluid into the closed loop through check valves, while bleeding-off a percentage of the hot
© 2006 by Taylor & Francis Group, LLC
Basic Hydraulic Pump and Circuit Design
10-19
fluid through a bleed valve. This hot fluid is then cooled by a heat exchanger and stored in a small reservoir before returning to the main system. The pressure in the supercharge circuit is limited to 100 to 300 psi by the supercharge relief valve. The requirements of the pump/motor combination and the operating conditions of the system determine the pressure setting of this valve. The cross-port relief valves on the motor are there only to protect the actuator from load induced pressure spikes. They are not meant to function like those found in open loop designs, since that would cause severe overheating of the circuit due to the insufficient fluid supply, inherent in a closed loop, to carry away this extra heat. In closed loop circuits, the pressure, the flow, and the directional motor control are all achieved by controlling the variable displacement pump. The advantages of a closed loop circuit are that high horsepower systems are compact, efficient, and require less hydraulic fluid storage. The high efficiency of this circuit is due to the pump control being designed to supply only the fluid flow required by the actuator to operate at the load induced pressure. The pump is the heart of the system and controls the direction, acceleration, speed, and torque of the hydraulic motor, thus eliminating the need for pressure and flow control components. A major disadvantage of a closed loop circuit is that a single pump can only operate a single output function or actuator. In addition, this type of hydraulic circuit is generally used only with motor actuators. 10.2.2.4.3 Half-Closed Loop Circuit Figure 10.25 [4] is an illustration of a “half-closed” loop circuit. This circuit is similar to the closed circuit except that it can be used with cylinder actuators having differential areas. As can be seen from the figure, during cylinder extension, the pump must generate a larger flow from its left-hand port than is being returned to its right-hand port from the cylinder. The extra fluid needed by the pump is supplied by its left-hand inlet check valve, which is an integral part of the pump. When the pump control strokes the pump over center, the flow from the pump is reversed and the cylinder begins to retract. During retraction, the differential area of the cylinder piston causes a larger flow than needed at the inlet of the pump. This excess flow is directed to the reservoir through the “unloading valve.” The unloaded fluid is filtered and cooled prior to its return to the reservoir. In this way, a portion of the closed loop fluid is filtered and cooled in an open loop circuit each time the cylinder is cycled. 10.2.2.5 Contamination Considerations Every hydraulic system will have particulate contamination entrained in the circulating fluid. These contaminant particles can enter the clearance space of every component, but more specifically the hydraulic pump [7]. Contaminants will enter the system from reservoir breathers, seals, etc., as well as from the wear of internal surfaces. There are three ways of addressing the contamination problems in a hydraulic system. This can be accomplished by using return line and reservoir breather filters. One way is to prevent particles from entering the system. Since the particles, which enter the hydraulic system, will cause wear on the internal surfaces of the components, the elimination of these particles will also reduce the production of wear debris. The second approach is to filter the particles, which become entrained in the circulating fluid, while the third approach is to select components that effectively resist the contaminant attack. In selecting the pump, none of these approaches are completely adequate. Since the development of contaminant sensitivity test procedures at the Fluid Power Research Center, formerly located at Oklahoma State University, it is now possible to evaluate the efficiency of seals [8,9] and breathers in preventing the entrance of particulate contamination as well as the removal efficiency of hydraulic filters [10,11]. In addition, contaminant sensitivity test procedures are available to evaluate the resistance or tolerance of a hydraulic component, such as a pump, to entrained particulate contaminants [12–14]. Therefore, knowing the ability of the seals and breathers to prevent the entrance of contamination and the effectiveness of the filter in removing that contaminant, which does enter the system, a reasonable selection of the pump can be made to produce the desired service life. Table 10.4 shows the level of cleanliness that can be achieved as a function of degree of filtration as it applies to various hydraulic components. This subject is covered in much greater detail in Chapter 12.
© 2006 by Taylor & Francis Group, LLC
10-20
Handbook of Lubrication and Tribology
Unloading valve
Crossport relief valves Filter
Variable displacement pump
Heat exchanger
Suction check valves (incorporated in pump)
FIGURE 10.25
Oil reservoir
Illustration of a half-closed loop circuit. TABLE 10.4 Filtration and Cleanliness Guidelines for Various Hydraulic Components Filtration (µm)
Cleanliness (class)
1–5 10
0–1 2–4
20–25 40
4–5 6
Hydraulic application Servo valves Piston pumps and motors, flow controls, relief valves Gear and vane pumps Infrequent duty cycle and noncritical components
10.2.2.5.1 Preparation of Pipes and Fittings When installing pipes and fittings on a hydraulic system, it is imperative that they be as clean as possible. The following steps are recommended to prepare metal pipes and fittings prior to installation [15]: 1. Ream inside and outside edges of pipe or tubing and clean with a wire brush to loosen and remove any particles.
© 2006 by Taylor & Francis Group, LLC
Basic Hydraulic Pump and Circuit Design
10-21
2. Sandblast short pieces of pipe and tubing to remove any rust and scale. In the case of longer pieces or short pieces having complex shapes, they should be first cleaned of all grease and oil in a degreasing solvent and then pickled in a suitable solution until all rust and scale is removed. 3. After pickling, rinse all parts thoroughly in cold running water and then immerse parts in a tank containing neutralizing solution at the proper temperature and length of time as recommended by the manufacturer. 4. Rinse parts in hot water and place into another tank containing an antirust solution. If parts are not to be immediately installed, they should be left to air-dry with antirust solution remaining on them. If pipes are dry and will be stored, they should be capped to prevent dirt from entering. Before using any pickled part, it should be thoroughly flushed with a suitable degreasing solvent. 5. Cover all openings into the hydraulic system to prevent dirt and any foreign matter from entering the system. 6. Inspect all threaded fittings and remove any burs or metal slivers. 7. Before filling or adding hydraulic fluid to the reservoir, make sure that the fluid is as specified and that it is clean. 8. When adding hydraulic fluid to the reservoir, use a fluid filtration cart to pre-filter the fluid as it enters the reservoir. Never add fluid directly from the storage container or drum without filtering. 10.2.2.6 Pump Performance Characteristics The performance of the pump is extremely important in the overall success of any hydraulic system. A pump exhibits mechanical type losses as well as volumetric losses [16]. Mechanical losses are due to the motion of the working element within the pump due to friction. As shown previously (Equation [10.10]), the theoretical torque required to drive the pump is equal to the product of the pump displacement and the pressure differential across the pump. Obviously, the actual torque is greater than this theoretical value in order to make up for the mechanical losses within the pump. As stated earlier the mechanical efficiency is the ratio of the theoretical torque to the actual torque (Equation [10.16]). The effectiveness of a pump in converting the mechanical input energy into output hydraulic energy must be measured by tests. Theoretically, at low pressures the output flow of a positive displacement pump is equal to the product of the displacement and the shaft rotational speed as shown by the following equation:
FLOW (gpm) =
DISPLACEMENT (in.3 /rev ) × SPEED (rpm) 231
(10.30)
As the differential pressure across the pump increases, the flow through the clearances or leakage paths within the pump will increase to create slip flow, which will be subtracted from the output flow. Therefore, the volumetric efficiency is the ratio of the actual flow to the theoretical flow and this parameter reflects the magnitude of the volumetric losses (Equation [10.15]). The overall efficiency is equal to the volumetric efficiency multiplied by the mechanical efficiency (Equation [10.17]). Figure 10.26 illustrates the relationship between volumetric, mechanical, and overall efficiency of a hydraulic pump. To properly select a hydraulic pump for a given application, the efficiency information is extremely important. This data must be acquired by the pump manufacturer by testing and is normally reported in a graph, which resembles that shown in Figure 10.27. The upper curves in this figure show that at a given load pressure the volumetric efficiency increases with speed. This is due to the fact that all fluids have a property known as viscosity and that at greater and greater speeds there is insufficient time available for the fluid to leak across to case or slip past clearances in the pump. In addition, the lower curves indicate that at a given load pressure the input power requirement increases as speed increases. This is due to the increased flow output of the pump as rotational speed increases (Equations [10.12] and [10.13]). It should be noted that this data is usually run at one temperature using one fluid. No viscosity or density effects are taken into consideration in these data. The cost penalty paid
© 2006 by Taylor & Francis Group, LLC
10-22
Handbook of Lubrication and Tribology
Volumetric efficiency h
Mech
anica
Operating range
FIGURE 10.26
l effic
Over all ef
iency
h
ficien
cy h
Typical hydraulic pump efficiency curves.
80
500 PSI
60
1500
40
60
40
20
Input power (hp)
Input power (kW)
80
0 2500
60
60
40
50
20
40
0
30
120
20
100
10
40
80
0
0
60 40
2500 PSI 1500 500
20 0
0
0
500
1000
1500
2000
Speed (rpm)
FIGURE 10.27
Typical performance data reported by a pump manufacturer.
© 2006 by Taylor & Francis Group, LLC
2500
3000
200 160 120 80
Delivery (L/min)
80
20
Delivery (ppm)
Overall efficiency
100
Volumetric efficiency
100
eo
re
tic
al
10-23
Th
Discharge flow
Basic Hydraulic Pump and Circuit Design
Increased viscosity Higher density Lower inlet pressure
Shaft speed
FIGURE 10.28 density.
Discharge flow of a positive displacement pump as a function of shaft speed, fluid viscosity, and
for poor efficiency can be significant. Therefore, care must be taken to evaluate the efficiency characteristics of the pump during the selection process. The mechanical strength of the pump to survive the pressure duty cycle expected in the application is reflected by the proof and burst pressure information. This information is obtained through failure tests conducted by the pump manufacturer. The proof pressure is normally 1.5 the rated pressure and the burst pressure is 2.5 times the rated pressure. The ability of the pump to operate for sufficient time at the actual pressure duty cycle is determined by conducting endurance tests at conditions agreed upon by the industry. 10.2.2.7 Pump Inlet Condition In theory, a positive displacement pump will produce flow in direct proportion to the shaft speed (Equation [10.12]). However, if the fluid cannot be supplied to the pumping chambers of the pump, this relationship will not hold and the pump is said to cavitate. The flow vs. shaft speed for a typical hydraulic pump will be linear up to the point that fluid can no longer enter the pumping chambers of the pump as these chambers are opened and closed due to shaft rotation. When this occurs the chambers will only partially fill and the outlet flow will reduce. Under these conditions, the pump will be starved for fluid. The speed that this starvation will occur depends upon the viscosity and the density of the hydraulic fluid as well as the physical configuration of the pump inlet and the connecting lines. This phenomenon is illustrated in Figure 10.28. In considering the starvation of a positive displacement pump, there is normally very little that can be done by the system designer with the configuration of the pump inlet. Moreover, the fluid being used in the system is generally selected for reasons other than the pump inlet conditions, such as for hightemperature operation, fire resistance, or for biodegradability. Therefore, it is necessary to size the inlet piping and position the pump relative to the reservoir such that the inlet pressure to the pump is positive. The pressure at the inlet to the pump is normally called the “net positive suction head” (NPSH) and may be calculated in terms of absolute pressure [17]. The entire system from the fluid level in the reservoir to the inlet port of the pump must be taken into account when determining the NPSH (Figure 10.29) [17]. The primary factors in determining the NPSH are: • The atmospheric head or the atmospheric pressure at the particular location (HA ). • The friction head or the pressure needed to overcome the losses due to friction when the fluid is flowing through the pipe, fittings, valves, and area changes (HP ). Table 10.5 [18] shows data for various hose and pipe diameters.
© 2006 by Taylor & Francis Group, LLC
10-24
Handbook of Lubrication and Tribology
HA Hvp HR
HP
FIGURE 10.29
Illustration of the “net positive suction head” parameters. TABLE 10.5 Pressure Drop P (psi/ft) and Flow Rates (gpm) for Various Hose and Pipe Diameters at Typical Flow Velocities P (gpm) Hose/pipe I.D. (in.) 0.500 H 0.750 H 0.875 H 1.00 H 2.00 H 3.00 H 4.00 H 0.493 P 0.742 P 0.884 P 1.049 P 2.067 P 3.068 P 4.026 P
2 ft/sec
4 ft/sec
10 ft/sec
15 ft/sec
20 ft/sec
0.176 (1.22) 0.0782 (2.75) 0.0575 (3.75) 0.0440 (4.90) 0.0110 (19.6) 0.00489 (44.1) 0.00275 (78.3) 0.181 (1.19) 0.0799 (2.70) 0.0563 (3.83) 0.0400 (5.39) 0.0103 (20.9) 0.00468 (46.1) 0.00272 (79.4)
0.352 (2.45) 0.156 (5.51) 0.115 (7.50) 0.0880 (9.79) 0.0220 (39.2) 0.0186 (88.1) 0.0130 (157) 0.362 (2.38) 0.160 (5.39) 0.113 (7.65) 0.0800 (10.8) 0.0206 (41.8) 0.0181 (92.2) 0.0129 (159)
0.880 (6.12) 0.391 (13.8) 0.432 (18.7) 0.365 (24.5) 0.154 (97.9) 0.0925 (220) 0.0646 (392) 0.905 (5.95) 0.400 (13.5) 0.426 (19.1) 0.344 (26.9) 0.147 (105) 0.0900 (230) 0.0641 (397)
1.32 (9.18) 1.06 (20.7) 0.878 (28.1) 0.743 (36.7) 0.312 (147) 0.188 (330) 0.131 (588) 1.36 (8.92) 1.08 (20.2) 0.867 (28.7) 0.700 (40.4) 0.300 (157) 0.183 (346) 0.130 (595)
2.92 (12.2) 1.76 (27.5) 1.45 (37.5) 1.23 (49.0) 0.517 (196) 0.311 (441) 0.217 (783) 2.97 (11.9) 1.78 (27.0) 1.43 (38.3) 1.16 (53.9) 0.496 (209) 0.303 (461) 0.215 (794)
• Static inlet head or the vertical distance from the centerline of the pump inlet to the free surface level in the reservoir (HR ). • The vapor pressure of the fluid (HVP ). The above parameters can be used to calculate the NPSH according to the following equation: NPSH = HA + HR − HP − HVP
© 2006 by Taylor & Francis Group, LLC
(10.31)
Basic Hydraulic Pump and Circuit Design
10-25
psi 15 14 13 12 11 10 9 8 7 6 5 4 3 2 1 0 0
5
10
15
20
25
30
35
40
45
50
55
60
65
70
75
80
85
90
95 100
Temperature (°C)
FIGURE 10.30 Water vapor pressure curve.
It is interesting to note that the term HVP , though negligible for mineral oil-based hydraulic fluids, can be the most dominant term where volatile fluids are concerned, such as with water-based fire resistant hydraulic fluids. Figure 10.30 shows a plot of water vapor pressure as a function of temperature [17]. There are equations for each of the factors that are involved in establishing the NPSH. However, numerous graphs and nomographs have reduced the burden of calculating the NPSH for a given pump and inlet condition. One such graph is given in Figure 10.31. With this graph, one can determine the approximate NPSH at the pump inlet from the flow and speed of the pump. As an example, a pump having a displacement of 0.05 gal/rev. running at 1800 rpm would, according to the nomograph, require a NPSH of ∼24 ft of oil pressure, or 0.35 × 24 = 8.4 psi (for mineral oil), to prevent cavitation. This graph provides information concerning the minimum NPSH and is fairly accurate for viscosities below about 200 SUS and specific gravities of about 0.9. Care should be taken when dealing with a fluid with a high specific gravity, such as many of the synthetic fluids (polyol esters, phosphate esters, etc.) and water-based fluids. In addition, complicated or exceptionally long inlet lines should be considered as special cases when determining the NPSH. Friction head or the losses due to the pipe friction acting upon the fluid flowing through the pipe can be calculated using Darcy’s equation as given below: P = λ
L ρQ 2 2DA 2
(10.32)
where P is pressure loss due to friction, λ is friction factor, L is length of pipe, ρ is density of fluid, Q is flow through the pipe, D is pipe diameter, and A is cross-sectional area of pipe. The friction factor, λ, can be obtained from a modified Moody diagram as shown in Figure 10.32 [5]. It should be noted that the solid line on the left-hand side of the graph shown in Figure 10.32 is for fully developed laminar flow, while the solid line on the right-hand side is for fully developed turbulent flow. The dashed lines show the changes that occur when the laminar flow is not fully developed and at very
© 2006 by Taylor & Francis Group, LLC
10-26
Handbook of Lubrication and Tribology 1500 gpm 100 70 50
0
rpm
0 90
0 10
00
7
70
18
50
5
00
3
30
12
10
2
20
0
90
15
NPSH (ft)
00 10
0 30
0 15
00
24
10
1.5
0
7
60
0
45
5
1
0 20
15
00 48 0 0 42 0 0 36
0 50
20
0 70
00
60
30
3
0.7
0
2
30
1.5
0.5
1 0.7
0.3
0
15
m
gp
0.2
50 rpm
0
0.15
10
0.1 0.001
0.002
0.005
0.01
0.02
0.05
0.1
0.2 0.3
0.5
1
Displacement (gal/rev)
FIGURE 10.31
Minimum pump head pressure estimation nomograph.
0.10 0.08
l=
75 R
Friction factor l
0.06 l=
0.04
l=
0.160 R 0.182
64 R
0.02 l= 0.01 103
2
4
6
0.3164 R 0.25
104 Reynolds number
FIGURE 10.32
Friction factor vs. Reynolds number (Moody diagram).
© 2006 by Taylor & Francis Group, LLC
2
4
6
105
2
Basic Hydraulic Pump and Circuit Design
10-27 Diameter (mm) 100
Fluid velocity (m/sec)
Pressure loss (bar/m) 100 50
100 50 40 30
50 40 30
40 30
20
1000
20
20 10
500 400 300
10 5 4 3
5 4 3
10
2 1 5
1.0 0.5 0.4 0.3
4
50 40 30 20
3
0.2
10
0.2
2
0.1
5 4 3
0.1 0.05 0.04 0.03
200 100
2
0.5 0.4 0.3
Viscosity (cP)
0.05 0.04 0.03
1
0.02
2 1
0.02 0.01 0.5
0.01
FIGURE 10.33 Pressure loss per unit pipe length for laminar flow.
high Reynolds Numbers. Equation (10.32) can be rewritten in terms of head loss as follows: h=
P ρG
(10.33)
where h is head loss (ft) and G is Gravitational constant (ft/sec2 ). The frictional losses can also be found using nomographs. The nomograph shown in Figure 10.33 [5] can be used to find the pressure loss in a pipe due to friction under conditions of laminar flow. An example is shown in Figure 10.33. In this example, the fluid velocity is 2.0 m/sec and the fluid viscosity is 30 cP. A straight line is drawn between these two points. The pipe diameter is 20 mm. By drawing a straight line between the pipe diameter of 20 mm through the intersection of the first line drawn with the turning line, one will find the pressure loss per foot of this pipe at these flow conditions to be ∼0.06 bar/m. Then
© 2006 by Taylor & Francis Group, LLC
10-28
Handbook of Lubrication and Tribology Diameter (mm) Pressure loss (bar/m) Fluid velocity (m/sec) 100 50 40 30
100 50 40 30
50 40
20
30
20
10
20
10
5 4 3 2
0.5 0.4 0.3 0.2 0.1 0.05 0.04 0.03 0.02
500 400 300 10
200 100
2 1.0
Viscosity (cP) 1000
5 4 3
1 0.5 0.4 0.3
5 4 3
0.1
2
0.05 0.04 0.03
50 40 30 20
0.2
10 5 4 3 1
0.02
2 1
0.01
0.01
FIGURE 10.34
100
0.5
Pressure loss per unit pipe length for turbulent flow.
multiply this number by the total length of the pipe to find the total pressure loss. The nomograph shown in Figure 10.34 [5], for turbulent flow, is used in exactly the same way as that shown in Figure 10.33 for laminar flow.
10.2.3 Hydraulic System Components There are probably as many different hydraulic systems and component designs as there are designers. However, a fundamental hydraulic system consists of the following components and circuits in addition to the pumping component: • Flow control • Pressure control • Rotary and linear actuators
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Basic Hydraulic Pump and Circuit Design
10-29
• Accumulators • Piping and hose • Reservoir Each of these components has a unique mission in the operation of a hydraulic system. 10.2.3.1 Flow Control Flow control in a hydraulic system is commonly used to control the rod velocity of linear actuators or the rotary speed of hydraulic motors. There are three ways to accomplish flow control. One is to vary the speed of a fixed displacement pump while another is to regulate the displacement of a variable displacement pump. The third way is with the use of flow control valves. Flow control valves may vary from a simple orifice to restrict the flow to a complex pressure compensated flow control valve and on to flow dividers. 10.2.3.1.1 Uncompensated Flow Control Needle Valves The simplest uncompensated flow control is the fixed area orifice. Normally, these orifices are used in conjunction with a check valve so that the fluid must pass through the orifice in one direction, but in the reverse direction the fluid may pass through the check thus bypassing the orifice. Another design incorporates a variable area orifice so that the effective area of the orifice can be increased or decreased (usually manually). One style of the variable area orifice with a reverse flow check valve is shown in Figure 10.35 [5]. These uncompensated flow control valves are used where exact flow control is not critical. Flow through an orifice is proportional to the pressure drop across the orifice. Therefore, if the pressure differential increases, flow will also increase. To avoid this, a compensated flow control valve must be utilized. 10.2.3.1.2 Compensated Flow Control Needle Valves A very simple compensated flow control valve is shown in Figure 10.36 [5]. In this valve, the force opposed by the spring is a function of the pressure drop across the fixed orifice, not the pressure drop across the entire valve. As the pressure differential across the valve from the inlet to the outlet increases, flow will also attempt to increase. However, any increase in flow will be accompanied by a resulting increase in the pressure drop across the fixed orifice. When this pressure differential becomes larger than the spring preload, the valve spool will shift and the outlet port will be restricted. There are compensated flow control valves, which are much more complex than the one shown; however, most operate in the same way since the pressure drop across the control orifice is held constant by utilizing a secondary variable orifice.
FIGURE 10.35 Illustration of an uncompensated flow control needle valve.
© 2006 by Taylor & Francis Group, LLC
10-30
Handbook of Lubrication and Tribology Control orifice
Fixed orifice
Outlet
Inlet
FIGURE 10.36
Illustration of a compensated flow control needle valve. Hydraulic pump and circuit design CD values
Flow
0.62 0.98
0.61
0.82
0.97
0.54
0.72
1 1
2
3
4
2 3
CD = 0.8 0.63 0.7 0.8
4
Series circuit
Parallel circuit PX
P1
1 2 3 4
P1
1
P2
P3
2
P3
P4
3
P4
P5
4
P5
P2
Series / Parallel circuit
Partial parallel circuit
FIGURE 10.37 “CD” values and circuit definitions for orifice calculations.
The following equations can be used to calculate the flow rate (Q) through a needle valve, or a series of valves, at a given system pressure. Refer to Figure 10.37 [18] for the “orifice coefficient” (CD) values and circuit definitions. Parallel circuits Q = 29.81 ×
P × [CD 1 × D12 + CD 2 × D22 + · · · ] SG
(10.34)
where Q is flow rate (gpm), P is pressure drop (psig), SG is specific gravity of fluid, CD is orifice coefficient, and D is orifice diameter (in.).
© 2006 by Taylor & Francis Group, LLC
Basic Hydraulic Pump and Circuit Design Priority outlet
10-31 Secondary outlet
Inlet from pump
Fixed orifice Priority flow divider
FIGURE 10.38 Illustration of a priority flow divider valve.
Series circuits Q=
P × 29.812 SG × ([1/(CD1 × D12 )]2 + [1/(CD2 × D22 )]2 + · · · )
(10.35)
Series/parallel circuits Qx =
29.81 × CD x × Dx2
×
P SG
(10.36)
Partial parallel circuits 29.81 × CD 1 × D12 × P1 − P2 + CD 2 × D22 × P1 − P3 + · · · Q= √ SG
(10.37)
10.2.3.1.3 Flow Dividers Flow dividers are a form of flow control valves. There are at least two types of flow dividers. One is called a priority flow divider while the other is a proportional flow divider. The priority type of flow control provides flow to a critical circuit at the expense of other circuits in the system. For example, many of the earth-moving machines are equipped with power steering. From a safety standpoint, the steering system is a very critical function. Figure 10.38 [5] illustrates a priority flow divider. In operation the flow will enter the priority flow divider from the right-hand end as shown [5]. When the flow reaches a value such that the pressure drop across the fixed orifice produces a force larger than that provided by the spring, the spool will move to the left. This action will begin to close the priority outlet port and open the secondary outlet. When the flow is below the designed priority flow the spool will be all the way to the right and the secondary will be closed and the priority will be wide open. The proportional type flow divider follows the same principle as the priority flow divider, except that two orifices are used and the spool is normally spring loaded to a particular flow split ratio. 10.2.3.2 Pressure Control The primary pressure control valves are relief valves. There are direct acting relief valves and pilot operated relief valves. In addition, the pressure-reducing and the counterbalance valves fall under the pressure control category. There are a great many more valves that would fall into the pressure control category that will not be discussed here.
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10-32
Handbook of Lubrication and Tribology
Outlet
Inlet
FIGURE 10.39
Illustration of a direct acting relief valve.
10.2.3.2.1 Pressure Relief Valves There are two major kinds of pressure relief valves. One is described as a direct acting valve and the other is pilot operated. The direct acting relief valve is shown in Figure 10.39 [5]. The model shown is actually adjustable but not all direct acting relief valves are externally adjustable. In operation, the flow enters from the bottom of the valve shown in Figure 10.39. When the inlet pressure reaches the value such that the pressure times the exposed area of the ball is greater than the spring setting, the valve will begin to pass hydraulic fluid. Notice that the spring must be compressed in order for the seat (ball) to move and provide greater flow area. Therefore, the pressure will increase as the flow through the valve increases. The pressure at which the valve first begins to open is called the cracking pressure, while the pressure at rated flow is termed the full flow pressure. In the case of the direct acting relief valve the difference between the cracking pressure and the full flow pressure could be large. This difference is called the override pressure. The pilot operated relief valve is shown in Figure 10.40 [5]. The pilot operated pressure relief valve increases pressure sensitivity and reduces the pressure override normally found in relief valves using only the direct acting force of the system pressure against a spring element. In operation, fluid pressure acts upon both sides of piston 1 due to the small orifice “C” through the piston and the piston is held in the closed position by the light bias spring 2. When the pressure increases sufficiently to move the pilot poppet “4” from its seat, the fluid behind piston 1 will be directed to the low-pressure area, such as the return line. The resulting pressure imbalance on piston 1 will cause it to move in the direction of the lower pressure, compressing spring 2 and opening the discharge port. This action will effectively prevent any additional increase in pressure. The setting of the pilot operated relief valve is adjusted by the preload of the poppet spring 3. 10.2.3.2.2 Pressure-Reducing Valves Pressure-reducing valves are used to supply fluid to branch circuits at a pressure lower than that of the main system. Their main purpose is to step the pressure down to the requirements of the branch circuit by restricting the flow when the branch reaches some preset limit. The pressure-reducing valve is illustrated in Figure 10.41 [5]. In operation, a pressure-reducing valve permits fluid to pass freely from port “C” to port “D” until the pressure at port “D” becomes high enough to overcome the force of spring 2. At this
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Basic Hydraulic Pump and Circuit Design
10-33
3
Drain
4
D
2 C
1 Outlet
Inlet
Discharge
FIGURE 10.40
Illustration of a pilot operated relief valve.
Drain 2 1
C D
E
FIGURE 10.41
Illustration of a pressure-reducing valve.
point, the spool will move, obstructing flow to port “D” and thus regulating the downstream pressure. The direction of flow is irrelevant with a pressure-reducing valve since the spool will close when the pressure at port “D” reaches the set value. If free reverse flow is required a check valve must be used. 10.2.3.2.3 Counterbalance Valves The normal use of counterbalance valves is to maintain backpressure on a vertically mounted cylinder to hold vertical loads such as encountered in hydraulic presses. A typical circuit using a counterbalance valve is shown in Figure 10.42 [5]. Counterbalance valves can be operated by either a direct pilot or a remote pilot. As shown in Figure 10.42, when a direct pilot is utilized the pressure on the rod side of the cylinder must reach the valve setting before it will open and permit flow. When a remote pilot operates the valve, the pilot line can be connected to the pump outlet. In this case, the valve will open when the inlet pressure
© 2006 by Taylor & Francis Group, LLC
10-34
Handbook of Lubrication and Tribology
From pump
Optional remote pilot
Direct pilot To tank Counterbalance valve
FIGURE 10.42 A typical counterbalance valve hydraulic circuit.
Outlet
FIGURE 10.43
Inlet
Illustration of a simple check valve.
to the cylinder reaches some value. There will be very little rod side pressure in this arrangement. Reverse flow will not pass through the counterbalance valve as shown in Figure 10.42. Therefore, a bypass check valve must be included to permit the cylinder to be raised. 10.2.3.3 Check Valves Check valves are normally used to control the direction of fluid flow. However, their operation is similar to that of a direct operated relief valve. Figure 10.43 [19] shows a simple check valve and a cross-sectional illustration of the parts. The valve consists of a seat, a poppet, and a spring. The valve remains closed against flow until the pressure at its inlet creates sufficient force to overcome the spring force. Once the poppet leaves its seat, hydraulic fluid is permitted to flow around and through the poppet to the valve outlet port. For this reason a simple check valve can only allow flow in one direction. Like direct operated relief valves, simple check valves have a cracking pressure. By changing the spring, cracking pressures can be had between 5 and 75 psi. For special applications, a “no spring” version is also available. For load holding and in decompression type hydraulic press circuits, a pilot operated check valve is used. It performs the same function as the simple check valve described above. However, in contrast to
© 2006 by Taylor & Francis Group, LLC
Basic Hydraulic Pump and Circuit Design
10-35
Pilot section
Check valve X
B
A
FIGURE 10.44 Illustration of a pilot operated check valve. To actuator
A
B
Spool lands Spool
Under cut
T
P
To tank
From pump
Valve housing
FIGURE 10.45 Illustration of a spool type directional control valve.
the simple check valve, a pilot operated check valve can be piloted “open” when a reverse flow is required. Figure 10.44 [19] illustrates the components of a pilot operated check valve. The valve has two distinct sections, the check valve section and the pilot section. The check valve section allows free fluid flow from port “A” to port “B” while preventing reverse flow from “B” to “A” without leakage. However, if a pilot pressure signal is supplied to port “X,” then a force is applied to the pilot piston, which forces the piston rod against the check valve poppet. This force then unseats the poppet allowing free flow of fluid from port “B” to port “A.” 10.2.3.4 Directional Control Valves In typical hydraulic systems, there may be rotary or linear actuators present. These actuators normally have two ports. If oil is pumped into one of the ports while the other is connected to tank, the actuator will move in one direction. In order to reverse its direction of motion, the pump and tank connections must be reversed. The sliding spool type directional control valve has been found to be the best way to accomplish this change. Figure 10.45 [20] shows an illustration of a sliding spool type directional control valve. The valve has a cylindrical shaft called a “spool,” which slides in a machined bore in the valve housing. The housing has ports to which the hydraulic pump, return line to tank and lines for the actuator are connected. The number of ports designates the valve type. For example, a valve with 4 ports is referred to as a “4-way” valve, while a valve with 3 port connections would be called a “3-way” valve. Furthermore, spool valves can be classified as “2-position” or “3-position” valves.
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10-36
Handbook of Lubrication and Tribology
Typical flow paths available with 3-position valve spools
FIGURE 10.46
Closed center
Restricted open center
Open center
Regenerative end closed center
Tandem center
B Blocked P & A → T
Float center
P & B → T A Blocked
Regenerative center
P & B Blocked A → T
Restricted float center
P & A Blocked B → T
Typical flow paths available for 3-position spool valves.
A “2-position” valve can only be shifted fully left or fully right. A common use for such a valve would be in a cylinder application that only requires the cylinder to extend or retract to its fullest positions. Another application would be in hydraulic motors, which only run in forward or reverse directions. A “3-position” valve is similar in operation to a “2-position” valve except that it can be stopped in a third or “neutral” position between ports “A” and “B.” While in the centered or neutral position, flow may or may not be possible depending on the spool design of the center position. Figure 10.46 [20] shows some common “3-position” spool designs. Most valve manufacturers test their valves for flow capacity and develop charts that plot valve flow rate vs. pressure drop ( P). From these plots a flow factor denoted “CV ” can be determined for each valve. The “CV ” factor then can be used to calculate the flow characteristics of the valve at other conditions. For example, a valve with a CV = 1 will flow 1 gpm at a 1 psig pressure drop using a 1.0 specific gravity (SG) fluid. Figure 10.47 [18] lists the circuit definitions used in the following equations. The “CV ” factor is calculated from the following expression: √ Q SG CV = √ P
(10.38)
where CV is flow factor, Q is flow rate (gpm), SG is specific gravity of fluid, and P is pressure drop (psig). Parallel circuits Q = [CV 1 + CV 2 + · · · ] ×
© 2006 by Taylor & Francis Group, LLC
P SG
(10.39)
Basic Hydraulic Pump and Circuit Design
10-37
1 1
2
2
3
4
3 4 Parallel circuit
Px
P1
1 2 3 4
Series circuit
P2
P1
P3 P4 P5
Series / Parallel circuit
FIGURE 10.47
1
P2
2
P3
3
P4
4
P5
Partial parallel circuit
Circuit definitions for valve flow “CV ” factor calculations.
Series circuits Q=
P SG × [(1/CV 21 ) + (1/CV 22 ) + · · · ]
(10.40)
Series/parallel circuits √ Px − P 1 Qx = CV x × √ SG
(10.41)
Partial parallel circuits Q=
CV 2 ×
√
P1 − P2 + CV 3 × √ SG
√
P1 − P 3 + · · ·
(10.42)
10.2.3.5 Rotary and Linear Actuators Rotary motors and linear cylinders are used to convert the energy in the hydraulic circuit to either rotary torque and speed or linear force and velocity. Rotary actuators or motors can be gear, vane, or piston design and will operate very similar to a pump except that flow and pressure are inputs and torque and rotation are outputs. These are normally referred to as continuous rotation actuators. Another type of rotary actuator is the limited rotation design and is sometimes called a rotary cylinder. In this design, the output shaft is limited, usually to less than 360◦ of rotation. By far the most prevalent actuator found in hydraulic systems is the linear actuator or cylinder. Cylinders are either single- or double-acting. Hydraulic cylinders are normally constructed of a barrel, piston assembly, piston rod, end caps, ports, and seals as shown in Figure 10.48 [21]. The piston provides the effective area against which the fluid pressure is applied and supports the piston end of the rod. The opposite end of the rod is attached to the load. The cylinder bore, end caps, ports, and seals maintain a fluid tight chamber into which the fluid energy is connected. Whether the rod will extend or retract in a double-acting cylinder depends upon which port fluid is directed. In a single-acting cylinder there is
© 2006 by Taylor & Francis Group, LLC
10-38
Handbook of Lubrication and Tribology
Rod end cap
Barrel
Piston seals (Packings) Cap end
Piston rod
Rod eye Piston assembly Rod bearing and seal
FIGURE 10.48
Components of a typical double-acting cylinder.
TABLE 10.6
Cylinder Size, Load, and Pressure Data
Bore dia. (in.) ROD (2.5
)
Mode
Eff. area (in.2 )
1,000 (psi)
2,000 (psi)
3,000 (psi)
5,000 (psi)
4 5 6 6
PULL PULL PULL PUSH
7.66 14.7 23.4 28.3
7,700 15,000 23,000 28,000
15,000 29,000 47,000 56,000
23,000 44,000 70,000 85,000
38,000 74,000 120,000 140,000
Load (lb) at
only one port which when adequately pressurized will extend the rod. The single-acting cylinder depends upon external force such as weight and gravity to retract the rod. Hydraulic cylinders are normally sized to accommodate the load requirements (Table 10.6) [18]. For example, if the load requirements are such that the cylinder must move a load of 20,000 lb at a speed of 20 ft/min in the extended direction; this information will determine the size of the cylinder, the necessary fluid pressure, and the input flow rate, as was shown earlier in this chapter (Section 10.2.1.5). 10.2.3.6 Accumulators The purpose of an accumulator in a hydraulic system is to store or provide fluid at pressure to minimize short duration pressure spikes or to reach a short duration high-flow demand. Most accumulators used in hydraulic systems are the spring loaded or the gas charged type. The spring-loaded accumulator simply uses the spring force to load a piston. When the fluid pressure increases to a point above the preload force of the spring, fluid will enter the accumulator to be stored until the pressure reduces. The gas charged accumulator can be either a piston type or a bladder type as shown in Figure 10.49 and Figure 10.50, respectively [22]. In the gas charged accumulator, an inert gas such as dried nitrogen is used as a precharge medium. In operation, this type of accumulator contains the relatively incompressible hydraulic fluid and the more readily compressible gas. When the hydraulic pressure exceeds the precharge pressure exerted by the gas, the gas will compress allowing hydraulic fluid to enter the accumulator. 10.2.3.7 Components of a Hydraulic Circuit Diagram The proper planning of any hydraulic system should start with a properly drawn hydraulic circuit, using ISO-1219 approved graphic symbols. Figure 10.46 [20], Figure 10.51 [15], and Figure 10.52 [15] show the most common symbols found in circuit diagrams. An example of a simple hydraulic circuit, being used by hydraulic fluid manufacturers, is the ASTM D2882-83 “Standard Method for Indicating the Wear Characteristics of Petroleum and Nonpetroleum Hydraulic Fluids in a Constant Volume Vane Pump.” This pump test is currently the only one that has ASTM status. Figure 10.53 is an illustration of the hydraulic
© 2006 by Taylor & Francis Group, LLC
Basic Hydraulic Pump and Circuit Design
10-39
Gas valve
Gas
Piston
D Oil port
FIGURE 10.49 Illustration of a piston type gas charged accumulator.
Gas valve
Nitrogen
Bladder
Protection valve
Oil
FIGURE 10.50 Illustration of a bladder type gas charged accumulator.
circuit diagram for this test. The electric motor supplies mechanical energy to a Vickers V-104 vane pump, which in turn converts mechanical energy into hydraulic energy. The pump outlet pressure is monitored by a pressure gage. Relief valve is adjusted to induce a load pressure of 2000 psi as measured by pressure gage. At the outlet of the relief valve is at low pressure below 20 psi. The fluid then passes through filter, then through flow meter, which measures the flow rate to be ∼5 to 6 gpm (8 gpm at no load). The fluid then
© 2006 by Taylor & Francis Group, LLC
10-40
Handbook of Lubrication and Tribology Lines
Valves
Line, working (main) Line, pilot (for control)
Check On–Off (manual shut-off)
Line, liquid drain
Hydraulic pumps Fixed displacement
Variable displacement
Hydraulic flow, direction of pneumatic
Pressure relief
or
Lines crossing
Hydraulic motors Pressure reducing Lines joining
Fixed displacement
Line with fixed restriction
Flow control Adjustable — Noncompensated
Line, flexible
Flow control, adjustable (temperature and pressure compensated
Station, testing, measurement or power take-off Variable component (run arrow through symbol at 458)
Two position Two connection
Variable displacement
Hydraulic cylinders Single acting
Two position Three connection
Pressure compensated units (arrow parallel to short side of symbol)
Two position Four connection
Double acting Single rod Double rod
Three position Four connection
Temperature cause or effect
Valves, capable of infinite positioning (horizontal bars indicate infinite positioning ability)
Vented reservoir pressurized Line, to reservoir above fluid level Below fluid level Vented manifold
FIGURE 10.51
Basic hydraulic symbols-1.
passes through a heat exchanger and then into the reservoir. After the reservoir, the fluid passes through a 60-mesh filter (this filter is usually inside the reservoir at the outlet to the pump). A thermoregulator valve is used to maintain a constant reservoir temperature (set at 80◦ C for oil or 65◦ C for water-based fluids). Based on the known parameters of speed (1,200 rpm), pressure (2,000 psi), and flow (8 gpm, theoretical); other parameters such as power, torque, and heat can be readily calculated for this circuit as follows: GPM × PSI 8.0 × 2,000 = = 9.3 hp 1714 × Et 1714 × 1.0 HP × 63,025 9.3 × 63,025 = = = 490 in.-lbs RPM × Et 1,200 × 1.0
Powerinput = Torqueinput
BTU/hr = 1.5 × GPM × PSI = 1.5 × 8.0 × 2,000 = 24,000
© 2006 by Taylor & Francis Group, LLC
(10.43) (10.44) (10.45)
Basic Hydraulic Pump and Circuit Design Miscellaneous units Electric motor
10-41 Operation methods Spring Manual
Accumulator, spring loaded
Push button Push–Pull lever
Accumulator, gas charged
Heater
Pedal or treadle Mechanical Detent
Cooler
Temperature controller
Pressure compensated Solenoid, single winding Servo control
Filter, strainer
Pressure switch
Pilot pressure Remote supply Internal supply
Pressure indicator
Temperature indicator
Component enclosure Direction of shaft rotation (assume arrow on rear side of shaft)
FIGURE 10.52 Basic hydraulic symbols-2.
It should be noted that these calculations have assumed the pump to be 100% efficient (Et = 1.0). In the real world this is never the case. Typically, vane pumps have Et < 0.9. The effect of this would be to increase the power requirement of the pump to deliver the desired flow rate at a given load induced pressure. However, from actual experience with this ASTM test, the Vickers V-104 pump delivers only ∼5 to 6 gpm at 2000 psi. This is because the test requires running the pump at a 2000-psi load pressure, which is 1000 psi higher than the designed maximum pressure for this pump. Therefore, the volumetric efficiency Ev = ∼6/8 = ∼0.75, assuming the mechanical efficiency Em = 0.9, we have the total efficiency Et = 0.75 × 0.9 = 0.68 as a more reasonable value for the total efficiency of this pump at 2000 psi.
© 2006 by Taylor & Francis Group, LLC
10-42
Handbook of Lubrication and Tribology Heat exchanger
Flow meter 15 gal reservoir
Filter (25 mm)
Relief valve
Drain
Filter (60 mesh) Temperature
Thermoregulator
Pressure
Pump Drain
Drain
FIGURE 10.53 ASTM D2882 pump test hydraulic circuit diagram.
10.2.4 Basic Hydraulic System Design 10.2.4.1 Pipe and Hose Sizing While it is necessary to connect the various components in a hydraulic system with some kind of piping, such piping will produce flow resistance and therefore cause parasitic losses in the hydraulic system. To avoid as much loss as possible, the piping or hose must be sized properly. The internal diameter (I.D.) of the hose is extremely important because the fluid velocity at any given flow rate will depend upon that diameter. In fact, the fluid velocity will equal the flow rate divided by the internal area of the pipe as follows: V =
0.3208 × Q A
(10.46)
where V is velocity (ft/sec), Q is flow rate (gpm), and A is internal pipe area (in.2 ). The fluid velocities recommended for hydraulic systems are given in Table 10.7. Pressure drop calculations for piping or hose can be made using the equation presented earlier in this chapter(Equation [10.32]). While most texts refer to the flow regime present in the pipe when making such calculations, the Moody diagram will take the flow regime into consideration since it relies on the Reynolds number (Nr ), which depends upon fluid velocity, fluid viscosity, and the inside diameter of the pipe as shown by
© 2006 by Taylor & Francis Group, LLC
Basic Hydraulic Pump and Circuit Design TABLE 10.7
10-43
Recommended Hydraulic Circuit Flow Velocities Suction line
Viscosity (SUS)
Pressure line
Velocity (ft/sec)
700 465 230 140 140–700
Pressure (psi)
2.0 2.5 4.0 4.3 4.3–2.0
Velocity (ft/sec)
Return line Velocity (ft/sec)
8.2–10.0 11.5–13.0 14.5–16.5 16.5–20.0 20.0
5.5–15.0 5.5–15.0 5.5–15.0 5.5–15.0 5.5–15.0
365 725 1450 2900 >2900
TABLE 10.8
Equivalent Length Values Equivalent length (Le /D)
Device Check valve 90◦ Standard elbow 45◦ Standard elbow Close return bend Standard tee-run Standard tee-branch
150 30 16 50 20 60
the following equation: Nr =
3162 × Q µd
(10.47)
where Nr is Reynolds number, µ is viscosity (cSt), and d is pipe I.D. (in.). Fittings and valves must be handled somewhat differently than straight runs of pipe. The easiest way to calculate the losses due to fittings and valves is to use the equivalent length method to estimate the effect by treating it as if it were an additional length of pipe. Table 10.8 lists some common devices and their equivalent length values, which are given as the length-to-diameter (Le /D) ratios so that they can be used directly in the modification of the Darcy equation as follows: hr = λ ×
Le v2 × D 2g
(10.48)
where hf is equivalent length, λ is friction factor, Le /D is equivalent length values, v is fluid velocity, and g is gravitational constant. The analytical methods presented here to calculate pressure losses in hydraulic piping and fittings are accurate but can be very time consuming. A method that is less accurate but provides a reasonable estimate of pressure losses in hydraulic systems involves the use of tables, which are available from pipe manufacturers and in various handbooks concerning fluid flow (Table 10.5). 10.2.4.2 Reservoir Design A typical design for an industrial reservoir is shown in Figure 10.54 [21]. Several features can be seen in this figure. The overall dimensions should enclose a sufficient volume of oil to permit air bubbles and foam to escape during the resident time of the fluid in the reservoir. The depth must be adequate to assure that during peak pump demands the oil level will not drop below the pump inlet level. The pump should be mounted below the reservoir so that a positive head pressure is available at all times. This is very critical when water based hydraulic fluids are used, since these fluids can have a higher specific gravity as well as a
© 2006 by Taylor & Francis Group, LLC
10-44
Handbook of Lubrication and Tribology Mounting plate Suction line Filler / Breather cap
Oil level gauge
Dished bottom
Cleanout cover
Return line Drain line
Drain plug
Baffle plate
FIGURE 10.54 A typical design for an industrial reservoir.
much higher vapor pressure than mineral oil-based fluids (Section 10.2.2.7). The reservoir should be sized to afford adequate fluid cooling. Baffles are provided to prevent channeling of the fluid from the return line to the inlet line. The bottom of the return line is usually cut at a 45◦ angle to assist in redirection of the fluid away from the inlet. A clean-out plate is provide to promote cleaning and inspection. Sight gages are normally used to monitor the fluid level. A breather system with a filter is provided to admit clean air and to maintain atmospheric pressure as fluid is pumped into and out of the reservoir. With water-based hydraulic fluids a pressurized reservoir is recommended. Special breather caps can be purchased to vent between 1 and 15 psig. If one of these is used make sure that it has a vacuum brake to vent at ∼−0.5 psig (Note: Not all pressure caps have a vacuum brake). This is important so that when the reservoir is cooling down no appreciable vacuum develops in the reservoir tank. This feature will minimize pump cavitation upon start-up and also prevent a possible tank implosion. 10.2.4.3 Natural Frequency and Time Response When designing any hydraulic system, especially when heavy masses are moved quickly, there is one very important design factor that needs to be considered. That factor is known as the “natural frequency” (ωo ) of the system. Knowledge of this frequency is important because it determines how fast one can accelerate a given load and thus its maximum achievable velocity. From the physical laws of motion, the natural frequency of a hydraulic system can be found by taking the square root of the effective spring constant divided by the effective moving mass. ωo =
C M
(10.49)
where ωo is natural frequency, C is effective spring constant, and M is effective moving mass. This is a simple statement; however, determination of the effective spring constant and effective moving mass is not so simple. The effective spring constant not only includes the compressibility of the trapped hydraulic fluid between the valves and the actuators, but also the movement of any hoses or piping as well as structural vibrations. The effective mass of the system is the combination of all the moving loads, including the mass of the trapped fluid between the valves and actuators as illustrated in Figure 10.55 [23].
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Basic Hydraulic Pump and Circuit Design
10-45
M1
M2 Mechanical component
Hydraulic spring-mass system
Valve
FIGURE 10.55 Illustration of the hydraulic spring-mass system.
Ae Ab V1
M V2
FIGURE 10.56 Illustration of the natural frequency parameters.
In the simplified case of a linear cylinder in a closed circuit (Figure 10.56) [23], the natural frequency can be calculated using the following expression: ωo =
Ab2 × β A2 × β + e V1 × M V2 × M
or
Fo =
ωo 2π
(10.50)
where ωo is natural frequency (rad/sec), Ab is cylinder blind end area (in.2 ), Ae is cylinder extending end area (in.2 ), β is bulk modulus of fluid, V1 is cylinder blind end volume (in.3 ), V2 is cylinder extending end volume (in.3 ), M is effective moving mass (lbs-sec2 /ft, slugs), and Fo is natural frequency (Hz). There are computer programs available that can be used to determine the frequency response of a hydraulic system by using the impulse method. The time response of a hydraulic system is the synergistic result of the response times of all of the components used in the system [24]. Therefore, most component manufacturers will provide information relative to the responsiveness of their components. Unfortunately, the information derived from the component manufacturers is not consistent. The ability to understand and utilize the response information obtained from component manufacturers using a second-order system depends upon the definition of several aspects of the response subject as follows: • Delay time: the time required for the output to reach 50% of the steady output. • Rise time: the time required for the output to rise from 10% to 90% of the final output value. • Maximum overshoot: the time at which the maximum overshoot occurs. • Settling time: the time for the system to reach and stay within a stated plus-and-minus tolerance band around the steady-state output.
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10-46
Handbook of Lubrication and Tribology Maximum overshoot time
c(t) 1.5
Unit step input
Tolerance band 6d
1+d 1.0 1–d 0.90
Steady-state error (t → `)
0.5 Delay time
0.1 0.0 Rise time Settling time
FIGURE 10.57
Time
Step response of a second-order system.
S
(S – d )
d
Ae Ab L 1 V1
FIGURE 10.58
M V2 L 2
Illustration of a typical natural frequency calculation.
A graph, which illustrates these parameters, is shown in Figure 10.57. Control technology can be used to evaluate the response of a complete hydraulic system if all of the component information is given in consistent and correct terms. 10.2.4.3.1 Calculation of Natural Frequency, Acceleration, Maximum Velocity, Acceleration Pressure, and Flow Rate For economic reasons, it is often desirable to operate a hydraulic system as fast as possible. This is especially true on automated assembly lines where hydraulics is used to move parts. As an example of a simple calculation, consider the following application where one needs to determine the maximum speed and shortest cycle time to perform a repetitive task. By way of a single-rod hydraulic cylinder (1.5
bore, 1
rod), a proportional directional control valve is used to accelerate a 1000 lb load (M ) to a constant velocity over a distance of 30 in. in 1 sec and then decelerate the load to a stop. The load is then retracted in the same manner to start the cycle over again (Figure 10.58) [23]. To solve this problem, the natural frequency must first be calculated so that the time to accelerate the load can be determined.
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Basic Hydraulic Pump and Circuit Design
10-47
Then the maximum velocity, acceleration pressures, and required flow rates can be calculated for both the extending and retracting modes. The following information is given: w = 1000 lbs. (Load) Ts = 1.0 sec (Stroke time) Ab = 1.76 in.2 (1.5
Cylinder bore) Ae = 0.98 in.2 (1.5
bore area − 1
rod area) S = 30 in. (Stroke distance) β = 200,000 lb/in.2 (Bulk Modulus of oil) L1 = 46.50 in. (Cylinder blind-end pipe length) L2 = 38.75 in. (Cylinder rod-end pipe length) D = 0.62 in. (Pipe I.D.)
Pipe Size = 3/4
O.D. × 0.065
wall The first step is to calculate the pipe-trapped volumes between the control valve and the cylinder blind-end inlet (V3 ) and the rod-end inlet (V4 ) in Figure 10.58 [23] from the following equation: π D2 × L1 4 π(0.62)2 = × 46.5 4
V3 =
= 14.04 in.3
(10.51)
V4 = 0.30 × 38.75 = 11.7 in.3 Next, calculate the dimension “d,” using the above values along with the given parameters, using the following expression: [(Ae × S + V4 )/ Ae3 ] − (V3 / Ab3 ) d= √ √ (1/ Ae ) + (1/ Ab ) √ √ {[(0.98 × 30) + 11.7] / 0.983 } − (14.04/ 1.763 ) = √ √ (1/ 0.98) + (1/ 1.76) = 20.6 in.
(10.52)
Next, calculate the total trapped volume between the valve and cylinder blind-end (V1 ) and cylinder rod-end (V2 ), using the following relations: V1 = V3 + (Ab × d) = 14.04 + 1.76 × 20.6 = 50.3 in.3 V2 = V4 + Ae (S − d) = 11.7 + 0.98(30 − 20.6) = 20.9 in.3
© 2006 by Taylor & Francis Group, LLC
(10.53)
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Handbook of Lubrication and Tribology
Convert the load to units of mass as follows: M=
1000 w = = 2.59 (lbs-sec2 /in.) g 386
(10.54)
where M is effective moving mass (lbs-sec2 /in., slugs), w is load force (lbs), and g is gravitational constant (in./sec2 ). Then substitute the known quantities into Equation (10.50) to obtain the natural frequency (ωo ) of this system as follows: ωo =
1.762 × 200,000 0.982 × 200,000 + 50.3 × 2.59 20.9 × 2.59
= 91.1 rad/sec
(10.55)
In calculating “ωo ” we have not taken into consideration other factors which contribute to the spring constant of the system, namely, hoses and other mechanical components. However, it has been shown over the years, that a good approximation to determine the useable acceleration is to divide the calculated natural frequency by three [23]. This simplification avoids a much more complex mathematical analysis, which would have required variables, which are difficult, if not impossible to define. Therefore, the useable frequency (ω) can be estimated as follows: ω=
ωo 91.1 = = 30.4 rad/sec 3 3 or
F=
(10.56)
30.4 ω = = 4.8 Hz 2π 2π
The acceleration time (T ) or the time for one complete oscillation can now be calculated: T =
1 1 = = 0.033 sec ω 30.4
(10.57)
However, it has been determined that this period is too short for acceleration to stabilize using proportional valves. Generally, for stable acceleration, the time allowed must be a minimum of four to six times the time period for one oscillation [23]. Therefore, the acceleration stabilizing time (Tb ) is calculated as: Tb = 6 × T = 6 × 0.033 = 0.20 sec
(10.58)
From the stroke distance (S), the acceleration time (Tb ) and the stroke time (Ts ); the maximum velocity (Vmax ), acceleration (Amax ), and the acceleration force (Fa ) can be easily calculated from the following expressions: S 30 = = 37.5 in./sec Ts − T b 1.0 − 0.2 Vmax 37.5 = 188 in./sec2 = Amax = Tb 0.20 w 1000 × 188 = 487 lbs Fa = MA max = Amax = g 386 Vmax =
© 2006 by Taylor & Francis Group, LLC
(10.59) (10.60) (10.61)
Basic Hydraulic Pump and Circuit Design
10-49
Before we can calculate the acceleration pressure at the blind-end (Pb ) and rod-end (Pr ) of the cylinder, the frictional force that the load imposes on the system needs to be determined. For this calculation it is assumed that the coefficient of friction (µ) equals 0.58, we then can determine the force due to friction (Fµ ) and the total force (Ft ) as follows: Fµ = µw = 0.58 × 1000 = 580 lbs
(10.62)
Ft = Fµ + Fa = 580 + 487 = 1067 lbs
(10.63)
Ft 1067 = = 606 psi Ab 1.76 Ft 1067 = = 1089 psi Pr = Ar 0.98
(10.64)
Pb =
One should note that for a single-rod cylinder, the rod-end pressure is always greater than the blindend, but only with double-rod cylinders having equal rod diameters will the pressure be the same at both ends. Finally, the flow rate required at the blind-end (Qb ) and rod-end (Qr ) may be calculated as follows: 37.5 × 1.76 × 60 Vmax × Ab × 60 = = 17.1 gpm 231 231 Vmax × Ar × 60 37.5 × 0.98 × 60 = = 9.6 gpm Qr = 231 231
Qb =
(10.65)
10.2.5 Hydraulic Fluid Considerations 10.2.5.1 Foaming Most hydraulic fluids have an antifoaming agent as an additive. These additives have caused discussions among hydraulic system designers and users. Most of the additives used to control the foaming tendencies of hydraulic fluids accomplish this task by increasing the surface tension of the fluid. When the surface tension increases the size of air or vapor bubbles, which will coexist in the fluid, become smaller and therefore are less likely to rise to the surface and cause a foaming situation. However, when the air is allowed to remain in the fluid, the compressibility of the fluid increases, or stated in another way the bulk modulus of the fluid decreases. The suspension of air or vapor in the circulating fluid of a hydraulic system is a fault of the system. That is, a well-designed system will not permit air or vapor to become entrained in the fluid. Some expert designers of hydraulic systems have said that they would rather not have an antifoaming agent present. Without the addition of the antifoaming agent a system, which is poorly designed, will be readily apparent and can be fixed. Details on foaming, air entrainment, and air release are provided in Chapter 2. 10.2.5.2 Bulk Modulus The bulk modulus of a fluid is a term used to describe the compressibility of the fluid. In fact, the bulk modulus is inversely proportional to the compressibility. The purpose of a hydraulic system is to raise the potential energy of the system by increasing the pressure of the fluid. This potential energy can then be converted into kinetic energy, which will due useful work. However, a fluid with a low bulk modulus will be very compressible and the energy necessary to raise the pressure must also be sufficient to compress the fluid. Most hydraulic fluids have a very high bulk modulus in the pristine condition. However, when air is present, the effective bulk modulus will be low and the system fluid will need to absorb the heat generated when the compression takes place. Calculation procedures for bulk modulus and fluid compressibility are described in more detail in Chapter 2.
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Handbook of Lubrication and Tribology
S
DS DL
V1
A1
To tank
VP
From pump
FIGURE 10.59
Illustration of a cylinder meter-in circuit.
10.2.5.2.1 Fluid Compressibility and Cylinder Lunge Fluid compressibility has a great effect on cylinder performance. Especially when the fluid type is changed, such as changing from a mineral oil to a water-based or synthetic fluid. Hydraulic cylinders are especially sensitive to changes in bulk modulus. In critical operations it is often necessary to extend the cylinder smoothly and at a very constant velocity. If the load changes, the compressibility of the hydraulic fluid will have a negative influence on the constant velocity. Also, any change in the volume ( V ) of the fluid under compression will translate into a change in cylinder stroke ( S) defined as “lunge.” The following expressions can be used to calculate “lunge” ( S) and the resultant velocity change ( v): S =
[Vp + (A × S )] × L A2 β S × 60 v = τ
(10.66) (10.67)
where S is lunge (in.), Vp is volume in pipe (in.3 ), A is effective piston area (in.2 ), S is stroke (in.), L is load change (lbs.), β is bulk modulus of fluid, v is velocity change (in./min), τ is load change time (sec). We will now apply these equations to the meter-in (Figure 10.59) [25] and the meter-out (Figure 10.60) [25] circuits under the following conditions: A1 = 4.9 in.2 (Blind-end area) A2 = 2.5 in.2 (Rod-end area) S = 24 in. (Stroke) L1 = 3000 lbs (Full load) L2 = 1000 lbs (Reduced load) L = 2000 lbs (Load change, L1 − L2 ) τ = 1 sec (Load change time) Vp = 36 in.3 (Oil line volume) β = 200,000 lb/in.2 (Bulk modulus of oil)
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Basic Hydraulic Pump and Circuit Design
10-51
DS
S
V2 DL
A2 From pump VP
To tank
FIGURE 10.60
Illustration of a cylinder meter-out circuit.
For the meter-in mode (Figure 10.59) using Equations (10.66) and (10.67) we calculate: [36 + (4.9 × 24)] × 2000 = 0.064
(4.9)2 (200,000) 0.064 × 60 = 3.8 in./ min v = 1
S =
(10.68) (10.69)
For the meter-out mode (Figure 10.60) using Equations (10.66) and (10.67) we calculate: S =
[36 + (2.5 × 24)] × 2000 = 0.154
(2.5)2 × 200,000 0.154 × 60 = 9.2 in./ min v = 1
(10.70) (10.71)
As you can see from the examples above, the degree of “lunge” is directly proportional to the load change and inversely proportional to the bulk modulus of the fluid. In addition, cylinder lunge is greater in the meter-out mode than in the meter-in mode. This is due to pressure intensification in the rod-end of the cylinder as discussed earlier in this chapter (Section 2.1.5).
References [1] Frankenfield, T.C., Using industrial hydraulics, Chapter 1, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [2] Blackburn, J.F., Reethof, G., and Shearer, J.L., Fluid Power Control, The M.I.T. Press, Cambridge, MA, 1960. [3] Fitch, E.C. and Hong, I.T., Hydraulic Component Design and Selection, BarDyne Inc., Stillwater, OK, 1997.
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[4] Frankenfield, T.C., Using industrial hydraulics, Chapter 6, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [5] Sullivan, J.A., Fluid Power Theory and Applications, 2nd ed., Reston Publishing Company, Reston, Virginia, 1982. [6] Mannesmann Rexroth, Pump and Controls — Open Loop, The Rexroth Training Center, Bethlehem Vocational Technical School, Bethlehem, PA, June 1995. [7] Fitch, E.C., Fluid Contamination Control, FES Inc., Stillwater, OK, 1988. [8] Tessmann, R.K. and Howsden, J.M., Environmental influence upon wiper seal performance, BFPR Journal, P75-5, FPRC/OSU, Stillwater, OK, October 1975. [9] Tessmann, R.K. and Howsden, J.M., Service life of wiper seals, BFPR Journal, P76-31, FPRC/OSU, Stillwater, OK, October 1976. [10] Fitch, E.C. and Tessmann, R.K., Modeling the performance of filter assemblies, BFPR Journal, P73-CC-12, FPRC/OSU, Stillwater, OK, October 1973. [11] Fitch, E.C. and Tessmann, R.K., The filter selection graph — a basic contamination control tool, BFPR Journal, P74-55, FPRC/OSU, Stillwater, OK, October 1974. [12] Wolf, M.L., Contaminant Particle Effects on Pumps as a Function of Size, Type and Concentration, M.S. Thesis, FPRC/OSU, Stillwater, OK, 1965. [13] Tessmann, R.K., Contaminant wear in hydraulic and lubricating systems, BFPR Journal, P75-4, FPRC/OSU, Stillwater, OK, October 1975. [14] Bensch, L.E., Verification of the pump contaminant wear theory, BFPR Journal, 11, FPRC/OSU, Stillwater, OK, October 1977. [15] Hydraulic Hints and Troubleshooting Guide, No. 694, Vickers, Incorporated, Troy MI, August, 1996. [16] Fitch, E.C., Fluid Power Engineering, FES Inc., Stillwater, OK, 1982. [17] Mackay, R.C., Pump suction conditions, Pumps and Systems Magazine, 20, May 1993. [18] Paul-Munroe, Lightning Reference Handbook, 8th ed., Rucker, Inc., 1994. [19] Frankenfield, T.C., Using industrial hydraulics, Chapter 4, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [20] Frankenfield, T.C., Using industrial hydraulics, Chapter 5, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [21] Norvelle, F.D., Fluid Power Technology, West Publishing Company, New York, NY, 1995. [22] Frankenfield, T.C., Using industrial hydraulics, Chapter 10, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [23] Frankenfield, T.C., Using industrial hydraulics, Chapter 9, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [24] Hong, I.T. and Tessmann, R.K., What time do you have? Proceedings of the National Conference on Fluid Power, Vol. II, National Fluid Power Association, pp. 23–25, 1996. [25] Frankenfield, T.C., Using industrial hydraulics, Chapter 3, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [26] Wilson, W.E., Rotary-Pump Theory, Transactions of the A.S.M.E., 68, 371–384, May 1946. [27] Wilson, W.E., Clearance Design in Positive-Displacement Pumps, Machine Design, 127–130, February 1953. [28] Handbook of Chemistry and Physics, The Chemical Rubber Co., 49th ed., p. D109.
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11 Hydraulic Fluids 11.1 Functions of Hydraulic Fluids . . . . . . . . . . . . . . . . . . . . . . . . 11.2 Types of Hydraulic Fluid . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
11-1 11-2
Fluids in General • Fire-Resistant Fluids • Environmentally Compatible Fluids • Special Fluids
11.3 Properties of Hydraulic Fluids . . . . . . . . . . . . . . . . . . . . . . . .
11-7
Viscosity • Density • Compression Modulus • Load-Bearing Capacity and Antiwear Properties • Setting Point/Pour-Point • Gas Solubility • Air Release Capacity and Foaming Tendency • Aging Behavior • Material Behavior • Detergent and Dispersant • Thermal Capacity and Thermal Conduction • Flammability • Biodegradability and Toxicity
H. Murrenhoff and O.-C. Göhler Institute of Fluidpower Drives and Controls (IFAS) RWTH Aachen University
T. Meindorf Argo-Hytos GmbH
11.4 Fortifying Hydraulic Fluids . . . . . . . . . . . . . . . . . . . . . . . . . . .
11-18
Oxidation Inhibitors • Metal Deactivators • Wear Protection • Friction-Reducing Agents, Friction-Modifiers • Viscosity Index Improvers • Setting Point Depressants • Antifoaming Agents • Detergents and Dispersants • Corrosion Inhibitors
11.5 Impurities in Hydraulic Fluids . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
11-22 11-23
Hydraulic fluid is a complex component which is part of the hydraulic system. Apart from its primary function of transferring energy, the medium also has to perform a number of other functions and is predominantly used as a lubricant. The characteristic profile required for this can no longer be fulfilled by a base fluid based on mineral oil, as it used to be in the early days of oil hydraulics. The technical performance capability is now supplemented or extended by special chemical additives. Other base fluids can be used to meet additional requirements where a fluid is required to be fire resistant, environmentally compatible, or capable of withstanding extreme loads, for example.
11.1 Functions of Hydraulic Fluids The primary function of a hydraulic fluid is to transfer energy from a pressure generator to the consumer. A volume connection is established for this, which is analogous to the form closure or frictional connection in a mechanical transmission concept. After meeting this functional requirement, the hydraulic medium has to perform other functions. It acts as a lubricant, thereby reducing the friction and wear of parts moving against one another in 11-1
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11-2
Handbook of Lubrication and Tribology
Power transmission Pressure transmission
Volume connection Functions of hydraulic fluids
Heat removal
Corrosion protection
FIGURE 11.1
Wear reduction
Elastomer compatibility
Functions performed by hydraulic fluids.
bearing arrangements, for example. The components of the hydraulic system must be protected against corrosion and other chemical reactions. The losses attributable to friction and throttling are discharged by the hydraulic medium in the form of thermal energy. Figure 11.1 shows an overview of the functions performed by the fluid within the hydraulic system. These functional properties of the fluid must be assured over a wide temperature range. The characteristics of the medium itself must not be impaired by impurities, such as rubbed-off particles or water. After all, there are both economic and ecological arguments in favor of a long lifetime for the medium with its properties remaining unchanged. Safety aspects give rise to additional requirements in situations where the fluid must not be inflammable or must have a reduced evaporation tendency for reasons related to working safety. Increasing importance has been attached to ecological aspects for some years now. As far as mobile applications in particular are concerned, there are growing demands for fluids which, unlike conventional media, decompose quickly and completely in the event of leakage and are less harmful in toxicological terms. The requirements to be met by hydraulic fluids are summarized in Figure 11.2.
11.2 Types of Hydraulic Fluid Hydraulic media generally comprise a base fluid which is doped to produce a ready-to-use formulation with the required properties by adding other substances, which are referred to as additives. The proportion of these additional substances ranges from just a few percent in lightly doped mineral oils right through to 50% in a few fire resistant fluids. The nature of the base fluid, which in itself may be a mixture of heterogeneous molecules of the same type, for example ester molecules with a varying degree of saturation of the fatty acid, essentially determines the application for which the hydraulic fluid is used.
11.2.1 Fluids in General Accounting for approximately 80 to 85% of hydraulic fluids, those based on mineral oil constitute the most significant group of hydraulic media in terms of quantity and cost effectiveness. They are capable of meeting a broad spectrum of requirements and therefore find universal application in stationary and mobile systems. Refined by many years of experience with these media — accompanied by as many years spent in their development — the properties of these media are clearly defined and can be reproduced at any time.
© 2006 by Taylor & Francis Group, LLC
Hydraulic Fluids
11-3
High EP characteristics
Low compressibility
High aging stability Demands on hydraulic fluids
High heat capacity and conductivity
Non corrosive
FIGURE 11.2
High biodegradability
Low flammability
Requirements to be met by hydraulic fluids.
Table 11.1 shows the mineral oil sales figures for Germany for 1997–2003, which have been published by the Association of the German Oil Industry, Mineralöl Wirtschafts Verband (MWV). The base fluids are produced by distilling and refining crude oil and usually comprise a mixture of C20 to C35 molecules, consisting of normal paraffin, isoparaffin, naphthene, and isolated aromatic compounds. The essential properties of the fluid, such as viscosity, flash point, aging stability, viscosity/temperature behavior, and response to low temperatures are determined by varying the stages in the process [1]. Special properties not demonstrated by the base oil, or not to an adequate extent, are achieved by means of additives. The types of doped additives are distinguished according to the classification of the fluid. This means, for example, that hydraulic fluids which have no active ingredients and are hardly used nowadays in practical terms belong to Group H. Group HL contains hydraulic fluids with active ingredients which improve aging stability and corrosion protection and Group HLP contains hydraulic fluids with additional active ingredients which reduce wear and increase the capability of withstanding load. Hydraulic fluids containing additional active ingredients which increase the viscosity/temperature index (VI) to a value in excess of 140 are found in Group HVLP. This elevation can also be achieved in the base oil by means of a special refining process known as hydro-cracking. This classification is specified by DIN 51 524 “Pressure fluids — Hydraulic oils; Minimum requirements,” which also stipulates the minimum requirements to be met with respect to viscosity characteristics and aging stability, for example, as well as corrosion protection and antiwear properties. Additives with detergent and dispersant effects are added to the fluids for special applications. These types of fluid are identified by the letter D in the nomenclature. The additives in an HLP-D fluid, for instance, render it capable of binding up to 5% of water in the form of an emulsion. The water is enclosed in the form of tiny droplets in such a way as to avoid obstructing the formation of a lubricating film, with the result that no corrosion occurs on the metallic surfaces and there is no intensified aging of the medium. These fluids are given precedence in maritime applications, where there is a particularly high risk of the medium being contaminated by water. Table 11.2 shows an overview of the currently applicable mineral oil standards.
11.2.2 Fire-Resistant Fluids Fluids which are classified as being fire-resistant have been developed for use in the mining, aircraft construction, die-casting, and rolling mill industries. These fluids have a much higher inflammation temperature than mineral oils and their use is mandatory for certain applications. This means, for example, that the fire prevention regulations for underground mining forbid the use of mineral oil. A distinction
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Handbook of Lubrication and Tribology
TABLE 11.1
German Lubricant Sales Figures Classified According to Grades, 1997–2003
Groups of substances
1997
1998
1999
2000
2001
2002
Engine oil Compressor oil Turbine oil
411.5 9.3 4.5
381.6 9.9 5.2
384.9 11.1 5.2
369.9 10.9 3.6
343.7 10.5 2.7
348.7 12.0 2.6
343.9 11.6 2.3
Transmission fluid Automotive Industrial Hydraulic fluid
66.9 24.1 153.9
64.0 26.7 158.2
66.9 28.9 150.3
66.0 28.0 154.6
68.1 24.8 146.0
70.3 22.1 143.9
68.0 21.1 135.5
46.2 31.4 3.0 9.9
48.2 30.7
45.4 29.8
45.6 32.8
47.2 29.7
50.4 28.1
10.3
9.5
9.2
7.2
7.6
49.6 27.9 2.0 7.9
42.4 9.0 99.3 18.8 50.5 59.6
38.6 9.4 80.3 19.2 34.4 59.9
49.3 10.0 85.4 15.1 35.7 54.9
44.4 10.9 83.3 10.9 30.8 53.4
38.6 16.8 84.8 12.6 30.3 55.0
34.8 17.9 85.8 13.7 32.2 51.7
45.4 11.2 108.9 9.1 31.4 48.0
32.4 45.9 49.4
34.4 44.4 63.5
35.7 45.2 83.0
30.8 45.9 72.0
30.3 43.7 55.2
32.2 44.2 79.0
35.9 33.8 73.1
1168.0 75.0
1146.8 73.0
1159.9 51.8
1122.3 77.1
1057.7 96.1
1076.6 110.2
1066.8 101.6
Metal working oil Straight Water-miscible Hardening oil Anticorrosive agents Paraffin oil Medical Technical Other process oils Electrical insulating oil Machine oil Other industrial oils/fluids not used for lubrication Lubricating grease Extracts from lubricating oil refinery Base oil Total Including quantities of recycled waste oil (already included in the respective groups)
2003
Source: Federal Office for Trade and Industry.
TABLE 11.2
Mineral Oil Standards
DIN 51 524
ISO 6743-4
H
HH
Without special additives (base oil)
HL
HL
HLP
HM
HVLP
HV
HLPD
(-)
With active ingredients to increase corrosion protection and aging stability. DIN 51 524, Part 1 As for HL but with further additives to reduce fretting wear with mixed friction. DIN 51 524, Part 2 As for HLP but with further additives for improve the viscosity/temperature behavior. DIN 51 524, Part 3 As for HLP but with further additives to dissolve deposits (detergent) and waterbearing to a certain extent (emulsifying/ dispersing)
Composition
Fields of application Systems with no special requirements (infrequent) Systems with moderate pressures, but high temperatures. Satisfactory separating capacity Systems with high pressures and temperatures. High-quality, widely used hydraulic fluid, particularly HLP 46 Wider temperature range than HLP with low initial values as a consequence of shallow viscosity characteristic Systems where water may ingress to fluid fill (condensation, cooling lubricants for machine tools, mobile systems)
is made between aqueous and nonaqueous hydraulic fluids. VDMA Guidelines 24 317 and 24 320 classify these fluids as follows (the designations are in accordance with CETOP RP 77 H, ISO 6743, Part 4, and DIN 51 502). The HFA group contains oil in water emulsions or highly aqueous solutions, the HFB group contains water-in-oil emulsions, the HFC group contains aqueous solutions, and the HFD group contains nonaqueous fluids.
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Hydraulic Fluids
11-5
The concentrate component of HFA fluids usually accounts for between 1% and a maximum of 5%, the rest is water. The permitted 20% limit is not exploited, for reasons related to costs. The concentrate component may be either an emulsifiable substance or a medium that dissolves in water. The concentrate contains additives which improve anticorrosion and antiwear properties or increase lubricity. It also contains biocides which prevent the formation of bacteria, fungi, and yeast in the medium. This type of fluid is primarily used to support mine workings hydraulically. Other applications include hot-working and manufacturing facilities in the automotive industry. HFB fluids comprise 40% water and 60% mineral oil. Unlike the situation in British collieries, this type of fluid is not used in Germany as it does not pass one of the fire tests prescribed by the German mining authorities. HFC fluids comprise 35 to 55% water, and viscosity-increasing substances and other additives make up the remaining 45 to 65%. In most cases, polyakylene glycols are used as thickening agents in order to increase the viscosity to a level that is similar to that of mineral oil. Anticorrosives and antiwear substances are used as the additives. This type of fluid is primarily used in high-performance underground mining systems and in hot-working plants. HFD fluids are nonaqueous, synthetic media. Phosphate esters and chlorinated hydrocarbons, the chloroaromatic compounds, are most frequently used. Diesters and silicone oils, polyphenylene ethers, polyglycols, silicate esters, and fluorocarbons are also used for special applications. These fluids are identified more precisely by an extra letter added to the “HFD.” Because of their environmental impact, fluids containing chlorinated hydrocarbons (PCBs) are used only under exceptional circumstances. The applications for this type of fluid include hot-working plants, such as die-casting machines, as well as mining. An overview of the currently applicable standards for fire resistant hydraulic fluids is given in Table 11.3.
11.2.3 Environmentally Compatible Fluids Efforts to substitute mineral oil products with renewable, plant-based raw materials were already being made — predominantly in Finland — during the first oil crisis at the beginning of the 1970s. However, the development of ecologically friendly hydraulic fluids is primarily oriented to a high level of environmental compatibility, and these fluids are being used to an increasing extent in mobile systems as well as stationary installations. In Germany, VDMA Guideline 24 568 specifies the minimum requirements for four classes of fluid: HETG (hydraulic oil environmental native triglycerid), HEES (hydraulic oil environmental synthetic ester), HEPG (hydraulic oil environmental polyglycol), and HEPR (hydraulic oil environmental polyalphaolefins and related products). The oils obtained from rapeseed and cognate plants possess very good technical properties, have a high VI, and, as such, constitute an ideal basis for HETG “native” hydraulic fluids. Like all vegetable oils these contain an ester molecule, which comprises a trivalent glycerine as the alcohol component and three
TABLE 11.3
Standards for Fire Resistant Fluids
ISO 6743/CETOP Lux. Ber./VDMA HFA HFB HFC
Composition Oil-in-water emulsion or synth. aqueous solution containing max. 20% concentrate Water-in-oil emulsion containing max. 60% oil Aqueous polymer solution containing 35 to 55% water
HFDU
Carboxylic ester (nonaqueous, synthetic)
HFDR
Phosphate ester (nonaqueous, synthetic)
© 2006 by Taylor & Francis Group, LLC
Fields of application Mining, hydraulic presses, temperature range: 5 to 55◦ C Mining, temperature range: 5 to 60◦ C Mining, foundries, moderate pressure, environmental protection, temperature range: −20 to 60◦ C Temperature range: −35 to 100◦ C, more widely used than HFDR Motor vehicles, aerospace industry, temperature range: −20 to 150◦ C
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Handbook of Lubrication and Tribology
fatty acids, such as oleic acid, linoleic acid, or erucic acid. The typical product for a plant is determined by the proportionate quantities of the fatty acids. The term “esterification” is used to refer to the reversible process which takes place during the reaction of alcohols and acids, as given in a simplified manner in the following equation: Alcohol + Acid Ester + H2 O
(11.1)
Esters are polar in character. The resulting adhesion to metal surfaces is accompanied by excellent antiwear properties and the surface wetting process also provides satisfactory protection against corrosion [1–5]. The ester compounds and the characteristic double carbon bonds of the fatty acids are changed relatively easily by means of oxidation (adding oxygen), hydration (adding hydrogen), or hydrolysis (breaking down into alcohol and free fatty acids). As a result, the native oils demonstrate low aging stability and give rise to aggressive cleavage and reaction products under the very conditions, with the associated high temperatures, which are usually encountered in service. On the other hand, these properties also enable rapid and virtually complete biodegradation of any fluid leaking into the ground, bodies of water, or sewage plants. This means that native oils must be fortified by suitable substances to prevent the fluid aging while in service. Of course, the additives used for this must not hinder the natural biological decomposition and must be nontoxic in themselves [1,4]. Synthesis can be used to produce esters which are similar to vegetable oils in terms of chemical structure and ecological characteristics but demonstrate much better aging stability. In this respect, there are diverse possible combinations of alcohols and acids which offer a means of selectively formulating products which meet the requirements for HEES hydraulic media [4]. Natural or petrochemical products are chosen as the starting materials and these are combined in different ways in the subsequent stages of the process. However, molecular stabilization is generally accompanied by a reduction in the rapid biodegradation capacity. Additives must be mixed into these fluids as well, and the additives are also subject to the ecological restrictions. A distinction is made between two main groups of synthetic esters which constitute suitable base fluids for hydraulic media. Natural fatty acids, and oleic acid (C 18 : 1) in particular, are used for unsaturated esters, such as TMP (trimethylolpropane) esters. Although their double bonds guarantee good lowtemperature properties, they detrimentally affect aging stability. The chemical structure is very similar to that of the native oils. High-quality, completely saturated esters without double bonds have predominantly petrochemical origins. Their structure differs from that of the native oils to a considerable extent. Rapidly degradable, nontoxic base fluids can be produced with properties that even surpass those of mineral oils. However, the manufacturing process is very complex and very expensive. Polyglycols have been used as high-performance synthetic lubricants for several decades now. Of this group, the polyalkylene glycols present significantly less environmental hazard potential than mineral oils and are therefore used as base fluids for the HEPG class. In the petrochemical synthesis process, the viscosity can be influenced by varying the relative molecular mass, which means that there are different viscosity classes available. The viscosity’s temperature dependence is less pronounced for ester fluids than it is for mineral oils with a VI of 185 to 215 without any fortification. The tribological properties are equivalent to or better than those of the mineral oils. Unlike oils, polyglycols are water soluble. This plays an important role in their biodegradation, which can take place in the solution rather than in the boundary surface, as is the case for oils. Rain water quickly carries any leakages into the deeper, lifeless earth layers where there is less oxygen so that the polyglycols can get into the groundwater without being decomposed beforehand. This is why they are no longer regarded as being environmentally compatible in Austria. Fluids which are predominantly synthesized from polyalphaolefins (PAOs) and cognate hydrocarbons are classified as belonging to a relatively new group with the designation HEPR. While PAOs are frequently used in the transmission fluid sector, the HEPRs are not yet playing a significant role in hydraulic systems.
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Hydraulic Fluids
11-7
Nor has any long-term experience been gathered with this type of oil. One of the unclarified aspects is their compatibility with the materials usually used in hydraulic systems. One factor that must be regarded in a critical manner is that the HEPRs may include multifarious formulations with virtually any combinations of PAOs and hydrocarbons, which makes it difficult to agree on a choice of materials. When environmentally compatible fluids are used, the fact is that they may react differently with plastics and certain metals than mineral oil. The materials used for seals and hoses, as well as sliding bearings and filter elements, for example, must therefore be matched to the fluid. If a polyglycol is used, it is usually necessary to remove the coating on the inside of the container as polyglycols dissolve paint and lacquer. The relatively high prices of synthetic esters, particularly those of saturated synthetic esters, are still preventing the HE media from being used in any significant quantities. A study of all lubricants and hydraulic media showed that bio fluids accounted for just 1.8% of the total lubricant market in Germany in 2003. Only around 36% of these belong to the group of biogenic fluids, that is, those based at least 50% on renewable raw materials. Although the remaining proportion of bio fluids are rapidly biodegradable and ecotoxicologically harmless, these are still based on beef suet and petrochemical products to the greatest extent. If the focus is on the hydraulic fluids alone, around 6.9% of the lubricants and hydraulic fluids can be described as bio oils. Biogenic oils account for 2.4% of the market as a whole. The bio oils (>90%) are almost exclusively found in the mobile hydraulic sector (approximately 60,000 Tonnes per annum). Only a very small proportion are used in stationary installations. In 2003, 17% of mobile hydraulic systems were using bio oils, while biogenic fluids accounted for 6%. Official stipulations and requirements, financial benefits resulting from the use of less water-polluting substances, and the users’ increasing environmental awareness are leading to noticeable expansion in the market.
11.2.4 Special Fluids Apart from the aforementioned media which can be used for more or less any purpose within their respective field of applications, there are many fluids that have been developed for specific, primarily mobile applications. The high temperatures prevailing in car and truck braking systems require highly stable base fluids, usually polyalkylene glycol. These fluids must be capable of absorbing a lot of water as undissolved water can boil prematurely in the brake line, drastically reducing the braking effect. The hydraulic systems, steering, transmission and wet brakes of a farming tractor are all supplied from a common fluid reservoir. The universal tractor oils used for this, with the designations UTTO and STOU, therefore unite the properties of a hydraulic medium with those of a transmission lubricant. The automatic transmission fluids (ATFs) used in torque converters and automatic transmissions fulfill similar requirements. They have been tuned to such an extent as to achieve constant frictional behavior as stick-slip effects are not wanted in multi-plate clutches. Flow improvers are added to the ATFs to ensure that they maintain their full functional capability during the cold winter months. Preference is given to the use of clarified water as a hydraulic medium where hygiene is important in the food industry. In these cases, the components have to perform virtually all of the tribological functions. Problems are also encountered as a result of the high cavitation tendency and the internal leakage caused by the low viscosity. The performance level has been improved here by the most recent developments in the material and manufacturing technology sectors.
11.3 Properties of Hydraulic Fluids This section describes the properties and parameters which are characteristic of a hydraulic fluid. The approximate values for various groups of fluids are given in the form of Table 11.5 at the end of the chapter.
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Handbook of Lubrication and Tribology
11.3.1 Viscosity The dynamic viscosity is the most important parameter in describing the load-bearing capacity of a hydraulic fluid. Conventional viscometers do not measure the dynamic viscosity directly, but rather its relationship to the density of the medium. This kinematic viscosity is defined as ν=
η ρ
(11.2)
The fluids are divided into certain classes according to their respective kinematic viscosity values. This classification is based on a mean viscosity at 40◦ C in accordance with DIN 51 519, which is the ISO viscosity classification for industrial liquid lubricants. This standard lays down a permanently specified series of numbers for the viscosity values. This series of numbers is as follows: 10, 15, 22, 32, 46, 68, 100 … mm2 /sec. The viscosity may deviate from this value by a certain amount, for example, by ±10 mm2 /sec for viscosity class ISO VG 100. As far as the unit of notation is concerned, the old unit centiStokes (cSt) is still used, as well as mm2 /sec. A fluid of viscosity group (VG) 46, for instance, is identified by adding the number 46 to its designation (e.g., HLP 46). The viscosity is highly dependent on the temperature for the majority of hydraulic fluids. The viscosity is plotted according to Ubbelohde in Figure 11.3. The double logarithm of the viscosity is plotted in mm2 /sec + 0.8 along the y-axis and the logarithm of the temperature is plotted along the x-axis. This type of graph gives rise to straight lines for mineral oil and HFC fluids over a relatively wide temperature range. The shallower the characteristic, the more favorable the cold-start behavior and operating viscosity. This applies to a particular extent in mobile applications which have to cover a broad temperature range between startup at temperatures below freezing and full-load operation during the hot summer months. The characteristic for HFD media rises very progressively in the lower temperature range, where its viscosity is more highly dependent on temperature than that of the other fluids. Shallow in other respects, the characteristic for HE media rises noticeably at low temperatures due to partial crystallization of the molecules. The ascending gradient of the fluid characteristic in the Ubbelohde diagram serves as a judgment scale for the viscosity/temperature (VT) behavior, which is referred to as the VI. A high VI indicates a low dependence on temperature. The VI can be increased by means of special additives, the VI improvers. 10000 [mm2/sec]
HLP HFC/HEES HFD HETG
Kinematic viscosity h
1000
100
10
3 253
FIGURE 11.3
273
313 293 Temperature q
Kinematic viscosity as a function of the temperature.
© 2006 by Taylor & Francis Group, LLC
333
353 [K]
373
Hydraulic Fluids
11-9 10000 [mPasec] 5000
HFD HLP
Dynamic viscosity h
2000 1000 HETG/ HEES
500 200 100
HFC
50 20 10 0
1000
2000
3000 [bar]
4000
Pressure p
FIGURE 11.4 Dynamic viscosity of various hydraulic media as a function of pressure with the same initial viscosity.
HLP has a viscosity index of around 100; HV has a VI of up to 400; the VI for HFC is around 150 and the VI for HFD is even less than zero. Hydraulic media based on vegetable oil have a very high inherent VI of around 200. The viscosity/pressure behavior of a medium is essentially responsible for the load-bearing capability of a liquid lubricating film. A common property shared by all media is that the dynamic viscosity increases when they are subjected to pressure. Figure 11.4 shows this behavior as demonstrated by different hydraulic media. The following equation applies to the pressure dependence of the dynamic viscosity at a constant temperature: η = η0 · eb·p
(11.3)
where η0 is dynamic viscosity at atmospheric pressure, b ≈ 1.7×10−3 bar−1 for mineral oil, b ≈ 3.5×10−4 bar−1 for an HFC medium, b ≈ 2.2 × 10−3 bar−1 for an HFD medium, and b ≈ 1.1 × 10−3 bar−1 for an HETG or HEES medium. Assuming an increase in pressure from 0 to 2000 bar, the viscosity of the HFC fluid increases by factor of 2, that of mineral oil by factor of 30, and that of the HFD fluid by factor of 80. The increase in viscosity as the pressure rises exerts a positive influence under high bearing loads as the lubricating film undergoes a self-reinforcing effect. The behavior shown in Figure 11.4 is one of the reasons why rolling bearings have a relative short service life when water-based fluids (HFA, HFC, clarified water) are used. The viscosity cannot be increased to any significant extent by means of additives.
11.3.2 Density The losses in pipelines and the flow channels of components are directly proportional to the density of the hydraulic medium. The density is also required to calculate the dynamic viscosity from the kinematic viscosity, as described above. The density is temperature-dependent as the volume of a fluid expands when the temperature increases. The coefficient of expansion (γ ) is defined as γ = where ϑ is temperature.
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1 ∂V V ∂ϑ
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Handbook of Lubrication and Tribology
This gives rise to an initial volume V0 with an increase in temperature by ϑ for a change in volume V : V = V0 γ ϑ
(11.4)
The density of the fluid decreases in proportion to the increase in volume. m V0
ρ0 =
m V0 + V
ρ= give rise to the following equation for the density: ρ=
ρ0 1 + γ ϑ
(11.5)
This equation relates to 15◦ C for the normal range of applications. Values for ρ15◦ C and γ can be found in the table at the end of this chapter. Figure 11.5 shows the density of various standard hydraulic fluids as a function of the temperature. The coefficient of expansion for mineral oil amounts to 7 × 10−4 K−1 . This means that the volume expands by 0.7% when the temperature increases by 10◦ C.
11.3.3 Compression Modulus The density of a hydraulic fluid is determined by the pressure. This compressibility is very important for the dynamic performance of hydraulic systems. The compressibility coefficient β is defined as 1 ∂V 1 β=− = (11.6) V0 ∂p ϑ EFl where EFl is the compression modulus.
1200 HFD
[kg/m3] 1100
Density r
HFC HFA
1000
HEES/HETG 900 HLP
800
FIGURE 11.5
0
20
40
60 80 Temperature q
Density of hydraulic fluids as a function of the temperature.
© 2006 by Taylor & Francis Group, LLC
100
120 [°C]
Hydraulic Fluids
11-11
The density of the fluid increases according to the reduction in volume when the pressure rises. ρ=
m V
V = −V0 β p gives rise to ρ=
ρ0 1 − β p
(11.7)
where ρ0 is the density at atmospheric pressure. This ratio applies to the usual pressure and temperature conditions in hydraulic systems. A mean constant compressibility coefficient β can be used for calculation here, even though β actually decreases as the pressure rises and increases with the temperature. Analogous to the modulus of elasticity for solids, the reciprocal value of β is referred to as the compression modulus EFl : EFl =
1 dp = −VFl,0 β dVFl
(11.8)
Accordingly, the compression modulus increases with the pressure and decreases with the temperature. Figure 11.6 shows the true adiabatic compression modulus of a mineral oil as a function of the pressure with the temperature plotted as a parameter. The term “true adiabatic compression modulus” is used to refer to the ascending gradient of the volume/pressure curve at the respective pressure values. Undissolved air components of 5 to 10% by volume are often found [9], particularly in mobile systems with short circulation times and unfavorable tank designs. As shown by the calculation below, these gas bubbles exert a very strong influence on the compressibility of the fluid and therefore on the stiffness of the system under load and its dynamic performance. The volume of a fluid/air mixture V0 = VFl,0 + VL,0 is subjected to a change in pressure dp, where VFl indicates the volume of the fluid and VL is the volume of dissolved air. The equivalent compression modulus EG of the mixture is given by EG = −
VFl + VL (dVFl /dp) + (dVL /dp)
(11.9)
True adiabatic compression modulus EFl [104 bar]
4 °C 10 30
3 50 70 90 110
2
1
0
700 Pressure p
[bar]
FIGURE 11.6 True adiabatic compression modulus of an HLP 46 mineral oil.
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Handbook of Lubrication and Tribology
The compression modulus of the fluid up to a pressure of 700 bar is approximately linearly dependent on the pressure (also refer to Table 11.3) EFl = E0 + m · p
(11.10)
Equation 11.8 and p0 = 0 bar give VFl = VFl,0 · e−(1/EFl )(p−p0 ) VFl,0 dVFl =− dp E0
1+
m·p E0
(m+1)/m
(11.11)
The air is assumed to be an ideal gas, so that the following is given for polytropic changes in state: n p · VLn = const = p0 · VL,0
VL,0 dVL =− dp n · p0
p0 p
(n+1)/n
(11.12)
The equivalent compression modulus of a fluid/air mixture is therefore given by EG =
VFl,0 (1 + (m · p)/E0 )−1/m + VL,0 (p0 /p)1/n (VFl,0 /E0 )(1 + (m · p)/E0 )−(m+1)/m + (VL,0 /(n · p0 ))(p0 /p)(n+1)/n
(11.13)
If α is then defined as the air content of the mixture in its initial state: α=
VL,0 V0
the equivalent compression modulus can be calculated from: EG =
(1 − α)(1 + (m · p)/E0 )−1/m + α(p0 /p)1/n (1/E0 )(1 − α)(1 + (m · p)/E0 )−(m+1)/m + (α/(n · p0 ))(p0 /p)(n+1)/n
(11.14)
n = 1 for slow changes in state with isothermal characteristics. Figure 11.7 shows the equivalent compression modulus characteristic for a mixture of mineral oil and air compared with air-free fluid for α = 0.1, 1, and 10%, E0 = 15,000 bar, and m = 10. This calculation does not allow for the air-dissolving capacity as a function of the pressure. NFPA Standard T2.13.7R1-1997 provides a procedure for determining the compressibility of a fluid [10]. If the gas content is low, the mixture already reaches the compression modulus of the fluid at around 50 bar. A higher gas content leads to an increase in compressibility throughout the entire normal operating pressure range. Another effect can also occur here, and that is, that the change in state — for example, when the fluid is being pumped — takes place very quickly and is therefore approximately adiabatic. As a result, the gas bubbles suddenly heat up to a high temperature. In extreme cases, the temperature reaches the inflammation temperature of the fluid and the fluid is damaged by what is referred to as the “micro-diesel effect.” The gas release and redissolving behavior of the various types of media and other influences are such that the actual characteristic of the compression modulus deviates from the values calculated above. Experiments must therefore be carried out in order to be able to calculate the dynamic performance of the system exactly.
© 2006 by Taylor & Francis Group, LLC
Equivalent compression modulus EG [104 bar]
Hydraulic Fluids
11-13 2 Proportion of air a: 0.1%
1.6 1%
10%
1.2
0.8 Compression modulus for the fluid Isothermal change in state Adiabatic change in state
0.4
0 0
50
100
150
200
250 [bar]
300
Pressure p
FIGURE 11.7 Compression modulus of a mixture of mineral oil and air.
11.3.4 Load-Bearing Capacity and Antiwear Properties A high load-bearing capacity is one of the most important requirements to be met by a hydraulic fluid, and this also implies good antiwear properties. The dynamic viscosity is the most important parameter for the antiwear properties for hydrodynamic lubrication. If the forces acting at low sliding velocities in the mixed friction zone are not sufficient to separate the mating frictional surfaces completely, then the antiwear properties of a fluid are determined by its ability to wet a metal surface and form friction-reducing reaction layers on the mating faces. The wetting capability is also referred to as “oiliness.” The load-bearing capacity of a hydraulic fluid in the mixed friction zone can be improved by means of antiwear additives and substances which reduce the coefficient of friction.
11.3.5 Setting Point/Pour-Point The setting point of a fluid is determined by the temperature at which the medium just ceases to flow under certain testing conditions. By comparison, the pour-point corresponds to the temperature at which the medium just continues to flow. This is around 6 to 8◦ C higher than the setting point. Determination of the pour-point alone is not admissible for an evaluation of the low-temperature characteristics of ester-based hydraulic media. The slow crystallization processes occurring here are such that the time dependence must be taken into consideration as well as just the temperature, and this is determined by means of special test procedures.
11.3.6 Gas Solubility All hydraulic fluids are capable of dissolving a certain proportion of gas. This gas solubility is proportional to the pressure up to around 300 bar. Henry’s law applies: VG = VFl αV
p p0
(11.15)
where VG is volume of gas dissolved at the reference pressure; VFl is volume of fluid; p0 is atmospheric pressure, reference pressure; p is absolute pressure; αV is Bunsen coefficient. The Bunsen coefficient indicates the percentage of gas by volume which is dissolved in a volume unit of the fluid under normal conditions (1.013 mbar, 20◦ C). This value must be determined for every gas or gas mixture. The Bunsen coefficient of air is only slightly dependent on temperature and viscosity. Values for
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Handbook of Lubrication and Tribology
Constant geometric volume
Variable geometric volume
Closed system
q q
x·A q q
Displacement chambers in pumps motors cylinders
Closed tank Closed pipeline
Open system
QIn
QOut Q In
Suction line etc. vIn
vOut
Control edge, screw joint, etc.
FIGURE 11.8
x·A
Displacement chambers in pumps motors cylinders
Buildup of negative pressure in hydraulic systems.
various fluids are given in Table 11.5 at the end of this chapter. The Bunsen coefficient may be determined by ASTM D3827. Under normal circumstances, dissolved air does not exert any influence on the properties of the hydraulic fluid. It may bleed out of the fluid, however, if a low static pressure is applied locally, particularly if the fluid is simultaneously subjected to shearing stresses. The process is referred to as cavitation. The word “cavitation” literally means the formation of cavities. As shown in Figure 11.8, a negative pressure builds up in constant or variable geometric volumes and in open or closed systems. If, for example, a system is shut down with a volume of fluid enclosed in a line, thermal contraction causes the pressure to drop. Where piston port and valve controlled displacement chambers widen in pumps, motors, or cylinders, a partial vacuum negative pressure builds up if the fluid does not continue to flow through to an adequate extent. The gas released as a consequence of this is not redissolved until the pressure increases, when it is dissolved again quickly. Cavitation is most likely to occur in an open system with a constant volume. Such systems include pump suction lines and intake ports, where flow losses are caused by narrow cross-sections, filters, manifolds, and excessive suction height. The consequences are disturbances in the delivery behavior, noise, and an increase in wear due to inadequate lubrication. A low absolute pressure may prevail in flow resistors, for example, throttles, orifices, control edges, even if the system pressure is high. The narrowing cross-section causes the pressure to be converted into a high kinetic energy with simultaneous shearing stresses. These circumstances frequently give rise to so-called cavitation erosion, whereby the resulting gas bubbles implode due to a sudden increase in the pressure acting on their surfaces behind the flow resistors. The continuous stress acting on the material causes fatigue, and particles are broken away [7]. Flow resistance also gives rise to loud noises, as well as instability in throttle controllers. If the fluid–gas mixture is led back into the tank downstream of a resistor without being pressurized, foam is produced as a result of the low redissolving velocity. Cavitation and cavitation erosion, which presents a serious problem, particularly with respect to water-based fluids, can be reduced by suitable design measures, such as the selection of special materials
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Hydraulic Fluids
11-15
for surfaces which are susceptible to erosion, guiding the cavitation stream into uncritical areas away from the walls, or diminishing the differential pressure at one resistor by fitting several resistors, one behind the other.
11.3.7 Air Release Capacity and Foaming Tendency Air bubbles may be entrained where fluid flows into the tank in an unimpeded stream. Apart from this, air may also get into the fluid as a result of leakages in the system, eddy currents in the tank, or cavitation. This air must be released again on the surface before it can be sucked into the pump. As shown by the following calculation, the air bubbles’ rate of ascent depends on their diameter and on the viscosity and density of the fluid. The bubbles’ lifting force amounts to FA = 43 π(ρFl − ρL ) · r 3 g
(11.16)
where r is the radius of the bubble, ρFl is the density of the fluid, and ρL is the density of the air. According to Stokes, the flow resistance of spherical bodies for very low Reynolds’ numbers is given by FW = 6π ηvr
(11.17)
This means that in the case of equilibrium, the rate of ascent is given by v=
2 r 2g (ρFl − ρL ) 9 η
(11.18)
Fluids like these have a good air release capacity for a given operating viscosity, which allows the undissolved air to coagulate into larger bubbles and therefore rise more quickly. This may be boosted by surface-active additives. Unfavorable currents in the tank may slow the air release process down considerably. Wherever possible, design measures must be implemented to prevent air getting into the fluid and facilitate the release process. ASTM D3427 provides experimental details for determining air release properties of fluids. One negative characteristic of hydraulic fluids is the formation of foam on the surface as a consequence of the released air. This behavior may be prevented by the use of suitable additives, but these also have a detrimental influence on the air release capacity.
11.3.8 Aging Behavior The term “aging” includes changes that take place in the composition and chemical structure of a hydraulic fluid. Aging is brought about by such chemical reactions as oxidation, hydrolysis, polymerization, and thermal decomposition, or by mechanical influences, such as shearing action. Oxidation refers to the reaction with O2 producing left-over acids. Polymerization refers to the enlargement of hydrocarbons resulting from the formation of side chains or macromolecules. This process produces waste products such as sludge or resin-like coatings on components. Hydrolysis refers to the cracking of esters when they come into contact with water. The aging process breaks down or destroys the additives and, at the same time, changes the molecules of the base fluid. It is accelerated by high operating temperatures and contamination through extraneous air, water, and metallic catalysts, predominantly copper, copper alloys, and iron. One example of a measure for the aging condition of a fluid is the acid number (AN). The acid value indicates the acid content of a fluid by defining how many milligram of caustic potash solution would be required to neutralize 1 gram of a sample: AN =
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mgKOH gFl
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Handbook of Lubrication and Tribology 0.6 Without contamination 3 vol. % undissolved air 8 vol. % undissolved air 2.5 vol. % water
[mg/g]
TAN
0.4
250 bar 70°C E
0.2
0
FIGURE 11.9
0
250
500 Test period
750
[h]
1000
Influence of air and water on the aging of undoped mineral oil.
The aging stability of a hydraulic fluid can be determined by means of easy laboratory tests, for example, using a Baader device in accordance with DIN 51 554. Results with more practical relevance can be obtained from test-rig experiments [8] as shown in Figure 11.9. The figure shows the way in which air and water influence the increase in the acid value of an unfortified mineral oil over 1000 h under a constant high load. A study of ready-doped hydraulic fluids shows that the TAN drops initially as a result of the breakdown of the acid antiaging additives with a subsequent transition to a progressive ascending gradient. The aging stability of mineral oils, vegetable oils, and HFD fluids can be improved by means of certain additives. As far as water-based fluids are concerned, the base fluid does not age. However, a reduction in the water content must be anticipated during the fluid’s useful life due to evaporation, which results in an increase in the additive concentration.
11.3.9 Material Behavior Hydraulic fluids should not attack metallic materials. Some HFC fluids react aggressively with tin and cadmium. HFD fluids attack aluminum and aluminum alloys in the presence of friction stresses. Far greater difficulties are caused by the ways in which the fluids react with plastics of the types used for seals, hoses, paints, lacquers, and varnishes. Virtually all of the material currently available on the market can be used in conjunction with mineral oils. The only materials which can be used for HFC fluids are silicone rubber and Teflon materials, while only Viton and Teflon materials can be used with HFD fluids. The polar character of environmentally friendly ester fluids leads to a noticeable swelling in conventional standard elastomers. Furthermore, if the compatibility with aged fluids is not known to an adequate extent, a number of adapted materials now exist which permit comparatively safe operation. More serious problems are being encountered with hoses or contamination with water in isolated cases. In addition to this, the products of aging may give rise to problems. Epikote and DD paints are resistant to HEPG, HFC, and HFD fluids to a certain extent. In this respect, it is a good idea to avoid painting the inside surfaces of tanks and to choose suitable materials to avoid the corrosion problem.
11.3.10 Detergent and Dispersant The separation of water and solids in the tank offers a satisfactory means of keeping the fluid clean, at least for low circulation rates. Mineral oils and HFD fluids usually have a good separating capacity, which is referred to as their detergent properties. The precipitation rate of water follows Stokes’ law, analogous
© 2006 by Taylor & Francis Group, LLC
Hydraulic Fluids
11-17
to air bubbles’ rate of ascent (Equation 11.18): v=
2 r 2g (ρFl − ρw ) 9 η
(11.19)
where ρw is the density of the water. Because water has a higher density than mineral oil, it separates out at the bottom of the tank, but it separates out at the surface of HFD fluid because it has a lower density. Here too, the detergent behavior is boosted by the coagulation of smaller drops into bigger ones. In special cases where there is a very high risk of contamination, water and solids are finely distributed and kept in suspension by dispersant substances. This eliminates the risk of malfunctioning valves and detrimental effects on the antiwear properties and loadbearing capacity. The polar ester fluids have a tendency to form stable suspensions and are also capable of dissolving relatively large quantities of dirt particles and water. However, the hydrolytic effect of water on ester compounds is such that appropriate design measures must be implemented to prevent the ingress of water into the hydraulic system where native and synthetic esters are used.
11.3.11 Thermal Capacity and Thermal Conduction The specific heat and thermal conductivity parameters influence the steady-state temperature in a stationary hydraulic installation, as well as changes in temperature with alternating loads. These parameters are needed for the dimensioning of heat exchangers. Because of the higher specific heat and better thermal conductivity of water-based fluids, the steadystate temperature of systems using these fluids can always be kept lower than the temperature of systems using mineral oils or HFD fluids under otherwise identical conditions. This is also necessary due to the higher vapor pressure of water.
11.3.12 Flammability The flash point is the temperature at which fluid vapor which has developed in a test vessel ignites for the first time when approached by a flame (ISO 2592). It is lower than the inflammation temperature of the fluid. The inflammation temperature is the temperature at which droplets of fluid ignite spontaneously under certain test conditions. This temperature is used as a criterion for the hydraulic fluids being fire resistant. The ignition delay is the period between applying 40 ml of a fluid onto a molten mass of hot aluminum at 800◦ C and the fluid bursting into flames. This period is substantially longer for the fireresistant fluids than it is for mineral oil. This procedure and the designation “ignition delay” are not standardized. Reference 11 provides a summary of a range of fire-resistance tests that may be used to evaluate hydraulic fluid flammability potential.
11.3.13 Biodegradability and Toxicity The biological degradation of a substance involves the use of micro-organisms in an aqueous environment to convert the substance into CO2 and biomass. From the point of view of environmental protection, it is important that degradation takes place quickly and that it is complete. As shown in Figure 11.10, native oils have decomposed by almost 100% after 21 days under very idealized test conditions, whereas mineral oils merely achieve a degradation rate of 25%. Paraffin oils, which do not contain any aromatic compounds, reach a rate of 40%. The bandwidth of synthetic esters is the result of their multiple possible variations. A hydraulic fluid must have a degradation rate of 80% for the environmental symbol (“Blue Angel”) 79. The toxic influence of hydraulic fluids on mammals, plants, and bacteria is essentially determined by the additives. According to the German water resources act (Verwaltungsvorschrift wassergefährdende Stoffe), it is evaluated according to three water pollution classes (WGK 1 to 3): slightly hazardous to water, hazardous to water, and very hazardous to water. Environmentally compatible fluids (vegetable oils and some synthetic esters) may not be regarded as being hazardous to water or are put into the WGK 1 class
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11-18
Handbook of Lubrication and Tribology 100%
WGK (1–3)
FIGURE 11.10
1
1–2
1–2
1–3
Paraffin oils
Mineral oils
Synthetic esters
Polyglycols
Vegetable oils
50%
Biological degradation as per CEC-L-33-T82 (21 days)
1–2
Bandwidths of degradation rates and water pollution classes (WGK) for various base fluids.
according to the additives used. Standards for test procedures of biodegradability and toxicity can be found in ASTM D7044, ISO 9439 and OECD procedures OECD 201, 202, 301B and 301B.
11.4 Fortifying Hydraulic Fluids Native and synthetic fluids, as well as natural hydrocarbon oils, are not always capable of meeting the requirements of modern hydraulic components and systems. The quality of the base fluid can be improved only to a certain extent by modifying the manufacturing processes, which means that additional chemical substances must be used to improve the fluid’s performance. These substances are referred to as additives. The chemical degradation of the additives during the fluid’s life cycle is such that a high-grade base oil with few additives should be given precedence over a lower-grade base oil with a high additive content to ensure a long useful life [1,5]. Additives are classified into those which influence the physical and chemical properties of the base fluids, such as VT behavior, crystallization tendency, and aging stability. At the same time, other additives act on the boundary surface between the fluid and components or impurities and thereby improve the frictional and wear behavior, prevent corrosion, or keep particles in suspension. Hydraulic media are made up of a base fluid, which is doped with a so-called additive package to produce a ready-to-use formulation. The effects brought about by the chemical actions of the various additives may be synergetic or even antagonistic. Many additives perform several functions which reduces the possibility of reciprocal interference by individual additives. This group of substances is referred to as the group of “multipurpose additives.” Hydraulic fluid classifications are provided in ISO 6734/4 and ASTM D6158 and D7044.
11.4.1 Oxidation Inhibitors The oxidation reactions occurring in a hydraulic fluid as a result of atmospheric air at higher temperatures cause the fluid to age. Metal ions, such as copper, iron, and lead, may also exert oxidative or reductive influences and accelerate the aging process [6]. Modern hydraulic fluids require a balanced number of so-called oxidation inhibitors in order to counteract these undesirable effects. At the end of the refining process, base mineral oils contain natural inhibitors in the form of sulfur and nitrogen compounds. Frequently inadequate, the resulting oxidation stability is increased by adding other specific compounds. The most important representatives of this group are sulfur compounds, phosphor compounds in the form of phenol phosphate derivatives, compounds of sulfur and phosphor in the form of zinc-dialkyldithiophosphates, phenol derivatives in the form of sterically hindered polyalkyl phenols, and amines.
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Hydraulic Fluids
11-19
11.4.2 Metal Deactivators If a catalytic acceleration of the oxidation process — more precisely the autoxidation process — in hydraulic media caused by metal ions, and copper and iron in particular, is to be prevented, these ions must be “masked out.” Suitable additives are what are referred to as chelating agents, for example, N -salicylidene ethylene diamine, which are effective in very low concentrations, binding ions in the form of complexes. Film-forming media produce a passivating protective film on metal surfaces, thereby preventing the transfer of catalytic ions to the fluid, as well as the oxidative attack by oxygen and oil aging products on the surface.
11.4.3 Wear Protection A high load-bearing capacity must be given to a hydraulic medium where larger forces are to be transmitted with a low rate of wear. The fluid may be doped with so-called high-pressure or EP (extreme pressure) additives for this. EP additives act on sliding faces which are subjected to high pressure and high temperature loads at the transition between hydrodynamic lubrication and mixed friction. They form metal compounds on the surfaces of the sliding, mating friction faces, which are solid under normal circumstances but liquid to slippy under wear conditions. This prevents the surface from becoming worn. Important EP additives include organic sulfur and phosphor compounds and combinations of these elements. Zinc dithiophosphates (ZnDTP), in particular, are used in a wide range of applications. There is trend toward fluid formulations that are free of heavy metals, however, which means that ZnDTP will play a less significant role in the future. Methods to evaluate wear characteristics of hydraulic fluids are summarized in ASTM D7043, D6158 and D6973. Classification Standards are provided in Sections 11.4 and 11.5.
11.4.4 Friction-Reducing Agents, Friction-Modifiers The mixed friction zone is crossed when sliding metal faces run in and out, which means that mild highpressure additives are used for many applications, to prevent stick-slip or noise and reduce the frictional forces and thereby save energy. Also referred to as “friction modifiers,” these additives generally work by forming thin layers on the mating faces by means of physical adsorption; they comprise polar substances, such as fatty alcohols, fatty acids or fatty-acid esters, amides, or salts. HE media offer sufficient lubricating properties as a result of the polar base fluid.
11.4.5 Viscosity Index Improvers VI improvers are additives that improve the viscosity/temperature behavior of fluids, that is, they diminish the reduction in viscosity as a function of the temperature. The VI improvers usually used today are made up of linear polymer molecules, which are effective in that they increase the viscosity of a fluid to a different degree at different temperatures. This affects the flow behavior of the hydraulic medium in such a way that it can no longer be referred to as a Newtonian fluid. The dynamic viscosity changes with the shear stress. Another very important aspect of VI improvers is that they become more sensitive to mechanical stresses as their molecular mass increases. Shear stresses that occur when the fluid flows through control edges, for example, irreversibly break polymer molecules down into fragments, which ultimately causes a reduction in viscosity at higher temperatures. HE fluids and water-based media have a naturally high VI and do not therefore require any VI improvers.
11.4.6 Setting Point Depressants Where machines and installations are used in locations or regions where the ambient temperature is well below freezing point, the flowability of the medium must be assured during a cold start. As they cool down and reach their solubility limits, mineral-oil-based hydraulic media precipitate n-paraffin hydrocarbons in crystalline form as needles and plates, which form a matted network and stop the oil flowing; that is, it sets.
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11-20
Handbook of Lubrication and Tribology TABLE 11.4
Types and Causes of Contamination in Hydraulic Fluids Types
Particles
Liquids Molecular
Chips, scale, etc. Dust, sand Rubbed-off parts Water Extraneous liquids Sludge, resin, acids Metal ions Oxygen
Causes/sources Production, assembly Storage, installation, maintenance, drawn in during operation Wear Storage, maintenance, drawn in during operation Combined systems, maintenance Aging products Friction, wear Dissolved air
The low-temperature properties of these oils can be improved by means of exhaustive deparaffinization. This is an expensive process, however, so deparaffinization is carried out only to a setting point of around −15◦ C. Further improvements can be achieved by using setting point depressors based on the products or polymerization and condensation. Typical examples of these include polymethacrylates, alkylphenols, and copolymers of vinyl acetate and ethylene.
11.4.7 Antifoaming Agents Heavy foam formation exerts a detrimental effect on the lubricating properties of hydraulic fluids, promotes their oxidation, and can cause air to be sucked into the pump. As far as pure mineral oils are concerned, the stability of the foam is a function of the viscosity and the surface tension. According to Stokes’ law, the speed at which the air bubbles precipitate is proportional to the square of their diameter and inversely proportional to the viscosity. Foam with large bubbles which disappear quickly is produced in a low-viscosity fluid, whereas finely distributed small bubbles form in highly viscous fluids and these make the foam highly stable. Additives derived from liquid silicones (polydissethyl siloxanes in particular) have proven to be the most effective antifoaming agents.
11.4.8 Detergents and Dispersants Owing to their large-scale use in engine oils, the surface-active detergents and dispersants have become the most significant types of additives, accounting for approximately 50% of the market share. Their function is to keep oil-insoluble substances, resinous and bituminous oxidation products, and water in suspension, or to accelerate their sedimentation in order to prevent deposits forming on metal surfaces, thickening of the fluid, precipitation of sludge, and corrosion. The dispersant or detergent effects of many of these additives depend on the respective concentration. Dispersants are ash-free organic compounds which prevent the flocculation or coagulation of colloidal particles. The oil-soluble or finely dispersible metal salts of organic acids known as detergents are reputed to have good dirt-dissolving properties. Both types of additives complement one another with respect to their effective characteristics, which are supplemented by the ability to neutralize acid products, thereby inhibiting oxidation. The triple action of the dispersant, cleaning, and neutralizing substances is such that relatively large quantities of these HD (heavy duty) additives are needed and used. If water gets into hydraulic fluids, relatively stable water-in-oil emulsions with disturbing properties may be produced, which can frequently be de-emulsified only by changing the interfacial surface tension. Basically speaking, all types of surface-active compounds are suitable as de-emulsifying agents. De-emulsifying agents increase the foam formation tendency and may therefore be added only in very low concentrations. Emulsifiers are particularly important in their capacity as emulsifying aids for fire resistant hydraulic fluids. Because of their hydrophobic-hydrophilic molecular structure, emulsifiers have surface-active
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Hydraulic Fluids
11-21 Number of particles per 1 ml more than up to
Maximum number of particles per 100 ml in the specified particle size range Class 5–15 mm 15–25 mm 25–50 mm 50–100 mm >100 mm 125 250
22 44
4 8
1 2
0 0
0 0
500
89
16
3
1
1
1,000
178
32
6
1
2
2,000
356
63
11
2
3
4,000 8,000 1,6000
712 1,425 2,850
126 253 506
22 45 90
4 8 16
4 5 6
32,000
5,700
1,012
180
32
7
64,000
11,400
2,025
360
64
8
128,000 256,000
22,800 45,600
4,050 8,100
720 1,440
128 256
9 10
512,000 91,200 1,024,000 182,400
16,200 32,400
2,880 5,760
512 1,024
11 12
Scale number
2,500,000 1,300,000 640,000 320,000 160,000
2,500,000 1,300,000 640,000 320,000
> 28 28 27 26 25
80,000 40,000 20,000 10,000 5,000
160,000 80,000 40,000 20,000 10,000
24 23 22 21 20
2,500 1,300 640 320 160
5,000 2,500 1,300 640 320
19 18 17 16 15
80 40 20 10 5
160 80 40 20 10
14 13 12 11 10
2.5 1.3 0.64 0.32 0.16
5 2.5 1.3 0.64 0.32
9 8 7 6 5
0.08 0.04 0.02 0.01 0.00
0.16 0.08 0.04 0.02 0.01
4 3 2 1 <1
Purity classes in accordance with NAS 1638
Degree of contamination as per ISO 4406 Count result Particle size
Number per 1 ml
Particle size
Number per 1 ml
Scale number
>4 mm
1452.53
18
>4 mm
1452.53
>6 mm
274.4
15
>6 mm
274.4
>14 mm
18.51
11
>14 mm
18.51
ISO 4406:1999 18/15/11
Count result Particle size >5 mm >15 mm
Number per 1 ml 186.2 14.45
>25 mm
4.9
>50 mm
0.72
>100 mm
0.08
Particle range
Number per 100 ml
Class
5–15 mm
17,175
7
15–25 mm
955
5
25–50 mm
418
6
50–100 mm
64
6
>100 mm
8
5
NAS 1638 Class 7
Example for classification of a count result
Recommendations for degrees of purity Type of hydraulic system
Degree of purity
Sensitive control systems, laboratory, aerospace systems
11/8
High-performance servo systems, high-pressure systems with long life
14/11
Reliable, high-quality systems, general machines
16/13
Medium pressure range, general machines + automotive
18/14
Low pressure range, general/heavy-duty machines, automotive
19/15
Low-pressure systems with large tolerances
21/17
FIGURE 11.11 Degrees of purity for hydraulic fluids as per ISO 4406 and NAS 1638.
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11-22
Handbook of Lubrication and Tribology
properties and reduce the interfacial surface tension of the water to facilitate the formation and stability of the emulsion. A distinction is made between anionic, cationic, and nonionic emulsifiers.
11.4.9 Corrosion Inhibitors Corrosion occurs when a metal surface is exposed to oxygen (or another aggressive substance) and moisture at the same time. The corrosion caused by electrolytic processes can be prevented to a great extent by the formation of a nonmetallic protective layer. Effective inhibitors should adhere firmly to the metal and produce a film that is impermeable to water and oxygen. Nitrogen compounds, fatty-acid amides, phosphoric acid derivatives, sulphonic acids, sulfur compounds, and carboxylic acid derivatives are particularly important in this respect.
11.5 Impurities in Hydraulic Fluids There are many sources of contamination or impurities in hydraulic fluids. Refer to Table 11.4. These are already found during the manufacturing and assembly processes in the form of metal chips, grinding dust, welding beads, sand, scale, etc. The initial contamination of freshly supplied fluid is often substantially greater than is permitted for normal operation and this can increase further if the fluid is not stored properly. Under normal operating conditions, dust, fine sand, condensation water, and rainwater from the environment are drawn into the tank through the air vent and into the system by means of deposits on the piston rod. Rubbed-off metal parts are found in the system in the form of particles and released metal ions, along with rubbed-off parts of seals as a consequence of wear. The chemical change brought about in the fluid by temperature, pressure, and shear stresses leads to aging products, such as sludge, resins, and acids, which are usually caused by oxidation with the dissolved atmospheric air. Contamination not only affects the aging of the medium but also the useful lives and functions of the components. Initiated by the introduced energy, several contaminants frequently act together, which means, for instance, that water, oxidized oil molecules, metal ions, and used additives combine to produce sludge. Dirt particles are transported to all parts of a system with the fluid. They may lead to a direct failure if particles cause the slide in a valve to jam, for example, or block control nozzles. A far more significant influence is brought about by the impurities; however, additional wear in the components. Abrasion is caused by particles caught between mating faces, and erosion takes place on edges and surfaces which are exposed to the fluid flowing at high velocity. Solid contaminants are described according to purity classes, which define the maximum permitted number of particles of a particular size. Recommendations are given for the applicable purity class according to the components used, Figure 11.11. In this respect, apart from the components’ sensitivity to wear, the gap widths specified for the design and the control nozzle diameter also play particularly important roles. Continuous maintenance of the fluid by filtering and separating water out, as well as preventive measures, such as thorough cleaning of the components during the assembly process and using filter systems TABLE 11.5
Characteristic Values of Hydraulic Fluids
Density at 15◦ C (g/cm3 )
Kinematic viscosity at 40◦ C (mm2 /sec) Mean compression modulus E (N/m2 ) Viscosity/temperature index Specific heat at 20◦ C (kJ/kgK)
© 2006 by Taylor & Francis Group, LLC
HLP
HFA (3%)
HFD
HETG
0.87 10 to 100 2 × 109 100 2.1
1.0 0.7 2.5 × 109 — 4.2
1.15 15 to 70 2.3–2.8 × 109 <0 1.3 to 1.5
0.92 32 to 48 2.5 × 109 210 2.1
ATF 0.87 36 to 40 2 × 109 150 2.1
Hydraulic Fluids TABLE 11.5
11-23
Continued
Thermal conductivity at 20◦ C (W/mK) Viscosity/temperature index Specific heat at 20◦ C (kJ/kgK) Thermal conductivity at 20◦ C (W/mK) Volumetric expansion coefficient (1/K) Operating temperature range (◦ C) Maximum temperature range (◦ C) Flash point (◦ C) Inflammation temperature (◦ C) Setting point (◦ C) Bunsen coefficient αv at 20◦ C for air Vapor pressure at 50◦ C (mbar) Cavitation tendency Relative costs of fluid (%) Market share (%)
HLP
HFA (3%)
HFD
HETG
0.14 100 2.1 0.14 7 × 10−4 −10 to 80 −40 to 120 210 310 to 360 −18 6.8 4 × 10−2 Slight 100 85
0.6 — 4.2 0.6 1.8 × 10−4 5 to 50 0 to 55 — — 0 — 100 Great 10 to 15 4
0.11 <0 1.3 to 1.5 0.11 7 × 10−4 10 to 70 −20 to 150 245 500 −24 to 6 — 10−2 Slight 200 to 400 2
0.17 210 2.1 0.17 7.5 × 10−4 0 to 70 −20 to 90 315 350 to 500 −25 4.6 3 × 10−3 Slight 150 to 300 3
ATF 0.14 150 2.1 0.14 7 × 10−4 −20 to 100 −40 to 120 190 300 −40 9.4 4 × 10−2 Slight 300 —
when filling with fluid, are very important contributory factors where efforts are concentrated on achieving a satisfactory degree of purity on a long-term basis with the associated reliable functioning and long service life.
References [1] Mang, T. and Dresel, W. (publisher) Lubricants and Lubrication, Wiley-VCH, Weinheim, 2001. [2] Bartz, W.J. (publisher) Hydraulikflüssigkeiten: Eigenschaften, Norumung und Prüfung, Anwendung [Hydraulic fluids: properties, standards and testing, application], expert-Verlag, RenningenMalmsheim, 1995. [3] Krstic, M. Umweltfreundliche Schmier- und Druckflüssigkeiten: Vorteile und Auswahlkriterien für die Anwendung [Environmentally compatible lubricants and hydraulic fluids: advantages and selection criteria], Verlag Moderne Industrie, Landsberg/Lech, 2000. [4] Schmidt, M. Untersuchung und Ansätze zur modellhaften Beschreibung der Alterung auf Estern basierender Zwischenstoffe für den Einsatz in umweltverträglichen Tribosystemen [Investigation of and approaches to the use of models to describe aging behavior of ester-based precursors for use in environmentally compatible tribological systems], dissertation, RWTH-Aachen, 2003. [5] Bartz, W.J. (publisher) Additive für Schmierstoffe [Additives for lubricants], expert-Verlag, Renningen-Malmsheim, 1994. [6] Remmelmann, A. Die Entwicklung und Untersuchung von biologisch schnell abbaubaren Druckübertragungsmedien auf Basis von synthetischen Estern [The development and investigation of rapidly biodegradable pressure media based on synthetic esters], dissertation, RWTH-Aachen, 1999. [7] Backé, W. Grundlagen der Ölhydraulik [Fundamentals of oil hydraulics], publication to accompany a lecture at a RWTH-Aachen, 10th ed., 1994. [8] Remmelmann, A., and Murrenhoff, H. “Chapter 11-in ASTM” Fuels and Lubricants Handbook, Eds. G. E. Totten, S. Westbrook and R. J. Shah, ASTM Manual Series: MNL37WCD, 2003, ASTM International, West Conshohocken, PA. [9] Murrenhoff, H. Grundlagen der Fluidtechnik [Fundamentals of fluid technology], publication to accompany a lecture at the RWTH-Aachen, 4th ed., 2005. [10] NFPA Standard T2.13.7R1-1997: Hydraulic Fluid Power. Petroleum Fluids — Prediction of Bulk moduli, National Fluid Power Association, Milwaukee, WI, 1997. [11] Fire Resistance of Industrial Fluids, Eds. G. E. Totten and J. Reichel, ASTM STP 1234, June 1995, American Society for Testing and Materials, West Conshohocken, PA.
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12 Coolants and Lubricants in Metal Cutting 12.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12.2 Classification of Coolants and Lubricants . . . . . . . . . . .
S.C. Veldhuis and G.S. Fox-Rabinovich Department of Mechanical Engineering, McMaster University
L.S. Shuster Ufa Aviation Institute
12-1 12-2
Gaseous Lubricants • Cutting Fluids • Plastic Lubricants • Solid Lubricants
12.3 Fundamentals of Fluid and Lubricant Application for Cutting Operations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12-4 Functionality and Service Properties of Cutting Fluids
12.4 Future Trends . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
12-11
New Lubrication Techniques • Dry Machining
References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
12-14
12.1 Introduction Cutting fluids represent a significant portion of total manufacturing costs. In some processes the initial investment and maintenance costs for the cutting fluids can account for up to 15–17% of the total manufacturing costs and close to half of the machine tool costs [1,2]. Cutting fluids also play a vital role in lubricating and removing heat as well as swarf from a process. Hence the proper selection and application of cutting fluids is a major economic as well as engineering issue for high performance machining. The application of coolants and lubricants is obviously one of the most affordable and beneficial ways to improve machining performance. The successful application of coolants and lubricants can often result in cutting tool life improvements on the order of 1.2–4 times, intensification of cutting parameters by 20–60%, and productivity increases of 10–50%. The majority of the coolants and lubricants that are used in practice today are fluids. It is well known from the practice of machining that a flood of liquid is directed over the tool to act as a coolant or lubricant, or both. Some typical objectives and properties of these fluids follow: • To prevent overheating of the machine, tool, and workpiece during cutting (cooling effect) • To reduce friction and the amount of heat generation; this results in tool life and surface finish improvements (lubricating effect) • To reduce the seizure and galling phenomena (antifriction properties) • To clear chips from the cutting zone (flushing effect) 12-1
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12-2
• • • • • • •
Handbook of Lubrication and Tribology
To enhance fluid penetration to the cutting zone (good wetting properties) To provide a degree of corrosion protection To allow chips and other metal debris to settle out (low viscosity) Resistance to the formation of a sticky or gummy residue on parts as well as on machine tools Chemical stability to provide a safe work environment Antibacterial properties Overall economical efficiency
Application of efficient coolants and lubricants results in a better surface finish for final machined components, especially during machining of hard to machine alloys. These alloys are becoming increasingly topical for the North American market. A large variety of cutting fluids are currently available on the market. These fluids all try to improve cutting performance on one hand while meeting numerous requirements such as those outlined above on the other. The proper selection and development of advanced fluids are possible based on an in-depth understanding of the mechanism associated with the interaction between the tool and the workpiece material during cutting. However, at this time the selection of fluids depends mainly on extensive testing and practical experience. The fundamentals of a fluid’s impact on friction in general are based on well-known tribological studies [3–10]. The result of these studies for a machining application leads to the wide use of surface active materials in cutting fluids.
12.2 Classification of Coolants and Lubricants All coolants and lubricants can be divided into a few groups based on the state of aggregation of matter: gaseous, fluid, plastic, and solid. Coolants and lubricants can also differ depending on their application as universally applied solutions or as solutions tailored for specific applications. Universal coolants and lubricants can be used for a wide range of machining operations. Specific coolants and lubricants, such as gaseous, plastic, and solid ones as well as liquid lubricants are developed and used for a relatively narrow domain of application.
12.2.1 Gaseous Lubricants Lubricants of this type can be either inert (nitrogen, argon, helium) or active (oxygen-containing: air, oxygen, carbon dioxide) ones. Active gases have the potential to both cool and protect the surface through the formation of oxide films. Machining (turning and drilling) of hard to machine materials such as Inconel as well as sharpening of high speed steel (HSS) and carbide cutting tools in an oxygen environment provides some beneficial results. The use of CO2 is also effective and clean but expensive in comparison with compressed air. Gaseous coolants are not widely used. The cooling limitation of gaseous coolants can be illustrated by the application of liquid nitrogen [11,12]. In this case the temperature in the cutting zone was found to drop only slightly. Even extreme cooling action could not prevent the localized overheating of the rake surface and thus intensive crater wear.
12.2.2 Cutting Fluids There are two major types of cutting fluids: • Oils • Water-based fluids In addition to these two, some metal melts can also be considered as cutting fluids. 12.2.2.1 Cutting Oils Typically, cutting fluids consist of petroleum products or animal and vegetable oils. These oils usually contain so-called extreme pressure or antiweld additives such as sulfur and chlorine. These additives react
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Coolants and Lubricants in Metal Cutting
12-3
under pressure and heat to give the oil better lubricating properties. The obvious advantages of oils are good lubricity, antiseizure, and antigalling properties together with a high level of corrosion protection. Oil usage is mandatory in many cutting operations such as tapping of hard to machine alloys, free machining, deep hole drilling, and gear hobbing. They are practically irreplaceable in these applications. Sometimes synthetic oils or their mixture with mineral oils are used as fluids. These fluids have some advantages; however, they are typically not price competitive when compared with natural mineral oils. In summary, oil-based fluids increase cutting tool life, improve workpiece surface finish, and protect the machine tool against corrosion. Oils without additives are used for machining of magnesium, copper and copper based alloys (such as brass and bronze), and sometimes carbon steels under nonaggressive cutting conditions. Unfortunately these oils have a low degree of efficiency during machining of hard to machine materials and under high performance machining conditions owing to a number of disadvantages inherent in their use. Disadvantages of these fluids include the following: • • • •
Relatively low cooling ability Low thermal stability Increased evaporation and odor Relatively high price
12.2.2.2 Water-Based Cutting Fluids These fluids typically contain emulsifiers, mineral oils, water, spirits, and inhibitors to prevent corrosion and growth of bacteria, antigalling as well as antifoaming additives, electrolytes, and other organic and nonorganic additives. The fluids can also contain some fat or oils to improve lubricity. Recently, some surface active materials or chlorine additives have been used. The lubricating properties of the fluids (i.e., water-based emulsions as well as oils) can also be improved by the addition of chlorine and sulfur. These fluids are widely used as emulsions or water solutions during the cutting of a wide range of steels and nonferrous metals. The main advantage of the water-based fluids is their higher efficiency as coolants. These fluids are also cheaper and more environmentally friendly. The disadvantage of the water-based fluids is their relatively low lubricity, their low efficiency under specific cutting conditions, enhanced corrosion (especially if the mixture separates), and the relatively low stability of their properties over time. Water-based fluids can be divided into a few groups: emulsions, semisynthetic, synthetic, and electrolyte solutions. An emulsion is usually a mineral oil that has been dissolved in water. These fluids consist of suspensions of oil droplets in water. These are the most widely used cutting fluids because they combine the lubricity of oils and the cooling properties of water. With these properties they can be used for a wide range of cutting operations under various machining conditions. Emulsions are mixtures of mineral oils and emulsifiers, as well as corrosion and growth bacteria inhibitors together with other additives, which often serve as stabilizers. The content of mineral oils in cutting fluids can be up to 85%. Emulsions are typically used as 1 to 5% water solution. Semisynthetic cutting fluids are similar in composition to emulsions but the concentration of the components is different. The base of the semisynthetic cutting fluids is typically water at approximately 50%, with emulsifiers making up approximately 40%. One of the important properties of mineral oil is a low viscosity around 3 to 10 vv/sec at 50◦ C. The semisynthetic cutting fluids can also contain some additives. This type of fluid is used as a 1 to 10% water solution. Synthetic cutting fluids are the mixture of water-soluble polymers, surface active materials, additives, and water. To improve the lubricity of this type of fluid some antiwear and antigalling additives are used. Synthetic coolants are used as a 1 to 10% water solution. Synthetic fluids consist of inorganic or organic materials dissolved in water to improve the lubricity of these cutting fluids. The primary functions of these fluids are to cool as well as inhibit rust and corrosion. The synthetic fluids contain no oils and usually have low lubricating values. Their lubricating properties can be improved through the addition of active agents, which lower the surface tension of water and hence increase its wetting ability through the formation of
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colloidal aggregates among the surface active molecules. The disposal of these fluids is a relatively easy task compared with that of other lubricants. These fluids have considerable potential. Semisynthetic fluids have recently become less popular because of the significantly improved lubricity provided by state of the art synthetic fluids. 12.2.2.3 Metal Melts Metal melts such as tin, cadmium, bismuth, and zinc are used for the machining of difficult to machine metals, Inconel, titanium, and titanium-based alloys. These melt applications result in significant tool life improvements. However, it is very difficult to use the metal melts under automated conditions because of challenges associated with applying them in the cutting zone.
12.2.3 Plastic Lubricants Plastic lubricants are used under conditions typical of small lot size production in tapping, drilling, and reaming operations. The area of application of these lubricants is narrow due to the difficulties related to penetration into the localized cutting zone, as well as environmental issues associated with cleaning and recycling. The plastic lubricants can be divided as follows: • Lubricants that are based on hydrocarbon thickening agents (paraffin, wax, and a few polymers) • Lubricants that are based on soap thickening agents (sodium, lithium, calcium, barium, aluminum, lead-based, and others) • Lubricants that are based on inorganic thickening agents (silica gel, clay, molybdenum disulfide, mica, and asbestos)
12.2.4 Solid Lubricants Solid lubricants for cutting applications can be divided into three groups based on their chemical composition: inorganic products (talcum, graphite, mica, molybdenum disulfide, and others), organic compounds (wax, soap, solid fats, polymers), and soft metals (tin, lead, copper). These lubricants are used under extreme cutting conditions when heavy loads and temperatures are applied and if the other types of lubricants do not perform well enough. Solid lubricants are deposited as coatings on the surface of cutting tools as well as the machined part. The methods employed to deposit these lubricants depend on the type of solid lubricant used. Recently some of these coatings have started to be deposited using advanced plasma vapor deposition (PVD) techniques.
12.3 Fundamentals of Fluid and Lubricant Application for Cutting Operations Cutting is one of the most “ancient” methods of metalworking [13]. This method comes in many forms: turning, milling, drilling, tapping, broaching, reaming, gear hobbing, grinding, and so on. Metal cutting involves surface plastic deformation, fracture, and metal removal as a chip forms during cutting. During machining, a chip leaves the cutting zone together with some particles of the worn tool. The tool’s actual wear volume is small; however, over time the cutting edge profile changes and eventually the tool loses its functionality. Wear occurs both on the rake and flank surfaces of the tool during cutting and is a very complicated phenomenon. The major wear modes observed during cutting include the following: • Abrasive (similar to grinding) • Adhesive (formation of welded asperity junctions between the chip and the tool face, followed by the subsequent fracture of these junctions by shear, which leads to microscopic fragments of the tool material being torn out and subsequently adhering to the chip or the workpiece) • Diffusive or chemical wear (a complicated process involving material transfer from one metal to another passing through the interface) and their combination
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There are a few other mechanisms of cutting edge failure observed during service such as fatigue fracture (crack formation and chipping) and plastic deformation of the cutting edge. When a chip leaves the cutting zone a juvenile surface of the workpiece forms that possesses some very special features. Ideally, a clean juvenile surface has very high chemical reactivity and emits an intense flow of electrons. It also has a very high coefficient of friction. These features of the juvenile surface enhance adhesive interaction and diffusion at the cutting tool–workpiece interface. Furthermore, the juvenile surfaces that are generated during friction act as catalysts for chemical reactions at the cutting tool–workpiece interface, especially in the presence of lubricants. The interaction of the juvenile surface with active molecules in the cutting fluids, such as oxygen, iodine, and chlorine, results in the formation of chemical films, and it can decrease both the friction forces and the intensity of adhesive and diffusive processes during cutting tool wear. In these cases even a small amount of fluid reaching the cutting zone can result in significant changes in the efficiency of the cutting process. This is why significant attention is paid to the issues associated with getting the cutting fluid to penetrate into the cutting zone. One of the major features of the cutting phenomenon is the intensive seizure occurring at the cutting tool–workpiece interface. When intensive seizure takes place, which is typical for machining operations such as turning of ductile materials, there is limited or almost no access of the cutting fluids into the workpiece–cutting tool interface. Given the confines of the geometry in the cutting zone, fluid access is typically precluded from this region, especially along the plastic part of the tool–chip contact length [13]. Despite this, cutting fluid application obviously improves such critical parameters of cutting as tool wear rate and the surface finish of the machined components. Direct observations were made that support the role of cutting fluids to reduce the tool–chip contact length [2]. For this reason the theory considering cutting fluids as boundary lubricants is still commonly used [14]. Furthermore, boundary lubrication conditions could take place when seizure does not result in the formation of continuous metal to metal contact layers at the workpiece–cutting stool interface. When the seizure is island-like or the area of seizure is localized, which is typical for some milling operations or gear hobbing conditions, the cutting fluid application can be very effective. There are a great number of publications related to cutting fluids, but only a few of them have been aimed at understanding the role of cutting fluids on the mechanism of cutting [15–21]. Major paths of cutting fluid penetration into the contact zone are the pores and capillary network that form at the cutting tool–workpiece interface (Figure 12.1[a]) and the voids that are formed as a result of the build-up edge tearing off (Figure 12.1[b]). Vibration during machining also enhances the lubricant penetration into the contact zone (Figure 12.1[c]). Experimental studies show that the contact of the chip and cutting tool is discontinuous during cutting under low and moderate cutting speeds. In fact some researchers intentionally introduce small amplitude high frequency vibration between the tool and the workpiece to reduce the contact time between the tool and the workpiece and to facilitate the introduction of coolant into the cutting zone [22]. The formation of micron-sized capillary networks during cutting ensures the steady supply of fluid and mist into the cutting zone. Herein fluid velocity reaches 3.5 to 4.0 m/sec, and thus the permanent supply of the lubricant to the cutting tool–workpiece interface is ensured. The depth of cutting fluid penetration depends on the size of the capillary network at the cutting tool–workpiece material interface, the surface tension at the phase boundary, and the density of the lubricant.
(a)
(b)
(c)
(d)
FIGURE 12.1 Major paths of cutting fluid penetration into the workpiece–cutting tool contact zone: (a) pores and capillary network that form at the cutting tool–workpiece interface; (b) voids that are formed as a result of the build-up edge tearing off; (c) vibration during machining; (d) distorted lattice structure within the deformed layer of the chips.
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Under high performance machining conditions a nearly continuous seizure zone is formed on the rake surface of the cutting tool. As the size of the capillaries decreases dramatically, there is only one possible way for fluids to penetrate into the cutting tool–workpiece interface; this involves the penetration of the fluids in a gaseous-like state. In this case the actual lubrication at the cutting tool–workpiece material interface is performed by the products of the fluids’ thermal dissociation. These products are interacting with the friction surfaces or with oxygen from the environment and after that with the machined metal surface. The other possible path of penetration for the smallest particles of fluids such as ions, atoms, or molecules to the interface is through the distorted lattice structure within the deformed layer of the chips (Figure 12.1[d]). It has been determined that cutting fluids in these forms are penetrating through the sides of the chips. The driving force of the lubricant penetration to the cutting tool–workpiece interface is the force of the chemical interaction, adsorption, vibration, and external electrical and magnetic fields. The fluid’s penetration to the actual zone of cutting is also a result of the vibration of the cutting tool and the workpiece, that is, the stiffness of the machine tool–holder–cutting tool/workpiece system or externally driven vibration. Among the most important parameters that determine the efficiency of the cutting fluid’s penetration are the following: • • • •
A method of fluid supply to the cutting zone (free fluids flow, forced lubrication, spraying) Viscosity of the fluids Surface properties Chemical nature and size of molecules, atoms, and ions
Cutting fluids act as coolants, lubricants, dispersants, and cleaning agents. The different roles of the cutting fluids are developed at the specific segments of the cutting tool–workpiece contact zone (Figure 12.2). Segments I and II are the cooling zones and zones of embrittlement, respectively. Whole segments III and IV are lubricating zones due to the formation of protective films.
12.3.1 Functionality and Service Properties of Cutting Fluids 12.3.1.1 Lubrication Lubricants in cutting fluids change the friction conditions at the cutting tool–workpiece interface. The major goals of the cutting fluid lubrication are as follows (1) to reduce the intensity of seizure at the cutting tool–workpiece interface as well as (2) to reduce the heat generation due to friction; (3) to change the tool-chip contact length; (4) to decrease the cutting force; and (5) to affect the build-up formation. Under high performance machining conditions the efficiency of the cutting fluid is lower because any fluid access is supposed to be precluded, especially along the plastic part of the tool–chip contact length at the cutting tool–workpiece interface.
I II
III
IV
FIGURE 12.2 Specific segments of the cutting tool–workpiece contact zone. Segments I and II are the cooling zones and zones of embrittlement; segments III and IV are lubricating zones due to protective film formation.
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The enhanced lubricity of a cutting fluid leads to: • • • •
Tool life improvements due to the decrease of the adhesion and diffusion at the interface Productivity increase Surface finish improvements of the workpiece Lowering of the residual stresses in the machined component due to a decrease in the intensity of adhesion and the shearing process’s localization • Reduction in cutting forces and heat generation At the same time the improvement in a cutting fluid’s lubricity could lead to some negative results: • Decrease of the tool-chip contact length on the rake surface of cutting tools, which leads to a growth in the stress concentration at the cutting edge, so that as a result, cutting edge chipping can occur. • Lowering the protective function of the build-up material. • Growth of the cutting forces during the machining of some metals (e.g., cutting forces during machining of aluminum, lead, and copper using high speed steel cutting tools are higher in air than in vacuum; the cause of this phenomenon is the formation of surface tribo-films with a strength higher than that of the bulk material). High temperatures in the cutting zone, heavy applied loads, and plastic deformation occur and lead to the formation of strongly reactive juvenile surfaces. The molecules of surface active materials in the cutting fluids are adsorbed onto the surface. The presence of the moisture and oxygen in the environment accelerates the processes associated with chemosorption. Temperature plays a significant role in these processes. The adsorbing and chemisorbing films have low temperature stability. Hence they have low efficiency during the machining of hard to machine materials and materials that generate significant amounts of heat during cutting. The lubricating film-forming additives in the cutting fluids, which act as antiwear and antigalling additives, have higher efficiencies under these conditions. The molecules of these additives interact with each other, oxygen from the environment, and the juvenile surfaces of the frictional bodies. As a result, the molecules of the additives dissociate and form atoms and radicals that interact with the metals, resulting in the formation of a surface-lubricating layer. There are different data on the composition and properties of the lubricants that exhibit high temperature stability and efficiency for various machining operations. Most efficient are the compounds such as oxides, sulfides, chlorides, and complex compounds that have a shear strength that is lower than that of the contacting frictional bodies. The tribo-film formation is based on these compounds preventing the direct interaction of the frictional bodies, which results in a decreased level of adhesive and fatigue wear rate for the cutting tools. The positive role of oxygen in the formation of protective tribo-films with improved lubricity is known from the literature [23]. However, an excess of oxygen in the cutting zone leads to the formation of relatively thick and brittle oxide films on the surface of the cutting tools that can result in a wear intensity increase. That is why it is important to control the process with the aim to accelerate, or if necessary, depress the reactions of the frictional surfaces with oxygen as compared with cutting in a natural air environment. The major goal of this control is to inhibit or more realistically to reduce the adhesion and diffusion processes at the cutting tool–workpiece interface. In addition to the oxygen and the carbon that are supplied by fluids to the cutting zone, the other elements such as boron and nitrogen could play a similar role. The amount of these elements within the cutting zone as well as the amount of oxygen and carbon should be optimized to get the best performance. In general, the mechanism of lubrication from a cutting fluid is a result of the following processes: • Depression of the reactivity of the juvenile surfaces that are formed during cutting • Formation of the boundary tribo-films, protecting materials of the cutting tool and workpiece against mechanical and physicochemical destruction • Lowering of the surface energy of the workpiece material as well as a decrease in the shearing strength of the surface layer (Rebinder’s effect [24]) • Lowering of the tool-chip contact length and a decrease in the frictional forces
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• Formation of the wedge pressure in the microcracks that are forming in a workpiece material during cutting and inhibition of the welding of the formed crack sites Lubrication of the cutting fluids could be evaluated during cutting and tribological tests. Lubrication during cutting could be evaluated based on such parameters of the cutting process as (1) cutting forces and work, (2) friction parameter on the rake and flank surfaces, (3) chip compression ratio, (4) wear rate data, and (5) build-up formation. The results of the cutting test are also dependent on the cleaning, cooling, and dispersant properties of the cutting fluids; however, the tribological test data are typically not sensitive to these properties. Sometimes the lubricating properties of cutting fluids are evaluated using special pieces of tribological apparatus that model the conditions of cutting. However, the data obtained using this apparatus can only indirectly characterize the lubricity as well as the antiwear, antigalling, and antifrictional properties of the cutting fluids because such parameters of the cutting process as cutting data, chip thickness, and chip compression ratio strongly affect the experimental values obtained. That is why the lubricity of cutting fluids is evaluated most often using universal tribometers, where cutting is not modeled and the efficiency of lubrication is estimated based on the friction parameter and the wear rate data. Temperature dependent properties of the cutting fluids are the most important parameters to characterize the lubricity of these materials. Based on the knowledge of the actual temperature in the cutting zone and the temperature-dependent properties of the cutting fluids, the proper selection of the cutting fluids can be made for specific applications. The role of antifriction fat-based additives in cutting fluids consists of the formation of adsorbed lubricating films on the surface of the cutting tools that lead to a wear rate decrease. The concentration of these fat-based additives in cutting fluids is in the range of 0.5 to 20%. Antigalling additives inhibit the adhesion between the cutting tool and the workpiece material and promote a wear rate decrease under high temperature and heavy load conditions. The action mechanism of these additives includes the chemical interaction of the dissociation products of these materials with metal surfaces in contact during cutting. The chemical compounds that are formed as a result of this process have lower shear strength and melting temperature than the bulk material. Chemical lubricating films decrease cutting forces and prevent intensive adhesion and diffusion. The concentration of the antigalling additives in cutting fluids can be in the range of 0.5 to 50%, depending on the cutting conditions. 12.3.1.2 Cutting Fluids as Coolants There are two major sources of heat generated during the cutting operation. The first one is the primary shear plane and the second one is the tool–workpiece interface. Usually, the majority of the heat is generated due to the intense shearing action within the workpiece material. Conversely, the heat that is generated during friction makes a minor contribution to the overall temperature rise during cutting. Unfortunately, coolant has no direct access to these heat sources; yet it is well known that water-based coolants efficiently reduce the temperature in the cutting zone. The removal of the heat from the primary shear zone only slightly changes the cutting tool’s performance. In comparison to this, removing the heat generated during friction at the tool–workpiece interface has a large impact on the tool’s life and the wear behavior of the tool. In many cases cooling can be effectively achieved through the tool using internal passages, which better direct the coolant into the cutting zone. It should be mentioned that the flow zone of the chip at the rake surface of the cutter with highest actual temperature is an area of seizure of the workpiece material to the tool surface. It has been shown experimentally that the application of coolant could not prevent the generation of high temperatures at the cutting tool–workpiece interface because the interface is inaccessible to the direct action of the fluid. But the application of the coolant significantly reduces the volume of the tool surface affected by overheating. In practice overall flood cooling is most widely used. The majority of the mechanical energy generated during cutting is transformed into heat and only a small part of the energy is spent on the structural transformation of the workpiece surface layer. The heat generation during cutting leads to temperature growth within the cutting zone including the tool, workpiece, chips, and the environment. In general, the heat generated due to the cutting action
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increases with cutting speed and to a lesser extent with feed and depth of cut. Heat generation during the machining of ductile materials such as steels is higher than that during the machining of brittle material such as cast iron. The cooling action of cutting fluids impacts: • • • • • •
Cutting tool life Accuracy of machining (thermal deformation) Surface finish Residual stress formation in the surface layer of the machined component Chip formation (chips bending, tool-chip contact length changes) Machining productivity
Diffusion and adhesion processes that take place during cutting, as outlined above, largely determine the cutting forces. These processes are the temperature-dependent ones. Cutting fluid applications lead to a significant decrease of the workpiece, cutting tool, and chip temperature. Thus cutting fluids could impact the temperature distribution during cutting. The enhancement of a cutting fluid’s cooling properties leads in general to tool life improvements during cutting operations with intensive heat generation. The machining of steels is usually associated with the use of fluids. The use of coolant is very important when HSS cutting tools are involved. Cutting fluids lower the temperature in the cutting zone when the machining is performed at speeds of up to 150 m/min. At higher cutting speeds cutting fluids just stabilize the temperature of the workpiece. The use of coolant is quite efficient for turning operations. But for some operations, such as drilling of shallow and deep holes, the use of coolants is critical because it significantly drops the temperature down in the cutting zone to allow cost-effective machining. For heavily loaded cutting operations such as tapping and gear hobbing the lubrication properties of fluids largely determine the performance of the cutting operation. On the other hand the cooling of the cutting zone could have some negative results such as intensive cutting edge chipping during interrupted cutting operations (milling, planing) especially at elevated cutting speeds. This is very topical when ceramic and even cemented carbide tooling is used because the coolant application results in intensive thermal cycling, which leads to cutting edge chipping. In addition to this, the intensive cooling of the workpiece can sometimes lead to the formation of tensile internal stresses that result in poor surface properties of the machined part. 12.3.1.3 Washing Ability of Cutting Fluids Chips and other waste materials such as wear debris, oxide particles, dust, dirt, and products of the cutting fluid’s breakdown are formed during cutting. The washing ability of the cutting fluid is a combination of the chemical phenomena that leads to the cleaning of the cutting surface and the machine tool as well as the workpiece. This characteristic of the cutting fluid ensures the cleaning of the workpiece, cutting tool, and machine tool of chips and waste material. We have to emphasize that the term “washing ability” of a cutting fluid covers two meanings (1) its washing ability in the proper sense of the word and (2) its ability to flush out and evacuate chips or metal debris as well as inhibit the formation of carbon build-up on the surface of the cutting tools and workpiece. Carbon build-up is a result of two processes occurring during cutting (1) adsorption of carbohydrate molecules present in the coolant collecting on the hot machined surfaces and (2) oxidation from the oxygen that is in solution in the coolant. The initial stage of the cutting fluid washing action is the moistening of the dirty surfaces, both the metal and the nonmetal particles. The next stage of washing action is the formation of the stable suspension of the waste material in the fluids. The molecules of the surface-active material adsorption lead to the dispersion of debris due to their surface strength decrease as well as a wedging action. In addition to the surface-active molecule’s action there is also a so-called “colloidal solution” of the dirt forming in the cutting fluids.
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Environment
Pyrolisis
Formation of active molecules, radicals, ions Boundary lubrication
Cooling action of environment Physical adsorption
Chemosorption
Formation of a chemical film
Friction forces decrease
FIGURE 12.3
Heat generation
Diffusion
Chip plasticization
Embrittlement action
Wear rate decrease
Cutting force decrease
The scheme of cutting fluids action.
The washing of parts and the evacuation of chips is one of the most important functional properties of a cutting fluid. The washing ability of a coolant is highly dependent on the quantity of the cutting fluid used, its directed supply in the cutting zone, flow velocity, method of fluid supply as well as the concentration of the washing compounds in the fluid. The washing ability of a coolant is becoming more important as increased levels of automation are being put into place that require clean parts for robotic gripping and part fixturing. In industrial practice the washing ability of the coolant is evaluated visually. The quantitative estimation of this property is very time consuming and expensive. Based on the properties of the cutting fluids outlined above we can explain the main action mechanisms and some of the ways that the cutting fluid’s efficiency can be enhanced using the scheme suggested in Figure 12.3. 12.3.1.4 Impact of Cutting Fluids on Tool Life, Cutting Forces, Surface Finish, and the other Parameters of Cutting The main criteria for evaluating the major service properties of a cutting fluid are as follows: • Wear of cutting tools • Tool life • Surface finish (dimension accuracy and roughness of the machined surfaces) Additional criteria that could characterize the service properties of cutting fluids are as follows: • Torque • Cutting forces
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Temperature distribution at the tool surface Micro-profiles of the cutting tool–workpiece material interface Vibration resistance during cutting Cutting tool chipping
The most comprehensive evaluation of all the service properties of cutting fluids can be carried out through testing under real cutting conditions. However, for these tests to be statistically relevant requires tightly controlled conditions on a large number of machines with the significant consumption of cutting tools, workpiece materials, and cutting fluids. These industrial tests are both very expensive and time consuming. One of the main challenges in conducting testing of this nature in a production environment is dealing with the large number of changes that take place over the period of the test. For this reason developing new cutting fluids for specific applications requires methods of accelerated testing using specialized apparatus, such as tribometers, in addition to controlled testing on machine tools. We have to emphasize that cutting fluid selection for a specific application is a difficult technological task. Tables 12.1 and 12.2 [25] could be used as a guide for cutting fluid selection for specific applications. The choice is influenced by many machining parameters. Table 12.1 shows the application of cutting fluids for the major machining operations, such as turning, milling, drilling, and tapping, for machined materials that are widely used in industry. The trend in the past decade has been toward the more intensive usage of cemented carbide tools. That is why the recommendations presented are based mainly on carbide cutting tools. Tapping is still one of the main exceptions as the majority of these tools are still made from high-speed steels.
12.4 Future Trends 12.4.1 New Lubrication Techniques Recently, cutting fluids have been viewed as a huge environmental and health liability. Governmental agencies in North America and Western Europe have classified cutting fluids as hazardous wastes and unions have been pushing to reduce the use of cutting fluids due to complaints from their members. The net effect is that regulations controlling lubricants have increased every year, driving up the final cost of using lubricants in a cutting process. Minimizing lubricant consumption is a major ecological and economic objective of machining operations. As a long term goal, dry machining should be considered. But dry machining has some obvious disadvantages associated with tool overheating, excessive seizure, and problems with chip evacuation from the cutting zone. Alternative methods of cooling and lubricating without the use of cutting fluids are now being considered. Some methods of machining using minimum quantities or no coolant [7,9,26] should be mentioned: • Cooling with compressive air • The use of so-called minimum mist lubrication or minimum quantity lubrication (MQL) However, the use of MQL is only acceptable if the main tasks of the cutting fluid are resolved [27], that is, cooling, lubrication, and chip removal in the cutting process [28]. Mist cooling holds the promise of combining gases having good penetration into the cutting zone with liquids having good lubricity. The basis of the MQL method is as follows: very small quantities of high lubricity oil are applied in a focused manner onto the point of interest. A semidry lubricant supply system is sometimes also referred to as near dry machining (NDM). In North America this technology had its origin in the aerospace industry with a view to cutting difficult to machine materials [2]. The benefit of this process is that dry metal chips are formed that are easy to recycle. In these applications the chip removal functionality comes from a vacuum system. Minimum quantities of oil can reduce the risk of fire; they also significantly reduce the administration costs for the lubricants, reduce the need for waste treatment, eliminate the problem of maintaining the lubricating oil, and greatly reduce waste oil disposal [29].
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Handbook of Lubrication and Tribology Recommendations for Cutting and Grinding Fluids and Application Methods for Carbide Tools Carbide
Machined materials
Type of cut
Turning
Face milling
End milling
Drilling
Carbon and low alloy steels
All
Low alloy steels
Rough
F 0, 3–3.12; 2–2.10 F 0, 1–21
F 0, 2–2.10; 1.10–1.12 F 0, 2–2.10
F 3.1–3.3, 3.8–3.12 F 2.1, 2.2
Low alloy steels
Finish
F 0, 3.1–3.13
F 0, 3.1–3.12
F 3.1–3.12
Cast iron
Rough
M–F 0, 2.3–2.10
M–F 0, 2.3–2.10
Cast iron
Finish
M–F 3.8–3.12
M–F 3.8–3.12
Die steels
Rough
Die steels
Finish
Stainless steel (Austenitic)
Rough
Finish
Nickel-based alloys
Rough
Nickel-based alloys
Finish
Magnesium
All
Aluminum
All
Cooper and alloys
Rough
Cooper and alloys
Finish
Titanium
Rough
Titanium
Rough
F 2.1–2.10, 1.28 F 3.1–3.12, 1.28 F 0, 2.2, 2.13 3.1, 3.2, 2.13 F 3.1–3.3, 2.13 F 0, 1.13–1.21, 1.29 F 0, 2.3–2.10, 1.29 M–F 0, 1.22,1.23, 1.26 M–F 3.1–3.3, 2.1, 1.26 F 2.1, 2.2, 1.25 F 2.1, 2.2, 1.25 F 1.24, 2.11, 1.27 F 1.24, 2.11, 1.27
F 2.1–2.10, 1.28 F 3.1–3.12, 1.28 F 0, 2.2, 2.13
Stainless steel
M–F 0, 2–2.10, 1.24 M–F 0, 3.1–3.3, 2.1–2.2 F 0, 1.10–1.12, 1,28 F 0, 3.1–3.12, 1.28 F 0, 2.1–2.10, 1.31 F 0, 2.1–2.10, 1.31 F 1.13–1.21, 1.29 F 2.3–2.10, 1.29 M–F 0, 1.23–1.26, 3.20 M–F 3.3, 2.1, 1.26, 3.20 F 2.3–2.10, 1.24, 1.25 F 2.3–2.10, 1.24, 1.25 M–F 2.10, 3.4, 2.11 M–F 2.10, 3.4, 2.11
F 3.4–3.12, 2.3–2.10 F 3.4–3.12, 2.3–2.10 F 3.4–3.12, 2.3–2.10 F 0, 2.1, 2.2
F 3.1–3.3, 2.13 F 0, 1.13–1.21, 1.29 F 0, 2.3–2.10, 1.29 M–F 0, 1.22,1.23, 1.26 M–F 3.1–3.3, 2.1, 1.26 F 2.1, 2.2, 1.25 F 2.1, 2.2, 1.25 M–F 1.24, 2.11, 1.27 M–F 1.24, 2.11, 1.27
F 3.1–3.3, 3.8–3.12 F, H 3.4–3.12, 1.10–1.16 F, H 3.4–3.12, 1.10–1.16 F, H 3.4–3.12, 2.3–2.10 F, H 3.4–3.12, 2.3–2.10 F, H, I 1.13–1.21, 1.29 F, H, I 1.13–1.21, 1.29 F 0, 1.22, 1.23, 1.26 F 2.1–2.2, 1.26 F, H 2.1–2.2, 1.25 F, H 2.1–2.2, 1.25 F, H, I 1.27 F, H, I 1.27
Application method: F, flow; M, mist; H, hand; I, immersion; P, high pressure. Source: Silliman, J.D., Cutting and Grinding Fluids: Selection and Applications, 2nd ed., SME, Michigan, 2002, 216.
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HSS tapping F 1–1.2 F 1-1.16 F 1–1.16 F, H 0, 3.4–3.12, 2.3–2.10 F, H 0, 3.4–3.12, 2.3–2.10 F, H, I 1.5–1.21 F, H, I 1.5–1.21 F, H, I 2.3–2.10, 1.13–1.21 F, H, I 2.3–2.10, 1.13–1.21 F, H, I 1.17–1.21, 1.29 F, H, I 1.17–1.21, 1.29 F–H 0, 1.22, 1.26 F–H 0, 1.22, 1.26 F, H 1.24 F, H 1.24 F, P, H, I 1.34, 1.27 F, P, H, I 1.34, 1.27
Coolants and Lubricants in Metal Cutting TABLE 12.2
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Cutting Fluids Codes for Table 12.1
Code
Fluid type
Code
Fluid type
0 1 1.1 1.2 1.3 1.4 1.5 1.6 1.7 1.8 1.9 1.10 1.11 1.12 1.13 1.14 1.15 1.16 1.17 1.18 1.19 1.20 1.21 1.22 1.23 1.24 1.25 1.26
Dry Cutting oils Sulfurized oil, light duty Sulfurized mineral-lard oil, light duty Sulfurized mineral oil, light duty Sulfurized lard oil with chlorine, light duty Sulfurized oil, medium/heavy duty Sulfurized mineral-lard oil, medium/heavy duty Sulfurized fat compounded oil, medium/heavy duty Sulfurized mineral oil, medium/heavy duty Sulfurized lard oil, medium/heavy duty Sulfo-chlorinated mineral-lard oil, light duty Sulfo-chlorinated mineral oil, light duty Sulfo-chlorinated lard oil, light duty Sulfo-chlorinated mineral-lard oil, medium duty Sulfo-chlorinated mineral oil, medium duty Sulfo-chlorinated lard oil, medium duty Sulfo-chlorinated oil, medium duty Sulfo-chlorinated mineral-lard oil, medium duty Sulfo-chlorinated lard oil, heavy duty Sulfo-chlorinated oil, heavy duty Sulfo-chlorinated mineral oil, heavy duty Highly chlorinated lard oil, heavy duty Straight mineral oil Straight oil Mineral-lard oil, medium/heavy duty Mineral-lard oil, light duty Oil specially recommended for aluminum, magnesium alloys Oil specially recommended for titanium alloys Oil specially recommended for high temperature alloys Oil specially recommended for nickel-based alloys Oil specially recommended for cobalt base Oil specially recommended for stainless steels
1.32 1.33 1.34 2 2.1 2.2 2.3 2.4 2.5 2.6
Chlorinated mineral-lard oil Honing oil Tapping oil Emulsifiable oils (water miscible), all Water miscible oil, light duty Water miscible oil, medium duty Water miscible oil, heavy duty Sulfo-chlorinated water miscible oil, heavy duty Chlorinated water miscible oil, heavy duty Sulfo-chlorinated water miscible compound, heavy duty Water miscible compound, active sulfur, heavy duty Water miscible mineral oil Fatter water miscible oil Extreme pressure water miscible oil, heavy duty Water miscible oil specially recommended for titanium alloys Water miscible oil specially recommended for high nickel cobalt alloys Chemical (synthetic) fluid Chemical emulsion, light duty Water based chemical, light duty Water miscible petrochemical, light duty Chemical emulsion, heavy duty Sulfurized water-based chemical, heavy duty Chlorinated water-based chemical, heavy duty Water miscible, heavy duty Chemical (synthetic) fluid Chemical solution Chemical solution, oil based Chemical and oil solution, heavy duty Chemical and organic compound solution Chemical with extreme pressure and wetting agent, water miscible
1.27 1.28 1.29 1.30 1.31
2.7 2.8 2.9 2.10 2.11 2.12 3 3.1 3.2 3.3 3.4 3.5 3.6 3.7 3.8 3.9 3.10 3.11 3.12 3.13
Source: Silliman, J.D., Cutting and Grinding Fluids: Selection and Applications, 2nd ed., SME, Michigan, 2002, 216.
Another promising method is a “high jet” technique for applying coolant. This involves a high pressure stream of compressed air or a cutting fluid that is applied at high velocity to the cutting zone. A significantly improved tool life is observed compared with the regular free flow of cutting fluids [29].
12.4.2 Dry Machining Dry machining is a very challenging area. To achieve this goal the cutting tool has to meet the requirements of the cutting fluid on its own. The major method to develop dry machining conditions is through a complex approach including cutting tool design and process parameter selection [30,31], advanced tooling material selection [32–34], and the application of advanced surface engineering [35–42]. This approach has been used by cutting tool manufacturers such as Titex Plus Company (http://www.titex.com/). The cutting tool design features used include: • Small friction forces through the use of large rake angles and the use of heavily rounded cutting edges • Low friction via narrow cylindrical lands and strong back taper • Good chip evacuation due to special flute profiles
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To achieve the design features specified above, special micrograin carbide materials with high cobalt content are typically used to improve toughness and edge strength. The use of advanced coatings is mandatory for high speed machining operations. They serve a number of critical functions for dry machining, that is, thermal isolation and separation of tool and workpiece as well as tailoring the friction to the application. Traditional hard coatings such as TiAlN could be used with high efficiency for these conditions. The lubricity of the coatings can be further improved by the • Deposition of low friction coatings such as DLC or metal doped DLC, Mo2 S, and WC/C as an outer layer for TiAlN coatings • Deposition of advanced adaptive coatings such as TiAlN/VN or filtered arc TiAlN [26] as well as TiAlCrN and AlCrN, which exhibit lubricating properties at high temperatures We can conclude that cutting fluids play a vital role in machining productivity. The environmental and health implications for cutting fluids are major issues for the future. The development of biofriendly cutting fluids based on synthetic oils and intensive research in the field of minimum and dry machining are becoming increasingly more topical.
References [1] Byrne, G. and Scholta, E., Environmentally clean machining processes — a strategic approach, Ann. CIRP, 42, 1993, 471. [2] Astakov, V.P., Tribology in metal cutting, in Mechanical Tribology. Materials, Characterization, and Applications, Liang, H. and Totten, G., Eds, Marcel Dekker, New York, chap. 9, 2002, 307. [3] Braun, E.D., Bushe, N.A., and Buyanovsky, I.A., Fundamentals of Tribology (Friction, Wear and Lubricants), Center Science and Technique, Moscow, 1995, 778. [4] Garkunov, E.D., Tribo-Engineering (Design, Manufacturing and Machine Service), 5th ed., MSCA Publishing, Moscow, 2002, 632. [5] Goracheva, I.G., Mechanics of Frictional Interaction, Nauka, Moscow, 2001, 478. [6] Mang, T.A., New Generation of Non Hazardous and Environmentally Safe Metal Working Oils, ITC, Melbourne, 1987, 336. [7] Chichinadze, A.V., Fundamentals of Tribology (Friction, Wear and Lubricants), Mashinostroenie, Moscow, 2001, 664. [8] Shkolnikov, V.M., Fuels, Lubricants, Engineering Fluids. Selection and Applications, 2nd ed., Chemistry, Moscow, 1999, 596. [9] Sholom, V.J., Crioni, N.K., Shuster, L.S., and Migranov, M.S., Influence of active additives on antigalling properties of lubricants, in Proceedings of the International Conference on Reliability, Quality in Industry, Power Generation and Transportation, Samara State University Publishing, Samara, 1999, 64. [10] Shuster, L.S., Adhesive Interaction in Solids, Gilem, Ufa, 1999, 199. [11] Wang, Z.Y. and Rajurkar, K.P., Cryogenic machining of hard-to-cut materials, Wear, 238, 2000, 169. [12] Hong, S.Y., Ding, Y., and Ekkens, R.G., Improving low carbon steel chip breakability by cryogenic cooling, Int. J. Mach. Tools Manuf., 39, 1999, 1065. [13] Trent, E.M. and Wright, P.K., Metal Cutting, Butterworth-Heinemann, Boston, MA, 2000, 446. [14] Bailey, J.A., Friction in metal machining — mechanical aspect, Wear, 31, 1975, 243. [15] Williams, J.A. and Tabor, D., The role of lubricants in machining, Wear, 43, 1977, 275. [16] Williams, J.A., The action of lubricants in metal cutting, J. Mech. Eng. Sci., 19, 1977, 202. [17] Doyle, E.D., Horne, J.G., and Tabor, D., Frictional interactions between chip and rake surface in continuous chip formation, Proc. R. Soc. Lond. A, 366, 1979, 173. [18] De Chiffre, L., Mechanics of metal cutting and cutting fluids action, Int. J. Mach. Tool Des. Res., 17, 1977, 225. [19] De Chiffre, L., Frequency analysis of surface machined using different lubricants, ASLE Trans. 27, 1984, 220.
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Coolants and Lubricants in Metal Cutting
[20] [21] [22] [23]
[24]
[25] [26]
[27] [28] [29] [30]
[31] [32] [33]
[34] [35] [36] [37] [38]
[39] [40] [41] [42]
12-15
De Chiffre, L., Mechanical testing and selection of cutting fluids, Lubr. Eng., 36, 1980, 33. De Chiffre, L., What can we do about chips formation mechanics? Ann. CIRP, 34, 1985, 129. Brinksmeier, E. and Glabe, R., Ann. CIRP, 50, 2001, 385. Fox-Rabinovich, G.S., Kovalev, A.I., and Weatherly, G.C., Tribology and the design of surface engineered materials for cutting tool applications, in Modeling and Simulation for Material Selection and Mechanical Design, Totten, G.E., Xie, L., and Funatani, K., Eds, Marcel Dekker, New York, 2004, chap. 5. Rebinder, P.A. and Likhtman, V.I., Effect of surface-acting media on strains and ruptures in solids, in Proceedings of the Second International Conference of Surface Activity, Butterworth, London, 1947, 563. Silliman, J.D., Cutting and Grinding Fluids: Selection and Applications, 2nd ed., SME, Michigan, 2002, 216. Fox-Rabinovich, G.S., Weatherly, G.C., Dodonov, A.I., Kovalev, A.I., Veldhuis, S.C., Shuster, L.S., Dosbaeva, G.K., Wainstein, D.L., and Migranov, M.S., Nano-crystaline FAD (filtered arc deposited) TiAlN PVD coatings for high-speed machining application, Surf. Coat. Technol., 177–178, 2004, 800. De Chiffre, L., Function of cutting fluids in machining, Lubr. Eng., 44, 1988, 514. Chen, N.N. and Pun, W.K., Stresses at cutting tool wear land, Int. Mach. Tool Manuf., 28, 1988, 79. Sreejth, P.S. and Ngoi, B.K.A., Dry machining: machining of the future, J. Mater. Process., 101, 2000, 287. Balzer, S.A., Haan, D.M., Rao, P.D., Olson, W.W. and Sutherland, J.W., Minimizing the quantity of cutting fluid required for specific machining processes through the manipulation of process input parameters, J. Mater. Proc. Technol., 79, 1998, 72. Diniz, A.E. and Micaroni, R., Cutting conditions for finish turning process aiming: the use of dry cutting, Int. J. Mach. Tools Manuf., 45, 2002, 899. Knolcke, F. and Eisenblatter, G., Dry cutting. Ann CIRP, 46, 1997, 519–526. Braga, D.U., Diniz, A.E., Miranda, G.W.A., and Coppini, N.L., Using a minimum quantity of lubricant (MQL) and a diamond coated tool in the drilling of aluminium — silicone alloys, J. Mater. Proc. Technol., 122, 2002, 127–138. Andrews, C.J.E., Feng, H.-Y., and Lau, W.M., Machining of an aluminum/SiC composite using diamond inserts, J. Mater. Proc. Technol., 102, 2000, 25–29. Tonshoff, H.K. and Micaroni, R., PVD-coatings for wear protection in dry cutting operations, Surf. Coat. Technol., 96, 1997, 88. Lahers, M. and Jorgensen, G., Properties and dry cutting performance of diamond-coated tools, Surf. Coat. Technol., 96, 1997, 198–204. Kustus, F.M., Fehrehnbacher, L.L., and Komandiri, R., Nanocoatings on cutting tools for dry machining, Ann. CIRP, 46, 1997, 39. Renevier, N.M., Liobindo, N., Fox, V.C., Teer, D.G., and Hampshire, J., Coating characteristics and tribological properties of sputter-deposited MoS2 /metal composite coatings deposited by closed field unbalanced magnetron sputter ion plating, Surf. Coat. Technol., 123, 2000, 84. Harris, S.G., Vlasveld, A.C., Doyle, E.D., and Dolder, P.J., Dry machining — commercial viability through filtered arc vapour deposited coatings. Surf. Coat. Technol., 133–134, 2000, 383–388. Gresik, W., Zalisz, Z., and Nieslony, P., Friction and wear testing of multilayer coatings on carbide substrates for dry machining applications, Surf. Coat. Technol., 155, 2002, 37. Lahres, M., Muller-Hummel, P., and Doerfel, O., Applicability of different hard coatings in dry milling aluminium alloys, Surf. Coat. Technol., 91, 1997, 116–121. Klocke, F. and Kreig, T., Coated tools for metal cutting — features and applications, Ann. CIRP, 48, 1999, 515–525.
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13 Lubricating Industrial Electric Motors 13.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.2 General Motor Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
13-1 13-2
Fixed Speed Motors • Variable or Adjustable Speed Drives
13.3 Motor Lubrication System Design . . . . . . . . . . . . . . . . . . . .
13-2
Oil Lubrication • Grease Lubrication • Grease Bearings Shields and Seals
13.4 Motor Lubricant Selection . . . . . . . . . . . . . . . . . . . . . . . . . . . .
13-6
Oil-Lubricated Motors • Grease-Lubricated Motors • Summary Specification for a Typical Electric Motor Grease
13.5 Motor Lubrication Maintenance. . . . . . . . . . . . . . . . . . . . . .
Drew D. Troyer Noria Corporation
13-12
Oil-Lubricated Electric Motors • Grease-Lubricated Electric Motors
References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
13-20
13.1 Introduction More than any other prime mover, electric motors, in their various embodiments, drive industry. These motors are employed to drive pumps, fans, hydraulic units, cranes, and a host of other applications too numerous to name. A number of factors influence motor reliability, including design, manufacture, selection, installation, operating environment, and context and maintenance. In all instances, however, to operate effectively, efficiently, and dependably, electric motors must be properly lubricated. Recent investigations by the IEEE and the Electric Power Research Institute (EPRI) reported that 44% and 41%, respectively, of all motor failures are bearing related. Many of these failures can be attributed in some form to ineffective lubrication, most notably overlubrication, underlubrication, contamination, improper lubricant selection, and mixing of incompatible lubricants. Designed to assist plant engineers to lubricate electric motors more effectively, this chapter provides a general introduction to electric motors, details the various ways in which lubrication systems are designed for electric motors, provides guidelines for selecting lubricants for electric motors, addresses lubrication maintenance for electric motors, including the deployment of condition monitoring techniques and technologies, and suggests strategies to ensure proper lubrication when it is time to rebuild the motor.
13-1
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Handbook of Lubrication and Tribology
13.2 General Motor Design Electric motors convert electricity into rotary mechanical power. While it is beyond the scope of this chapter to detail all the various embodiments of electric motor design, a few essential elements must be addressed. Generally, motors fall into two categories, fixed speed and variable speed drives. Fixed speed industrial motors may be further divided into single-phase induction motors, three-phase induction motors, and three-phase synchronous motors. Variable or adjustable speed motor drives (VSDs or ASDs) may be further divided into alternating current (AC) powered variable frequency drives (VFDs) and direct current (DC) powered drives. Any of these motors may be mounted horizontally, vertically, or at angles in between. In all instances, motor components include a stator, or the stationary element of the motor, the rotor, which is the moving element of the motor, including the shaft, bearings, and a frame. Depending upon the design, the motor may also have slip rings, brushes and their rigging, a commutator, and a brushless exciter.
13.2.1 Fixed Speed Motors By far the most common motor employed in typical industrial facilities is the fixed speed AC induction motor. While numerous variations exist, most AC motors can be categorized as single-phase induction motors, three-phase induction motors, or three-phase synchronous motors. Induction motors operate on the principle that the flow of electricity into the motor creates flux lines that produce a magnetic field, which results in shaft rotation. The greater the number of flux lines, called flux density, the greater the force of rotation. The AC induction motor, the most common motor in industry, consists of a rotor mounted conducting loop. In a single-phase AC induction motor, electrical current is fed into the loop by brushes through slip rings. A soft iron core creates a magnetic field around the loop, which causes the loop to turn. As the current’s polarity reverses, momentum carries the loop until the electromagnetic force is reinitiated upon reaching the opposite polarity of the AC cycle. In other words, the loop is motivated by the magnetic field twice during each 360◦ cycle. In a three-phase AC induction motor, the loop is motivated by the magnetic field six times per 360◦ cycle, twice for each phase. Three-phase synchronous motors are employed when precise control over speed is required.
13.2.2 Variable or Adjustable Speed Drives Variable frequency drive AC motors and DC motors may be employed to overcome the fixed speed limitation of conventional AC motors. DC motors operate by first converting AC electricity to DC electricity using a rectifier, which controls the phase angle and, thus, the power and speed of the drive. The DC current is then applied to a DC motor that employs a split ring commutator that switches the direction of the DC current every half turn to maintain shaft rotation. Modifying the DC current alters the motor’s speed. Employing solid state electronics and microprocessors, VFDs convert fixed frequency AC power input into a variable frequency AC output. This provides users with variable control over the speed at which the motor runs, which is important in some applications. More complex than DC-powered ASD controllers, the VFD motor controller must first convert the input AC power into DC power, after which an inverter changes the DC power into adjustable frequency and voltage AC power. This provides both the benefit of variable speed operation and the simplicity and reliability of using an AC motor. Recent research suggests that approximately 21% of industrial motors are variable speed (the research does not differentiate between VFD AC motors and DC motors).
13.3 Motor Lubrication System Design Bearings are the critical motor components to the lubrication engineer. Some industrial motors are equipped with sleeve bearings (also known as plain bearings, journal bearings, or bushings), but
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Lubricating Industrial Electric Motors
Bearing type and lubrication method
13-3
Horizontal direct connected 200 hp and 200 to below 2000 hp
Above 2000 hp
Horizontal belt connected 200 and below
200 to 200 hp and 2000 hp below
Vertical 200 to 2000 hp
Above 2000 hp
Rolling element ball type Rolling element roller type Sleeve bearing, oilring-lubricated Sleeve bearing, oil bath-lubricated Sleeve bearing/forced lubrication Thrust bearing, rolling element, ball oil-lubricated Thrust bearing, rolling element, ball, grease-lubricated Thrust bearing, tilting pad Thrust bearing, disc type
FIGURE 13.1 The selection of bearing configuration, lubricant, and lubrication method depends upon the motor’s position and horsepower rating.
most — more than 80% according to recent market research — are equipped with rolling element or antifriction bearings, which may be either ball-type bearings or roller-type bearings. The choice of bearing type and lubricant delivery mechanism depends upon application details, including motor horsepower, connection to the driven components (either belt or direct coupled), shaft orientation (vertical or horizontal), and the manner in which the bearings are lubricated (Figure 13.1).
13.3.1 Oil Lubrication As with grease, there are advantages and disadvantages associated with lubricating with oil. In general, oil-lubricated bearings can operate at higher speeds than grease-lubricated motor bearings and they last longer than grease-lubricated bearings because oil typically contains fewer contaminants than grease (due to settling and filtration) and oil-lubricated bearings usually run cooler than their grease-lubricated counterparts. However, oil is more difficult to contain in the housing, which poses volume management problems as well as the risk of leakage on the product. Further, if set up properly, oil analysis for machine condition monitoring can be employed on oil-lubricated motors. Oil-lubricated systems typically require more complex designs and more routine maintenance than grease-lubricated motor bearings. And, in most instances, oil runs off of the bearings easily when the motor is not running, so the risk of dry start wear is higher with oil-lubricated systems. Other advantages and disadvantages associated with oil vs. grease lubrication relate to the lubricant application method. The most common oil lubricant application methods are discussed below. The discussion is not exhaustive and many variations on the themes presented are possible. 13.3.1.1 Bath Oil Lubrication As the name implies, bath-lubricated bearings operate in a bath of oil. The movement of the bearing itself serves to distribute the oil to the bearings. Maintaining the proper level is critical in bath-lubricated motors. If the level is too low, the bearings are starved of lubrication. If the level is too high, foaming, aeration,
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and frictional heat generation occur. Bath-lubricated motors are often equipped with constant-level oiler devices that automatically replenish the oil if the level drops. Bath oil lubrication is most commonly employed in vertical motors. 13.3.1.2 Ring Oil Lubrication Ring-lubricated motors employ a slinging device — either a ring, which is not affixed to the shaft, or a collar, which is affixed to the shaft. The ring or collar serves to “sling” oil onto the bearings. Maintaining the proper oil level is critical in ring- or collar-lubricated motors. If the level is too low, the bearings are starved of lubrication. If the level is too high, foaming, excessive aeration, and frictional heat generation occur. Constant-level oil devices, when properly maintained, can ensure that the proper level is maintained in ring- and collar-lubricated motors. 13.3.1.3 Force Circulation Lubrication Forced oil lubrication is preferred in many industrial applications, especially for large horsepower motors driving critical application machines. Force circulation lubrication provides the bearings with a supply of lubricating oil, typically from a reservoir. This lubrication method is often referred to as dry sump lubrication because the bearing housings themselves do not store oil. Rather, transient oil is supplied to the bearing to lubricate and cool the bearing and remove contaminants, after which it returns to the tank. If properly designed with respect to flow, tank size, and design and to temperature and contamination control devices, force circulation lubrication provides maximum bearing life. Due to the incorporation of a tank, pipes or hoses, lube pumps and motors, cooling systems, contamination control system, ans so on, force circulation-lubricated motors are more complex and expensive to design, install, and maintain, and they occupy a larger footprint in a plant. 13.3.1.4 Oil Mist Lubrication Oil mist is a sophisticated form of motor bearing lubrication whereby an oil mist generator atomizes, or nebulizes, oil into tiny droplets that can be supported by air. The mist “fogs” to the machine through piping that connects the oil mist system to the lubricated component. Upon reaching the component, the mist is reclassified for application to the components. Mist systems may be dedicated to a single motor or employed to lubricate multiple components. Typically, the mist is a once-through lubrication, but it can be regenerated. The oil is prefiltered before being atomized. Because the oil is fresh and clean, mist provides excellent lubrication. Likewise, mist generates a minimal amount of fluid friction-related energy losses. Mist lubrication does not cool bearings as well as forced circulation lubrication, and there is a concern about stray mist. Stray mist is atomized oil mist that manages to escape the bearing housing, which can pose health concerns if inhaled, and can gum up the motor’s housing if excessive. Likewise, as with any centralized lubrication system, if a single mist system is employed to lubricate multiple, otherwise independent systems, the mist system becomes a critical system from a reliability engineering perspective and must be carefully monitored and maintained. Mist systems are employed widely to lubricate pumps in petrochemical applications.
13.3.2 Grease Lubrication Easily the most common method of motor lubrication, grease lubrication offers several advantages, including immediate availability of grease during startup, reduced risk of lubricant escaping the bearing to contaminate the product (particularly important in food and pharmaceutical applications), and better exclusion of contaminant ingestion past the housing seals. Key disadvantages include lower bearing speed limitations, increased heat generation, inability to remove contaminants once ingress has occurred, increased compatibility concerns due to the thickener, thickener performance degradation, and generally shorter oxidative life of the base oil. Further, because of the difficulty in obtaining a representative sample, oil analysis is rarely employed for grease-lubricated motors. Motors may be greased manually or automatically utilizing a single-point or multiple-point lubrication system. Typically, motors equipped
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Lubricating Industrial Electric Motors
13-5 Zerk
Bearing Supply tube Inner bearing cap
Grease reservoirs
Drain tube
Plug or pressurerelief vent
FIGURE 13.2 In grease-lubricated motors, grease enters the supply tube through a grease fitting, usually a Zerk, and flows into the exterior grease reservoir, where it is made available to the bearing. Excess grease relieves through the drain tube.
with rolling element bearings are grease-lubricated (both horizontally and vertically mounted). It is unusual to encounter greased-lubricated motors equipped with sleeve bearings. As depicted in Figure 13.2, when applied, grease flows from its source (a grease gun if manual lubrication is employed) to the bearing through the supply tube via an access port, typically a Zerk-type or buttonhead-type fitting. The grease enters the outer reservoir, where it is stored before entering the bearing itself. The shields may be mounted either facing the grease supply or facing the motor. Also, some bearing housings are equipped with metering plates. After passing through the bearing, grease is stored in the inner reservoir of the housing, where it is contained by a seal, often a labyrinth. The motor is typically equipped with a pressure relief tube, which facilitates the exit of excess grease. The relief tube itself may be equipped with a relief-type plug set to relieve at 1 to 5 lb/in.2 to avoid the need to remove and replace the plug (these are not allowed in some areas requiring explosion proofing). If the housing is not equipped with a relief tube, it may be equipped with a relief-type Zerk fitting to avoid the need to remove the grease fitting after applying grease.
13.3.3 Grease Bearings Shields and Seals In addition to selecting antifriction bearing type, ball or roller, the choice of seal or shield configuration is also important. Options include open bearings, shielded bearings, and sealed bearings (Figure 13.3). 13.3.3.1 Open Bearings Open bearings, as the name suggests, have no seals or shields. The open bearing configuration runs cooler than either sealed or shielded bearings, poses no risk for shield collapse into the motor, and is easy to regrease. However, open bearings pose the greatest risk for grease churning, entry of contaminants and hardened thickener into the bearing, and entry of grease into the motor’s windings, which compromises
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Open bearing
Sealed bearing
Shielded bearing
Outer race
Outer race
Outer race
Inner race
Inner race
No shields or seals
Seal (soft elastomer) fixed to outer race, with lip seal contact at inner race
FIGURE 13.3
Shield Inner race (steel) Gap where grease can enter bearing (0.005 to 0.015 in.) (125 to 375 mm)
Motors may be configured with open, sealed, or shielded bearings.
heat dissipation and decreases the life of motor insulation. A recent study revealed that among those industrial respondents stating a preference, 9% prefer an open bearing configuration for their industrial motors. 13.3.3.2 Shielded Bearings Shielded bearings are equipped with overhung metal shields that are affixed to the outer race section of the bearing. At the inner race, a gap, or annulus, of approximately 125 to 375 µm allows grease to enter and exit the bearing cavity. Bearings may be equipped with a single shield facing the grease supply, a single shield facing the motor, or double shields, one on each side of the bearing. Advantages of shielded bearings include reduced churning of grease, regulated flow of grease to the bearings, restricted entrance of contaminants and hardened thickener, and reduced risk of grease entry into the motor’s windings. However, shields cause the bearings to run hotter than open bearings and, for shields that face the grease supply, there is a risk of pushing the overhung shield into the bearing, so extreme caution must be applied when regreasing shielded bearings. Approximately 64% of industrial respondents indicated a preference for shielded bearings, particularly single-shielded bearings with the shield facing the motor. 13.3.3.3 Sealed Bearings Sealed bearings are equipped with elastomer seals that completely enclose the bearing. Sealed bearings must be prelubricated before being sealed and they may not be relubricated. These bearings are often called “lubed-for-life” bearings. Sealed bearings will not last as long as properly maintained motor bearings that enable relubrication because the limiting factor of the bearing’s life is the oxidative life of the base oil in the grease. Sealed bearings do not facilitate the release of heat, so they tend to run hotter than both open and shielded bearings. Of those responding, 27% of industrial users stated a preference for sealed bearings, primarily because less routine lubrication maintenance is required for relubrication and there is no risk of overlubrication. The use of sealed, lubed-for-life bearings is typically limited to small motors, usually less than 10 Hp. However, they have in some instances been employed on larger motors.
13.4 Motor Lubricant Selection As with any lubricant selection decision, motor bearing lubricant selection must be performance property driven. Applicable performance properties may depend upon the bearing type (sleeve vs. antifriction), the operating context (including load, speed, and temperature), and lubricant delivery mechanism (such as grease, mist, force circulation, etc). Performance property requirements relative to the application in turn drive the choice of lubricant, including base oil type and viscosity, additive system requirements and thickener, in the case of grease. The first decision, of course, is the decision to employ grease or oil. As previously discussed, there are advantages and disadvantages associated with each. However, depending upon the bearing size, speed, and type, conventional grease may not be an option. The angular velocity, or Dm N , of the bearing is
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an important variable in the decision. The Dm N may be easily calculated using the following equation: Dm N =
N (Di + Do ) 2
Where Di is bore diameter in millimeters; Do is outer diameter in millimeters; N is speed in rotations per minute; and Dm is mean diameter in millimeters. A motor equipped with radial ball bearings may be lubricated with conventional grease if the speed factor is less than 340,000. If the motor is equipped with spherical rolling element bearings, conventional grease is suitable if the speed factor less than 145,000. If the speed factors exceed the limit for the application, the bearing will run hot and/or the grease will be damaged. If the bearing cannot be lubricated with conventional grease due to speed factor limitations, a specialty high Dm N grease may be an option. Otherwise, it must be lubricated with bath oil, force circulation oil, mist, drip oil, or spray, all of which provide higher speed factor limitations. The Dm N is versatile calculation. In addition to the decision about whether to lubricate with grease or oil, it is employed in the selection of base oil viscosity and grease consistency and is a factor in the base oil selection decision. A simpler version of the angular velocity calculation, called the DN , can be simply calculated by multiplying the bearing bore diameter (mm) by the speed (rpm). However, I prefer the Dm N variation.
13.4.1 Oil-Lubricated Motors Whether it be for sleeve-type or antifriction-type bearings, the following performance properties are important in the choice of lubricant for an oil-lubricated motor (not in order of importance). Where appropriate, applicable standards from the American Society of Testing and Materials (ASTM) and the International Organization for Standardization (ISO) are identified. 13.4.1.1 Air Release (ASTM D3427) Oil-lubricated motors, particularly bath-, ring-, and collar-lubricated motors, agitate the oil. To maintain good film strength and lubrication, it is important that the air release characteristics of the oil be exceptional. Air release characteristics are primarily influenced by the oil’s viscosity surface tension. Highly refined mineral oils and synthesized hydrocarbon base oils typically offer high natural surface tension. Typically, the formulated electric motor lubricant will be enhanced with an additive to increase surface tension (and interfacial tension). The presence of water and other impurities will impede the release of air. 13.4.1.2 Foam Stability (ASTM D892) As with air release, the tendency of motor bearings to agitate the oil requires that the lubricant possess excellent foam suppression characteristics. The oil’s resistance to foaming is dependent upon the oil’s viscosity and its surface tension. Again, highly refined mineral oils and synthesized hydrocarbon base oils offer high surface tension and are a good choice for typical electric motor oil lubrication application. Likewise, water and other polar impurities serve to induce foaming. Industrial electric motor oils are typically formulated with an additive that enhances foam suppression characteristics. 13.4.1.3 Oxidative Stability (ASTM D943 and ASTM D2272) Industrial motors often operate for long periods of time. In order to complete long run cycles, the oil’s oxidation stability must be exceptional. Likewise, bath-, ring-, and collar-lubricated motors often run hot and highly aerated, further increasing the need for oxidative stability. Highly refined mineral base oils and synthesized hydrocarbon base oils make a good choice for industrial motor oils. The formulated lubricant should be enhanced with an oxidation inhibitor, which increases the lubricant’s oxidative life.
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13.4.1.4 Rust and Corrosion Protection (ASTM D665 and ASTM D130) While the oil in industrial motors is typically free of water contamination, the lubricant should be formulated to provide protection against rust and acid corrosion. This requirement is particularly important in those applications where water contamination is a risk. 13.4.1.5 Thermal Stability (ASTM D2070) Thermal stability refers to the lubricant’s ability to resist thermal breakdown that leads to coking and varnishing. Thermal failure occurs when the oil contacts hot surfaces. Bearings, particularly high-speed bearings, often have sufficient internal temperatures to induce localized thermal failure. Thermal stability is primarily a function of the base oil type. For industrial electric motors, highly refined mineral oils and synthesized hydrocarbons usually offer sufficient thermal stability. However, in extreme applications, specially formulated oils that employ more exotic base oils, such as silicone, silicate ester, or polyphenyl ether, might be required. 13.4.1.6 Antiwear/Extreme Pressure (ASTM D2782 and ASTM D2783) Typically, oil-lubricated motor bearings do not require antiwear (AW) or extreme pressure (EP), or antiscuff, enhancement. Motors fitted with sleeve bearings usually run on a hydrodynamic fluid film. Occasionally, oil-lubricated motor bearings require AW or EP enhancement if the bearing loads are high or the speed is slow. The presence of additives that enhance antiwear and extreme pressure performance can adversely affect the air release and foam stability properties of the oil. 13.4.1.7 Filterability (ISO 13357) Filterability refers to the ease with which a lubricant passes through a filter membrane. Like air release and foam stability, filterability is influenced by the oil’s surface tension/interfacial tension. Again, highly refined mineral base oils and synthesized hydrocarbon base oils offer the best filterability performance. Filterability is most important on motors with force circulation lubrication where a filter is employed for contamination control purposes. 13.4.1.8 Other Performance Properties Other performance properties may be important in oil-lubricated electric motor applications on a caseby-case basis, including hydrolytic stability, biodegradability, fire-resistance rating, food-grade rating, and aqueous toxicity rating. 13.4.1.9 Base Oil Type, Viscosity, and Viscosity Index (ASTM D445 and ASTM D2270) The most common base oil choice for oil-lubricated electric motors is mineral oil, typically highly refined mineral oil that offers excellent oxidative and thermal stability and low volatility. Synthesized hydrocarbonbased base oils are also commonly used, particularly where longer oxidative life, high viscosity index, and extremely low volatility are required. Occasionally, other base oils are employed where unusual operating conditions warrant their use. A detailed investigation of base oil viscosity selection for oil-lubricated industrial motors is beyond the scope of this chapter. However, the rules for viscosity selection are the same as for any sleeve or antifriction bearing applications. Load, speed, and operating temperature dictate the selection. For antifriction bearings, the speed factor, or DN, is used to estimate the required viscosity. Typically, oil-lubricated industrial motors (sleeve and antifriction bearing) require viscosity grades ranging from 32 to 150 cSt (mm2 sec−1 ) at 40◦ C. Occasionally, however, lower or higher viscosity oil is required where warranted by operating conditions. If the viscosity is too low, the dynamic film may be insufficient to provide separation component surfaces. If the viscosity is too high, cavitation and oil whip may be observed in sleeve bearings. In rolling element bearings, excessive viscosity can result in ball/roller skidding and viscous drag. In both cases, excessive viscosity results in frictional energy losses that manifest in the form of heat, which shortens the life of the bearings and the lubricant. High viscosity index (VI) is important only where the motor operation involves wide temperature changes or extremely
© 2006 by Taylor & Francis Group, LLC
Lubricating Industrial Electric Motors Desired property Compatibility with machine and other lubes
13-9 Primary formulation factor
Additives
Film strength, friction, and wear control Thermal and oxidation stability
Thickener type and concentration
Corrosion control
Influences six important grease properties
Consistency, shear strength, mobility
Base oil type
Examples EP Antioxidants Tackifiers Solid additive VI improvers Rust inhibitor Lithium complex Barium Polyurea
Mineral PAO Di-ester
Separation and resistance to G forces Base oil viscosity and VI Resistance to water
ISO VG 100 ISO VG 220 ISO VG 460
FIGURE 13.4 Grease is formulated using a combination of base oil, additives, and thickener, which all contribute to performance properties. Focus on performance properties when selecting a lubricant for an industrial electric motor.
low-temperature startup. No negative ramifications are associated with the selection of a high VI base oil motivated by oxidative stability, thermal stability, air release performance or other reasons. 13.4.1.10 Additive System Industrial electric motor oils are typically formulated with a simple rust and oxidation (R&O) additive system that includes a foam suppression component. Occasionally, AW or EP additives are required in oil-lubricated electric motors equipped with antifriction bearings, but employ caution when making such a decision.
13.4.2 Grease-Lubricated Motors With few exceptions, grease-lubricated electric motors are equipped with antifriction bearings. As with the selection of oil, the selection of electric motor grease must be dictated by performance requirements. In grease, performance properties are influenced by the additive system, the thickener type and concentration, the base oil type, and the base oil viscosity and VI (Figure 13.4). Grease performance properties as they pertain to lubricating electric motors are discussed below. 13.4.2.1 Oxidation Stability (ASTM D942) As with oil-lubricated electric motors, oxidation stability is critical in grease-lubricated electric motor bearing applications. Greased motor bearings run hotter and in a more aerated state than oil-lubricated motor bearings due primarily to the thickener, which causes the lubricant to retain heat and air. For motors that are regreased, the lubricant must resist oxidation long enough to get to the next lubrication cycle. If the base oil degrades, the bearing’s reliability is put at risk. Likewise, the remnants of oxidation will put the fresh oil at risk. It is impossible to clear out the old grease without disassembly. For sealed bearings that are lubed for life, the oxidative life of the grease is the limiting factor that determines the life of the bearings, and hence the motor. For motors with sealed bearings, or where the regrease interval is long, select a grease that is formulated with a base oil that has excellent resistance to oxidation, such as a synthesized hydrocarbon or a highly refined mineral oil that is formulated with an oxidation inhibiting additive. The thickener also influences oxidation stability. For electric motors, polyurea and
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lithium complex offer excellent resistance to oxidation. Aluminum complex thickener also offers excellent resistance to oxidation, but it is less commonly used in motors due to its generally poorer mechanical shearing performance. 13.4.2.2 High-Temperature Performance (ASTM D566 and ASTM 2265) In grease-lubricated motor bearings, there are two major aspects of high-temperature performance, base oil performance and thickener performance. First, the base oil should resist thermal degradation and evaporation. In most industrial applications, grease formulated from highly refined mineral and synthesized hydrocarbon base oils offers adequate thermal stability performance. However, in some extreme temperature applications, grease formulated with more exotic base oils, such as silicone, silicate ester, or polyphenyl ether, might be required. The grease thickener must posses a high dropping point temperature, which is the temperature at which the grease begins to transition from a semisolid to a liquid. The dropping point, to a large extent, determines the temperature limitation of the grease. For electric motors, lithium complex, polyurea, and calcium sulfonate thickeners offer the best high-temperature performance. Barium complex thickener also offers excellent high-temperature performance but is less popular for motor applications due to its poor shear stability and bleed resistance. Aluminum complex thickener offers good high-temperature stability but tends to harden when exposed to heat over time. 13.4.2.3 Antiwear/Extreme Pressure (ASTM D2509 and ASTM D2596) While the film thickness is determined by the base oil, enhancements in lubricity are almost entirely attributable to the additive system. If the ratio of viscosity at operating temperature to required viscosity for the bearing, which is referred to as the coefficient Kappa, is less than one, an AW or EP additive may be required in the grease formulation to reduce wear in electric motors with antifriction bearings. However, due to the speed at which typical industrial motors operate, an aggressive EP is rarely required. Typically, the viscosity of the base oil is sufficient to lubricate the bearings effectively. Exercise caution to avoid unnecessary use of EP additives, which can cause corrosion, in electric motor applications. Some thickeners, such as calcium complex, offer intrinsic antiwear protection. 13.4.2.4 Rust and Corrosion Protection (ASTM D1743) Resistance to rust and corrosion is of greatest importance in grease-lubricated motors that run intermittently. In these applications, water can condense onto bearing surfaces and cause localized corrosive etching. Motors that run continuously tend not to accumulate water. Rust and corrosion inhibition is provided primarily by the lubricant’s additive system, 13.4.2.5 Bleed Resistance (ASTM D1742 and ASTM D1263) Bleed, or separability, is the term used to describe separation of the base oil from the thickener in the grease. This is a critical performance property in electric motor applications. When the oil bleeds easily from the thickener, it can leak out of the bearing housing, often past the labyrinth and into the motor. If the grease has poor bleed resistance, hard, caked-up thickener gets left behind, making it difficult or impossible to relubricate the bearing. Bleed resistance is determined primarily by the thickener and the base oil viscosity. Higher viscosity base oil tends to resist bleeding better than lower viscosity base oil. Aluminum complex thickener offers excellent bleed resistance, but tends to harden over time. Polyurea, lithium complex, and calcium sulfonate thickeners offer good bleed resistance performance, polyurea being slightly better. 13.4.2.6 Shear and Worked Stability (ASTM D217, ASTM D217a, and ASTM D1831) Shear and worked stability refers to the grease’s ability to resist softening over time. Over time, grease softens as a result of the mechanical shearing of the thickener structure. Electric motor grease must retain its mechanical integrity to avoid running out of the motor. Shear and worked stability are influenced both by the base oil and the thickener, but the thickener is the primary contributor. Lithium complex and polyurea thickeners offer the best resistance to shear and mechanical worked stability.
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13.4.2.7 Other Performance Properties On a case-by-case basis, other performance properties may come into the choice of electric motor grease, including hydrolytic stability, biodegradability, fire-resistance rating, food-grade rating, aqueous toxicity performance, and conductivity. 13.4.2.8 Base Oil Type, Viscosity, and Viscosity Index (ASTM D445 and ASTM D2270) As in the case of oil-lubricated industrial motors, the base oils most commonly used in industrial motor grease are highly refined mineral oil and synthesized hydrocarbon synthetic oil, which offer excellent thermal and oxidative stability. In some instances, grease formulated with specialty base oils is required, particularly in high-temperature applications or environmentally sensitive areas. Viscosity index is important only in applications where the motor is subjected to frequent starts and stops. A detailed discussion about viscosity selection is beyond the scope of this chapter, but the guidelines for selecting the viscosity for any antifriction bearing generally apply for motors so equipped. The speed factor, or Dm N formula, previously discussed is typically employed to guide the base oil viscosity selection. For most grease-lubricated industrial electric motors, the appropriate viscosity is typically 100 to 150 (mm2 sec−1 ) cSt at 40◦ C. However, this can vary based upon the speed factor and the operating context of the motor. Many industrial users are unaware of the base oil viscosity of the grease they use to lubricate electric motors, thinking that consistency is the most important property on which to focus. However, it is the base oil and the additive that lubricate the bearings in electric motors. The thickener, which is the main determinant of consistency, primarily serves to deliver the base oil and additives to the bearings. As a result, it is common to see multipurpose grease used to lubricate electric motors. Often, multipurpose grease is formulated with a higher viscosity than is required to lubricate electric motors, which run at a much higher speed than other grease-lubricated industrial equipment that employs antifriction bearings (e.g., conveyor systems). As a result, the bearings run hot, which wastes energy, shortens the grease life, and reduces bearing life. 13.4.2.9 Thickener Type and Consistency Several thickener types can be used in electric motor grease. Polyurea- and lithium-complex-thickened greases generally offer the most balanced overall performance with respect to the important characteristics previously discussed. Almost 80% of those responding to a recent market survey into industrial motor lubrication practices use either polyurea- or lithium-complex-thickened grease. The percentages were split approximately evenly between the two. Lithium complex offers slightly better shear stability performance and is conductive. Polyurea offers slightly better resistance to oxidation and is nonconductive. Both are excellent allround greases. Polyurea is the preferred choice in food-grade applications, where lithium complex is not suitable. Lithium complex grease offers better water washout performance. Motor grease typically has a worked penetration, per ASTM D217, of between 350 and 250 (each ASTM D217 penetration number equals 100 µm or 0.1 mm), which equates to a National Lubrication Grease Institute (NLGI) number 1.5 to 3. The most common consistency is NLGI #2, which has a worked penetration of between 265 and 295. The consistency of the grease is affected by both the thickener and the base oil viscosity, but the thickener has the primary influence. The consistency number does not directly relate to the dynamic film thickness that will be achieved in the bearing. This is determined primarily by the base oil viscosity, VI, and operating temperature.
13.4.3 Summary Specification for a Typical Electric Motor Grease In summary, a typical electric motor grease that is suitable for most industrial applications possesses the following characteristics: • Lithium complex or polyurea thickener • High-quality mineral or synthesized hydrocarbon base oil
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• Consistency of between 250 and 350 ASTM D217 penetration points (NLGI #1.5 to #3, with NLGI #2 being most common) • Base oil viscosity of between 100 and 150 cSt (mm2 sec−1 ) at 40◦ C • High dropping point temperature, at least 200◦ C (∼400◦ F) • Low bleed characteristics • Excellent resistance to high-temperature oxidation • Good low-temperature torque characteristics • Good AW performance, but not EP 13.4.3.1 Initial Fill Volume The literature related to initial fill volume for grease-lubricated motors with antifriction bearings is divergent. However, the versatile speed factor can again be employed to derive general guidelines. If the speed factor is less than 50,000, fill the bearing to capacity. If it is 50,000 to 200,000 fill the bearing to 50 to 70% of capacity. If the speed factor exceeds 200,000, fill the bearing 10 to 40% of capacity. These are general guidelines, so use them judiciously. Stay on the low side of the range when filling motors equipped with spherical roller bearings or thrust-type ball and roller bearings.
13.5 Motor Lubrication Maintenance Once the motor is selected, filled, and installed in the application, our attention must turn to maintenance. If properly maintained, motors will provide years of trouble-free operation. If neglected, motors will serve as a continuous source of expense and frustration. Clearly, lubrication maintenance is a critical part of any effort to manage electric motor reliability. This section provides guidelines for proper maintenance of oil- and grease-lubricated electric motors, including instructions for the motor rebuild shop. In all instances, it is necessary to get the right lubricant in the motor. A standard best practice is to tag all the machines and dispensing devices to identify the appropriate lubricant. Tags that employ color or shape coding schemes are most effective because of their intuitive nature. Also, tools and devices required for performing lubrication maintenance should be kept clean, in an orderly fashion, and in good working condition. As with all maintenance activities, select lubrication tasks carefully based upon the machine type, criticality, failure history, operating context, and environmental conditions. Clearly define the lubrication maintenance plan in detailed written procedures. These procedures serve to codify the lubrication or reliability engineer’s intentions, assure that practices are completed with consistency, enable continuity regardless of who is in the lube tech role by defining expectations and training requirements, and serve as the basis for continuous improvement.
13.5.1 Oil-Lubricated Electric Motors While oil-lubricated electric motors represent the minority of all motors in an industrial facility, they are disproportionately represented in high-horsepower and high-criticality applications. They tend to drive the most critical, unspared industrial processes. Likewise, the motors themselves are often not stored in the parts warehouse due to high cost; lead time can be long if a new one must be ordered and expeditious rebuilds require careful planning. Sometimes, removing these motors requires cranes, rooftop access, etc. Suffice to say that good lubrication maintenance for these motors provides a healthy economic payback. 13.5.1.1 Sensory Inspections The most basic form of care is the routine inspection. Inspections should be purposeful and systematic, employing checklists that clearly define nonconformance for each of the applicable conditions. I prefer inspection checklists that employ simple yes or no answers (e.g., lube oil pressure is between 22 and
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24 psi — yes or no). The nonconforming condition should always be either yes or no to avoid confusion associated with the changing polarity of questions. 13.5.1.1.1 Level The most basic inspection for an oil-lubricated electric motor is to ensure that the oil is at the appropriate levels in the sump for bath-, ring-, and collar-lubricated motors, and in the tank for circulating oil systems. A simple level indicator or bulls-eye is sufficient to accomplish this inspection. If the level is low, the technician should be directed to adjust the level and record the approximate volume of oil required to restore the sump or tank to its appropriate level. The technician should note abnormal darkening of the oil, which can indicate thermal or oxidative failure, as well as haziness or cloudiness, which might suggest water contamination or excessive aeration. 13.5.1.1.2 Vent Breather Condition If the tank or sump employs a vent filter to exclude contamination (which it should), check it periodically to ensure that the breather is not full or saturated (in the case of desiccant-type breathers). Ideally, the vent breather will be equipped with a vacuum indicator, which alerts the technician that the element requires replacement. Desiccant changes color when it becomes saturated, which alerts the technician to service the element. 13.5.1.1.3 Bottom Sediment and Water Bath-, ring-, and collar-lubricated motors may be equipped with a bottom sediment and water (BS&W) bowl to collect water, sludge, and debris. A simple inspection of this device alerts the technician to abnormal thermal and oxidative lubricant failure, water contamination, and aggravated wear debris generation. Most of these devices are equipped with a needle gauge so water can be cleared immediately while the technician searches for the cause of ingress. 13.5.1.1.4 Ring and Collar Operation Most motors are equipped with an inspection port to ensure that rings and collars are effectively slinging lubricant to the bearings. When conducting this inspection, the technician should look for evidence of excessive aeration or foaming. Likewise, he or she should inspect the gasket on the inspection port and service it if required. 13.5.1.1.5 Fill Cap The fill cap should be inspected to ensure that it is securely in place. It should not be removed, however, unless necessary to adjust the oil level or change the oil. When it is removed, inspect it to ensure that its gasket or seal is in good condition and service as required. 13.5.1.1.6 Pressure Gauge For circulating oil systems, check the oil pressure to the motor bearings routinely. Loss of pressure could indicate a leak or diminished volumetric efficiency of the lube oil pump. 13.5.1.1.7 Temperature Check the temperature each time the motor is inspected while it is operating. This can be accomplished with a dedicated temperature gauge, a noncontact thermometer (be sure you are consistent with the measurement procedure), or a thermography camera. The thermography camera is the most expensive and time consuming, but in the hands of a skilled technician, provides a great deal of useful information. 13.5.1.1.8 Pressure Differential Indicator For motors with force circulation lubrication, filters are typically employed inline or in a kidney loop configuration. Routinely inspect the pressure differential (psid) indicator or gauge and change the filter as required. If the filter has been in service for an unusually long period of time with no movement in the differential pressure, there is a good chance that the element is not functioning properly. Replace the element and inspect the used one. If the problem is chronic, contact the filter supplier. The element may not be appropriate for the application.
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13.5.1.1.9 Magnetic Chip Collector The motor can be equipped with a magnetic chip collector or magnetic drain plug. Inspect the collector for abnormal accumulation of wear debris. Usually, this inspection must be done when the motor is not operating to avoid damage resulting from loss of the lubricant upon removal of the plug. 13.5.1.1.10 Leakage Record any evidence of abnormal leakage of oil out of the bearing, which might indicate that the seals are worn and might require replacement. Or, for circulating oil systems, excessive leakage can be caused by an excessive flow rate. 13.5.1.1.11 Flow For a circulating system, confirm flow to the bearings, either visually or with a flow metering device. 13.5.1.1.12 Constant Level Oiler Verify that the constant level oil is filled to the appropriate level and that it is adjusted to the appropriate level (if adjustable). If the constant level oiler fails to maintain the right level, its adjustment device may be malfunctioning, installed incorrectly, or installed on the wrong side of the sump. The motor bearings should turn in the direction of the constant level oiler to reduce the risk that a vacuum will disrupt the constant level oiler’s hydraulic lock and result in overfilling the motor bearing’s sump. 13.5.1.2 Oil Analysis Oil analysis is a powerful diagnostic tool that is useful for monitoring the health of oil-lubricated electric motors. It provides important information about the health of the lubricant, the manner in which the machine is interfacing with its environment to ingest contaminants, and the health of the bearings themselves. The decision to employ oil analysis should be dictated by the motor’s criticality, history of problems, sump or tank volume, etc. 13.5.1.2.1 Sampling Interval The sampling interval for electric motors can range from continuous monitoring to complete exclusion from the program. It is beyond the scope of this chapter to discuss all the influencing variables in detail. However, unless the motor is equipped with force circulation lubrication, the oil is probably going to be sampled at the time the it is changed. One might ask why we are interested in analyzing the oil from a motor if we only intend to dispose of it. Our primary objective when analyzing oil following a change is to determine whether the oil has degraded abnormally (become unusually contaminated with debris, water, chemicals, the wrong oil, etc.) and whether the bearings are generating an abnormal amount of wear debris, which can be analyzed elementally and morphologically to evaluate the nature and severity of the wear mechanism. 13.5.1.2.2 Test Slate Once a motor is included in the oil analysis program, an appropriate test slate must be structured. The slate must reveal information about the health of the oil and its additives, contamination level, and wear debris generation. A common test slate for routine oil analysis of oil-lubricated electric motors includes viscosity in cSt (mm2 sec−1 ) at 40◦ C, acid number (AN), Fourier transform infrared (FTIR) analysis, moisture concentration test, particle count, distribution and ISO 4406 cleanliness code, elemental spectroscopy for wear metals, additive metals, and contaminants and ferrous density. Other tests may be required on an as-needed basis, including microscopic wear debris analysis (WDA). 13.5.1.3 Top-ups Periodically, oil-lubricated electric motors require the addition of oil to restore proper level. When doing so, ensure that the appropriate oil is added to the motor and that it is free of contamination. Ideally, add the oil through a filter into the tank or sump using a quick-connect-type fitting to avoid opening the system. If the system must be opened to add oil, do so carefully to minimize contamination. If the oil cannot be pumped directly into the sump or tank through a filter, add prefiltered oil using only cleanable
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and closable containers made from nonreactive material. Be sure the containers are kept clean. Funnels, where required to top-up oil, are a major source of contamination. Because most motors are not equipped with filters, extra care must be taken to keep contamination out in the first place. Doubling the average particle contamination level in a rolling element bearing, which is extremely easy to do with careless top-ups, will cut a rolling element or sleeve bearing’s life by almost 20%, on average. In the event that the water contamination level doubles as a result of careless handling, the life of a sleeve bearing will be cut by almost 20%, while the life of an antifriction bearing will be reduced by about 30%. If both dirt and water are allowed in due to careless handling, the reduction in expected life caused by the two contaminants is combined. 13.5.1.4 Oil Drain and Fill For a bath-, ring-, or collar-lubricated electric motor, the oil drain schedule will normally be time based, not condition based. However, a sample may be drawn for analysis at the time of the drain to uncover any unusual problems or to optimize the drain interval relative to the rate of lubricant degradation or contaminant ingress. For force circulation oil-lubricated motors, the drain interval will more likely be triggered by oil analysis results. In either case, the key to an effective oil change is to completely eliminate the old oil and the contaminants. To do so, drain the oil as quickly as possible following equipment shutdown. If the lubricant has experienced severe degradation or extremely high levels of contamination, a flush of the system is suggested. After draining the old oil and flushing if required, fill the tank with clean new oil. Ideally, transfer the oil through a filter cart or other pumping device equipped with a filter. If this is impossible, prefilter the new oil and transfer it using the same practices suggested for performing a top-up. 13.5.1.5 Flushing If the oil is oxidized prematurely or heavily contaminated, or if a different (possibly incompatible) lubricant is introduced into the system, flushing may be necessary. Depending upon the lubricant application method, the following flushing methods might apply to oil-lubricated electric motors. 13.5.1.5.1 Drawdown Filtration/Separation During Motor Shutdown By connecting an off-line filtration system (typically portable) the oil can be cleaned over time — typically 5 to 10 turns of the sump volume. For bath-, ring-, and collar-lubricated electric motors that have a small sump volume, it is best to carry out this flush while the machine is shut down using a higher flow pump. High flow will create turbulence in the housing, which will help to scrub particles off component surfaces and force them out of their hiding places. Often, this strategy can replace a scheduled oil change, particularly when the objective of the oil change is simply to decontaminate the lubricant. As long as the lubricant’s performance properties meet specification, there is no reason to dispose of the oil, and the drawdown flush typically results in cleaner oil than the oil drain and fill approach. The filtration system can be equipped with a water removal mechanism if required. Ideally, the motor’s bearing sumps will be equipped with quick-connect fittings to enable easy, nonintrusive access to the lubricant. After decontamination, remove the pumping system, close all valves, reset the oil levels, and replace all caps and vent breathers before startup. 13.5.1.5.2 Drawdown Filtration/Separation During Motor Operation If the motor is equipped with a tank and circulating system, drawdown while the machine is operating is quite simple. The portable decontamination rig, equipped to remove particles and water, if required, is attached to the circulating oil tank using quick-connects, after which the oil is filtered long enough to turn the tank volume over a minimum of 5 to 10 times. For best results, separate the suction and return lines of the pumping system as much as possible. This strategy can be employed to decontaminate the lubricant in electric motors with bath, ring, or collar lubrication. However, it is a much higher-risk proposition to decontaminate the lubricant in a motor with small bearing sumps when it is operating than when it is shut down. Exercise extreme caution to ensure that the sump level does not change during decontamination and to ensure that the flow from the pumping system does not cause the oil to foam
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Handbook of Lubrication and Tribology
or become excessively aerated. Dropping the level and causing foaming and aeration can result in serious bearing damage. Occasionally, other flushing techniques may be required, including turbulent line flush for motors with force circulation lubrication systems, solvent, chemical, or detergent cleaning, pulsating or reverse flow flushing, mechanical cleaning, etc. A detailed discussion of these techniques is beyond the scope of this chapter.
13.5.2 Grease-Lubricated Electric Motors While the largest horsepower and, often, more critical application electric motors are lubricated with oil, the vast majority of industrial electric motors are lubricated with grease. Greasing correctly will, in my opinion, have a significant positive impact on motor reliability and total cost of ownership of industrial electric motors. Overgreasing is common due to poor control. Undergreasing is common due to neglect or fear of overgreasing. The key is to apply the correct amount of lubricant at the appropriate interval. 13.5.2.1 Regreasing Once the right lubricant is selected for the application, there are three objectives for a motor relubrication program — the right volume, the right interval, and the right practices. 13.5.2.1.1 Volume The right amount of grease can be estimated using tables based upon motor frame size or shaft diameter, but the calculated volume is just as simple and generally more accurate. The objective, of course, is to apply an amount of grease that will replenish the bearing without causing churning, heat generation, or escape of grease into the motor’s windings. The following equation is generally accepted as being appropriate for determining the required volume of grease to apply during lubrication: Gq = 0.005DB where Gq is grease quantity in grams; D is bearing outside diameter in millimeters; B is bearing total width in millimeters. Note: For grease quantity in ounces, Gq = 0.114DB, with all dimensional measurements in inches. 13.5.2.1.2 Interval While several general tables and guidelines exist for determining how often to grease electric motor bearings, a calculated interval is more precise. One equation, depicted in Figure 13.5, starts with a basic computation based upon the bearing’s size and speed, and requires the input of the following adjustment factors to produce a net regreasing interval in operating hours: • Temperature. Motors that run hot require more frequent regreasing because the grease tends to get used up faster and is more like to run out. Hotter-running bearings are also more apt to cake up with old, used-up thickener. More frequent regreasing helps to reduce this effect. • Contamination. Motors are often regreased with greater frequency when the risk of dirt contamination is high. The purpose is to make sure there is sufficient grease to keep housing seals from drying out. • Moisture. Moisture causes serious problems for grease. It reduces the lubricant’s film strength, causes hydrolysis of the base oil, additives, and thickener, increases the rate of oxidation, and, in some cases, washes the grease out of the bearing. Increase the regrease frequency when water is present. • Vibration. Vibration contributes to shearing of the grease thickener and simply shakes the grease out of the bearing. More frequent relubrication is required to compensate for the vibration.
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Lubricating Industrial Electric Motors
13-17 Grease interval correction factors
T=K
[(
)
]
14,000,000 – 4d n(d 0.5)
where T = Time until next relubrication (hours) K = Product of all correction factors Ft × Fc × Fm × Fv × Fp × Fd (see table) n = Speed (rpm) d = Bore diameter (mm)
Note: ips = inches/second 0.2 inches/second = 0.5 mm/sec
Average Correction operating range factor 1.0 Housing below 150°F Temperature 0.5 150 to 175°F Ft 0.2 175 to 175°F 0.1 Above 200°F Light, nonabrasive dust 1.0 Contamination 0.7 Heavy, nonabrasive dust Fc 0.4 Light, abrasive dust 0.2 Heavy, abrasive dust 1.0 Humidity mostly below 80% Moisture 0.7 Humidity between 80 and 90% Fm Occasional condensation 0.4 Occasional water on housing 0.1 Less than 0.2 ips velocity, peak 1.0 Vibration 0.2 to 0.4 ips 0.6 Fv above 0.4 0.3 Horizontal bore centerline 1.0 Position 45° bore centerline 0.5 Fp Vertical centerline 0.3 1.0 Ball bearings Bearing design 5.0 Cylindrical and needle roller bearings Fd 1.0 Tapered and spherical roller bearings Condition
FIGURE 13.5 The calculated regrease interval must consider the operating temperature, moisture and particle contamination levels, vibration, shaft position, and bearing type in addition to bearing size and speed.
• Shaft Position. Gravity causes grease to flow out of the motor’s bearings and their housings regardless of position. However, the rate of loss increases as the shaft moves away from the horizontal position. Motors with nonhorizontal shafts require more frequent relubrication. Vertical shafts, of course, require the most frequent relubrication. • Bearing Design. Tapered and spherical rolling element bearings use up grease 10 times as fast as radial ball bearings. So, a motor’s regrease interval must be adjusted to account for the various consumption rates for different bearing designs. 13.5.2.1.3 Best Practices After establishing an appropriate regrease volume, the following list of best practices will help improve the overall quality of electric motor lubrication: • Dedicate a grease gun to a grease type and tag it using intuitive color or shape coding in addition to clear lettering. While pneumatic or electric-powered grease guns are convenient, I prefer manual grease guns because an experienced lubrication technician can feel problems in motors, such as excessive back pressure, when manually greasing. This important feedback mechanism is lost when externally powered grease guns are employed. • Be mindful of grease compatibility problems and avoid mixing greases with incompatible base oils, additives, and thickeners. Thickeners represent perhaps the greatest risk. Be sure to extend your compatibility control to the motor rebuild shop (discussed later in this chapter). • Confirm the amount of grease the application gun expels per shot. Normally this is 0.10 ounces per shot. But some grease guns expel as little as 0.05 ounces per shot and some expel as much as 0.33 ounces per shot. Periodically check the output of each grease gun to ensure that it is expelling the appropriate amount. Adjust the number of shots if the gun’s output changes or simply discard the gun and replace it with a new one.
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Handbook of Lubrication and Tribology
• Remove plug from relief pipe before applying grease. Leave the plug out for 10 to 30 min following relubrication so that the motor has time to expel excess grease. Do not be concerned if you do not see grease exit from the purge pipe; this is not unusual. • Apply grease slowly, allowing 3 to 5 sec per shot, so the grease will distribute evenly in the bearing and the grease reservoir. Applying grease quickly can cause the new grease to channel, reducing the effectiveness of relubrication. • Do not push through back pressure. Lever-operated manual grease guns can develop more than 10,000 psi of pressure. Pushing too hard can damage housing seals. If the bearing is equipped with a shield facing the grease supply, it is possible to push the overhung shield into the bearing if too much pressure is applied and there is no other relief. This causes a great deal of damage to the bearing. • Discontinue greasing immediately if you see grease exist past the housing seals or the purge pipe. This indicates that the grease reservoir has been overfilled. • Leave a dollop of grease on the end of the grease fitting after the motor has been lubricated. This allows dirt to collect above the surface of the fitting as opposed to in its entry way. Wiping away the dollop and its collected dirt before greasing the motor avoids the slug of dirt that otherwise enters the bearing during relubrication. • If a motor has not been regreased for more than two years, it may be difficult to effectively lubricate it. The exact timing depends upon the operating temperature. Exercise caution when greasing these bearings. Feedback tools, which are discussed in the next section, can help. 13.5.2.1.4 Feedback Mechanisms Technology continues to influence the manner in which we maintain industrial equipment. Greasing motor bearings is no exception. Several technologies are available to provide feedback to the lubrication technician while he or she is regreasing bearings. These technologies add precision to the calculationbased regrease volume and interval techniques and the best practices based on rule-of-thumb previously outlined. • Grease-Gun-Mounted Pressure Gauge. Perhaps the simplest of all available feedback tools used for greasing motor bearings is the pressure gauge. In the previous section, the importance of “feeling” back pressure on the grease gun was discussed. A pressure gauge mounted on the grease gun adds precision. In fact, if the technician is consistent with the rate of volume, the back pressure on the bearing can be recorded and trended. It is a tell-tale sign that the bearing may be building up a mass of caked-up thickener and may be due for cleaning. • Grease-Gun-Mounted Flow Meter. The formulas used to calculate the required regrease volume typically report grease by weight in ounces or grams, despite the fact that the inputs for the equation are dimensional. This is for convenience purposes. The true objective is to apply a sufficient volume of grease to fill the bearing. A flow meter attached to a grease gun adds another level of precision in the pursuit of control. • Vibration Analysis. When a motor bearing is properly lubricated, the vibration level is typically low. However, as the lubrication film becomes strained due to lack of grease, the bearing’s surface asperities begin to contact, which results in low amplitudes, but high-frequency physical displacement, or vibration, which alerts the lubrication technician that the bearing requires relubrication. The reestablishment of the lubrication film should reduce or eliminate the high-frequency vibrations. If the condition is not rectified, asperity contacts will become more severe and eventually produce impacts. When impacting occurs, damage to the rolling elements and or the raceways occurs, which is detectable as bearing fault defects, which normally are lower-frequency, higher-amplitude vibrations. Vibration analysis can supplement conventional calculation-based regrease interval estimates to refine motor lubrication practices. • Acoustic Emissions Analysis. As with vibration analysis, asperity contact caused by the initial stages of lubricant starvation also manifests in the form of acoustic emissions. Unlike vibration analysis,
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Lubricating Industrial Electric Motors
13-19
where the frequency of vibration produced is a function of running speed, ultrasonic acoustic emissions associated with the lack of lubrication always occur at approximately the 25 to 35 kHz frequency range. The amplitude of the emissions, which is reported in microvolt-based decibels (dB µV), however, is dependent upon running speed. When a motor bearing requires grease, the amplitude of the ultrasonic signal will increase, indicating a need to relubricate. If lubrication is neglected and bearing damage occurs, both ultrasonic and sonic acoustic emissions are produced. When grease reaches the bearing, a reduction in the amplitude of acoustic emissions is detectable by watching a decibel meter or by listening with headphones (ultrasonic emissions must be modulated to an audible frequency for listening, a process referred to as heterodyning). Like vibration analysis, ultrasonic analysis can be employed in conjunction with and to refine calculated relubrication intervals and to ensure that grease is reaching the bearings. Likewise, if the required regrease interval shortens, it can indicate seal failure, thickener incompatibility, or other systemic problems. • Thermographic and Thermometric Analysis. Motor bearings run hot when they are starved or when they are overlubricated. Clearly, one does not wish to wait until the bearings are burning up to decide to apply lubricant, so these methods are not particularly suitable for determining when to lubricate. However, they are useful for determining whether the bearing has been overgreased. None of the other feedback tools is as effective at identifying overgreasing. The only limitation, of course, is a lag time between the act of overgreasing and the bearing heating up. It is too late by the time overgreasing is detected — the damage is done. However, over time, the method can be employed to refine the regrease volume estimate produced by the previously discussed formula for a particular motor or class of motors (Figure 13.6). • Grease Analysis. The challenge with routine grease analysis is the difficulty in obtaining a representative sample. However, grease analysis is a useful technique for diagnosing problems, such as the combination of incompatible thickeners. Also, wear debris analysis techniques can be employed for grease samples, particularly visual evaluation of wear particles. 13.5.2.1.5 Automatic Greasing Systems In some instances, automatic greasing systems are preferred to manual greasing for motors. If the motor is difficult or unsafe to reach for relubrication, or if the application demands frequent regreasing, automatic lubrication should be considered. A variety of automatic lubrication systems, both single-point and multiple-point, are available. Regreasing temperature profiles
Temperature
Overcharged
Properly regreased Not regreased
Time
FIGURE 13.6 When grease is applied, the bearing’s temperature may temporarily rise, then fall below the pregreasing temperature. However, if the bearing is overgreased, the temperature will rise and remain high.
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Handbook of Lubrication and Tribology
13.5.2.1.6 Instructions to Motor Rebuild Shop Eventually, all motors must be rebuilt or replaced. Increasingly, industrial sites are relying upon outside motor rebuild shops in lieu of in-house rebuilds. It is important to provide clear instructions to motor rebuild shops, either internal or external, regarding lubrication. • Insist on Detailed Feedback. Even if lubrication was not the reason for the motor rebuild, direct the rebuild shop to report all lubricant-related details, including signs of varnish and scorching, excessive caking of thickener, and grease on the windings. Often, motor rebuilders will not provide useful, detailed feedback unless specifically requested to do so. • Precisely Specify the Required Grease. A common mistake and cause of abbreviated motor life is incompatibility between the grease originally packed in the motor bearings and the grease used at the industrial site for regreasing. For instance, rebuild shops commonly initially grease motor bearings with polyurea grease. If the site regreases with lithium complex grease, softening due to incompatibility is likely, resulting in the grease running out of the bearings — often into the motor’s windings. A note of caution: not all polyurea thickeners are compatible. Be thorough in reviewing the compatibility of the greases used at the motor rebuild shop and at the site. Ideally, use the same grease for both. • Proper Fill Volume. Define the initial fill volume for the bearings, considering bearing type, speed, and motor design, and leave the supply pipe filled with grease. Properly selected, installed, operated, and maintained, oil- and grease-lubricated industrial electric motors can deliver years of trouble-free service. Selecting the appropriate lubricant, applying it effectively, and managing its condition and contamination levels are critical to achieving electric motor reliability.
References [1] Application Guide for AC Adjustable Speed Drive Systems. National Electrical Manufacturers Association, Rosslyn, VA, 2001. [2] Bloch, H., Practical Lubrication for Industrial Facilities. Marcel Dekker, New York, 2000. [3] Electric Motor Predictive and Preventive Maintenance Guide. Electric Power Research Institute (EPRI) report number NP-7502. Report available from: Corporate Communication Division, Electric Power Research Institute, 3412 Hillview Ave., P.O. Box 50490, Palo Alto, CA, 94304. [4] Fitch, J., D. Troyer, and M. Barnes, Oil Analysis II: Course Book. Noria Corporation, Tulsa, OK, 2004. [5] Gebarin, S., Industry Practices Relating to Electric Motor Bearing Lubrication. Noria Research Analysis Report, Noria Corporation, Tulsa, OK, 2004. [6] Information Guide for General Purpose Industrial AC Small and Medium Squirrel-Cage Induction Motor Standards. National Electrical Manufacturers Association, Rosslyn, VA, 2002. [7] Leugner, L., The Practical Handbook of Machinery Lubrication, 2nd ed. Maintenance Technology, International, Inc./Oil City Press, Edmonton, Alberta, Canada, 2000. [8] Lubricating Grease Guide, 4th ed. National Lubricating Grease Institute, Kansas City, MO, 1996. [9] Nelson, C., Millwrights and Mechanics Guide. Macmillan General Reference, New York, 1989. [10] SKF Bearing Maintenance Handbook. SKF Corporation, Denmark, 1996. [11] Troyer, D., “Condition-Based Greasing: How Much and How Often.” Proceedings of the Lubrication Excellence 2004 Conference, Nashville, TN, March 23–25, 2004. [12] Troyer, D., “Bearings, Motors and Grease”, MRO Today, 2004a, Oct./Nov., p. 35; “Inspecting Oil Lubed Motors”, MRO Today, 2004b, Dec. (2004)/Jan. (2005), p. 32. [13] Troyer, D. and J. Fitch, Oil Analysis Basics. Noria Publishing, Tulsa, OK, 2000.
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[14] Trujillo, G., D. Troyer, and J. Fitch, Machinery Lubrication I: Fundamentals Course Book. Noria Corporation, Tulsa, OK, 2004. [15] Trujillo, G., D. Troyer, and J. Fitch, Machinery Lubrication II: Applications Course Book. Noria Corporation, Tulsa, OK, 2004. [16] Underwood, J., “Electric Motor Lubrication — A New Perspective.” Workshop presented at the Lubrication Excellence 2004 Conference and Exhibition, Nashville, TN, 2004.
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14 Effects of Radiation on Lubricants 14.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14.2 Mechanism of Lubricant Radiolysis . . . . . . . . . . . . . . . . . . 14.3 Specific Radiation Effects . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
14-1 14-2 14-3
Base Oils • Hydrocarbons (Petroleum Oils and Some Synthetics) • Other Synthetic Base Oils • Additives • Radiation Damage Inhibitors • Antioxidants • Viscosity Index Improvers • Lubricity, Antiwear, and Extreme Pressure Agents • Foam and Rust Inhibitors • Grease Gelling Agents (Thickeners) • Grease Data • Other Data on Finished Products
14.4 Other Environmental Factors . . . . . . . . . . . . . . . . . . . . . . . . .
14-11
Radiation Type and Dose Rate Effects • Roles of Temperature and Exposure Atmosphere
Robert O. Bolt EPRI/NMAC Consultant, Lubricants of Lubrication
14.5 Look at Elastomers First . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14.6 Summary Principles . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
14-12 14-13 14-14
14.1 Introduction Why should one be interested in what radiation, as in nuclear radiation, does to lubricants? We in the United States get some 18% of our electric power from nuclear power plants; some other countries are much more dependent on this power source. Some lubricants used in such plants are exposed to radiation. So, when selecting these lubricants, radiolysis (interaction with radiation) must be considered along with traditional enemies of lubricants such as temperature and oxidation. This is particularly so because lubricants vary 1000-fold in their resistance to radiation. Incident radiation is important only if energy from it is absorbed. Thus, exposures are expressed in terms of the rad — 100 ergs absorbed/gram of absorber (1 rad = 0.01 gray = 0.01 J/kg). The rad will be expressed here in powers of 10, for example, 107 rads, etc., or shorter, E7, E8, rads, etc. Lubricants are used in three general regions in and around nuclear power plants, as shown in Figure 14.1. Area 1 is inside the reactor and behind the biological shield where radiation and temperature are maximum. Area 2 is the heat exchange region, where temperature and radiation are moderate to low. Area 3 is outside the biological shielding. Here, radiation levels are low and temperatures are those usually found in and around conventional turbine generators. High radiation levels may also prevail in handling spent fuel away 14-1
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14-2
Handbook of Lubrication and Tribology Control rod
1 2 Turbine
3
Generator
Coolant Pump
Pump Biological shields
FIGURE 14.1 Lubrication areas in a nuclear power plant. (From Carroll, J.G. and Bolt, R.O. (eds), Radiation Effects on Organic Materials, Academic Press, New York, 1963, chap. 1. With permission.)
TABLE 14.1 Radiation Dose Rates for Components in Nuclear Power Plantsa Equipment item Control rod drives Fuel handling devices Primary coolant pumps and blowers Auxiliary pumps and motors Block and throttle valves Turbines
Dose rates (rad/year) E2-E9 E2-2E9 E3-2E9 E3-E8 E7-2E9 0-E4
a Multiply by maintenance interval to get total dose.
from the reactor, but temperatures are low. The range of dose values in Table 14.1 results from differing requirements of various reactor types. Radiation changes organic substances by two competing routes: cross-linking (or polymerization) and cleavage. Some cleavage to smaller molecules always occurs. Gas evolves and is generally the first thing noted on radiolysis. Hydrogen (if present originally), low-molecular weight hydrocarbons, and other derivatives make up the gas phase. Reactive sites remain in the organic molecule after this cleavage. Subsequent reactions at these also depend upon the organic structures involved. Cross-linking shows physically in organic liquids by viscosity increase; in solids by increased hardness and brittleness. Cleavage results in less viscous liquids and softer solids. Continued exposure to radiation eventually turns all organic lubricants into hard, brittle solids.
14.2 Mechanism of Lubricant Radiolysis In passing into or through any material, radiation transfers energy at different rates, depending on the type of radiation, its intensity, the energy of the radiation, and the absorber. Energy is transferred in small increments; violent events are rare. Gamma rays and fast neutrons, of most interest here, deposit energy in organics almost entirely by interaction with electrons. Figure 14.2 illustrates this interaction with a high energy (1 MeV) gamma
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Effects of Radiation on Lubricants
14-3 0.5 MeV gamma ray
1 MeV gamma ray
Spur Delta ray Photoelectron
Compton electron
FIGURE 14.2 Interaction of a gamma photon with organic matter. (From Carroll, J.G. and Bolt, R.O., Advances in Petroleum Chemistry and Refining, Vol. 8, McKetta, J.J., ed., John Wiley & Sons, New York, 1964, chap. 1. With permission.)
ray. In the initial encounter an electron is ejected. The gamma ray gives up about half its energy in this encounter and is scattered in the process. It can then proceed to have other encounters with other atoms. The ejected electron can cause ionizations in subsequent reactions. The electron is often derived from an inner orbit of the original atom, leaving a vacancy to be filled from an outer orbit. The energy lost by the new electron in dropping to the lower orbit is given off in an x-ray and eventually in another electron (photoelectron). The original ejected electron (Compton electron) can also cause ejections of secondary electrons from atoms along its path. These secondaries are called delta rays if they are energetic enough to ionize other atoms. Ionizations and excitations from all electrons occur in small clusters called spurs. The spurs are more dense at the end of the tracks, that is, at low energy levels. The low-energy electrons from this process are eventually absorbed. Fast neutrons affect organics by a similar mechanism. However, they have no charge and do not cause ionizations directly. Instead, they transfer energy by elastic (or billiard-ball-like) collisions with nuclei and form recoil protons. These protons are the charged particles that cause ionizations. In hydrocarbons, the recoil protons are more likely to come from hydrogen atoms, but some come from carbon nuclei in the molecule. The proton produces relatively high spur density. In some cases, the spurs actually overlap. Thus, more interactions are possible among reactive substances in the spurs, rather than interactions between these substances and the surrounding, less reactive, original species. This can result in different products from fast neutron than from gamma radiolysis. Slow (thermal) neutrons do not have enough energy to cause elastic recoils that will break bonds in organic molecules. They produce no chemical effect except from secondary emissions of rays, for example, a sodium-23 atom (in a grease-gelling agent) accepts a slow neutron to become sodium-24; this radioactive atom gives off radiation that can cause ionizations and excitations.
14.3 Specific Radiation Effects Organic lubricants (including greases) consist of base oils and additives, the latter introduced to enhance existing desired properties or to provide new properties. Both the additives and the base oil play major roles in determining radiation resistance.
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14-4
Handbook of Lubrication and Tribology 1011
1010
Polyphenyls
109
Polyethers
108
Siloxanes Silicates Diesters
107
Phosphates
Poly(Phenyl Ethers) Alkyltaromatics
Mineral oils polyalkenes
Methyl silicones
106 Radiation dose, Rads
FIGURE 14.3 Ranges of radiation doses leading to failure of various base fluids. (From Carroll, J.G. and Bolt, R.O. (eds), Radiation Effects on Organic Materials, Academic Press, New York, 1963, chap. 1. With permission.)
14.3.1 Base Oils Lubricants based on mineral oils are the most common in the marketplace, but products based on synthetic oils such as polyalkenes (polyolefins, SHCs), esters, ethers, and silicones are also seen. As shown in Figure 14.3, base oil radiation resistance depends directly on chemical nature. The most important factor is aromatic content because aromatic rings have the unique ability to absorb energy and dissipate it (resonance) while minimizing carbon–carbon bond breakage. In general, radiation stability is reduced by any substituents on the aromatic ring. However, all base oils undergo decomposition on radiolysis — it is only a matter of degree. The removal of hydrogen or other fragments from the original molecule leaves changes in the surviving species. This results in decreased oxidation stability and in cross-linking or scission, etc. The varying degrees to which these events occur determine the usability of lubricants in the presence of large amounts of radiation.
14.3.2 Hydrocarbons (Petroleum Oils and Some Synthetics) Table 14.2 summarizes some effects of radiation exposure on a highly refined mineral oil (white oil) consisting primarily of paraffinic and naphthenic (cyclic) structures. After about 108 rads, damage was minor; after 9 times this exposure, it was very large. The first thing to occur on radiolysis is color change. Gas evolution also takes place early, as noted by decreases in flash point and increases in vapor pressure. Viscosity increases occur because molecular weight rises. Rearrangement of the molecular structure is indicated by changes in viscosity index and pour point. Petroleum oils are generally described in terms of the refining processes by which they are made. Processing does not always remove natural inhibitors, for example, sulfur compounds, that are effective in reducing radiation damage. This complicates the intercomparison of mineral base oils. Table 14.3 shows that such factors are noted at radiation doses of about 108 rads in bulk exposures.
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Effects of Radiation on Lubricants
14-5
TABLE 14.2 Changes in Highly Refined Naphthenic Mineral Oil (White Oil) on Radiolysis in Bulk Gamma dose at 24◦ C, E8 rads in air ASTM color (estimate) Viscosity at 99◦ C, cSt Viscosity at 38◦ C, cSt Viscosity index Pour point, ◦ C Flash point, ◦ C Vapor pressure, Pa at 204◦ C
0
0.9
9.0
0.5 7.4 71 6S −34 224 267
8.6 88 67 −34 204 1067
2 41 101 80 −21 202 133
TABLE 14.3 Viscosity Change of Petroleum Oils on Irradiation in Bulk Gamma doses at 24◦ C, E8 rads in air Viscosity at 38◦ C
0 cSt
25 VI naphthenic (pale) 85 VI paraffinic (bright stock) 85 VI paraffinic (neutral) 95 VI paraffinic (white)
10 93 11 76
0.7–1.1 4–6 % Increase 10 11 8 46
72 75 316 75
Modern base oils are classified by the American Petroleum Institute as follows: • • • • •
Group I — solvent-refined mineral oils Group II — mildly hydrogen-refined mineral oils Group III — severely hydrogen-refined mineral oils Group IV — polyalkenes [poly(alphaolefines) SHCs] Group V — all other base oils
The last three oils in Table 14.3 are similar in properties to Group I, Group I, and Group II, respectively. The neutral oil and bright stock still have some of their natural inhibitors remaining after the solventrefining process. The white oil would contain far fewer of the sulfur and similar compounds as a result of its hydroprocessing. Consequently, the more highly refined oil exhibits less radiation resistance, as shown in Table 14.3. The naphthenic (pale) oil is even less refined and contains more aromatic and sulfur- and nitrogen-containing materials. As a result, it is the most resistant to radiolysis. It must be remembered that a lubricant’s base oil, though important to the finished product’s performance, does not tell the whole story. The additives play an even more important part in performance, as will be detailed later. 14.3.2.1 Polyalkenes Polyalkenes (API Group IV) are the most common synthetic hydrocarbon base fluids seen in the marketplace. They would be expected to act like paraffinic mineral oils on radiolysis. However, two factors affect this performance. First, the synthetics contain no natural inhibitors. These oils do, however, respond well to additives. Second, the polyalkenes can sometimes contain residual unsaturation (double bonds) that is radiation sensitive. Thus, it is important to stabilize such synthetic bases by hydrogenation. 14.3.2.2 Alkylaromatics Aromatic ring compounds that contain no substitutions represent the ultimate in hydrocarbon radiation stability. Such materials, though, are limited as lubricant base fluids because of their poor rheological properties, for example, high melting points. Improved rheological properties can be achieved by attaching
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14-6
Handbook of Lubrication and Tribology TABLE 14.4
Other Synthetic Oils and Radiolysis Effects
Class
Example
Ethers
Polyglycols
Esters Silicones
Poly(phenyl ethers) Diesters: Silicates; phosphate [–Si(R2)O–]
Halogenated organics
CC1(F)C(F2)–
Radiolysis effects Large gas evolution and dilution; low viscosity increases; temperature sensitivity high Low gas evolution; good at high temperatures Splitting to form acids Good oxidative and thermal stability; tendency to gel Give off halogen acids; corrode metals
alkyl side chains to the aromatic rings. Short chains improve lubricant properties somewhat, yet permit good radiation resistance. Longer alkyl chains further improve viscosity and viscosity index, but at a sacrifice in radiation stability. Eventual choice of an alkyl chain for an optimized alkylaromatic base oil is necessarily a compromise. Specially synthesized alkylaromatics are expensive and their use is limited. Less optimized products can be used instead — materials that are by-products from other chemical operations. Their physical properties are less favorable from the lubricant standpoint than the optimized products, but their radiation resistance is similar [4,6].
14.3.3 Other Synthetic Base Oils Table 14.4 lists radiolysis properties of some other synthetics used for applications requiring their special physical and chemical properties. As with hydrocarbon base oils, synthetics are made more radiationresistant by the presence of aromatic rings, for example, aryl silicones are more stable than alkyl silicones.
14.3.4 Additives Most lubricant additives are organic compounds used in less than 30 mass %, usually less than 5%, and frequently less than 1%. Additives are used not only to reduce radiation damage but, more important, to (1) decrease oxidation, (2) improve viscosity-temperature properties, (3) reduce wear and improve load-carrying ability, (4) suppress foaming, and (5) gel oils to form greases. Sometimes an additive serves more than one of these purposes.
14.3.5 Radiation Damage Inhibitors Some additives inhibit lubricant property changes that occur on radiolysis. Their effect is frequently noted above 109 rad, but they are also of benefit at lower doses in lubricants of low or intermediate radiation resistance. Many of these additives are also oxidation inhibitors; others function whether or not oxygen is present. Alkylaromatic, dialkyl diester, and aliphatic polyether base oils were included in early experiments on radiation damage inhibitors with reactor radiation in the presence of air. A dialkyl selenide was increasingly effective in amounts of up to 10 mass percent. Organic sulfur compounds were also of use. Amines and phenols were much less beneficial. Radiation damage inhibitors were most potent in the radiation-sensitive base fluids, as shown in Figure 14.4 [1,23]. Diester oils can be protected with selected aromatic materials, Reference 10 as illustrated in Table 14.5. Physical mixtures of aromatic and aliphatic compounds can be equal in radiation resistance to specially tailored alkyl aromatics. For example, octadecylbenzene was compared with a mixture of mineral oil and 1-methylnaphthalene; each fluid contained the same aromatic carbon content. After irradiation, changes in viscosity with the physical mixture were equivalent to those with the “chemical mixture.” Thus, from a radiation-damage standpoint, exotic synthetic compounds are not required; simple blends may do as well. However, the blends do not have the rheological properties of the fully synthetic materials.
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Effects of Radiation on Lubricants
14-7
1400
Kinematic viscosity at 38°C, % of original
1200 Without additive With 5 mass % additive
1000
800 Phenyl methyl silicone 600
400 C16–18–Alkylbiphenyl
200
0
Phenyl methyl silicone 0
2
4
6
Gamma dose at 24°C in air
8 108
10
rad
FIGURE 14.4 Influence of didodecyl selenide on radiolytic viscosity change. (From Carroll, J.G. and Bolt, R.O. (eds), Radiation Effects on Organic Materials, Academic Press, New York, 1963, chap. 1. With permission.)
TABLE 14.5 Protection of Di(2-ethylhexyl) Sebacate with Naphthalene Derivative from Exposure in Bulk to Fast Electrons Dose of 0.9 E8 rads at 24◦ C under nitrogen Tert-butylated naphthalene (mass %) 0 5 20
Increase in viscosity at 38◦ C (%)
Increase in neutralization no. (mg KOH/mg)
45 24 16
7.0 3.7 1.2
14.3.6 Antioxidants Even without radiation, small amounts of oxygen (air) hasten the degradation of lubricants. Antioxidants retard this effect by reaction with free radicals formed in the oxidation process. Irradiation vastly increases the concentration of these and other reactive species and accelerates oxidation. Thus, a major burden is placed on the antioxidant to scavenge such species, and the antioxidant is “used up” rapidly during radiolysis in the presence of air. Additives can also become inactivated by suffering direct radiation damage themselves. Antioxidants are conventionally used in 0.1 to 2 mass percent in oils. Such low quantities are not sufficient to scavenge the reactive materials resulting from radiolysis alone, even disregarding those from oxidation. Calculations Reference 10 show that 13% of phenothiazine is needed to scavenge all the reactive species formed in a diester blend by 108 rads under nitrogen. Such high quantities of antioxidant (1) are normally above its solubility limits in a base oil and (2) often have adverse effects on the base oil’s physical properties. However, 5 mass percent of some antioxidants can be used to decrease radiolytic oxidation and improve general radiation resistance.
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14-8
Handbook of Lubrication and Tribology 25 30 mass % poly(a-Methylstyrene)
Viscosity at 38°C, cSt
20
15 8 mass % polybutene
4 mass % poly(dodecyl methacrylate)
10
Base fluid
0
0
1
2
3
4
5
8
Gamma dose at 24°C in air 10 rad
FIGURE 14.5 Change of the viscosity of polymer solutions by irradiation. Base fluid: mixed alkylbenzenes (mol wt 250) containing 5% didodecyl selenide and 0.1% quinizarin. (From Carroll, J.G. and Bolt, R.O. (eds), Radiation Effects on Organic Materials, Academic Press, New York, 1963, chap. 1. With permission.)
14.3.7 Viscosity Index Improvers “Multigrade” oils for automotive and industrial applications employ high-molecular weight polymers (viscosity index [VI] improvers) added to low-viscosity base oils. The polymer improves the viscositytemperature properties (VI) and raises the viscosity of the base oil. Polyesters and polyolefins are examples of conventional improvers. The effects of radiolysis on an alkylbenzene base oil containing various polymers are shown in Figure 14.5 [1,6,11,23]. While the polyester and polyolefin VI improvers broke down, the polystyrene (thickener — not a VI improver) cross-linked to increase the viscosity of the solution. An oil whose viscosity remains relatively constant is suggested from Figure 14.5. Ingredients are a low viscosity oil, a polymer which breaks down [4,11], and a second polymer which cross-links on radiolysis. This is the basis of a line of commercial products [12].
14.3.8 Lubricity, Antiwear, and Extreme Pressure Agents Lubricity and antiwear additives give mild protection where an oil film may be only intermittently present; extreme pressure (EP) agents are a more drastic treatment. Fatty acids (e.g., oleic acid) are examples of lubricity agents, and phosphate esters (tricresyl phosphate, zinc thiophosphates) represent antiwear agents. Organic compounds, for example, paraffins, fats, and esters, containing chlorine or sulfur are examples of EP additives. Because of increased stresses, hypoid gear assemblies, differentials, planetary and reduction gearing, screw thread actuators, and so on, require EP materials. All these agents interact with metal surfaces through physical adsorption or chemisorption (lubricity materials) or chemical reaction (antiwear and EP materials) to form a protective coating. In general, radiation stability of lubricity, antiwear, and EP agents is governed by that of the basic molecule involved. For example, a sulfurized ester would behave much like an ester. This stability, however, is reduced by the sulfur or other substituents. Some of these agents can become increasingly acidic on irradiation above the 107 rad level [5,6]. Corrosivity is a problem, particularly with the chlorine-containing materials which evolve HCl.
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Effects of Radiation on Lubricants
14-9
14.3.9 Foam and Rust Inhibitors Methyl silicone fluids are widely used in less than 0.01 mass percent for foam inhibition. Organic polymers are increasingly in vogue for this purpose as well. Irradiation generally increases foaming tendency in finished oils above about 107 rads [12] due to destruction of the antifoamer. The base oil employed in a lubricant has a major influence on foaming. For example, mixtures, like Group I mineral oils, are worse in foaming than relatively discrete diester-type lubricants. Many lubricants contain rust inhibitors to prevent rusting of ferrous surfaces to which water has access during lubrication. A half ester, dodecyl acid maleate, is an example of one type used in steam turbine oils. Doses of less than 108 rad can cause many steam turbine oils to fail the rust test. Other rust inhibitors, metallic or nonmetallic salts of petroleum sulfonates, are also used in heavy-duty motor and diesel engine oils. These additives are predicted to be somewhat superior in radiation stability to the acidic inhibitors.
14.3.10 Grease Gelling Agents (Thickeners) Lubricating greases are formed by dispersing a solid in an oil, thus providing the mixture a unique stay-inplace function. The solid (gelling agent) acts as a sponge that releases the oil as needed. Many materials, organic and inorganic, are employed as such agents. Table 14.6 lists common gelling agents and some of the important properties relating to their use [25]. The metallic soaps of fatty acids (e.g., lithium stearates) are the most used. These soap-thickened greases can be damaged by irradiation, as shown in Figure 14.6 [15]. At the lower dose, the soap crystallites are slightly smaller than they were originally and partially fractured. At the higher dose, they are disintegrated and cannot form a sponge structure. Silica, graphite,
TABLE 14.6
Grease Types and Performance
Property Gelling agent Calcium complex Calcium sulfonate complex Lithium soap Lithium complex Polyurea, polyurea complex Inorganics
Dropping point (◦ C)
Maximum temperature for prolonged use (◦ C)
Bleeding tendency
260–300 300–320 170–200 260–300 240–260 260+
120–150 150–175 120–135 150–175 150–175 120–140
Low Low High Medium Low Low
10 m
FIGURE 14.6 Effect of radiation on crystallite structure of sodium stearate–paraffinic mineral oil grease during gamma dose in air at 24◦ C. (From Hotten, B.W. and Carroll, J.G., Ind. Eng. Chem., 50, 217, 1958. With permission.)
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Handbook of Lubrication and Tribology
Percent of initial ASTM worked penetration (60 strokes)
180
Liquid
150
Sodium stearate (14 mass %)
100
50 Sodium n-octadecylterephthalamate (14 mass %)
Solid
0 0
5
10
15
20
25
30
Gamma dose, 108 rad
FIGURE 14.7 Effect of gelling agents on radiation resistance of mineral oil greases. (From Bolt, R.O., Carroll, J.G., Harrington, R., and Giberson, R.C., Proceedings of the Second International Conference on Peaceful Uses of Atomic Energy, Geneva, 1958, Vol. 29, United Nations, New York, 1959, 276. With permission.)
carbon black, organic dyes, and nonmetallic salt gelling agents are inherently stable chemically, and their radiation stability is increased by their nonfibrous nature. With organic gelling agents, aromatic materials enhance radiation resistance, as illustrated in Figure 14.7. Greases with the best radiation resistance are made from aromatic base oils, aromatic or inorganic gelling agents, and radiation damage-oxidation inhibitors such as those containing sulfur or selenium. Products usable to 3–5 × 109 rad can result.
14.3.11 Grease Data Most of the data cited so far have come from work to find the limits of lubricant resistance to radiolysis. What happens at radiation doses postulated to be seen in a reactor accident scenario? Doses there are much higher than those occurring during normal running but are most often less than the upper limits. Accident doses are generally around 2 × 108 rad. This comes from a small running dose plus the main accident dose plus a small continuing dose after the accident. The last component is small because the plant is no longer running and is winding down instead. Normal maintenance need not be considered either. The data in Table 14.7 involve commercial products of three grease types from Table 14.6. They were exposed in bulk in air first to 150◦ C for 300 h to simulate the expected accident thermal situation. Then, they were exposed to 2.2 × 108 rad of gamma radiation. The calcium complex grease exhibited a marked hardening on the thermal treatment followed by softening with the irradiation. The calcium sulfonate complex grease also softened significantly on irradiation. This softening on exposure to low to medium radiation doses is similar to the data in Figure 14.7 for a sodium soap grease. In Table 14.7 the two polyurea greases show very little effect from the irradiation disturbing their gel structures beyond what the thermal exposure did. The thermal treatment softened the polyureas significantly. Oxidation stability, a measure of the remaining useful life of a product, is also adversely affected by the thermal and radiation exposures. This is shown in Table 14.7 with high pressure scanning calorimetry (HPDSC) data (ASTM D-5483). Here, the effect of the thermal exposure is much less pronounced than that of irradiation. This illustrates the fact that irradiation not only produces reactive species that are more easily oxidized but can have an adverse effect on oxidation inhibitors as well.
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Effects of Radiation on Lubricants TABLE 14.7 Greases
14-11
Effects of Thermal Plus Radiation Exposures on Some Commercial
Ca complex
Ca sulfonate complex
Polyurea A
Polyurea B
Consistency ASTM 60-Stroke Pen. Start After 300 h at 150◦ C. After heat +2.2 E8 rad
334 204 318
329 341 399
296 392 383
283 341 353
Oxidation stability HPDSC at 210◦ C, Minutes Start After 300 h at 150◦ C. After heat +2.2 E8 rad
4 2 <1
48 27 8
9 11 7
30 27 15
Grease
With the greases in Table 14.7 there is no significant loss in antifriction, antiwear, or steam anticorrosion properties as a result of the thermal plus 2.2 × 108 rad exposures. Only the calcium sulfonate complex grease suffered any significant loss in dropping point (304 down to 222◦ C) as a result of the exposures. This is indicative of radiation damage to the gel structure similar to that shown in Figure 14.7 for a sodium stearate gel.
14.3.12 Other Data on Finished Products There are other data available on individual oils and greases from the manufacturers of these products. Notable among these studies is the work of Mobil [27]. Here, lubricants were exposed both thermally and to radiation, again to simulate reactor running and accident situations. Other studies involved only irradiation of products. The work of Texaco [28] and Gulf [29] is foremost among these. The reported results illustrate that greases generally soften on radiolysis before solidifying at very high doses and oils increase in viscosity slowly before reaching solidification at very high doses.
14.4 Other Environmental Factors 14.4.1 Radiation Type and Dose Rate Effects The same general mechanism of change on radiolysis applies whether exposure is to gamma rays or to fast neutrons (or x-rays, beta rays, etc.). Lubricants are altered due to chemical interactions in and around centers of ionization and excitation. Overlapping of these centers, because of higher linear energy transfer (LET) along the path of the ionizing radiation, can produce different products. Higher LET results from exposure to higher dose rates, or higher energy radiation, for example, fast neutrons vs. gamma rays. The most definitive work on effects of radiation type involves exposures of highly aromatic materials (polyphenyls) at high temperatures (about 315◦ C). A “swimming pool” reactor was used for separate exposures to nearly 100% reactor gamma rays and, with different absorbers, to about 90% fast neutrons. The conclusion from this and other work is that fast neutrons are two to three times more damaging (in polymer formation) than are gamma rays. Less definitive work on grease performance indicates that higher dose rates (for equal energy input) are more damaging (therefore, fast neutrons more damaging) [17]. Other experiments show cobalt-60 gamma rays alone are more damaging than reactor radiation. In this later reactor work, about 20% of the energy input was from fast neutrons. Molecular weight increases, viscosity increases, and so on, were greater for the gamma exposures, as shown in Figure 14.8 for viscosity with a naphthenic/paraffinic petroleum fraction containing about 15% aromatics. Exposures were at about 70◦ C in air.
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Handbook of Lubrication and Tribology
Viscosity at 50°C, cSt
300
200
100 70 50 40 30
0
400
800
Dose, 106 rads
FIGURE 14.8 Viscosity change (at 50◦ C) of naphtheniparaffinic oil with absorbed dose: CO-60, reactor radiation.
As a first approximation with these conflicting data, equal energy input creates equal effects. For more sophisticated evaluations, the aromatic content of the material must be considered, along with the radiation type and dose rate of the incident radiation.
14.4.2 Roles of Temperature and Exposure Atmosphere Oxidation is tremendously accelerated by temperature increases in the presence of radiation. If changes in a lubricant must be minimized and high temperatures are needed, inert gas blanketing is dictated. Temperatures need to be low if oxygen cannot be avoided. In radiolysis, there is usually a threshold temperature below which heat causes only a small amount of the total damage. Above the threshold, temperature is increasingly important. Illustrating this, a separate inhibited lubricants based on an alkylbenzene, an aromatic ester, an aliphatic ester, and an aliphatic polyether were exposed to 1.5 × 109 to 1.9 × 109 rad in a reactor. Air was present, and temperatures were 20 to 220◦ C. Viscosity change in these static exposures was little affected by temperatures below about 140◦ C; the effect was marked at 180◦ C. Thus, the temperature threshold is about 140◦ C under these conditions. With 5 × 108 rad gamma exposures at 24 and 204◦ C under helium, the temperature effect depends on the inherent stability of the base oil. Highly aromatic materials did not change much between the two temperatures; however, a anaphthenic white oil did. An effective inhibitor can alter this with white oil, as shown in Table 14.8.
14.5 Look at Elastomers First When trying to achieve maximum radiation resistance for a lubrication system, the elastomer used for seals, etc. is likely to be the weakest link. Lubricants are generally 2 to 10 times as radiation resistant as elastomers [19]. Figure 14.9 gives details of radiation effects on various elastomer types [20,21]. If an aromatic-based lubricant is dictated by radiolysis considerations, an elastomer is called for which is both resistant to radiolysis and compatible with aromatics. The nitrile, polysulfide, and urethane rubbers appear best for this dual purpose [20,22]. Note from Figure 14.8 that the best rubbers are only good for up to 5 × 106 rads. The best oils and greases, on the other hand, will withstand 10 times this level. To utilize lubricants to the fullest, elastomers should be avoided when high radiation doses are involved. Mechanical seals or reliance on greases may avoid seal problems altogether. Studies of elastomers show gamma rays to be far less damaging than fast neutrons. A factor of about 5 was measured for 10 common elastomers [21].
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Effects of Radiation on Lubricants TABLE 14.8 White Oila
14-13
Influence of Temperature and Inhibitor on Radiolysis of a Naphthenic
2Mass % didodecyl selenide
No inhibitor Irradiation temperature (◦ C) Increase in viscosity (%) At 99◦ C At 38◦ C Gassing (ml/g)
24
204
24
204
74 168 —
229 440 32
72 127 25
78 128 23
a Gamma dose of 5 × 10’ rad in helium.
Degree of change Elastomers
Nil
Mod.
High
Acrylics Butyls Fluoroelastomers Hypalons Natural rubbers Neoprenes Nitriles Polysulfides Silicones Styrenes Urethanes Vinylpyridines Others Polyethylenes Polyfluorocarbons Poly(vinyl chlorides) Silicone Resin — Glass Ceramics Metals 105
106
107 108 Gamma dose, rads
109
FIGURE 14.9 Radiation resistance of materials.
14.6 Summary Principles 1. Finished oil lubricants, in general, can be classified according to dose ranges as follows: • 106 rad or below: there are no unusual problems from radiolysis. • 106 to 107 rad: methyl silicone, aliphatic diester, and phosphate ester-based products become affected; polymers in solution (multigrade products) degrade. For most other cases, environmental factors other than radiation are controlling.
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Handbook of Lubrication and Tribology
• 107 to 108 rad: radiolysis renders diester- and certain mineral oil-based products marginal in performance because of physical property effects. Oxidation stability and thermal stability are adversely affected for all products. Some lubricants are usable and some are marginal in this range. • 108 to 109 rad: oxidation and thermal stabilities of most lubricants are seriously changed. Major effects occur in most physical properties. Products based on aliphatic ethers, aromatic esters, and certain mineral oils (carefully selected) may be used. • 109 to 1010 rad: products based on polyphenyls, poly(phenyl ethers), or alkylaromatics are needed. • 1010 rad and above: radiolysis causes extreme effects. Lubrication with even the best organic fluids is very restricted, and laminar solids, such as molybdenum disulfide or graphite, should be considered for use.
2. Additives normally used in lubricants, for example, antioxidants, antiwear, EP, and antifoam agents, themselves suffer radiation damage. Their depletion on radiolysis can cause complications at radiation levels below which the base oil itself degrades. 3. Radiation damage in base oils can be reduced by selected additives. They are most effective in the least stable fluids. However, greater gains in stability can be made from judicious choices of base stocks than from attempts to improve unstable fluids with additives. Oxidation can drastically reduce the life of a lubricant; radiation accelerates oxidation. The role of temperature is interrelated with those of oxygen, additives, and radiation dose. Above about 140◦ C, radiation damage increases significantly with temperature. 4. Under radiolysis, greases generally first soften because of damage to the gel structure then harden because of cross-linking of the oil component. Conventional greases are usable to about 107 rad with some going to 108 or so. A few special products are usable in the range of 109 to 5 × 109 rad. 5. Elastomers are 2 to 10 times more sensitive to radiolysis than are lubricants. Elastomer stability is the limiting factor in many lubrication systems.
References [1] Carroll, J.G. and Bolt, R.O., “Radiation Effects on Lubricants,” in Advances in Petroleum Chemistry and Refining, Vol. 8, McKetta, J.J., ed., John Wiley & Sons, New York, 1964, chap. 1. [2] Ferrie, J.S., Leinonen, P., Neil, B., and Wharton, E., “The Effects of Radiation on Lubricants in Nuclear Generating Stations,” 35th Annual Meeting, American Society of Lubrication Engineers, Anaheim, CA, 1980. [3] Mobile Oil, AG, “Lubrication of Nuclear Power Plants,” Technical Bulletin, Mobile Oil, AG, Hamburg, Germany, May 1974. [4] Bolt, R.O. and Carroll, J.G., “Effects of Radiation on Aircraft Lubricants and Fuels,” USAF Report WADC-TR-56-646, Pt. 2, Wright Air Development Center, Dayton, OH, April 1958. [5] Carroll, J.G. and Calish, S.R., “Some Effects of Gamma Radiation on Commercial Lubricants,” Lubr. Eng., 13, 338, 1957. [6] Carroll, J.G. and Bolt, R.O., “Development of Radiation-Resistant Oils,” Am. Soc. Lubr. Eng. Trans., 2, 1, 1959. [7] Carroll, J.G. and Bolt, R.O.,“Development of Radiation-Resistant Oils,” USAEC Report AECU-3764, Technical Information Service Extension, Oak Ridge, TN, June 1958. [8] Bolt, R.O. and Carroll, J.G., “Radiolysis and Radiolytical Oxidation of Lubricants,” Ind. Eng. Chem., 50, 221, 1958. [9] Hollinghurst, R., “Radiation-Resistant Fluids and Lubricants,” Inst. Pet., 52, 9, 1966. [10] Mahoney, C.L., Kerlin, W.W., Barnum, E.R., Sax, K.J., Saari, W.S., and Williams, P.H., “Engine Oil Development,” USAF Report WADC-TR-57-117, Wright Air Development Center, Dayton, OH, July 1957.
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Effects of Radiation on Lubricants
14-15
[11] Bolt, R.O., Carroll, J.G., and Wright, J.R., “Radiation-Resistant Lubricants — Their Development and Status,” USAEC Report TID-5186, Technical Information Service Extension, Oak Ridge, TN, June 1954. [12] Bolt, R.O. and Carroll, J.G., U.S. Patent 2 943 056, June 28, 1960; U.S. Patent 2 967 827, January 10, 1961. [13] Convair Division, General Dynamics Corp.,“Results of System Panels Test Number 2, Addendum 5,” USAF Report NARF-58-IT (Add. 5), Wright Air Development Center, Dayton, OH, September 1958. [14] Kottcamp, C.F., Nejak, R.P., and Kern, R.T., “The Effects of High Energy Ionizing Radiation on Turbine Oil Performance Characteristics,” Am. Soc. Lubr. Eng. Trans., 2, 7, 1959. [15] Hotten, B.W. and Carroll, J.G., “Radiation Damage in Lubricating Greases,” Ind. Eng. Chem., 50, 217, 1958. [16] Burns, W.G., Wild, W., and Williams, T.F., “The Effect of Fast Electrons and Fast Neutrons on Polyphenyls at High Temperatures,” in Proceedings of the Second International Conference on Peaceful Uses of Atomic Energy, Geneva, 1958, Vol. 29, United Nations, New York, 1959, 266. [17] Vaile, P.E.B., “Lubricants for Nuclear Reactors,” Proc. Inst. Mech. Eng., 176, 27, 1962. [18] Stukin, A.D., Shor, G.I., Gorbach, V.A., and Ryaboshapko, A.G., Khim. Tekhnol. Topl. Masel, 1, 47, 1969; transl. in Chem. Technol. Fuels Oils, 1–2, Consultants Bureau, New York, 1969, 62. [19] Carroll, J.G. and Bolt, R.O., “Radiation Effects on Organic Materials,” Nucleonics, 18, 78, 1960. [20] Harrington, R., “Elastomers for Use in Radiation Fields. IV. Effects of Gamma Radiation on Miscellaneous Elastomers and Rubberlike Materials,” Rubber Age, 83, 472, 1958. [21] Born, J.W., “Elastomeric Materials,” in Radiation Effects on Organic Materials, Bolt, R.O. and Carroll, J.G., eds, Academic Press, New York, 1963, chap. 7. [22] King, W.H., “A Fresh Look at Elastomers,” Mach. Des., 45, 108, 1973. [23] Carroll, J.G. and Bolt, R.O. (eds), “Radiation Effects on Lubricants,” in Radiation Effects on Organic Materials, Academic Press, New York, 1963, chap. 1. [24] Bolt, R.O., Carroll, J.G., Harrington, R., and Giberson, R.C., “Organic Lubricants and Polymers for Nuclear Power Plants,” in Proceedings of the International Conference on Peaceful Uses of Atomic Energy, Geneva, 1958, Vol. 29, United Nations, New York, 1959, 276. [25] Bolt & Associates, “Lubrication Guide: Revision 3” (Formerly NP-4916-R2), EPRI, Technical Report 1003085, Palo Alto, CA, 2001. [26] Bolt, R., Adams, H., and Herguth, W., “Comparative Analysis of Nebula and MOV Long Life Greases for Limitorque Main Gearbox Applications,” EPRI Final Report 1003483, Palo Alto, CA, December 2002; “Comparative Analysis of Polyrex and SRI Greases for Use in EQ Motor Bearings,” EPRI Final Report 1003484, Palo Alto, CA, December 2002. [27] Mobil Oil Corp. Sheets, M. (Nutech), “Class 1E Qualification of Five Mobil Oils and Five Mobil Greases,” 1989. [28] Texaco, Inc. “Nuclear Qualification of Texaco Lubricant Products,” Report T-1, 1988. [29] Schreuders, G., Gulf Research & Development Co., “Irradiation of Gulf Lubricants,” 2nd ed., Report 6316RS103, January 1985.
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15 Wire Rope and Chain 15.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.2 Wire Rope . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
15-1 15-1
Classification and Construction • Wire Strands • Lubrication and Maintenance
15.3 Chain Drives . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
15-5
Roller Chains • Silent Chains • Engineering Steel Chains • Components and Design Considerations • Lubrication and Maintenance
15.4 Lubricant Flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Paul Conley Chief Engineer Lincoln Industrial
15-10
Lubricant Characteristics • Types of Lubrication
References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
15-20
15.1 Introduction Wire ropes are important mechanisms for transmitting tensile loads where flexibility is needed. They are typically wound around sheaves, drums, or pulleys for transmitting force. They comprise of continuous wire strands wound around a central core. When a wire rope is bent around a sheave or a pulley, the individual wires and strands move in relation to each other to adjust themselves to the curvature assumed by the rope. In order to reduce or prevent the wear due to the metal strands rubbing against each other as they adjust themselves around the curvature, lubrication is needed. Chains are used to transmit power to drive machinery and to transport materials along conveyers. Chains can be described as a series of journal bearings connected by link plates or sidebars. Loads are transmitted though each link in tension through a series of pins and bushings. Lubrication is needed to prevent metal-to-metal contact between pins and bushings.
15.2 Wire Rope As a wire rope is bent around the axis of a pulley, drum, or sheave, the top portion of the wire rope strands tend to be in tension whereas the bottom portion of the wire rope strands tend to be in compression. This causes the wires to move in relation to each other, which will eventually cause wear. To reduce this wear, lubrication is needed. Lubrication is essential internally to prevent excessive metal-to-metal rubbing contact as well as externally to protect metal-to-metal contact with the outer surface of the wire rope (see Figure 15.1). The essential purpose of lubrication is to prevent metal-to-metal contact. Lubrication provides a protective film that separates two metal surfaces from each other. 15-1
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15-2
Handbook of Lubrication and Tribology Bending of rope induces internal wire rub against each other
External wear location
FIGURE 15.1
Position of cable around pulley. (Taken from Lincoln Industrial Corporation. With permission.)
TABLE 15.1 Wire Rope Classifications Based on the Nominal Number of Wires in Each Strand Classification 60 6 × 19 6 × 37 6 × 61 6 × 91 7 × 19 7 × 37 80 8 × 19 807 8 × 61 Rotation resistant ropes 8 × 19 190 19 × 19 350 35 × 19
Outer strands
Wires/strand
Maximum no. of outer wires
6 6 6 6 7 7 8 8 8 8
3–14 15–26 27–49 50–74 75–109 15–26 27–49 3–14 15–26 27–49 50–74
9 12 18 24 30 12 18 9 12 18 24
15a 17–19a 17–19a 26–36a 26–36a
15–26 6–9 15–26 6–9 15–26
12 8 12 8 12
a Total strands.
15.2.1 Classification and Construction Wire rope can be classified in the following ways (Table 15.1): 1. Size (diameter) and length. The nominal diameter of the wire rope determines its size. 2. Grade of steel. To handle large tension loads, high carbon steel is often used. Grade 120/130 extra improved plow type (EIPS) is used in the manufacture of wire ropes for special installations where
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Wire Rope and Chain
15-3
maximum rope strength is required and conditions of use permit some applications such as mine shaft hoisting where increased tonnage on existing skips and drum can be tolerated and where other conditions such as sheave and drum diameters are favorable to its use. Grade 110/120 improved plow (IPS) has a remarkable combination of high tensile strength, tough wearing qualities, and excellent fatigue resistance properties. By far the largest tonnage of wire rope is made from this grade. Grade 110/110 plow (PS) has lower tensile strength and resistance to wear than Grade 110/120 but retains high fatigue resistance properties. This grade can be used when strength is secondary to fatigue strength. Listed below are the tensile strength specifications for each grade:
Grade 110/130 EIPS 110/120 IPS 110/110 PS
Tensile strength (Gpa)
Tensile strength (psi)
1.7–2.35 1.5–2.05 1.35–1.8
245,000–340,000 220,000–300,000 195,000–258,000
Other metals used in wire ropes are aluminum-coated steel, galvanized coated steel, aluminum, bronze, and titanium. 3. Whether it is preformed. Wire strands can be preformed during manufacture. In a preformed wire rope, the wire strands are shaped in a helical fashion so that they line up naturally in position when meshed in the cable. Preformed wire rope will have better fatigue strength and is better for working over small sheaves. It is easier to handle, and the individual strands tend to remain in position after breaking, thereby reducing the tendency to protrude and damage other wires but making inspection more difficult. 4. By its lay. There are five basic types of lay (see Figure 15.2). The lay describes two types of features. The first applies to the direction of the helix in both the wires in strands and the strands in rope. Right lay strands rotate clockwise whereas left lay, counterclockwise. Regular lay indicates that the wire in the strand spiral opposite to the direction in which the strand spirals around the rope (see Figure 15.3). 5. The number of strains and the number of wires in each strand.
15.2.2 Wire Strands The core is the foundation of the wire rope. It is made of materials that provide proper support for the strands under normal bending and loading conditions. Core materials include either fibers or steel. A steel core consists of either strand or an independent wire rope. The three most commonly used core designations are fiber core (FC), independent wire rope core (IWRC), and wire strand core (WSC). A fiber core can consist of vegetable fibers or plastic materials such as polypropylene. Fiber core ropes have lower breaking strengths than do ropes with steel cores. These cores can prevent damage to the interior of the strands when wires come in contact with each other during flexing. Fiber cores can absorb lubricant, thus preventing damage due to wear. The result of this is longer life. Independent wire rope core is designated by the number of wires in a strand as well as the number of strands in the rope. IWRCs are normally 6 × 7 wire rope with a 1 × 7 strand resulting in a 7 × 7 wire rope. When this entire construction becomes the center of the wire rope, it is called an independent wire rope core. While the IWRC provides somewhat greater breaking strength than does a fiber core, lubricant service is more critical. In WSC core, a single wire strand forms the core. This is distinguished from IWRC that uses a multistrand wire rope. Ropes with wire strand cores have great strength and are normally employed in static applications. Internal lubrication in the field is a problem requiring the application to be carefully selected so that subsequent lubrication can be effective (see Figure 15.4).
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Handbook of Lubrication and Tribology
Right regular lay
Left regular lay
Right lang lay
Left lang lay
Right alternate lay
FIGURE 15.2
Five types of lay.
15.2.3 Lubrication and Maintenance During fabrication, ropes receive lubrication, the kind and amount depending on the rope’s size, type, and use if known. This in-process treatment provides the finished rope with ample protection for a reasonable time if it is stored under proper conditions. But when the rope is put into service, the initial lubrication will normally be less than needed for the full useful life of the rope. Because of this, periodic applications of a suitable rope lubricant are necessary. Wire rope lubricants should be free from acids and alkalis. The lubricant should have sufficient adhesive strength to remain on the rope. It should have a viscosity that is low enough to be capable of penetrating the interstices between wire and strands but large enough for sufficient film strength to limit metal-to-metal rubbing between the wire and strands. The lubricant should be soluble in the medium surrounding it under the actual operating conditions. Finally, the lubricant should resist oxidation. Before lubrication, dirt and other abrasive materials should be removed from the rope. Cleaning is accomplished with a stiff wire brush dipped in solvent, compressed air, or live steam. Immediately after cleaning, the rope should be lubricated. When it is normal for the rope to operate in dirt, rock, or other abrasive material, the lubricant should be selected with great care to make certain that it will penetrate and at the same time, will not pick up any of the material through which the rope must be dragged. As a general rule, the most efficient and economical means to do field lubrication is by using a method that will continuously apply the lubricant when the wire rope is in operation. Many techniques are
© 2006 by Taylor & Francis Group, LLC
Wire Rope and Chain
15-5
6 × 7 classification
6 × 7 FC 6 × 19 classification
6 × 19
6 × 21
6 × 25
6 × 26
6 × 31
6 × 36
6 × 36
6 × 31
6 × 41
6 × 41
6 × 46
6 × 49
6 × 57
6 × 61
6 × 37 classification
6 × 61 classification
6 × 52
FIGURE 15.3 Wire strand classification.
used: these include the continuous bath, dripping, pouring, swabbing, brush, or where circumstances dictated by the use of automatic systems that will apply the lubricant by a drip or pressure spray method (Figure 15.5). Below are some common types of lubrication methods used [1].
15.3 Chain Drives Chains are used as a medium in power transmission. In all types of chains, the main function is to transmit power. Features of chain include: • Controlled flexibility in only one plane. • The action of chain is positive over sprockets; no slippage occurs.
© 2006 by Taylor & Francis Group, LLC
15-6
Handbook of Lubrication and Tribology Core
Wire
Strand
FIGURE 15.4 Wire rope construction.
FIGURE 15.5
Common ways of applying lubrication.
• • • •
Chains can carry very heavy loads with negligible stretch. Chains are very efficient. Flexure takes place between bearing surfaces that are designed for wear resistance. Chains can operate satisfactorily in adverse surroundings such as elevated temperatures or where subject to moisture and foreign materials. • Chains have unlimited shelf life; they do not deteriorate with age or with sun, oil, or grease. Chains may be classified in different ways. In terms of how chains are manufactured, there are three types.
© 2006 by Taylor & Francis Group, LLC
Wire Rope and Chain
15-7
Pin link plate
Roller free to turn on outside of bushing
Roller link plate Bushing (press-fitted in link plate)
Sprocket
Pin (press fitted in link plate)
FIGURE 15.6
Roller chain construction.
15.3.1 Roller Chains The major purpose of roller chains is to reduce friction. The rollers in chains are used in two separate functions to do this; the same roller provides both the functions. The function of the rollers is to engage the sprocket teeth and thus transfer sliding action to the internal members of the chain, which are designated for that purpose. All roller chains are so constructed that the rollers are evenly spaced throughout the chain (Figure 15.6). In roller chains, two or more chains are joined, side by side, by means of pins that maintain the alignment of the rollers in the various strands. In power transmission applications, roller chains range from fractional horsepower drive to those requiring in excess of 1000 hp. A roller chain is made up of two kinds of links, roller links and pin links, alternately spaced throughout the length of the chain. The roller link is an assembly of steel shaped like the figure eight, with a hole at each end to receive the bushings. The center distance between the holes is kept within a close tolerance to maintain a uniform pitch in an assembled chain. Each bushing is a hollow cylinder, with the outside and inside surfaces hardened to resist wear. The rollers are also hollow cylinders, hardened and finished to a precise diameter. Each is of the proper length to permit it to turn freely between the link plates.
15.3.2 Silent Chains Silent chains are constructed to provide smooth and quiet operation at high speeds accompanied by long service life (see Figure 15.7). These types of chains consist of a series of toothed linked plates assembled on joint components in such a fashion that articulation occurs between adjoining pitches. These and longitudinal guide plates are the basic components of silent chains.
© 2006 by Taylor & Francis Group, LLC
15-8
FIGURE 15.7
Handbook of Lubrication and Tribology
Silent chain.
Silent chains consist of apertures of the link plates and guide plates. The link plates contain notches, flats, and projections in the apertures to position the joint in the link plates. Guide plates are provided in silent chains to prevent lateral movement. Four types of silent chains are described below: 1. Round pin with bushing. In this design, the joint components consist of a plain cylindrical pin with a bushing added to increase the bearing area of the chain joint. 2. Round pin and segmental bushing. The components of this style of joint are a plain cylindrical pin with segmented bushings. Each bushing covers less than half the circumference; lugs between the bushing segments prevent rotation in the links. 3. Rocker pin. In the rocker pin design, differently shaped and keyed components are in contact in such a way that a rocking motion occurs between the pins when the chains flexes at the joint. 4. Roller pins. The joint components are two pins with cylindrical surfaces that roll on each other.
15.3.3 Engineering Steel Chains In engineering steel drive chains the ability to bridge large spaces between parallel shafts is especially significant. Such chains are usually quite heavily loaded, but travel in relatively low speeds. Chains of this type are very rugged and have the toughness and elasticity to absorb heavy shock loads. Engineering steel chains are made of fabricated and machined steel. They are designed to perform a variety of other functions to meet known or anticipated operating conditions. Owing to their wide application, there are many types of engineering steel drives available and they are also classified broadly. Generally speaking, the types of engineering steel chains classified under ANSI standards as follows: 1. Straight sidebar steel chains with and without rollers. Straight sidebar steel fabricated roller chains have steel bushings and pins and are assembled using interference fits between those parts and the sidebars. The inside link, made up of two sidebars and two bushings, is called the bushing link. The outside link is called the pin link. 2. Offset sidebar steel chains with and without rollers. Each link of an offset steel fabricated chain is identical. The link components consist of sidebars, a bushing, and a pin in with a cotter. 3. Bar link chains. Bar link chains are sometimes called block and bar or steel block chains. This type of chain usually consists of two outer sidebars, one center bar, and two pins making up a two-link section. Bar link chain usually does not have bushings, the center bar flexing directly on the pin. The sprocket contact is with the ends of the center bars. Chains of this type are frequently used in tension linkage applications. These chains are used in slow-moving conveyors such as steel mill conveyors and are made in very large versions to raise and lower river and channel lock and dam gates. 4. Welded steel chains. Welded steel chains are similar to fabricated steel chains except that links are made as integral weldments rather than being held together by means of tight fits and locking surfaces. Welded steel chains are often used for slow-moving conveyor drives.
© 2006 by Taylor & Francis Group, LLC
Wire Rope and Chain
15-9
5. Open barrel steel pintle chains. This type is an economical lightweight conveyor chain. The basic chain consists of one piece offset formed steel links and pins. The pins are fixed against rotation by mechanical locks or interference fits. In terms on their applications, there are four types: • • • •
Drives for power transmission Conveyors Bucket elevators Tension linkages
Many standard chains are intended to be used for only one of these four applications. Roller chain and engineering steel chains are used in all areas of application in varying degrees. Silent chains are essentially employed in drive applications; however, they also find use in a few conveying applications.
15.3.4 Components and Design Considerations Listed below are the design factors that influence the applications of chains: 1. Chain tension. 2. Project bearing area. This is the contact surface between a pin and a bushing and is equal to the pin diameter times the bushing diameter times the bushing length. 3. Smoothness and hardness of the contacting surfaces. 4. Lubrication. 5. Frequency and degree of articulation of the chain joints. 6. Environment and vibration. Characteristic of wear. Wear is thought of as loss of material due to the chain components working either against each other or against some foreign material. The three most common types of wear are abrasion wear, adhesive wear, and fatigue wear. The actual amount of wear on a given chain depends on several factors including materials and heat treatment of the chain components, load and speed, lubrication, the presence of abrasive and corrosive substances, and the configuration of the sprockets. Most sliding or dragging chains have wear on the edges of the sidebars and often on the outer surfaces of the bushing of rollerless chains. In drive and elevator applications, wear in the chain joints is usually the limiting factor in determining the life of the chain. In bushed rollerless chains, wear is caused by the movement of the pins in the bushings and by the bushings rubbing on the sprocket teeth. In engineering steel chains with rollers, wear occurs on the outside diameter of the rollers, between the rollers and bushings and between the pins and bushings. Wear between the pin and bushing causes the chain to elongate (grow longer but not stretch) until the chain will not fit the sprockets correctly or will not maintain correct spacing or timing. Chain wear elongation usually progresses through three stages (see Figure 15.8). First, there is a short period of rapid initial, or run-in, wear. In this first stage, high spots are worn off, the pins and bushings and minor misalignments are quickly worn away. Second, there is gradual wear of the surfaces between the pins and the bushings. In this second stage, the pins are seated properly in the bushings and the bearing areas are normally well lubricated. In the last stage, there is another period of rapid wear. As the clearance between the pin and bushing gets larger, lubrication becomes less ineffective in such a way that the case hardening of the pins and bushings may have worn through, leaving a softer metal that will wear more dramatically. The result of excessive elongation on the sprocket may have caused loads on individual joints to increase dramatically [2]. Since a chain must articulate over a sprocket, wear occurs between the chain bushings and the pins and cause elongation of the chain or lengthening of the pitch. Sprockets are designed to accept a reasonable amount of elongation (3 to 6% for many styles) from wear, but when the chain elongates beyond this
© 2006 by Taylor & Francis Group, LLC
15-10
Handbook of Lubrication and Tribology Wear points
Worn chain — elongated
New chain
FIGURE 15.8
Chain wear locations.
point it no longer fits the sprockets and the system will not operate properly. Chain wear elongation can be reduced by: 1. 2. 3. 4.
Selecting a chain with reduced bearing pressure between the pins and bushings. Selecting a chain with increased hardness of the chain joint members. Improved lubrication. Increase sprocket size by increasing the number of teeth in the sprocket [3].
15.3.5 Lubrication and Maintenance Each joint in a roller chain is a journal bearing, so it is essential that it receives an adequate amount of proper lubricant to achieve maximum wear life (see Figure 15.9). Some low-speed roller chain drives operate successfully with only the initial factory lubrication. However, most roller chain drives must be either periodically or continuously relubricated to obtain their full potential service life. In addition to resisting wear between the pins and bushings, an adequate flow of lubricant smoothes the engagement of the chain rollers with the sprocket, cushions roller to sprocket impacts, dissipates heat, flushes away wear debris and foreign materials, and retards rust.
15.4 Lubricant Flow Lubrication of the pin and bushing surfaces, which articulate under load, is most important, but some lubrication between the roller and bushing is also necessary. The lubricant should be applied to the upper edges of the link plates in the lower span of the chain shortly before the chain engages a sprocket. Then, both gravity and centrifugal force will aid in carrying the lubricant to the critical pin and bushing surfaces. Surplus lubricant spilling over the link plate edges will supply the roller and bushing surfaces.
© 2006 by Taylor & Francis Group, LLC
Wire Rope and Chain
15-11
Location for lubrication
FIGURE 15.9
Lubrication points for chain.
15.4.1 Lubricant Characteristics Lubricants for roller chain drives should have the following characteristics: 1. Sufficiently low viscosity to penetrate into the critical internal surfaces. 2. Sufficiently high viscosity, or appropriate additives, to maintain the lubricating film under the prevailing bearing pressures. 3. Clean and free of contamination. 4. Capability to maintain lubricating qualities under the prevailing operating conditions. These requirements usually are met by a good grade of nondetergent petroleum base oil. Detergents normally are not necessary, but antifoam, antirust, and film strength improving additives are often beneficial. Low-grade or impure oils should be avoided. Low-grade oils cannot provide effective lubrication and acids or abrasive particles in the oil can damage the chain beyond repair. Heavy oils or greases should not be used because they are too thick to penetrate into the internal surfaces of the chain. The recommended oil viscosity for various surrounding temperature ranges is shown in Table 15.2.
15.4.2 Types of Lubrication The ANSI standards list three types of lubrication for roller chain drives. In ANSI B29.1, they are: • Type A: Manual or drip lubrication • Type B: Oil bath or slinger disc lubrication • Type C: Oil stream or pressure spray lubrication In ANSI B29.3, they are: • Type I: Manual, slow drip (4 to 10 drops per minute), or shallow bath lubrication. • Type II: Rapid drip (20 or more drops per minute), oil bath, or slinger disc lubrication. • Type III: Continuous lubrication, with slinger disc or circulating pump The recommended type is shown in the horsepower tables in the respective standards and is mainly influenced by the chain speed and the amount of horsepower being transmitted. The recommended types are minimum lubrication requirements. The use of a better type (Type B instead of Type A for example) is acceptable and may be required by operating conditions other than speed and power. Lubrication has a very significant effect on chain wear life, so it is vital to follow the lubrication recommendations in the horsepower rating tables. Consult a chain manufacturer when it appears desirable to use a lubrication type other than that recommended.
© 2006 by Taylor & Francis Group, LLC
15-12
Handbook of Lubrication and Tribology TABLE 15.2 Recommended Oil Viscosity for Chain Lubrication at Various Temperatures Recommended grade SAE 5 SAE 10 SAE 20 SAE 30 SAE 40 SAE 50
Temperature, ◦ F
(Temperature, ◦ C)
−50 to +50 −20 to +80 +10 to +110 +20 to +130 +30 to +140 +40 to +150
(−46 to +10) (−29 to +27) (−12 to +43) (−7 to +54) (−1 to +60) (+4 to +66)
When the temperature range permits a choice, the heavier grade should be used.
Lubricant reservoir
Wick packed distribution pipe
FIGURE 15.10
Drip lubrication.
15.4.2.1 Manual Lubrication Note: Manual lubrication is to be done only when the drive is stopped and power to the drive is locked out. For manual lubrication, oil is applied periodically with a brush or a spout can, preferably once each 8 h of operation. The time may be longer than 8 h, if it has proven adequate for that particular drive. The volume and frequency of oil application must be sufficient to prevent a red-brown (rust) discoloration of the oil in the joints. The red-brown discoloration indicates that the oil in the joints is inadequate. When the rust discoloration is found, remove, clean, relubricate, and reinstall the chain before continuing operation. 15.4.2.2 Drip Lubrication For drip lubrication, oil is dripped between the link plate edges by a drip lubricator, as shown in Figure 15.10. Drip rates range from 4 to 20 or more drops per minute, depending on chain speed. Here again, the drip rate must be sufficient to prevent a red-brown (rust) discoloration of the lubricant in the chain joints. Care must be taken to prevent windage from misdirecting the oil drops. The oil level in the reservoir should be checked after each 8 h of operation, and the reservoir refilled when needed.
© 2006 by Taylor & Francis Group, LLC
Wire Rope and Chain
15-13
Oil filler cap
Oil level
Oil gauge Oil drain
FIGURE 15.11
Oil bath lubrication.
For multiple strand chains, a distribution pipe is needed to feed oil to all the rows of link plates, and a wick packing is usually required to distribute oil uniformly to all the holes in the pipe. 15.4.2.3 Oil Bath Lubrication For oil bath lubrication, a short section of the lower strand of the chain runs through a sump of oil in the chain casing (see Figure 15.11). The oil level should just reach the pitch-line of the chain at its lowest point in operation. Long sections of chain running through the oil bath, as in a nearly horizontal lower span, should be avoided because they can cause oil foaming and overheating. 15.4.2.4 Slinger Disc Lubrication In slinger disc lubrication, the chain operates above the oil level. A disc on one shaft picks oil up from the sump and slings it against a collector plate. Then, the oil usually flows into a trough, which applies it to the upper edges of the link plates in the lower strand of the chain. The diameter of the disc should produce rim speeds between 600 ft/min (183 m/min) and 8000 ft/min (2438 m/min). Lower speeds may not pick up the oil effectively, while higher speeds may cause oil foaming or overheating (see Figure 15.12). In both oil bath and slinger disc lubrication, the temperature of the oil bath and the chain should not exceed 180◦ F (82◦ C). Also, the volume of oil applied to the chain must be great enough to prevent the red brown (rust) discoloration of lubricant in the chain joints. The oil level in the sump of both oil bath and slinger disc systems should be checked after each 8 h of operation, and oil added when needed. At the same time, the systems should be checked for leaking, foaming, or overheating. 15.4.2.5 Oil Stream Lubrication With oil stream lubrication, a pump delivers oil under pressure to nozzles that direct a stream or spray onto the chain. The oil should be applied evenly across the width of the chain, and be directed onto the lower span from inside the chain loop. Excess oil collects in the bottom of the casing and is returned to the pump suction reservoir. A pressureregulating valve may be used to divert excess pump discharge to the reservoir. Oil cooling may be by radiation from the external surfaces of the reservoir or by a separate heat exchanger. Oil stream lubrication is always recommended for chains operating at relatively high speeds and loads (Figure 15.13). It is absolutely essential for roller chains that operate in the indicated galling region for any extended period of time. The oil stream not only lubricates the chain, but also cools the chain and
© 2006 by Taylor & Francis Group, LLC
15-14
Handbook of Lubrication and Tribology
Oil filler cap
Oil disc
Oil level Oil gauge Drain plug
FIGURE 15.12
Slinger disc lubrication. Oil filler cap
Casing split
Flexible hose
Sight flow Valve and strainer
Motor Oil pump
Oil spray bar
Oil gauge
FIGURE 15.13
Oil stream lubrication.
carries away wear debris from a drive being run at or near full rated capacity. The minimum oil flow rate for the amount of horsepower being transmitted is shown in Table 15.3. Here again, the oil level in the sump should be checked after each 8 h of operation, and oil added when needed. At the same time, the system should be checked for leaking or overheating [4]. 15.4.2.6 Automatic Oil Spray Lubrication One of the most effective lubrication methods is to have the chain lubricated in minute amounts continuously or at frequent intervals. As each chain link travels through it’s cycle and returns to the spray nozzle it
© 2006 by Taylor & Francis Group, LLC
Wire Rope and Chain
15-15 TABLE 15.3 Required Oil Flow for Chain Drives Transmitted
Min. required
HP
(kW)
gal/min
50 100 150 200 300 400 500 600 800 1000 1500 2000
(37) (75) (112) (149) (224) (298) (373) (447) (597) (746) (1119) (1491)
0.25 0.5 0.75 1 1.5 2 2.5 3 3.75 4.75 7 10
(L/min) (1) (2) (3) (4) (6) (8) (9) (11) (14) (18) (27) (38)
Pressure switch Oil/grease reservoir
Controller Gauge
Pump
Injectors oil/grease
Brush
FIGURE 15.14 Automatic chain lubrication.
is replenished with new and fresh lubricant. The low pressure “high velocity” air forces the fluid/lubricant into the critical wear points and simultaneously provides a cleaning effect. The actual amount of lubrication needed for proper chain lubrication is very small. Automatic oil spray systems can deliver adequate lubrication with very little oil consumption. Most automatic chain lubrication systems operate by supplying oil to the chain wear points with the assistance of low-pressure air. Chain lubrication by automatic systems can come in the form of brush type, spray type, or precision spray type (see Figure 15.14). Brush type systems do not regulate or provide precise amounts of lubrication. For brush chain lubrication, a brush is in direct contact with the chain. The supply of oil to the brush can come directly from the pump or through metering valves. The nonprecision spray types use nozzles that spray oil in a mist or atomized fashion onto the application. Precision spray lubrication provides sufficient control so as not to create mist or atomize the lubricant. An example of a brush type chain lubrication system is discussed below.
© 2006 by Taylor & Francis Group, LLC
15-16
Handbook of Lubrication and Tribology System fault light Oil reservoir
Nozzle assembly Oil line
Pressure switch Filter regulator for main air supply
Injector air valve
A
Regulated nozzle air supply
A
Injector
Nozzle air valve
Pin link plate
Nozzle air regulator
Pin Roller Section A-A
FIGURE 15.15
Bushing Roller link plate
Precision oil spray system with reservoir and controls and injectors.
15.4.2.7 Precision Oil Spray Lubrication This system delivers oil in very small quantities, measured in fractions of an inch in either a continuous stream or intermittent pulse over a fixed period of time. The on-time is usually determined by the time it takes the chain to pass through one complete cycle. Precision oil spray systems are used in applications where a small film of oil is required for proper lubrication. Chains can be effectively lubricated using very small amounts of oil. The main components of the system include a controller, injectors, and spray nozzles (Figure 15.15). The key to the operation of these systems is the mixing chamber in the spray nozzle assembly. Regulated air flows from the valve stack to the spray nozzle, through it, and out the nozzle tip. Oil is held captive between the injector and the spray nozzle with check valves in each. When the injector cycles, it delivers a fixed amount of oil into the spray nozzle body and into the air stream. The objective is to use the greatest amount of air pressure that does not allow the lubricant to atomize (mist). To determine the proper air pressure for an application, simply increase the air pressure until the misting is visible at the spray nozzle, and then reduce it slightly until misting is no longer present. The air pressure can be easily adjusted by the air pressure regulator. Oil enters the nozzle body when the injector cycles, flowing over the check valve, and into the air stream. Below the air pressure threshold of atomization (misting) the lubricant will cling to the inner surface of the spray nozzle body. It then migrates slowly toward the tip in a sine wave fashion. At the nozzle spray tip, the particles are too large to become airborne resulting in a controlled spray pattern. In systems where atomization is present, it is a result of the nozzle air being too high (see Figure 15.16). The reason for the precision is that applications for this type of lubrication require small amounts of lubrication. One of the main applications of using precision oil spray is for chain lubrication. Also, traditional brush application and mist type oil spray produce an excessive amount of lubrication resulting in waste. There are three basic types of precision oil spray systems. One is a fully controlled and monitored system complete with reservoir. The controller is an integral part of the precision oil system. The second type of precision oil system is the valve stack and oil reservoir. No controller is provided. The controls of the system are provided by a Program Logical Controller (PLC) that is part of the Original Equipment
© 2006 by Taylor & Francis Group, LLC
Wire Rope and Chain
Oil supply bleed port
15-17 Injector output to nozzle Air supply to nozzle
Alternate injector output to nozzle Alternate air supply to nozzle
4000 ml reservoir Oil filter
Primary nozzle air regulator Nozzle air valve Injector air valve
Level switch
Filter/regulator for main air supply
0.10 PSIG check valve Pressure switch
FIGURE 15.16
Reservoir fill port
Oil supply inlet from reservoir
Precision oil spray system with oil reservoir and injectors.
Manufactures (OEM) equipment. The third system is just a valve (injector) stack where controls of the system are provided by a PLC that is part of the OEM equipment. In this system there is no oil reservoir. The oil reservoir would be part of the OEM equipment also. Figure 15.17 shows a system set up with eight separate injectors. Each injector can direct the lubricant to eight independent points of applications to the chain. Precision oil spray of the type shown can be configured between 1 and 16 injectors. The heart of the system is the injector and nozzle. A schematic of the injector/nozzle is shown in Figure 15.18. 15.4.2.7.1 How the Injector Works The injector is simple and straightforward. It is a positive displacement piston pump. The injector operates under a supply of air pressure. When an electronic controller opens an air solenoid, a supply of air is directed to act against a piston. The air pressure acting across the piston forces the piston to move forward. The piston is directly in contact with a plunger that is in contact with an oil supply (Figure 15.19). As the piston and plunger move forward, oil is dispensed. The outlet of the oil is then directed to the spray nozzle where it mixes with air and is ultimately dispensed to the application. When the controller shuts the air supply to the solenoid, the air pressure is relieved and the piston is moved back to its original position by a control spring. The frequency at which the injector operates determines the amount of oil being supplied. Table 15.4 indicates the typical outputs of precision oil spray lubrication. The precision oil spray lubrication of the type in this section can dispense a 12 drop (defined as 0.015 cc or ml) consistently over a 4-min time interval. The normal viscosity grades range from 10 to 460 cSt (100 to 200 SUS) for precision oil spray to be effective. 15.4.2.7.2 Design Considerations for using Precision Oil Spray The main consideration for using precision oil spray is to prevent excessive lubricant waste, to maintain cleanliness as the lubricating oil is directly applied without dripping, and for environmental safety as there are no air borne particles produced. During the lubrication event, the on-time is usually determined by the amount of time the chain takes to completely pass the lubrication point. For long chains, multiple lubrication nozzles can be dispersed throughout the length of the chain.
© 2006 by Taylor & Francis Group, LLC
15-18
Handbook of Lubrication and Tribology Oil supply bleed port Injector output to nozzle
Alternate injector output to nozzle
Air supply to nozzle
Alternate air supply to nozzle Oil supply inlet from reservoir Primary nozzle air regulator Nozzle air valve Injector air valve Main system air inlet port 110 PSIG. max. 60 PSIG. min.
FIGURE 15.17
Precision oil spray system.
Injector check plug W/Viton O-ring
Alternate injector output to nozzle Alternate regulated air supply to nozzle
Injector pin
Regulated air supply to nozzle
Injector closure plug W/Viton O-ring
Injector piston
Viton packing
Viton O-ring Injector return spring
FIGURE 15.18
Lubrication injector.
© 2006 by Taylor & Francis Group, LLC
Injector check assembly Viton O-ring Injector body
Injector output to nozzle
Wire Rope and Chain
15-19 Check Controlled oil volume
Injector Spray nozzle
Regulated air
Oil spray
FIGURE 15.19 Lubrication injector operation.
TABLE 15.4
Chain Lubricant Usage Chart
Injector cycling every second
ml per min
ml per hour
ml per 8 h
ml per day (24 h day)
ml per week (24 h day)
ml per month (24 h day)
ml per year (24 h day)
Gallons per year (24 h day)
0.900 0.450 0.300 0.225 0.180 0.150 0.129 0.113 0.100 0.090 0.060 0.045 0.036 0.030 0.026 0.023 0.020 0.018 0.016 0.015 0.014 0.013 0.012 0.011 0.011 0.010 0.010 0.009 0.009 0.008 0.008 0.008 0.007 0.007 0.007
54.000 27.000 18.000 13.500 10.800 9.000 7.714 6.750 6.000 5.400 3.600 2.700 2.160 1.800 1.543 1.350 1.200 1.080 0.982 0.900 0.831 0.771 0.720 0.675 0.635 0.600 0.568 0.540 0.514 0.491 0.470 0.450 0.432 0.415 0.400
432 216 144 108 86 72 62 54 48 43 29 22 17 14 12 11 10 9 8 7 7 6 6 5 5 5 5 4 4 4 4 4 3 3 3
1,296 648 432 324 259 216 185 160 144 130 86 65 52 43 37 32 29 26 24 22 20 19 17 16 15 14 14 13 12 12 11 11 10 10 10
9,072 4,536 3,024 2,268 1,814 1,512 1,296 1,134 1,008 907 605 454 363 302 259 227 202 181 165 151 140 130 121 113 107 101 95 91 86 82 79 76 73 70 67
38,880 19,440 12,960 9,720 7,776 6,480 5,554 4,860 4,320 3,888 2,592 1,944 1,555 1,296 1,111 972 864 778 707 648 598 555 518 486 457 432 409 389 370 353 338 324 311 299 288
473,040 236,520 157,680 118,260 94,608 78,840 67,577 59,130 52,560 47,304 31,536 23,652 18,921 15,768 13,515 11,826 10,512 9,461 8,601 7,884 7,278 6,758 6,307 5,913 5,565 5,256 4,979 4,730 4,505 4,300 4,113 3,942 3,784 3,639 3,504
98.86 49.43 32.95 24.71 19.77 16.48 14.12 12.36 10.98 9.89 6.59 4.94 3.95 3.30 2.82 2.47 2.20 1.98 1.80 1.65 1.52 1.41 1.32 1.24 1.16 1.10 1.04 0.99 0.94 0.90 0.86 0.82 0.79 0.76 0.73
1 2 3 4 5 6 7 8 9 10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120 125 130 135
Note: 0.015 ml = 12 drop; 1 gallon = 3.785 L = 3,785 ml; 0.030 ml = 1 drop; 1 gallon = 126,166 drops.
The next consideration is to supply the correct amount of air pressure so as not to create an oil mist. Each precision oil system comes complete with an air pressure and airflow regulator in the valve stack. As some of the lubrication points will be at further distances than others, the amount of air pressure and flow can be adjusted to compensate for respective back pressure associated with longer air/oil supply lines.
© 2006 by Taylor & Francis Group, LLC
15-20
Handbook of Lubrication and Tribology
Nozzle
Apply between inside/outside links to reach pin Pin (outside) link plate Pin
Roller
FIGURE 15.20
Bushing Roller (inside) link plate
Lubrication nozzle positioning.
The lube adjustments can take place by setting the timer frequency to the injectors. The higher the frequency that the injectors operate per minute, the more the lubricating oil that will be dispensed. The nozzle spray tips should be located no more than one inch from the spray surface. The closer the nozzle tip to the chain, the better. It is an advantage to locate the nozzles in an area where there is minimal chain deflection (see Figure 15.20).
References [1] Wire Rope Users Manual, 3rd ed., Wire Rope Technical Board, 1993. [2] Machinery Lubrication Magazine, Lubrication Basics for Wire Ropes, Issue July 2002. [3] Chains for Power Transmission and Material Handling, Design and Applications Handbook, American Chain Association; Mercel Dekker, 1982. [4] Identification Installation Lubrication and Maintenance of Power Transmission Roller Chains in ANSI B29.1 and ANSI B29.2. American Chain Association, 2003.
© 2006 by Taylor & Francis Group, LLC
16 Tribology of Hard Disk Drives — Magnetic Data Storage Technology 16.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16.2 Hard Disk Drive — Description . . . . . . . . . . . . . . . . . . . . . . 16.3 The Disk . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
16-1 16-4 16-7
Substrate • Hard Protective Coating • Disk Lubricant
16.4 The Head/Slider/Air Bearing Surface . . . . . . . . . . . . . . . . .
16-23
Hard Protective Coating • The Air Bearing Surface (ABS)
16.5 Contact Start–Stop. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
16-29
Stiction During Contact Start/Stop • Wear During Contact Start/Stop
16.6 Load/Unload . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16.7 Ultralow Flyability and Contact Recording . . . . . . . . . .
16-33 16-35
Ultralow Flying • Contact Recording — the Last Frontier
José Castillo Iomega Corporation
Bharat Bhushan Ohio State University
16.8 Closure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Acknowledgments. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Nomenclature and Terminology — Hard Disk Drive Industry . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
16-39 16-39 16-40 16-40
16.1 Introduction Hard disk drives (HDDs) have been the basis for the explosion of information available in the last decades. In 2002, approximately 5 exabytes1 of information were produced in the world and 92% were stored on magnetic media, mostly (>50%) stored in HDDs [1]. The estimated number of HDDs sold during 2003 reached 235 million, capable of storing more than 15 exabytes of data. That corresponds to a 15,000% increase in storage capability since 1995 [1]. At the same time, the cost per gigabyte (GB) has dropped dramatically, below $0.50/GB by 2005, making HDDs affordable for many other applications beyond computers. It is expected that 13% of all HDD shipped in 2005 will be used in consumer application 11
exabyte = 1018 bytes. (5 exabytes equals all the words ever spoken by human beings.)
16-1
© 2006 by Taylor & Francis Group, LLC
Handbook of Lubrication and Tribology
0
10
20
mm
30
40
50
16-2
0
10
20
30
40
50
mm
FIGURE 16.1 The smallest HDD presently manufactured with a capacity of 2 GB. Dimensions: 3.3 × 24 × 32 mm3 (0.13 in. × 0.94 in. × 1.26 in.) Weight: less than 10 g. A slightly thicker version (5 mm) with 4 GB capacity is also available. (With permission from Toshiba Corp.)
products others than computers, such as portable digital music players (like the iPod®), personal video recorders (Tivo®), and video game boxes (X-box®) [2]. The HDD has evolved dramatically in the last 20 years, but its conceptual design has not changed much since the first HDD introduced by IBM in 1956 (2006 will be the 50th year of the HDD). The first HDD consisted of 50 disks, each 24 in. in diameter, and stored about 5 MB of data. Today a popular laptop HDD stores up to 100 GB and weighs only 99 g. In 2005, Toshiba introduced a HDD smaller than a compact flash memory that holds 2 GB of memory (see Figure 16.1). The highest areal density (maximum amount of bits per unit area in the disk) will reach 133 Gbit/in2 by the end of 2005; this implies that one bit of information can be placed in a square of 70 × 70 nm2 . This is equivalent to storing more than 910,000 bits in the cross-section of an average human hair.2 It is expected that areal density will have reached the 1 Tbpsi3 mark before the end of the decade. Table 16.1 shows the different standard sizes available for today’s HDD and its most common applications. The magnetic recording process in HDDs involves relative motion between the disk (storage media) and the head (write and read element). During the process of writing and reading, the disk is always rotating under the head. By reducing the distance between the magnetic head and the disk, the signal from the head (or media) increases exponentially [3]. From the magnetic point of view no separation between the head and the magnetic layer on the disk is most desirable. However, the present state of the art requires that some separation is maintained.4 For corrosion and wear protections, the head and media are covered with a hard protective layer and, to avoid detrimental effects if these hard layers ever touch (head–disk contact), a molecular layer of lubricant is placed on the disk [4]. The number one tribological challenge in the HDD industry is to achieve the smallest possible separation between the head and the disk during relative motion, and control it in such a way that it does not change under any circumstance. As of today 2005 HDDs have already reached amazingly small separation between the head and the disk. The average flying distance is expected to be only 8 nm for the present leading HDD.5 This number is expected to average human hair is 75 µm in diameter. = Tera bit per square inch. Tera = 1012 bits. 4 The chemical sensitivity of the present magnetic materials requires that they be protected against corrosion by covering them with an overcoat, and it is imperative that this overcoat does not wear, otherwise the data would be destroyed. 5 A new approach where the reader-writer elements are protruded towards the media, only when needed is the latest (2005) approach to reduce the separation while keeping the flying height at “safer” distances most of the time. The separation could be reduced to 2–4 nm using this new approach. 2 An
3 Tbpsi
© 2006 by Taylor & Francis Group, LLC
Nominal dimension
Sizes and Target Applications of Today’s HDDs Physical dimension (height × width × depth)
Disk size
Max capacity∗
Application
3.5"
1" × 4" × 5.75" (25.4 × 101.6 × 146 mm3 )
95 mm (84 or 65 High end HDD)
500 GB (133 GB/Platter max)
2.5"
0.37" × 2.75" × 3.94" (9.5 × 70 × 100 mm3 )
65 mm
120 GB (60 GB/Platter max)
1.8"
0.315" × 2.12" × 3.09" (8 × 54 × 78.5 mm3 )
PC Card (PCMCIA)
1"
0.197" × 1.69" × 1.43" (5 × 42.8 × 36.4 mm3 )
CompactFlash
60 GB (30 GB/Platter max) 80 GB (40 GB/Platter max)∗∗ 8 GB
0.85"
0.13" × 0.94" × 1.26" (3.3 × 24 × 32 mm3 )
Ultra-small
4 GB (slightly thicker)
Standard platter size. The most common hard disk drives used in PCs and PVR and other nonportable electronics. Some High End (10K or 15K rpm) have smaller disks inside but the basic external dimension is the same) Laptop drives, small form factor PCs, High-end servers and automotive applications PC Card (PCMCIA) drives for laptops and some MP3 players such as the iPod Digital cameras, hand-held PCs, some MP3 players and other consumer electronic devices Traget aplications Cellphones and other consumer electronic devices
∗ as September 2005, ∗∗ to be available by the end of 2005.
16-3
© 2006 by Taylor & Francis Group, LLC
Tribology of Hard Disk Drives — Magnetic Data Storage Technology
TABLE 16.1
16-4
Handbook of Lubrication and Tribology
(a)
(b) Slider unloaded
Slider parked Stop position Slider
Slider
Ramp
FIGURE 16.2 Schematics of the two different technologies used to rest the heads while HDD is off. (a) CSS — The heads rest in a special textured zone and (b) L/UL heads rest in a ramp, outside of disk surface.
be only 3.5 nm to reach the 1 Tbpsi [5]. These distances are even more amazing if we consider that the relative velocity between the head and the disk could reach 47 m/sec (105 mph). The other big tribological challenge in the head–disk interaction involves starting and stopping the disk. The question is: where should the heads be when the rotation of the disk starts, and how does the head reach stable flying conditions without causing some damage to the disk or the head itself? Engineers have responded with two different approaches each of which has its intrinsic challenges (see Figure 16.2). In the majority of the desktop HDDs, the heads rest on a specially designed area on the inside of the disk. After the disk has reached steady rotational conditions (or just before that), the heads are moved to the data zone of the disk. This process requires that the head drags over the disk for a short time every time we start or stop the HDD, while enough air flows under the head to make it fly. These types of drives are called CSS (Contact Start–Stop) and they require passing tests of 50,000 CSS to be considered robust. The special challenges are: (1) to avoid heads from getting stuck on the disk (stiction), and (2) to avoid excessive debris generation during dragging for the life of the HDD. An alternative approach to having the head land over the disk surface during starts and stops is to have them stored outside the disks and bring them to the disk gently enough such that no significant damaging interaction occurs between the head and the disk. This approach is called load/unload (L/UL), since the heads have to be loaded and unloaded every time the disk starts and stops. This concept is standard in all the HDDs used in portable products, such as laptops, because having the heads separated from the media decreases the possibility of shock damage during nonoperating conditions. This chapter will focus on the state-of-the-art and tribology challenges of the head and disk in the HDDs. We will present the characteristics of the head and the disk, from the tribological point of view. Emphasis will be given to the lubricant placed on the disk. We will then examine the tribological challenges (1) wear and stiction for the CSS concept, (2) wear and dynamics for the L/UL systems, and (3) dynamic stability of the head while flying ultra low. In this last part, we will present the latest studies on contact recording as the last frontier of reducing the head–disk separation.
16.2 Hard Disk Drive — Description Before discussing the individual elements and interaction between them, we present an overall description of the HDD (see Figure 16.3). During operation, the disk is rotated by a small motor that also contains the spindle (in Figure 16.3, the rotation direction is CCW [counter clockwise]). The trend is to use fluid dynamic bearing (FDB) spindle motors, to decrease instabilities caused by ball bearing spindles, increase shock resistance, and reduce noise. Motor speeds range from 3,600 rpm, still common on small devices, to 15,000 rpm, used on HDD intended for server applications where response time is critical. The read/write element is located at the end of the slider, also called head slider as described in Figure 16.4. The side of
© 2006 by Taylor & Francis Group, LLC
Tribology of Hard Disk Drives — Magnetic Data Storage Technology
16-5
Cover mounting holes (cover not shown) Base casting Spindle Slider (and head) Actuator arm Actuator axis
Case mounting holes
Actuator
Platters Ribbon cable (attaches heads to logic board)
SCSI interface connector Jumper pins Jumper
FIGURE 16.3
Power connector
Tape seal
Inside of a HDD with its components described.
Disk Disk rotation
Read / write elements Actuator arm
FIGURE 16.4
Head slider
Schematic diagram of a HDD with one disk and two heads.
the slider facing the media incorporates a special shape called air bearing surface (ABS) that allows the head element to follow the disk surface at an established distance (<12 nm in today’s products) as shown in Figure 16.5. The head slider is attached to a suspension and actuator arm. The head is moved from one location to another on the disk by a voice coil motor, which uses a closed loop feedback from the read element to identify where the head is on the disk. Before this system can work, the disk has to be marked (servo marks) such that the head can check its location very frequently. In present HDDs, servo marks are read by the head every few degrees of rotation. It is common to find marks every 2.5◦ or even less than 2◦ . Each servo line (or servo wedge) has enough information to tell the head its exact location, as well as to help the head to position itself in the exact center of each track. Of course this information takes away valuable data storage space (∼2 to 4% of the disk), but without this information the HDD could not work as we know it. Presently, the distance between each track (circumferential) is smaller than 0.25 µm, and it will keep decreasing for the next generations.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
Read/write elements Slider Air bearing surface
Air flow Disk Arm Moving direction
FIGURE 16.5
Schematic diagram of head slider over the disk showing the ABS.
Write current
(a)
Write head Induced voltage Read head Flux lines from magnetized medium Medium motion Thickness
Wa ve l e
h
dt
ngt
h
k ac Tr
wi
Gap
(b)
Head core Spacing Magnetic medium Substrate
FIGURE 16.6 Sketch of the recording process in the HDD. (a) Illustration of the recording and reproduction process and (b) schematic of cross-sectional view showing the magnetic field at the gap.
© 2006 by Taylor & Francis Group, LLC
Tribology of Hard Disk Drives — Magnetic Data Storage Technology
16-7
The process of writing and reading the data is a very complex one, but it can be simplified as presented in Figure 16.6.6 To record or write data, a current is sent to the coil on the ring-shaped electromagnet, generating a magnetic field. This magnetic field travels through the ring and reaches an area with a gap located close to the magnetic medium. This gap is specifically designed such that the magnetic field has to travel through air as indicated in Figure 16.6(b). If this field is strong enough, it changes the magnetization direction of the magnetic medium on the disk. The direction of the magnetic medium is permanent until another large enough magnetic field passes through it. To read this data, Figure 16.6 shows the old inductive technology approach of using the reversed mechanism described above. By getting the gap of the ring-shaped electromagnet close enough to the constant magnetic field generated by the magnetic medium, a magnetic field starts traveling through the ring that induces a voltage on the coils placed around the ring as indicated in Figure 16.6. Today HDDs use a writing element very similar to the one described but the reading element uses the magneto-resistance (MR) effect of some materials. From the tribological point of view, it is important to understand that the distance between head and disk is the most important parameter to control magnetic signal amplitude; in other words, the amount of written data depends directly on the disk–head separation. While traveling over the disk, the head sliders are usually not aligned with the disk, so there is a skew angle that is formed between the slider and the disk. This skew angle affects how the air enters the interface between the ABS and the media, influencing the flying performance. The ABS has to be designed such that these changes of air direction, along with the differences of air velocity from the outside to the inside of the disk, are accommodated and the head element always stays at the desired distance from the disk. As we will see in Section 16.7 on flyability, this is a well-understood requirement and there have been different approaches to satisfy this objective. While not in operation (resting), the heads are located either over a very specific area on the disk for the CSS drives (landing zone) or removed from the disk and stored on the loading ramp (see Figure 16.2). In the case of the CSS drives, the biggest concern is that the heads get stuck on the media. This phenomenon, called stiction, is characteristic of very flat and polished surfaces when they are left in contact for a period of time (coefficient of friction can reach values >1). For this reason, special attention has been placed on the disk and the slider such that stiction does not happen. Most of the portable designs that require high robustness to shock store the heads in a special location outside the disk called a ramp. The challenge in these cases is to bring the head to the disk and take it out smoothly enough so that no damage occurs either to the head or the disk. Significant advances in L/UL technology, especially new ABS and suspension features, are allowing heads to load and unload for millions of cycles without any measurable damage.
16.3 The Disk The disks used in HDDs have evolved to the point where some of the basic characteristics, such as dimensions and material properties, have been standardized. The International Disk Drive Equipment and Materials Association (IDEMA) has been in charge of this standardization and has already produced a significant set of standards to guide the companies that manufacture these disks. Table 16.2 summarizes some of these specifications. The use of standards was of great value when the number of companies supplying disks was significant, but the standardization process has lost some relevance since there are now very few companies that can supply this highly specialized disks. This fact is evidenced by the arrival of a new generation’s selection of thicker and smaller disks that do not comply with established specifications. Today, the most common thickness of 95 mm media is 1.27 mm, and there are very small disks on the 6 The figure and explanation presented here correspond to Longitudinal Magnetic Recording (LMR) technology, that has dominated the industry as up today (2005). It is expected that next generations of HDD will be using Perpendicular Magnetic Recording (PMR) technology, where the magnetization direction on the disk is perpendicular to the surface, instead of parallel to the disk surface as shown in the Figure 16.6.
© 2006 by Taylor & Francis Group, LLC
16-8
TABLE 16.2
Disk Standard Dimensions (IDEMA)
Nominal diameter → Parameter
IDEMA reference
48 mm (1.8")
84 mm (3.0")
95 mm (3.5")
Tol.
Nominal
Tol.
Nominal
Tol.
Nominal
Tol.
Nominal
Tol.
[mm] [mm] [mm] [mm] [nm]
34 8
±0.05 +0.05, −0.00 max ±0.025 max max max max max max
48 12 0.025 0.635 2.5 25 5 13 1.3 4
±0.05 +0.03, −0.00 max ±0.02 max max max max max max
65 20 0.05 0.635 3 25 3 13 2 4
±0.10 +0.05, −0.00 max ±0.025 max max max max max max
84 25 0.01 0.8 0.40 −4.00 nm
±0.05 +0.05, −0.00 max ±0.025
8 Drive specific 5 2.5
max
95 25 0.01 1 2 20 8 Drive specific 5 2.5
±0.05 +0.05, −0.00 max ±0.013 max max max
[µm] [µm] [µm] [nm]
0.381 — — — — — — D16.93
Other dimensions available are 27.4 mm (1") and 21.6 mm (0.85").
© 2006 by Taylor & Francis Group, LLC
65 mm (2.5")
Nominal
D2.1-91
D13-93
D19-96
max max
D15-98
max max
Handbook of Lubrication and Tribology
Outside diameter (OD) Inside diameter (ID) Concentricity Thickness Roughness Ra Roughness Rp Flatness Runout (TIR) Clamping area flatness Clamping area roughness Ra
34 mm Units
0
View angle Light angle
A
0
A
–75
Digital instruments nanoscope Scan size 100.0 mm Scan rate 0.5003 Hz Number of samples 256 Image data Height Data scale 150.0 nm
16-9
75
Tribology of Hard Disk Drives — Magnetic Data Storage Technology
20.0 40.0 60.0 mm
80.0
A
A
20 40 60 80
mm
X 20.000 mm / div Z 150.000 nm / div
0 deg
FIGURE 16.7 Laser zone of the disk with profile of the bumps obtained using an AFM. Spacing between the bumps is 35 × 25 µm. Cross-section shows bumps of nominal diameter 7.5 µm and nominal height 20 nm. Liquid lubricant 1–2 nm Diamond-like carbon overcoat 2–4 nm Magnetic coating ~5 nm Underlayer Al–Mg/10 mm Ni–P, Glass or glass–ceramic 0.635–1.75 mm
FIGURE 16.8
Composition of the top layers of the disk.
market such as the 0.85 in. (21.6 mm) disk used in the newest Toshiba HDD (see Figure 16.1). Neither of these disk sizes is standardized. In the majority of the 3.5 in. HDDs (95 mm disk), the heads rest on a specific area toward the inside of the disk called the landing zone while the disk is not rotating, specially designed to reduce the static friction at the head–disk interface (stiction — see Section 16.5 on CSS). Figure 16.7 shows atomic force microscope (AFM) examinations of this zone on today’s disks. The “bumps” create a well-defined topography that reduces the contact area between head and disk. The process of generating these bumps is called “laser texturing” because it consists of a high repetition discharge of a laser beam while the disk is rotating, forming a spiral sequence of these bumps. To create the bumps, a laser melts a microscopic area of the disk substrate that solidifies in different shapes, dependent upon the duration and intensity of the laser beam. The height of these bumps has to be such that the heads can land after flying, so the height of these bumps has to be tightly controlled. A detailed discussion of these shapes will be given in Section 16.5 on CSS. To allow the head to fly low, the surface finish of the data zone has to be as polished as possible. Although the goal is for the surface to be as smooth as possible (values of 1.5 Å Ra are already obtained), the media is “textured” in the circumferential direction before the magnetic layer is sputtered to increase the magnetic orientation, hence enhancing the magnetic performance. The general composition of the media is illustrated in the cross-sectional diagram, in Figure 16.8. The mechanical characteristics are dominated at the macroscopic level (flatness, micro-waviness, flutter, and
© 2006 by Taylor & Francis Group, LLC
16-10
Handbook of Lubrication and Tribology
shock resistance) by the substrate and at the microscopic level (wear, chemical inertness, friction) by the overcoat and lubricant layer. The magnetic layer, illustrated as one layer, is in reality a complex system of layers designed to enhance the magnetic performance. Before sputtering the magnetic layer, a set of “seed” layers is placed over the substrate to create a good shield that will hold the magnetic material in place (avoid migration). In most of today’s products, the magnetic system consists of a three-atom-thick layer of ruthenium (a precious metal similar to platinum) sandwiched between two magnetic layers (disk manufacture, personal communication). Because the main components of the magnetic layer are cobaltbased alloys (e.g., Co–Pt–Cr, Co–Pt–Ni), it is important to avoid exposure of this material to oxygen in order to prevent corrosion. In this section we describe the substrate and focus on the tribological aspects of the hard protective coating and the lubricant layer.
16.3.1 Substrate Two basic configurations are available today: one based on Al–Mg substrate and the other on glass substrate. Each of these substrates has specific characteristics that have satisfied specific needs of the HDD industry for a long time [6, 7] and it seems that they will coexist in the future. The Al–Mg substrate has the advantage of cost when compared with the glass substrate, and more recently, its relatively softer characteristics have procured manufacturability benefits allowing better texturing that improve magnetic layer orientation. On the other hand, glass substrate disks have maintained superiority in applications where shock resistance is critical: in the event of head slapping the media during operation due to shock, the heads could plastically deform the Al–Mg media more easily than the glass substrate media. Another characteristic advantage of the glass substrate is its higher specific stiffness that allows the use of thinner disks to obtain similar vibration characteristics. Table 16.3 shows the main mechanical characteristics of these substrates. Other characteristics that have played a role in the applicability of these two types of substrate are their thermal expansion and thermal conductivity. Thermal expansion is critical in the manufacture of HDDs since they are expected to perform well from approximately 0◦ C to as high as 60◦ C.7 The design has to be such that thermal expansion does not contribute to the deformation of the disks, so a well-selected set of matching parts (disk clamping material) is necessary for different substrates. This fact alone has given the Al–Mg substrate an advantage over glass, since it has been easier (and less expensive) to match its thermal expansion with that of the parts that come in contact with the disk. The thermal conductivity between glass and Al–Mg has also been of importance in design considerations. In the event of high-speed contact between head and disk, the heat dissipation is important, especially because the head read-write element is hot during use. A substrate that allows a quick dissipation of this temperature spike is beneficial. Suk et al. [8], compared the behavior of glass and Al–Mg substrate under an impact event, and found that while the glass disk suffered from erasure of the magnetic information due to the temperature increase caused by the impact, the Al–Mg disk was deformed and the magnetic layer was damaged by physical (plastic) deformation. The erasure that occurred to the glass disk was rewritable since the magnetic layer was not damaged. Improvements to the shock events have been obtained by the use of lubricant (as a soft protective layer), by improving the thermal conductivity of the overcoat, and by reducing the head shock energy (smaller sliders). These improvements have significantly decreased the energy interaction between head and disk when contact occurs. One of the advantages that ushered in the use of glass substrate was its inherently good surface roughness. However, manufacturing processes have advanced in such a way that disk manufacturers have managed to obtain surface roughness on Al–Mg disk as good as that of glass substrate or even better. Also, Al–Mg has manufacture benefits by allowing easier orientation of the magnetic layer. An example of surface roughness measurements using AFM is presented in Figure 16.9. The flutter or axial run-out characteristic, controlled by substrate composition and thickness, is of increasing concern. The trend has been to increase the thickness to reduce the effect. Some HDD companies 7 This
range is larger (−30◦ C to 80◦ C) for recent applications like the automotive.
© 2006 by Taylor & Francis Group, LLC
Tribology of Hard Disk Drives — Magnetic Data Storage Technology TABLE 16.3
16-11
Selected Properties of Disk Substrate Material —- Comparison
Properties Density Modulus of elasticity Specific modulus Yield strength Coeff thermal exp Thermal conductivity Heat resistance Electrical resistivity Hardness, Knoop Toughness
Units
Al-Mg (NiP)a
Glass (Alumino-Silicate)
Glass (Crystallized)
[g/cc] [GPa] [GPa/(g/cc)] [MPa] [ppm/◦ C] [W/m/◦ K] [◦ C−1 ] [ohm-cm] [Kgf/mm2 ] [MPa-m1/2 ]
2.7 71 26.3 117 24.2 118 280 low 90b 20–25
2.2–2.6 85 32.7–38.6 600 9 3 560 10+13 620 1.0–2.0
∼2.42 92–94 38.0–38.8 240 7.0–8.0 2 650 10+15 680 1.0–2.0
nm
0 –5.0
Digital instruments nanoscope Scan size 10.0 mm Scan rate 0.5003 Hz Number of samples 512 Image data Height Data scale 10.0 nm
5.0
Reference: IDEMA Standard D20-96 a Generally heat-treatable aluminum-magnesium alloy AISI 5086 (95.4% Al, 4% Mg, 0.4% Mn and 0.15% Cr). NiP layer is electroless plated (90-10 wt%) b Depending on the amount of NiP layer this hardness can reach ∼600–800 kgf/mm2
0
2.00 4.00 6.00 mm
8.00
View angle Light angle
2
4 6 8
mm
X 2.000 mm / div Z 10.000 nm / div
0 deg
FIGURE 16.9 Data zone of a super-polished disk with profile obtained using an AFM. Surface roughness Ra = 2 Å.
are using 1.75 mm thick disks just to reduce this run-out, even though the need for larger disk mass and larger motors will increase costs.
16.3.2 Hard Protective Coating The magnetic layers in the disks are relatively soft and very susceptible to corrosion. That is the main reason why a hard coating is needed. In today’s hard disk, the dominant overcoat is amorphous hydrogenated and nitrogenated diamond-like carbon (DLC) (a-C : H and a-C : N). The main characteristics required for an overcoat are excellent topographical conformity, thickness uniformity, good adhesion to the magnetic layer, high density (no porosities), chemical inertness (no interaction with the magnetic layer or possible ambient molecules), but enough reactivity to form strong bonding with the lubricant, and high electrical resistivity for better electrochemical resistance [9]. All this has to be accomplished with only a few nanometers of coverage. Today (2005) disks are covered with only 2 to 3 nm of total overcoat. These
© 2006 by Taylor & Francis Group, LLC
16-12
Handbook of Lubrication and Tribology sp3 (diamond) Filtered cathodic arc ta–
C
Ion-assisted (PECVD, ion-beam)
a– C
a–
C– N
Laser ablation
Sputtered
:H
C I–
C:H
a–
Poly m
er
sp2 (Graphite)
FIGURE 16.10
Polymer –CH2–
Schematic illustration of various carbon films with respect to their sp2 –sp3 bond contents.
ultrathin overcoats have significantly increased the concern of corrosion because of the possibility of pin-holes that allow environmental exposure of the magnetic layer. The search for thinner overcoats has resulted in different sputtering and deposition techniques. The DLC coating can be deposited by DC/RF sputtering, RF-plasma enhanced chemical vapor deposition (RF-PECVD), electron-cyclotron resonance chemical vapor deposition (ECR-CVD), direct ion-beam deposition (IBD), and filtered cathodic arc (FCA) deposition. The structure and properties of these coatings depend on the deposition techniques and the deposition parameters, as well as incorporation of other elements such as nitrogen. The goal is to obtain a structure as similar as possible to diamond, or sp3 carbon. However this is not possible and the structure formed a random network of covalently bonded carbon in hybridized tetragonal — sp3 (diamond) — and trigonal — sp2 (graphite) — local coordination, with some of the bonds terminated to hydrogen. Figure 16.10 shows a schematic illustration of the various carbon films with respect to their sp3 /sp2 fraction. The larger the ratio the more diamond-like the structure. It has been reported that sp3 /sp2 fraction is in the following decreasing order: FCA deposition, IBD, PECVD, and sputtering [10,11]. Table 16.4 shows some characteristics of DLC deposited using different techniques. The density and hardness of the films have been related to the energy of deposition. High-energy deposition processes produce harder and denser coatings. FCA, ECR-CVD, and IBD processes are energetic, that is, carbon species arrive with energy significantly greater than that represented by the substrate temperature. The resultant coatings are, as indicated above, amorphous in structure, with hydrogen content up to 50%, and display a high degree of sp3 [12,13]. It is known that deposition of sp3 -bonded carbon requires the deposition species to have kinetic energies in the order of 100 eV or higher, well above those obtained in thermal processes such as evaporation (∼0.1 eV). The density of sputtered amorphous carbon films, using the most commonly used techniques, falls in the range of 1.5 to 1.7 g/cc, far shorter of either graphite (2.2 g/cc) or diamond (3.4 g/cc) [14]. The porosity of the films limits the thickness at which pin-hole-free coverage is achieved. Figure 16.11 shows a schematic comparison between sputtered and IBD of amorphous DLC [5]. Figure 16.12 shows a schematic representation of how the environment can attack the magnetic layer through the pin-holes. It has also been found that when the kinetic energy of the carbon species is too large, the excess amount of energy allows the film to relax into a more graphite-like structure, leading to a less dense structure or even island formation in extreme cases [5]. Another disadvantage of high deposition energy is the possibility of carbon atoms penetrating the magnetic film. As Gui [5] indicates, penetration can reach 0.8 nm at 100 eV/atom energy level. In the search for the best overcoat, several authors have reported the effects of doping the carbon with nitrogen [15–19]. They have found that by adding nitrogen to the amorphous carbon, there is an increase of the sp2 content, resulting in lower density and hardness. However, they have also found that
© 2006 by Taylor & Francis Group, LLC
Tribology of Hard Disk Drives — Magnetic Data Storage Technology TABLE 16.4
16-13
Selected Properties of DLC Coatings Obtained with Several Deposition Processes
Sample Cathodic-arc carbon coating (a-C) Ion-beam carbon coating (a-C : H) ECR-CVD carbon coating (a-C : H) DC sputtered carbon coating (a-C : H) Bulk graphite (for comparison) Diamond (for comparison) Single crystal silicon substrate
Hardnessa (GPa)
Elastic modulusa (GPa)
Fracture toughnessb Klc (MPa-m1/2 )
Critical loadc (mN)
Coefficient of friction during accelerated wear testingc
24
280
11.8
3.8
0.18
12.5
19
140
4.3
2.3
0.18
1.5
22
180
6.4
5.3
0.22
0.6
15
140
2.8
1.1
0.32
2.0
Very soft
9–15
—
—
—
—
10–104 11
900–1050 165
— —
— 0.5
— 0.55
— 0.02
Residual stressd (GPa)
a Measured on 100-nm thick coatings on single-crystal silicon substrate at a peak load of 0.2 mN with an indentation depth of 10 to 30 nm. b Measured on 100-nm thick coatings on single-crystal silicon substrates. c Measured on 100-nm thick coatings on single-crystal silicon substrates slid against 3-mm diameter sapphire ball at a normal load of 200 mN. d Measured on 400-nm thick coatings on single-crystal silicon substrates. Source: From Bhushan, B. (Ed.), Modern Tribology Handbook, CRC Press, Boca Raton, FL, 1413, 2001. With permission.
(a)
Energy 10 eV or less
(b)
Energy 40 – 100 eV
FIGURE 16.11 Schematic of a comparison between (a) sputter deposition and (b) IBD of an amorphous carbon film. (From Gui, J., IEEE Trans. Magn., 39, 716, 2003. With permission.)
H2O, O2
H2O, O2
H2O, O2
Pinholes Co-alloy magnetic film
FIGURE 16.12 Schematic of environmental attacks of magnetic thin film by water and oxygen through atomic channels, or “pinholes,” in the ultrathin protective overcoat. (From Gui, J., IEEE Trans. Magn., 39, 716, 2003. With permission.)
© 2006 by Taylor & Francis Group, LLC
16-14
Handbook of Lubrication and Tribology
the durability measured by scratch and micro-wear test increased with the amount of nitrogen (N content maximum of 21%) [15]. They also reported a maximum density of 2.1 g/cc and hardness of 25 GPa achieved with plasma beam source [16]. Even higher density and hardness were reported by using FCA: Yamamoto et al. [18] showed a more sp3 -like film (called tetrahedral amorphous carbon) with a density of 2.7 g/cc and hardness of 26 GPa. They also showed a superior performance of this more compact carbon compared with common a-C in both corrosion and micro-wear tests. Several studies present optimization of the deposition techniques to improve the resulting overcoat. Wang [20] showed that by lowering the sputtering temperatures (130◦ C) they produced a higher wear resistance than high-temperature sputtered carbon (212◦ C). Li et al. [21], showed that, by proper deposition conditions combined with substrate tilt and rotation of the disk during sputtering, the magnetron sputtering can produce a 1-nm thick CNx (nitrogenated carbon) overcoat with atomically smooth morphology and acceptable corrosion performance. Pirzada et al. [22] proposed a new combined system for the carbon overcoat: the lower layer would be an Ion Beam Carbon (IBC) layer and only the top 1 nm would be nitrogenated sputtered carbon. They called this system functionalized ion-beam (FIB) carbon. This overcoat system showed higher lubricant bonding ratio than IBC. A stronger carbon–lubricant interaction prevents lubricant degradation/desorption during CSS and flying, resulting in less lubricant pick-up by the head. The FIB carbon also showed lower resistivity and improved corrosion protection. Stiction was decreased in FIB carbon as compared to IBC. These thin overcoats (3 nm) have shown excellent tribological performance. At the present time IBD and PECVD are the preferred deposition techniques used today and they are already capable of supporting around 2 to 3 nm thin DLC with good corrosion and wear resistance. Even with this thin thicknesses, the trend is to vary the carbon deposition sequence, such that the bottom layer, on top of the magnetic layer has good adhesion and minimal effects to the magnetic properties, and the top layer has good bonding characteristics for the lubricant to be placed on top.
16.3.3 Disk Lubricant Although in the present HDD designs the head is expected to fly over the media during operation, in most of the HDDs the head actually rubs against the media every time the disk starts or stops. There are also many other opportunities for the head to touch or slide over the disk. During these events, it is necessary to have a soft layer that can absorb the energy of the impact or reduce the friction during sliding. This is the main reason why all the disks have a layer of very specialized lubricant. At the same time it is important that this lubricant layer be as thin as possible to reduce stiction issues (to be discussed later) and minimize the effect of physical separation between the head and the magnetic layer of the media. Figure 16.13 shows how essential this lubricant is for reaching acceptable disk life [23]. The study shows a rapid increase of coefficient of friction for the nonlubricated disk compared with the lubricated disk when a slider made of Al2 O3 –TiC slides over a disk surface. In the HDD industry, perfluoropolyethers (PFPEs), a lubricant family, has dominated for many years. PFPEs have been used for over 20 yr [24] because they have most of the desirable characteristics for the application: low volatility, acceptable thermal and oxidative stability, low surface tension and chemical inertness. There have been many variations around this backbone structure, with the intention of improving some of its weaknesses. Table 16.5 shows a summary of the lubricants used in the industry, most of them based on the PFPE backbone. The ideal lubricant for HDDs should be well attached to the media in a way that it will not be spun off with media rotation, and also that will not easily shear from the surface due to head sliding contact. In the case where the lubricant is sheared from the media by the energy of head contact, it is important that it is capable of moving back to the affected area quickly. This dual required characteristic of the lubricant system made necessary the combination of different molecules, some of them bonded to the surface and others more freely to move in case of sliding removal. It is typical that lubricant systems in today’s disks have a 50–50% to 70–30% bonded to not-bonded (or unbonded) ratio.
© 2006 by Taylor & Francis Group, LLC
Tribology of Hard Disk Drives — Magnetic Data Storage Technology
16-15
1.6
Friction coefficient
Unlubricated disk 1.2
0.8 Lubricated disk
0.4
0.0 0
2,000
4,000
6,000
8,000
10,000
Number of revolutions
FIGURE 16.13 Coefficient of friction as a function of the number of revolutions for slider on Al2 O3 –TiC unlubricated and lubricated disks. (From Chu, M.Y. et al., Tribol. Trans., 35, 603, 1992. With permission.)
Lubricant’s attraction to the disk is caused primarily by van der Waals (vdW) forces if no reactive or polar functional group is present. The vdW interaction energy of PFPE molecules with DLC surface is of the order of 2 kJ/mol or less [25]. Because vdW forces cannot sustain large forces, most lubricant systems include a reactive group of molecules with ending groups such as hydroxyl (–OH), carboxyl (–COOH), or piperonyl, –CH2 -phe=(O)2 =CH2 . These endings create hydrogen bonds with the DLC surface. These lubricants form bonded films and are effective in reducing friction and increasing durability [26, 27]. The addition of (–OH) end group to the lubricant molecule adds an interaction energy of the order of 20 kJ/mol, even if no chemical reaction occurs with the surface. The attachment of lubricant molecules with or without functional groups is enhanced after sliding, as long as lubricant chains are exposed to the disk surface [28]. Bonding strength can also be increased by heating, which causes the reaction of the alcohol group to the carbon surface. Figure 16.14 shows a schematic of the Z-DOL molecules absorbed on the carbon surface [29]. The carbon–oxygen bonds are known to be important for physical adsorption of functional as well as nonfunctional PFPE derivatives. Only few of the functional groups, carboxyl and hydroxyl, present on the carbon surface show a strong affinity for the lubricant. Disk drive lubricants, although molecularly thin, have also seen the need for additives. X-1P (see Figure 16.15) has been used as an additive to PFPE because it has been shown to improve its tribological performance [30]. It has been shown that by adding X-1P on PFPE, the mobility of the lubricant system increased [31]. This increase in mobility allows the lubricant to replenish faster into the sliding track, acting again as protection from hard surface contact. It has also been shown that X-1P passivates the head and prevents the catalytic reaction of the PFPE. By blocking the sites for PFPE degradation, the X-1P extends the life of the lubricant system. The large size of the X-1P molecule (∼1.8 nm per monolayer) could be seen as a limitation in present applications where the intention is to have only 1 nm of lubricant in order to decrease the total magnetic spacing. This fact, however, does not appear to be a deterrent to its use in today’s disks, implying that the functional thickness is actually lower than a monolayer, and it is related to the orientation of the molecule on the surface. The most recent lubricant that has gained popularity and is already used in commercial disks consists of cyclotriphosphazine-terminated PFPE [32, 33]. The terminal molecule (S-3500) is very similar to the common X-1P additive with the only difference being that all the benzenes are attached to CF3 terminal groups (see Table 16.5). The benefits found are durability and stability, although there is the concern of mobility. A study has shown that the mobility of the new lubricant is slower that Z-DOL lubricant [32,33]. Figure 16.16 presents a schematic of the conformation of these molecules to the carbon overcoat and a comparison with the standard Z-DOL chain.
© 2006 by Taylor & Francis Group, LLC
16-16
TABLE 16.5
Chemical Structure and Selected Properties of Lubricants and Additives Used on Magnetic Rigid Thin-Film Disk Formula X–O–(CF2 –CF2 –O)m –(CF2 –O)n –Y (unless indicated)
Lubricant Fomblin Z-25a Fomblin Z-15 Fomblin Z-03 Fomblin Z-DOL Fomblin Z-Tetraol Fomblin Z-TX Fomblin AM2001 FomblinYR
Krytoxc 143 AD
A20H-2000f A20H-2000 di-functional a
Specific gravity at 20◦ C
Kinematic viscosity (cSt) at 20◦ C
Surface tension at 20◦ C (dynes/cm)
Thermal stability (◦ C)
12800 9100 3600 2000 3000 — 2300
1.85 — — 1.81 1.75 — —
250 150 30 80 2200 — 80
25
370
— — —
— — —
[1] [2] [2] MW = 4000 also used, [3] [4] [4] [2]
6800
1.92
1600
21
370
[1]
2600
1.88
495 at 38◦ C
19
—
[1]
3600
1.87 1.46
75 at 38◦ C 2500
— —
∼420
[3] [5]
3000
—
—
—
—
[6]
4000
—
—
—
—
[7]
Fomblin family of lubricants is manufactured by Ausimont (Milan, Italy). 3,4-Methylenedioxybenzyl. Manufactured by Dupont (USA). d Manufactured by Diakin (Japan). e Partially fluorinated hexaphenoxy Cyclotriphosphazene. From Dow Chemical. f Product known as “Moresco.” From Matsumura Oil Research Corp. (http://www.moresco.co.jp/), also an A20H-4000 with MW = 5000 has been developed. g The Cyclotriphosphazene is also known by S-3500 and corresponds to the X -1P when n = 0. Molecular Weight of S-3500 is 1000 by itself. [1] Bhushan, B., Tribology and Mechanics of Magnetic Storage Devices, 2nd ed., Springer-Verlag, New York, 1996. [2] Bhushan, B. (Ed.), Micro/nanotribology and micro/nanomechanics of magnetic storage devices, in Handbook of Micro/Nano Tribology, 2nd ed., CRC Press, Boca Raton, FL, 754, 1999. [3] Bhushan, B. (Ed.), Macro- and microtribology of magnetic storage devices, in Modern Tribology Handbook, CRC Press, Boca Raton, FL, 1483, 2001. [4] Kasai, P.H. and Raman, V., Z-dol versus Z-tetraol: bonding and durability in magnetic hard disk application, Tribol. Lett., 16, 29, 2004. [5] Perettie, D., Morgan, T., and Kar, K., X-1P as a dual purpose lubricant for pseudo-contact recording, IDEMA Insight, IX, 6, 1996. [6] Tagawa, N. et al., Spreading of novel cyclotriphosphazine-terminated PFPE films on carbon surfaces, ASME J. Tribol., 126, 754, 2004. [7] Kobayashi, N. and Ikegami, M., Abstract. Environmental effect on properties of cyclotriphosphazene-terminated perfluoropolyether lubricant, Digest of TMRC, 2004. b c
© 2006 by Taylor & Francis Group, LLC
Observations/ reference
Handbook of Lubrication and Tribology
D-SAd X-1Pe
X and Y = CF3 X and Y = CF3 X and Y = CF3 X and Y = –CF2 –CH2 –OH X and Y = –CF2 –CH2 –O–CH2 –CH(OH)–CH–OH X and Y = –CF2 –CH2 –O–(–CH2 –CH2 –)1.5 –OH X = –CF2 –CH2 –O-piperonylb Y = –CF2 –O–piperonyl CF3 | CF3 –O–(C–CF2 –O)m –(CF2 –O)n –CF3 (m/n ∼ 40/1) | F CF3 | CF3 –CF2 –CF2 –O–(C–CF2 –O)m –CF2 –CF3 | F F–(CF2 –CF2 –CF2 –O)m –CF2 –CF2 –CH2 –OH (p–FC6 H4 O)n –N3 P3 –(m –OC6 H4 CF3 )6−n , where n = 0, 1, 2, . . . , 6 (typically n = 2) X = –CF2 –CH2 –Cyclotriphosphazene (see X-1P)g Y = –CF2 –CH2 –OH X and Y = –CF2 –CH2 –Cyclotriphosphazene (see X-1P)
Molecular weight
Tribology of Hard Disk Drives — Magnetic Data Storage Technology
16-17
FIGURE 16.14 Schematic for the Z-DOL molecules adsorbed on the carbon surface at 20◦ C. (From Bhushan, B. [Ed.], Modern Tribology Handbook, CRC Press, Boca Raton, FL, 1413, 2001. With permission.)
F CF3 O
O
N O
P
CF3
P O
N
N P
CF3 O
O
CF3
F
FIGURE 16.15
Chemical diagram of X-1P (n = 2) (see Table 16.5). S-3500
Cyclotriphosphazine
A2OH-2000
DLC films
Z-doI 2000
OH
FIGURE 16.16 Estimated existing conformation of the molecules of A20H, Z-DOL, and S-3500 on carbon surface. (From Tagawa, N. et al., IEEE Trans. Magn., 39, 2441, 2003. With permission.)
16.3.3.1 Study of the Effect of Degree of Chemical Bonding of Lubricant Film The effect of degrees of chemical bonding on static and dynamic friction and durability have been studied by several researchers [26,27,34,35]. Figure 16.17 shows normalized static friction force (Fs /W ) as a function of rest time for three lubricant systems (1) untreated, (2) partially bonded, and (3) fully bonded films of Z-DOL lubricant. The lubricants were placed in two different disks with different surface roughness. It was found that static friction increases with increasing resting time only in the case of the untreated lubricant film at the disk with composite roughness of 3.0 nm, and it levels off after a rest time of about 105 sec. The static friction force remains constant and does not change for all other types of lubricant films. The reason for a low static friction force, in the rough disk, even with the untreated film is that the lubricant thickness of 2 nm for a head/rough disk interface is not thick enough to produce a significant amount of meniscus force [25]. We will discuss more about this interaction in the section about stiction. In general it has been found that a partially bonded lubricant provides a balance of low stiction and high durability that it is not achievable with completely bonded or unbonded lubricants.
© 2006 by Taylor & Francis Group, LLC
16-18
Handbook of Lubrication and Tribology Lubricant Z-DOL 8 2 nm untreated 1.4 nm partially bonded 1.2 nm fully bonded
6
4
2
Fs / W
s = 3.0 nm 0 0.5 2 nm untreated 1.2 nm partially bonded 0.9 nm fully bonded
0.4 0.3 0.2 0.1
s = 4.7 nm
0 1
102
104
106
Rest time (sec)
FIGURE 16.17 Normalized static friction force as a function of rest time for disks lubricated with (a) untreated, (b) partially bonded, and (c) fully bonded Z-DOL lubricant films. (From Zhao, Z. and Bhushan, B., Wear, 202, 50, 1996. With permission.)
16.3.3.2 Study of Environment Effect Disk surfaces exposed to a humid environment can adsorb a layer of water film. The total liquid film on the disk surface consists of the lubricant and adsorbed water. Uniformity of the liquid film is necessary for a uniform value of friction [36,37]. It is known that water films can act as lubricants if there is no other film present. However, in the lubricated disks the effect of water is to displace the lubricant from its locations, increasing the possibility of corrosion of the disk by exposing any possible pin-hole in the carbon overcoat layer to oxidizing radicals. It is clear that the degree of bonded fraction and the hydrophobic characteristics of the lubricant are critical to its effectiveness against corrosion. 16.3.3.3 Study of Different Polar Ending Lubricants Bhushan and Zhao [36] compared Z-DOL and Demnum SA and found a significant difference in their behavior under several environmental conditions. A set of stiction and durability tests were performed at different relative humidity levels and with different levels of bonding (Figure 16.18). It is believed that the combined liquid from the lubricant film and water film are responsible for the increase in meniscus force (static friction). The disk lubricated with Demnum SA shows a lower value of static friction force (Fs /W ) than the disk with Z-DOL after the same rest time. This is believed to be due to the lower affinity value of the water to the Demnum SA lubricant. For the disks with untreated (less bonded) films of lubricant Z-DOL, the absorbed water at RH > 60% can displace the lubricant at some sites on the disk surface, resulting in a durability similar to that of the unlubricated disk.
© 2006 by Taylor & Francis Group, LLC
Tribology of Hard Disk Drives — Magnetic Data Storage Technology (b) 1.5
1
Fs / W
Unlubricated 2 nm of Z-DOL 2 nm of Demnum SA untreated
0.5
0.5
0.4
0.4
0.3
0.3
Fk / W
0 0.5
0.2
0.2
0.1
0.1
0 106
0 106
104
102 0
20
40
60
80
100
Relative humidity (%)
Unlubricated 1.8 nm of Z-DOL, bonded 1.8 nm of Demnum SA, bonded
1
0 0.5
Durability (revolutions)
Durability (revolutions)
Fk / W
Fs / W
(a) 1.5
16-19
104
102 0
20
40
60
80
100
Relative humidity (%)
FIGURE 16.18 Normalized static and kinetic friction forces and durability as a function of RH for unlubricated disks and disks lubricated with (a) untreated and (b) fully bonded lubricant films. The composite (σ ) roughness is 3.0 nm. (From Bhushan, B. and Zhao, Z., IEEE Trans. Magn., 33, 918, 1997. With permission.)
It is clear that for static and dynamic friction and durability, a disk with Demnum SA lubricant is relatively insensitive to a humid environment as compared with a disk lubricated with Z-DOL. Lubricant Demnum SA is insensitive to humidity because at one end it has an OH group that attaches to the carbon overcoat and the other end is nonpolar and unreactive to humidity. 16.3.3.4 Lubricant Deposition Process The standard procedure used for applying the lubricant on the disk is by dipping the disk in a bath with a mixture of the lubricant and solvent. This procedure is still the preferred method for lubricant deposition, even though it exposes the carbon overcoat to the environment and allows molecules of water in the environment to deposit in the carbon overcoat before the lubricant is deposited. The important factor in achieving a good uniform layer of lubricant is a tight control over the relative motion (rate) at which the interface between the disk and the lube/solvent mix separate. Any disturbance in the surface of the liquid can create a wave that results in a higher followed by a lower (or vice versa) thickness of the lubricant layer. It is not abnormal to find this phenomenon at the specific moment when the liquid level reaches the edge of the media holder (see Figure 16.19). Recently, several companies have been investigating the use of sputtering tools to deposit the lubricant. This procedure will avoid the exposure of the disk to the environment and it is claimed that it produces a more uniform layer and better bonding to the carbon overcoat. At the present time the author does not know of any company using this system in commercial disks. 16.3.3.5 Measurement of Lubricant in the Disk Several techniques have been developed to measure the film thickness: FTIR (Fourier transform infra-red), angle-resolved x-ray photoelectron spectroscopy (XPS), AFM, ellipsometry, capacitance, and specially developed tools such like the Optical Surface Analyzer (OSA) and the Surface Reflectance Analyzer (SRA)
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16-20
Handbook of Lubrication and Tribology
Lubricant + Solvent
FIGURE 16.19
Schematic of disk dip-coating lube process.
–18.7 Å
–17.6 Å
FIGURE 16.20 An example of a lube thickness mapping of a disk obtained using an optical surface analyzer (OSA). It shows a transversal line that corresponds to a ∼0.4 Å more lubricant that the rest of the disk. This line is found on dip-coating lube processed disks.
based on ellipsometry [25]. FTIR is used as an absolute reference tool, but it requires a substantial amount of time to map the entire surface of the disk. The same goes for XPS, AFM, and conventional ellipsometry. The capacitance technique can quickly, nondestructively, and accurately map lubricant films [38]. An adaptation of the ellipsometry technique, where the disk is rotated while the system performs the measurements, has been the adopted technique in the industry. The speed and extra utilities of this technique have made it the standard tool for lubricant measurement. Figure 16.20 shows a“lubricant map” of a commercial disk. The horizontal line observed represents less than 0.5 Å of lube increase due to nonuniformity of the solvent-lube mix during the deposition process (see Section 16.3.3.4). 16.3.3.6 Lubricant Loss and Degradation Mechanism One of the major areas of concern is the degradation or removal of the lubricant from the disk. Because the thickness is getting down to only a couple of atomic layers, it is critical that the lubricant does not evaporate or get removed from the disk, or that it loses its properties over its lifetime. The phenomenon of transfer to the slider has been studied and it has been found that the slider material plays a significant role in this process. Figure 16.21 presents a schematic model of the lubricant transfer process [39]. The transfer is enhanced by the affinity of the slider surface coating with the lubricant, which has been a consideration when selecting the best slider overcoat (see Section 16.4.1 on slider overcoat).
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Tribology of Hard Disk Drives — Magnetic Data Storage Technology
Slider
16-21
Mobile layer
Lubricant layer Disk
Adsorbed layer Initial contact of slider on lubricated disk
Transfer of lubricant from disk to slider
Formation of adsorbed layer on slider
FIGURE 16.21 Schematic model for lubricant transfer from a lubricated disk to an SiC slider. (From Xu, J. and Bhushan, B., Proc. Inst. Mech. Eng., Part J: J. Eng. Tribol., 211, 303, 1997. With permission.)
There have been several proposed mechanisms for the degradation of PFPE lubricants: thermal decomposition, catalytic decomposition by head slider or disk materials, triboelectric reaction, and mechanical scission. Bhushan and Zhao [27] performed a slider test using a high vacuum environmental tribotester that allowed the collection of the desorbed species. Using a mass spectrometer they found evidence of ZDOL degradation immediately when the sliding began. Although the test conditions do not correspond to the normal HDD conditions, the test makes evident that there is a level of slider energy that can decompose the lubricant. In this study it was also found that the fully bonded lubricant was the best for protecting the carbon overcoat, followed by partially bonded lubricant, and untreated (unbonded) lubricants. Thermal decomposition of PFPE has been reported at temperatures above 350◦ C. High flash temperatures can be generated by asperity contacts at high speeds during flying and those could initiate the decomposition, but there is no evidence of this high temperature occurring. The decomposition could occur at lower temperatures if the PFPE lubricant comes in contact with metal oxides. Kasai et al. [40] and Kasai [41] reported that Lewis acid sites present on Al2 O3 are catalytic and cause PFPE lubricant molecules to decompose at temperatures as low as 200◦ C. However, the experiments were conducted using Al2 O3 powder, and the kinetics of these experiments could be considered very different from those in the normal drive operation. Triboelectrical decomposition is used to describe the decomposition caused by low-energy electrons emitted by the rubbing or sliding of the surfaces. Zhao and Bhushan [42,43] reported that tribo-emission of electrons may be a significant factor in tribochemical reactions in the contact zone operating under
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Handbook of Lubrication and Tribology
lightly loaded conditions. Vurens et al. [44], studied the decomposition of PFPE lubricants by means of mass spectrometry and FTIR, and compared thermal and electron decomposition, finding that the gaseous products generated were different, suggesting different degradation pathways. Zhao and Bhushan [42,43] conducted studies using low-energy ultraviolet (UV) light to generate low-energy electrons and found that degradation of the lubricant was accelerated by this process. They concluded that the triboelectric reaction can be a dominant degradation mechanism. Mechanical scission can be produced by the shear rates generated during start–stop or isolated contact during flying. The relative velocity of the slider can reach 34 m/sec for the outer tracks of a standard 7200 rpm HDD. These high relative velocities, along with the typical lubricant thickness of 1 to 2 nm, produce shear rates of over 1010 /sec. At such rates significant energy can be imparted to the polymer chain. The effects of mechanical shear were studied by Zhao et al. [45], and Zhao and Bhushan [46], by changing the sliding velocity of the slider while dragging on the disk, and monitoring the partial pressure of selected species of Z-DOL lubricant. They found out that the higher the sliding velocity the stronger the peaks (see Figure 16.22). These results indicated that mechanical scission is also a dominant degradation mechanism of PFPE lubricant. Although several possible mechanisms for decomposition of PFPE are possible, the dominant mechanisms appear to be mechanical scission and triboelectric decomposition followed by thermal decomposition at high velocities. 16.3.3.7 The Future and Possible New Lubricants As Gui [5] indicated, it is expected that the contribution of the lubricant to the magnetic spacing can be no more than 0.8 to 1 nm to reach the 1 Tb/in2 areal densities. Because there is the possibility of the lubricant being transferred to the slider during use, it is important to select a combination of slider overcoat and lubricant that reduces this lubricant movement.
10 0 40
HCF2 100
0.15m/sec
0.6m/sec
0.45m/sec
0.3m/sec
20
0.9m/sec
30
0.75m/sec
Partial pressure (x 10–10 torr)
40
CF2O 200
300
1.05m/sec
Z-DOL, untreated, sliding for 9 m at each velocity
(a)
CFO
CF3 400
500
600
0.8
1.0
1.2
Time (sec)
Partial pressure (x 10–10 torr)
(b)
15 CFO HCF2 CF3 CF2O
10
5
0 0.0
0.2
0.4
0.6
Velocity of slider (m/sec)
FIGURE 16.22 (a) Partial pressure changes of selected species of lubricant Z-DOL at different velocities, and (b) partial pressure increase of selected species with sliding velocity for a sliding of 9 m at each velocity; lubricant film thickness is 2.2 nm. (From Zhao, X. and Bhushan, B., Proc. Inst. Mech. Eng., Part J: J. Eng. Tribol., 215, 173, 2001. With permission.)
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Tribology of Hard Disk Drives — Magnetic Data Storage Technology
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In the search for the perfect lubricant, some studies have focused on finding an alternative to the PFPEs that would overcome their few weaknesses at very low thicknesses, such as degradation with use and lack of stability (see Section 16.3.3.6). Several self-assembled monolayers (SAMs) have been proposed for some time as alternative to PFPE-based lubricants. Murayama et al. [47] presented two types of lubricants: stearyl amine (C18 H37 NH2 ) and stearic acid (C17 H35 COOH) and found that they “did not show excellent CSS performance.” It is understood that the lack of mobility of the SAM is the cause of its poor durability. More recently Choi et al. [48], have presented several alternatives. The organosilane monolayer was found to be effective in lowering the coefficient of friction of an amorphous carbon-coated disk, tested on a ball on flat tribotester, and compared with the unlubricated disk, but when compared with Z-DOL 4000 system with similar total thickness (1.5 nm), the coefficient of friction was ∼20% higher. Choi explained this result by arguing that the contact area of the ball is larger for the SAM molecules. They also show [49] that if the lubricant system is composed of a mixture of SAM as the bonded part and PFPE as the mobile phase, its durability is better. In another study, Choi et al. [48], presented as alternative a system based on P1D as bonded portion of the lube system deposited by plasma enhanced chemical vapor deposition (PECVD) with PFPE as a mobile phase. No durability data for this system was reported. In conclusion, it seems that PFPE-based lubricant systems will continue to be used in HDDs for the foreseeable future, and even at the incredible thicknesses of 1 nm there will be a combination of bonded and unbonded phases. The understanding of the factors controlling stability and degradability will be key to extend the durability over the lifetime of the system. The understanding of molecular interaction with the slider will also be of critical importance for very low flying sliders (<5 nm) and for succeeding on the pseudocontact recording approach.
16.4 The Head/Slider/Air Bearing Surface The heart of the HDD is the writer and reader elements (or head). But these could not possibly perform their function if they were not placed at the right distance from the disk (flying height) and at the correct location (servo control system). The fabrication process of these slider/heads is divided into two distinct steps: the wafer manufacture and the slider fabrication. The wafer manufacture yields the reader and writer elements (head). This process is very delicate, expensive, and time consuming and has to be performed in a clean room environment. It consists of a large set of sputtering steps, during which many different layers are deposited in order to create the structure of the reader first and the writer after, one on top of the other. Several thousand heads are manufactured this way on each wafer. The slider fabrication, on the other hand, although it is also a sophisticated process, is “very dirty” in comparison, and it is usually performed in completely different locations than those used for wafer manufacturing. The first step consists of slicing the wafer into rows. These rows expose the surface where the ABS will be created and that ultimately will face the disk surface while flying. The face that contains the reader and writer elements will become the end side of the slider. Figure 16.23 shows the schematics of the slider fabrication process [50]. The slider dimensions have been getting smaller with time, as with everything in the HDD (see Table 16.6). The first slider was 4 × 3.2 × 0.86 mm3 and the read and write functions were performed by the same element, manufactured with wire wound coils. During the 1990s, IDEMA set the goal of standardizing the slider dimensions, and standards for the Micro, Nano, Pico, and Femto sizes were established — some after the sliders were already on the market. The standardization process, as previously mentioned, has lost some significance because of the limited number of suppliers and HDD manufacturing companies, so that today different dimensions are starting to appear on the market that have not been standardized. The slider material has changed very little since the late 1980s. The common material used presently is a composite Al2 O3 –TiC (70 to 30 wt%). Several slider materials were studied in the past for durability [6,39], and SiC was found to be the best for a wide range of relative humidity levels. It was also indicated that the high thermal conductivity of SiC, and some machining advantages, made it a better choice. Another
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology Head manufacturing overview Wafer fab
HGA assembly
Slider fab
Air bearing surface Air bearing surface
FIGURE 16.23
Schematics of the slider fabrication process.
TABLE 16.6
Evolution of Slider/Head Dimensions, Names, and Introduction Year
Slider size Introduction year Dimensions, mm Length Width Height Slider mass, mg
Mini 100%
Micro 70%
Positive pressure Nano 62%
Negative pressure Nano 50%
Pico 30%
Femto 20%
1980
1986
1991
1994
1997
2003
4.00 3.20 0.86 55.0
2.80 2.24 0.60 16.2
2.50 1.70 0.43 7.8
2.00 1.60 0.43 5.9
1.25 1.00 0.30 1.6
0.85 0.70 0.23 0.6
recent study shows the concern that the effect of slider material has on the carbon oxidation. Ramirez et al. [51] investigated the effect of alumina and titanium carbide on the carbon gasification reaction using titanium carbide and alumina powder combined with graphite powder. Their findings support that slider material, as of today, may behave catalytically in the carbon oxidation process. The fact that the carbon overcoat on the slider is being reduced in thickness to the minimum possible will increase the importance of slider interaction, so as to guarantee a zero degradation of performance during the HDD lifetime. To perform its function, the slider is attached to a flexure that is placed on suspension and will carry the slider and head over the different parts of the disk (Figure 16.5). The shape processed on the slider, called ABS, is critical to assure that the reader and writing elements located at the end of the slider fly at the same distance from the disk at the different radius. In Sections 16.4.1 and 16.4.2 we will focus on the carbon overcoat and ABS developments for the slider, as critical tribological aspects of this part of the HDD.
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Tribology of Hard Disk Drives — Magnetic Data Storage Technology
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16.4.1 Hard Protective Coating While the disk is covered with a hard overcoat to protect the magnetic layer, the read/write elements need protection from the environment and the slider needs to be wear resistant to sustain the start–stop or load–unload cycles expected for the given product. It is also important that the surface is hydrophobic to reduce stiction in case of contact start/stop (see Section 16.5). Nowadays, this protection is given by a DLC layer somehow similar to those applied to the disk. Although most of what we talk about in the hard protective coating of the media applies (see Section 16.3.2), there are a few differences that need to be addressed. One, there is no need for carbon–lubricant compatibility. On the contrary, it is ideal that there is no transfer of the lubricant from the disk to the slider, since this layer increases the magnetic separation. It is believed, however, that the durability depends significantly on this lubricant transfer to the slider. This transfer of lubricant (Figure 16.21) was found to be very beneficial for wear protection in the study of SiC sliders by Xu and Bhushan [39]. Regarding the best deposition technique for applying the DLC layer over the slider, the basic issues stated on Section 16.3.2 apply: the best DLC will be very dense (no pin-holes), stable over time, and well attached to the slider. The techniques used to deposit DLC on the slider have been evolving, aiming to obtain the ideal cover with the minimum thickness. Zhao and Bhushan [52] evaluated four different deposition techniques: FCA, ion beam (IB), ECR-CVD, and sputtering (SP), over standard Al2 O3 –TiC slider. After performing drag and CSS tests, they found that IB coating showed the longest interface durability (Figure 16.24). However, the authors argue that FCA and ECR-CVD coating, which exhibit the highest hardness and fracture toughness, and the highest scratch resistance, promise superior tribological performance if a compatible DLC (same elastic modulus) is used on the disk. Although the trend is toward thinner overcoats in order to reduce the total magnetic spacing, approaches such as pseudocontact recording, or other approaches to control and reduce the flying height, could establish an increase such that no wear occurs during the life of the system. An example of this compromise is presented by Song et al. [53] who by using electrostatic voltage (±6 V) showed that FHt could be reduced by 2.5 nm. To avoid the possibility that this charge might cause electrostatic discharge (ESD) on the reader element, they show that DLC would have to be increased by ∼1.5 nm. Although this seems like a minor increase, a 1 nm absolute spacing reduction corresponds to more than 5% of the total magnetic spacing on present HDD products. Currently, ion beam deposition is commonly used for the deposition of the DLC. The type of carbon is hydrogenated carbon and a layer of Si (∼1 nm) is applied before the carbon deposition, with the purpose of improving the bonding of the carbon. The total thickness of the overcoat on today’s products is 2.5 to 4 nm. It is expected that companies use the latest FCA in future generations of heads.
16.4.2 The Air Bearing Surface (ABS) The read and write elements, in current HDD, fly over the disk at a distance of about 8–10 nm. The vehicle that makes this possible is the ABS.8 There have been significant theoretical developments in the area of fluid dynamics that have played a significant role in the ABS. Also, the advances in numerical methods and their adaptation to the specific problems in the HDD, have allowed the development of effective and efficient computer tools. Improvements in the manufacturing process have also been key to achieving the flying tolerance of less than ±2 nm in today’s products. The governing equation used in classical hydrodynamic lubrication theory is the Reynolds equation. It is a single differential equation relating pressure, density, surface velocities, and film thickness. Reynolds presented the equation in 1886 and it has been the foundation of hydrodynamic lubrication theory. The equation is developed from the Navier–Stoke equation, using the equations of continuity. In order to develop the equation, the following assumptions are necessary [54] (1) smooth surfaces, (2) Newtonian 8 The ABS
needs the adequate suspension to perform correctly.
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Handbook of Lubrication and Tribology FCA 80% RH
ECR-CVD 80% RH
(a) 2.0
(c) 2.0 5 nm
5 nm
1.5
1.0
1.0
0.5
0.5
0.0 2.0
Coefficient of static friction
Coefficient of static friction
1.5
10 nm
1.5 1.0 0.5 0.0 2.0
0.0 2.0 10 nm 1.5 1.0 0.5 0.0 2.0
20 nm
20 nm
1.5
1.5
1.0
1.0
0.5
0.5
0.0
0
5
10
15
0.0
20
0
5
CSS durability (k cycles) (b)
IB 80% RH
Coefficient of static friction
Coefficient of static friction
5 nm
1.0 0.5 0.0 2.0 20 nm
1.5
15
20
15
20
SP 80% RH
(d)
2.0 1.5
10 CSS durability (k cycles)
2.0 20 nm 1.5 1.0 0.5 0.0 0
5
10 CSS durability (k cycles)
1.0 0.5 0.0
0
5
10
15
20
CSS durability (k cycles)
FIGURE 16.24 CSS durability at 80% RH for the (a) FCA, (b) IB, (c) ECR-CVD, and (d) SP coatings. Arrows indicate that visible wear tracks on the disk surface were formed. IEEE Trans. Magn., 36, 2665, 2000. With permission.)
fluid, (3) laminar flow, (4) constant viscosity, (5) temperature of the bearing surfaces are equal and constant, (6) viscous shear forces dominate and inertia forces are small and neglected, (7) tension effects are negligible, (8) fluid film thickness is much smaller than other typical bearing dimensions, and (9) there are nonslip boundary conditions on both walls. Due to the small spacing between the slider and the disk, the assumptions of continuum and nonslip boundary conditions are not valid assumptions. The continuum assumption is considered not valid if the separation between the surfaces is less than ∼100 times the local mean free path of air (the fluid in this case). For air at normal conditions, the mean free path is ∼64 nm. Several theoretical developments in the 1980s and 1990s have focused on modifications to account for these violations of the Reynolds equation. The widely used Molecular Gas-film Lubrication (MGL) model [55] is the Reynolds equation with the slip correction based on Boltzmann’s equation, where the Poiseuille flow rate is calculated on the basis of a linearized BGK model of the Boltzmann equation [56]. The other area of great development has been the use of numerical analysis to allow the numerical solution of the equation by either finite differences or finite elements in an efficient and convergent way.
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Tribology of Hard Disk Drives — Magnetic Data Storage Technology
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1.25 mm Cavity
Trailing edge pad
Read / write element
0.8 Leading edge 1 mm
1.7 E–8 m 0.13 mm –1.4 E–6 m 1.0
1.3 mm 0.0
0.0
Main air bearing surface Shallow recess
FIGURE 16.25
Example of a pico sized ABS with its main parts and dimensions.
Most of the results presented in this section have been obtained using a software package developed at the University of California at Berkeley by the Computer Mechanical Laboratory (CML). Their software called CMLAir32 is already well known and widely used in the HDD industry. The design of ABS has evolved slightly differently in different companies, but all the present designs have some common characteristics. Figure 16.25 shows details of an ABS used in recent HDDs. The main air bearing surface corresponds to the areas that are located closer to the disk. An area located ∼0.1 to 0.2 µm farther from the disk is called shallow recess. The rest of the surface is recessed ∼1 to 2 µm and is called the main recess or cavity. It is customary to create a shallow transition wall between the main ABS and the shallow recess and cavity area. For the design shown, the angles are 2◦ and 20◦ . They affect the air flow and debris accumulation. Most of today’s designs have the reader and writer elements located in the center of the trailing edge (TE). The area closer to the reader and writer is called the TE pad. It is also standard for the final ABS to have a convex shape. The typical distance from the center of the air bearing to the edges in the Z direction is measured in the range 0 to 30 nm. Figure 16.26 shows photos of several designs used by different companies. The type of ABS used in most of the HDDs today is known as negative-pressure or subambient pressure design. All the designs in Figure 16.26 correspond to this kind of design. This design concept was first presented by Garnier [57] but it was only widely introduced in systems during the late 1990s. Its name comes from the generation of a subambient pressure region that attracts the slider against the disk, creating a much stiffer system. Figure 16.27 shows a pressure contour of the air bearing presented in Figure 16.25. The scale values correspond to “atmospheres” above or below the standard pressure. For this specific design the maximum pressure reaches 8.73 atm (885 kPa) at the location of the reader and writer gap, and the subambient pressure created has a maximum of −0.84 atm, or in other words an absolute pressure of 16.4 kPa. The total force applied by the suspension to the slider, typically −20 to 40 mN pushing the slider against the media, is compensated by the forces created by the air pressure distribution under the slider. The negative pressure also creates a force that attracts the slider to the media and it is compensated by the positive pressure regions. For the design shown, the negative pressure zone creates a force of the same order of magnitude as the slider initial force (∼−30 mN), and the positive pressure creates a separating force that balances the system (∼+60 mN). It is also important to minimize the flying sensitivity to altitude (air pressure). In the CSS designs, where the head lands on the disk during rest, it is important to minimize the effective contact area with
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Handbook of Lubrication and Tribology
FIGURE 16.26
Examples of negative pressure ABSs. Pressure contour 8.7
0.88
6.81
6.81
4.90
4.90
2.98
1.00
8.73
2.98
0.75 Width (mm)
8.73
Highest
0.63 0.50
Lowest pressure
x
x pressure
0.38 0.25
1.07 0.8 0.0 0.0
–0.84
–0.84 1.0
1.07 0.13 –0.84
0.00 0.00 0.15 0.31 0.46 0.62 0.77 0.93 1.08 1.24 Length (mm)
FIGURE 16.27 Example of pressure profile and pressure contour of the air bearing in Figure 16.25 obtained by simulation. The relative pressure reaches 8.73 atm at the trailing edge, and the negative pressure region drops to 0.84 atm below ambient.
the surface, optimize the take off velocity (relative velocity that will raise the ABS from the surface) and minimize the contact interaction during landing. In the L/UL designs, the ABS and suspension system have to be optimized to avoid large contact with the disk, every time the head is placed or removed from the disk. All these requirements have made the ABS the target of extensive analysis. Because the slider is mounted at the end of an arm that rotates in order to change the radial location on the disk (similar to old LP turntable arms), the angle of the slider and the disk is different at different radii (skew angle). This skew affects the direction from which air enters the space between the head and the disk. Besides this, the air velocity differential between the inside and the outside can be more than 100%. So the design of the ABS has to be such that, independently of these effects, it must be capable of maintaining the back of the slider, at the same distance within few ±1 to 2 nm in present disk drives. Figure 16.28 shows the air flow under the ABS for the inner, middle, and outer location of the slider on the disk.
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Tribology of Hard Disk Drives — Magnetic Data Storage Technology Mass flow 1.00
0.88
0.88
0.75
0.75
0.63
0.63
Width (mm)
Width (mm)
Rails 1.00
0.50 0.38
0.50 0.38
0.25
0.25
0.13
0.13
0.00 0.00 0.15 0.31 0.46 0.62 0.77 0.93 1.08 1.24
0.00 0.00 0.15 0.31 0.46 0.62 0.77 0.93 1.08 1.24
Length (mm)
Length (mm)
Mass flow
Mass flow
1.00
1.00
0.88
0.88
0.75
0.75
0.63
0.63
Width (mm)
Width (mm)
16-29
0.50 0.38
0.50 0.38
0.25
0.25
0.13
0.13
0.00 0.00 0.15 0.31 0.46 0.62 0.77 0.93 1.08 1.24
0.00 0.00 0.15 0.31 0.46 0.62 0.77 0.93 1.08 1.24
Length (mm)
Length (mm)
FIGURE 16.28 Examples of air flow lines at three locations on the disk. The direction of the lines show the relative skew of slider and the disk.
16.5 Contact Start–Stop In most of HDDs used in desktops and nonportable consumer electronics (such as personal video recorders), the heads rest on a special area of the disk while the disk is not in operation. While the head is flying during operation, it comes into physical contact with the disk every time the disk starts to spin and slows down to stop. Today drives are manufactured to perform at least 50,000 start–stops without any problem. The two possible issues that could occur during this process are (1) stiction: head gets stuck on the media surface while resting, (2) wear: plastic deformation and contact fatigue or both, occurs in the contact areas, while starting and stopping. While stiction can be minimized by reducing the contact area between the head and the disk while parking, wear is proportional to the stresses during contact or sliding and these can be minimized by increasing the area of contact. These conflicting requirements have resulted in the design of a special texturing system on the disk, known as laser zone texturing (see Figure 16.7). By controlling bump shape, height, and density, the texturing has been optimized to minimize stiction and maximize durability, while allowing the heads to fly over, even today at low flying heights.
16.5.1 Stiction During Contact Start/Stop Stiction is a recently coined term that comes from static friction [58]. It is mainly used to describe the failure mechanism that occurs in the HDDs when the slider/head gets stuck to the disk with enough
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Handbook of Lubrication and Tribology
Slider
Toe-dipping regime
Pillbox regime
Flooded regime
Immersed regime
FIGURE 16.29 Regimes of different liquid levels at the interface with a smooth slider in contact with a rough surface. (From Bhushan, B. [Ed.] , Modern Tribology Handbook, CRC Press, Boca Raton, FL, 1435, 2001. With permission.)
strength to keep the platters from spinning. It is also known that it typically occurs when a HDD has been turned off for a long period of time, especially if before being turned off it had been operating nonstop for a long time. More generally speaking, stiction is a phenomenon occurring when two well-polished solid surfaces having a matching profile are in contact for some period of time. Attraction forces created between the two produce large static friction forces that can not be easily overcome. In the case of the HDD, it has been found that capillary forces created by lubricant and water absorption on the mating surfaces and organic contaminants, or both, have been the main cause of stiction. Understanding how these variables interact has been the key to find the solutions. To analyze the stiction phenomenon we will consider a model of a smooth surface over a well-defined rough surface (see Figure 16.29). This model is adequate to describe the slider–disk contact with the laser bumps described above. Four distinct regimes can be formed [25]. First is the toe-dipping regime, where the liquid quantities are small, so only liquid bridges are formed at the tip of the contacting asperities. Higher amounts of liquid around the asperities allow the formation of the Pillbox regime, where liquid bridges are formed around one or more asperities. In the flooded regime, the entire surface where the asperities touch creates a liquid bridge. In these three cases, the liquid bridge forces are mainly functions of what is called meniscus effects, or in other words surface tension of the liquid. The viscosity of the liquid also affects stiction and is the main force in the immersed regime. In this case, the surfaces may be separated easily, provided the separation is carried out slowly. However, if the rate of separation is rapid, liquid must flow into the space between them and the viscosity of the liquid will be the determining factor in stiction. In the toe-dipping regime, the liquid adhesion force between a single asperity and a general surface can be modeled by a sphere with a composite radius of curvature in contact with a flat surface, with a liquid bridge in between. Total meniscus and viscous forces of all wetted asperity contact can be calculated by
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Tribology of Hard Disk Drives — Magnetic Data Storage Technology
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multiplying the number of contacts by the meniscus and viscous forces at a typical contact. If one assumes that surface asperity radii are constant and their heights follow a Gaussian distribution, the meniscus force can be expressed as [4]: FM ∼
W (E ∗ σp (σp /Rp ))/(16.6γ (cos θ1 + cos θ2 )) − 1
where FM is the force created by the meniscus, W is the normal load between the surfaces, E ∗ is the composite modulus of elasticity, σp and Rp are composite standard deviation and radius of curvature of the surface summits, respectively, γ is the surface tension of the liquid, and θ1 and θ2 are the contact angles of the liquid on the two surfaces. The flooded regime can be modeled by a liquid bridge between two flat surfaces. The forces due to meniscus and viscous effects can be expressed by: FM
Aa γ (cos θ1 + cos θ2 ) = h
and
FM
√ Aa η L α = √ h e
where h ∼ 1.4σp log
0.65 0.57ηE ∗ Rp σp (σp /Rp ) pa
Aa and pa are the apparent area and apparent pressure of contact, respectively, L is the distance the surfaces need to slide to become unstuck, η is the liquid viscosity, α is the start-up linear acceleration, and h is the separation between the two surfaces. In the toe-dipping regime, adhesion force is independent of the apparent area and proportional to the normal force (i.e., number of asperity contacts). However, the flooded regime shows the opposite tendencies. The pillbox regime is intermediate and can exhibit either behavior at the extremes. In all three regimes, adhesion force decreases with an increase in σp and a decrease in Rp , and is independent of η. Stiction increases with an increase in the time of contact and this effect is commonly observed in storage devices, especially at high humidity levels. Chilamakuri and Bhushan [59] developed a kinetic meniscus model to account for time and viscosity dependence of meniscus formation, which explains experimental trends. Typical results of their study are shown in Figure 16.30. Note that the equilibrium meniscus force is independent of the film thickness and the viscosity of the liquid on the surface and is dependent on the surface, tension, contact angle, and the interface geometry, whereas the rate of increase of meniscus force increases with decreasing viscosity of the liquid. The meniscus force, as well as the equilibrium time, is proportional to the asperity radius. The equilibrium time increases with decreases in film thickness. Many of these observations have been verified experimentally.
16.5.2 Wear During Contact Start/Stop Wear behavior of laser-textured disks depends on the bump size, bump shape, number of bumps, and “quality” of the bumps. The “quality” of the bumps is defined as the uniformity of diameter, smoothness, roundness, and spacing. Also critical is the height uniformity, which depends on the substrate finish, laser power variation, and substrate thickness [60]. When the number of bumps is too small or the nonuniformities create small real areas of contact, the stresses become too high and plastic deformation occurs during sliding, causing wear. On the other hand, when the number of bumps is too large stiction failure occurs. Several studies have focused on all the above parameters. Three types of bump shapes were developed and used in laser textured disks, but currently most of the disks (if not all) use what is called the W-type donut shape (see Figure 16.7). The typical diameter of today’s bumps is 6 µm and the typical spacing
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology 78
h = 3 nm
R = 250 mm h1 = 0.25 Pa sec
77
76
Meniscus force (mN)
78
h1 = 0.001 Pa sec
R = 250 mm
0.05
0.25
h =1 nm
77
76 15.7 15.6
R = 50 mm h =1 nm h1 = 0.25 Pa sec
15.5 15.4 15.3 15.2 100
102
104
106
108
Rest time (sec)
FIGURE 16.30 Effect of liquid film thickness (h), liquid viscosity (η1 ), and radius (R) of the sphere on the time dependent meniscus force. (From Chilamakuri, S.K. and Bhushan, B., J. Appl. Phys., 86, 4649, 1999. With permission.)
is 30 µm. This translates to more than 1000 bumps under the head during resting, but only ∼ 13 will be under the ABS main pads. The typical height of the bumps is 6 to 8 nm. The need to reduce the height of the bumps such that the head can fly over them at operational conditions has reduced the usefulness of the bumps to the point that they are no longer very effective in reducing stiction. This has made engineers study the migration of the bumps from the disk to the slider. The tribological consequences are obvious: the total rubbing time of each bump will be much larger, but it will allow an increase in the height of the bumps, by locating them toward the leading edge that during normal conditions fly much higher than the trailing edge. There are already products on the market in which the ABS has been modified with small bumps. This shows a new evolution in the tribology of the CSS systems.
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Tribology of Hard Disk Drives — Magnetic Data Storage Technology
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16.6 Load/Unload Although the CSS technology has dominated the HDD for many years due to its simplicity and relative cost advantages, the trend for lower flying heights and smoother disks is making it harder to implement in each new generation of HDD. The L/UL approach has been investigated for many years and used by companies that manufacture removable products (disk removed from the drive) such as Iomega and its Jaz product. Already in 1988, Yamada and Bogy [61] presented a study showing that heads could be safely loaded and unloaded even though the flying height (<1 µm) was many times smaller than the disk run out (vertical displacement of 25 to 50 µm). Today, most (if not all) HDDs used in mobile applications such as laptops or MP3 players use L/UL technology. L/UL provides two key advantages over CSS: • Improve shock resistance during nonoperation of the disk. • Reduce power consumption, critical in systems powered with batteries. Besides these important advantages, the L/UL technology eliminates the concern of stiction because the heads do not land on the media while stopped, and also eliminates the concern of wear during start and stop. By eliminating the stiction problem, the disks do not need to be manufactured with bumps, and this frees up a zone of the disk for data. However, this gain is reversed by the lost of data area at the outer zone of the disk where the head is loaded. By eliminating the concern of bumps wearing, the thickness of the carbon overcoat, specially formulated to resist the CSS sliding fatigue, can be reduced to the minimum such that it protects the disk against corrosion. Bhushan and Tambe [62] performed a comparison study between an air bearing under L/UL and similar ABS accommodated with pads to perform under CSS conditions. The test included particulate contaminants in order to understand the durability of both designs. They showed that the L/UL slider exhibits relatively good durability and that the pads in the CSS slider allowed the accumulation of the particles that ultimately were the precursors of the failure and wear on the head and disk. Although there are clear benefits in L/UL, there are new challenges: • Wear of the new contacting parts, ramp and lift-tap (see Figure 16.31), has to be understood and minimized so that there is no negative effect in HDD lifetime. • Loading and unloading has to be smooth enough so that head–media contact is minimized or eliminated. Current mobile HDDs claim to support at least 300,000 L/UL cycles, with some manufacturers extending the specifications to 1 million cycles. In order to avoid any wear of the ramp-lifter, significant experimental evaluations have been performed. It has been found that, by reducing the friction during load and unload and by selecting the proper material and surface roughness, wear can be reduced to acceptable levels. Hiller et al. [63] developed a tool to evaluate ramp/tab friction during dynamic L/UL and found a strong correlation between friction and wear. They also found that wear was reduced to a minimum by the use of acetal homopolymer (POM) and aliphatic polyketone (PK). These two materials behaved better than the ones more commonly used, such as LPC (liquid crystal polymer) + Teflon. They also investigated temperature, surface roughness, and sliding speed, finding that, as it is expected for polymers, friction increased for lower temperatures, increasing as much as ∼20% for LPC + Teflon when changing from operating temperatures of ∼30 to 0◦ C. With respect to sliding speed, they found out that the behavior depended on the material used: for POM the friction coefficient increased from ∼0.14 at 25 to ∼0.17 for 250 mm/sec. In the case of aliphatic polyketone, although the coefficient of friction was lower at all the sliding speeds tested, it showed an increase at lower speeds (<50 mm/sec) that appears to be due to stick-slip friction. Regarding roughness, the authors found that the smoothest surfaces, on both ramp and lift tab, always produced the lowest wear. They obtained no measurable wear when the lift tab roughness measured less than 0.6 µm P-V (peak-valley). In order to obtain this roughness, the tabs were coated (e.g., with hard, thermoset plastic), since the lowest roughness measured after tabs are coined (manufacture process used) was ∼ 1µm P-V. The authors also claim that there may be an optimum
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Handbook of Lubrication and Tribology (a)
(c)
Ramp
Lift tab
Dimple
Limiters
Gimbal Slider Disc (b)
(d)
FIGURE 16.31 Unloading process of slider showing the four stages (a) lift-tap starts contact, (b) dimple separates from the slider body, (c) limiters in the suspension touch the slider, and (d) slider completely separates from disk and a vibration is produced.
roughness, and that ultra smooth surface could produce increased adhesive wear, although they did not observe such effect in their experiments. The biggest challenge in L/UL technology is to place the head on the moving disk, at its flying distance of less than 15 nm without touching the disk, and after removing it smoothly enough back to the ramp, and perform all these more than 300,000 times, even in case of emergency unloading. Although the air bearing effect helps, working as a damper while loading, it has been found difficult to manufacture the system that, given all the inherent manufacturing tolerances, will allow the process to be repeated for thousands of times without any head–disk contact. It is obvious that a solution has been obtained because there are products on the market. However, it is the practice of the HDD manufacturers not to use the zone where the heads load and unload for any recording of data, for fear of degradation of the magnetic information. Suk and Jen [64] report that small disk scratches during L/UL could cause magnetic layer damage. They also found that the magnetostriction effect was negligible at least on their experimental system. Most of current HDDs use negative pressure air bearing surface (NP-ABS). Such ABS produces a subambient pressure that sucks the head against the disk. This effect has made it more difficult to unload the heads than to load them, therefore directing the efforts of many researches and designers to concentrate on the unloading process. Figure 16.31 shows the four stages of the unloading process (1) The lift-tab hits the load-ramp, at this moment the ramp starts applying an upward force toward the lift-tab that translates into a separation force on the slider; (2) the force on the lift-tab is large enough to bend the flexure and makes the dimple (location where the load is applied to the head) to separate from the slider, an undesirable situation the consequence of the subambient pressure on the slider; (3) by continuously moving up on the ramp, the uplifting load reaches a point that makes the limiters (special preformed arms designed specifically for this purpose) engage the suspension and produce an extra force to pull the head out of the disk; (4) at the moment of separation the sudden loss of suction force translates into an impulse on the slider that makes it hit the dimple, and depending on the situation, could bounce back and touch the disk. By increasing the suspension stiffness the bouncing of the head is reduced. Also by the correct design and location of the lifter, the impact force is reduced, thereby reducing the chance of the slider bouncing back to the disk. Tanaka et al. [65], performed a research where the ramp was substituted by a vertical lifter that allowed them to monitor the slider with a laser doppler vibrometer, while monitoring electrical resistance between the head and the lifter (to detect contact) and AE (acoustic emission) signal to detect any contact that
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occurs. They observed the four stages described above. Figure 16.31 shows the measurements at the four locations: lift tab, dimple, slider inner center, and slider inner trailing edge. Besides showing the different stages on the unloading process, they examined three different ABS designs with different levels of negative pressure force. They found out that as the negative pressure force of the slider decreases, the slider can be unloaded more easily and the amplitude of the slider vibration after unloading decreases. Because the optimization process is difficult by experimentation, a large effort has been made to create a computer model that can accurately predict the process on L/UL. In order to obtain a good simulation, the suspension has to be incorporated. Because the full FEA (finite element analysis) of the suspension slider could require significant time in computation power, some simplifications have been used and good agreement has been found between experimental data and simulation. Zeng et al. [66] incorporated a simplified L/UL suspension model inside their well-established ABS simulation software [67] and demonstrated by experimentation that good correlation could be established. Zeng and Bogy [68] extended their model to allow four degrees of freedom (DOF) and showed that the new model predicted the same unloading performance as the previous model but improved in the prediction of the loading process. Weissner et al. [69] presented a model that reduces the suspension system to 28 DOF, a reduction from more than 10,000 DOF used for a complete finite element model (FEM) suspension model. They believe that by this approach, they overcome some of the simplifications presented in previous models, improving the accuracy, avoiding the complexity and computation overhead of a complete FEM.
16.7 Ultralow Flyability and Contact Recording In the continued search for higher areal densities, the need for smaller magnetic spacing and consequently smaller physical spacing or flying height (FHt) is the norm. The present FHt of the read/write elements is already below 10 nm in current drives reaching 100 Gbit/in.2 . At these nominal distances sporadic interaction may occur between head and disk. However, the dominant forces that control the system still come from the air flow, and the effects on it, created by the ABS. By reducing the nominal FHt below 5 nm the interaction between the surfaces starts to play a significant role. Mainly vdW forces and electrostatic forces between the disk and the slider reach values that cannot be neglected. Because these forces are very sensitive to distance, they can easily create undesirable instabilities in the FHt of the elements and consequently affect the read and write capabilities. An alternative solution is to reduce the separation to zero by means of a constant contacting force. This alternative has the concern of wear and frictional forces that influence the head location along the track and dynamic instability (bouncing) if the force cannot keep the head touching the disk at all times. In both cases, there is interaction between the solid surfaces and it may be argued that both cases are different stages of contact. In this section, we explore the findings in these areas and the proposed solutions or approaches to reach this goal.
16.7.1 Ultralow Flying As surface roughness of the head and media have been reduced to less than 1 nm Ra [70, 71], it has been possible to fly heads at less than 7 nm. In those cases, effects have also been found that are not explainable by the hydrodynamic simulation equations. Of these effects, the one that causes the greatest concern is known as flying height modulation (FHM) and corresponds to instability of the flying height [70, 72]. A search to understand this phenomenon has generated several theories and approaches. Three explanations for this problem are (1) meniscus forces, (2) electrostatic forces, and (3) intermolecular forces including vdW forces and Casimir forces. Meniscus forces are known to be the dominant forces causing stiction, and it is believed that some time is needed for the meniscus to form. However, studies have shown immediate formation of this meniscus. To be able to measure precisely the meniscus force, Ono and Ohara [73], bounced a spherical slider to a nonrotating disk, lubricated with ∼2.5 nm of standard lubricant and found out that the ball
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needed to bounce back 8 nm before it was completely separated from the disk. The argument is that the meniscus formed around the ball immediately after touching the surface. The calculated contact time was only 20 µsec, indicating a fast meniscus formation. Another study by Kato et al. [74], presents an analysis of a head having the possibility of contacting the surface of the disk while flying. They argue that the system could be represented by a 3-DOF model with the possibility of four different regimes (1) flying regime (no contact), (2) diving regime (closer flying pad touch and penetration of the lubricant layer), (3) contacting regime (the pad displaces the lubricant layer completely and touches the disk’s hard overcoat layer, and (4) the jumping or detachment regimen when the pad bounces from the surface because of the energy of the contact but the meniscus forces contribute to the retardation of this detachment. By using this system model, along with the 3-DOF model the authors introduce critical frequencies in the system. At these frequencies, the FHM is larger. It also shows that by increasing the thickness of the lubricant layer and by increasing the surface energy of the lubricant the amplitude of the modulation can be reduced. Although no argument has been made that electrostatic forces cause FHM, it has been proven that electrical charge between the head and the disk can create an attraction that produces measurable results to the Flying Height Spacing [75,76]. Feng et al. [75] studied the effects of applying a DC voltage between the head and the disk, to the magnetic signal, quantified by the pulse width at 50% or PW50, known to be directly affected by the FH. They found that a mere 3 V could cause more than a 10% increase of the PW50. This roughly translates to more than 10% of the FHt that in this study was 12 nm. The authors, more concerned with the final FHt than with possible modulations, indicate that by having surface coating in the head and in the disk that create similar surface potentials, the effect of electrostatic attraction can be minimized. It is known that when the distance between bodies is small, intermolecular forces such as vdW and Casimir forces can produce an effect in the interaction between the bodies. Li et al. [77], proposed a model that includes both intermolecular forces, given that certain conditions occur. In their argument, the vdW force dominates for values below 10 nm, making the Casimir forces almost negligible and so will not be presented here. The vdW forces represented by the Lennard–Jones Potential equation in this case of slider and disk, can be written as: FvdW = −
A B + 6π a 3 45π a 9
where A is the Hamaker constant, B is a constant that depends on the atoms present in the interaction, and a is the distance between the slider and the disk. The first term on the right-hand side is the attractive vdW force and the second term is the repulsive intermolecular force. The strength of the attractive force increases with reduction in the spacing until it becomes small enough for the repulsive force to become dominant. The analysis of Li et al. concludes that, for the case of a 3.8 nm FHt, the effect of the vdW force produces ∼0.3 nm reduction on FHt, and this value could be reduced to less than 0.1 nm, if the trailing edge pad is reduced in area (no clear indication of the area reduction is provided, but it is believed to be linear). Thornton and Bogy [72, 78], presented a very clear case of flying instability that occurred between 2.5 and ∼7 nm (Figure 16.32) for the system analyzed. The instability is well explained by the existence of three equilibrium states between the 3.5 and 7.25 nm nominal FHt. Within this regime, the potential energy takes on a special form generally called a “double-well” potential. This creates an extremely complex dynamic system that can even be chaotic. By introducing intermolecular forces represented by the vdW equation (see above) they show that the instability stages can be predicted. The slider can fly stable at 4 or 5 nm until a perturbation of the system makes the molecular forces dominate and instability starts. The authors propose a set of recommendations to reduce this problem (1) reduce the total effect of the intermolecular force (smaller area or different molecules interacting), (2) optimize disk and slider morphology such that FHM (amplitude) is reduced, (3) increase ABS stiffness, and (4) increase ABS damping. The main conclusion is that there may be a fundamental lower FHt limit for a given slider–disk combination, below
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10 a Equilibria [nm]
8
f x*
6
1
4
b
2
x2*
x* 3
d
0 0
2
e
c 4
6
8
10
Nominal FH solution [nm]
FIGURE 16.32 Diagram showing equilibrium and stability at very low flying height for the specific studied system. It shows instability (- - - - -) between 2.8 and 7 nm. (From Thornton, B. and Bogy, D., IEEE Trans. Magn., 39, 2420, 2003. With permission.)
which the slider would not be able to fly due to dynamic instability caused by intermolecular adhesion forces.
16.7.2 Contact Recording — the Last Frontier Contact recording is regarded as the stage where no physical separation occurs between the head and the disk during their entire interaction. This implies continuous sliding. This also implies that the magnetic spacing is reduced to the sum of carbon overcoats in the head and disk and lubricant thickness. This is the ideal situation from the point of view of the magnetic signal, for writing and reading. Interest of industry in this approach is not new, and if it is implemented, it will not be the first time. However, conditions have changed and implementation now is a bigger challenge. In the early 1990s several different concepts were proposed to accomplish contact recording [79–81]. Hamilton et al. [79] presented a very innovative concept in which there was not ABS and the read/write elements were incorporated at the end of a flexible arm that contacted the surface with a force of about 300 µg — somehow similar to the tip of an AFM. It is not known whether this concept was ever implemented. Yeack-Scranton et al. [80], presented another approach that seems more feasible. By partially separating the read and write element location from the rest of the ABS and allowing them to be moved closer to the disk while the rest of the air bearing was still flying, allowed them to have contact recording only when required. They called the concept “practical” contact recording because 99% of the time the head was flying.9 A concept that was finally implemented was the tri-pad — tail dragging air bearing technology [82]. The idea consisted of dragging the ABS area close to where the read/write elements were located, and controlling the ABS forces in a way that only a small and constant load was applied at that location. Although it was thought that no wear was possible, it was later discovered that some wear occurred but was “acceptable.” This concept was implemented as a way to extend the life of the inductive head technologies that, by design, needed to get closer to the disk in order to capture the magnetic field produced by the recorded information. It was claimed that “tri-pad air-bearing design” approach provided the industry with a slider technology to carry inductive heads to 0.7 Gbpsi (bear in mind that today areal density is over 100 Gbpsi). The new magnetic head technology introduced by IBM, the MR head, allowed companies concerned with 9 This concept although not implemented as presented by the authors in 1991, it is being implemented by some HDD during 2005. The new design does not separate the read/write elements from the rest of the ABS, but rather “protrudes them,” toward the disk surface by only a few nanometers (∼2 to 6 nm) when needed by the read/write process. The system uses heat and thermal expansion properties of the materials. The specific technology is proprietary and confidential to each HDD manufacture. This concept seems to be the trend in the industry (2005).
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wear to go back to flying distances. Besides, the MR technology was very susceptible to contact (thermal asperities). It is the author’s opinion that reliability issues encountered by some companies in the late 1990s could have been a consequence of the use of the tail-dragging technology beyond the expected life of the system. However, wear studies continued to appear, indicating that wear was a concern but that it was manageable [83] and long life could be expected if media lubricant and carbon overcoat were optimized [84]. Other studies presented by Itoh et al. [85], and Kawakubo [86] concentrated on predicting the wear life of the system based on the wear generated. Today several concepts have been investigated to reduce the separation to zero. One novel approach consists of manufacturing the read/write elements protruding toward the disk encapsulated in a “WearIn-Pad” (WIP) at the trailing edge of the slider [87]. The idea is that as the slider glides over the disk, the slider overcoat and head protrusion are quickly worn off, resulting in the bottom of read and write elements clearing the disk roughness by only a few nanometers. Although this approach cannot be considered contact recording in the long run, it occurs during the run-in period. This approach not only reduces significantly the nominal spacing, but also reduces the distribution of spacing, thus eliminating manufacturing variability. As indicated by the authors, the new concern becomes the exposure of the read and write elements to the environmental elements since at the end of the process there is no overcoat to protect them from corrosion. The most promising concepts today are based on the same concept introduced by the tri-pad air bearing: Suspension load is supported mainly by the ABS with a very small load acting on the rear contact pad. Mate et al. [88] presented the results of this approach. In their design, less than 5 mN is supported by the read contact pad. They show that the new concept improves the magnetic performance significantly but the new problem becomes the dynamic behavior of the slider when the latest ultra smooth media is used. They found that surface roughness lower than 6 Å Ra produced instability of more than 8 nm (see Figure 16.33). Although no solution is presented, the following sets of approaches is recommended (1) decreasing the width of the contact pad and (2) decreasing the amount of mobile lubricant. These two recommendations although might reduce dynamic instability, may increase the possibility of wear. Experimentation will be necessary to determine this.
Trailing edge bounce (nm)
100 Bounce
10
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0
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Disk RMS roughness (Å)
FIGURE 16.33 Bounce of the slider trailing edge slider vs. the disk roughness: Squares are for sliders with 180-µm wide contact pad, circle for sliders with a 50-µm wide pad running against disks with no free lubricant (the free lubricant has been removed by rinsing with a perfluorohexane solvent). Bounce is the mean-to-peak (top 1%) displacement of the slider normal to the disk in the frequency range of 50 to 1000 kHz as measured using an LDV. Disks are made with isotropic textured substrates, standard magnetic layers, overcoated with sputtered CNx, and lubricated with 1 nm of Fomblin Z-DOL. The RMS roughness was measured by AFM over a 5 µm × 5 µm area. The sliding conditions are zero skew, speed = 9 m/sec, and 0.5 to 1 mN normal contact force. (From Mate, C.M. et al., IEEE Trans. Magn., 40, 3156, 2004. With permission.)
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Even if there is no wear there is the concern of how the contact interaction affects the magnetic recorded information. Two new terms have been quoted for these problems: tribo-decay and tribo-noise [89]. Tribodecay has been assigned to the study of how the sliding energy affects the stability of the magnetic field (data) in the disk. The friction energy is known to transform into heat and this heat can produce thermal decay as indicated by Weller [90]. Friction can also produce small local deformations that could affect the magnetic layer on the disk. Liu et al. argue that “tribo-decay of the recorded data will be an engineering limit of the achievable recording density before the thermal decay which is due to the physical size of the grains in magnetic media.” The tribo-noise deals with how the interaction energy affects the way the giant magneto-resistance (GMR) head senses the signal. Thermal asperities, as one kind of tribo-noise, have been studied and ways to deal with it have been found through the read channels. The concern increases if these thermal asperities can not be isolated and dealt with individually. It is obvious that the tribology challenges of contact recording are significant. Time will tell if the creativity of tribologists working for the HDD industry along with those working in universities and research institutes dedicated to this industry can find those narrow conditions to make contact recording work and also make feasible the manufacturing process.
16.8 Closure Tribology will continue to be one of the major areas of research for the development of the next HDDs. The dimensional limits will be dictated only by the laws of physics, and the interaction between physicists, chemist, material science, and mechanical engineers will be especially important for this application of tribology. The advances in tribology in this industry, will lead the advances in others industries, since it is expected that the need for smoother and smaller surfaces and parts is a trend. An example is the MEMS/NEMS where stiction is also a concern. Regarding where to find the latest information on HDD tribology, the author recommends the following publications10 : IEEE Transactions of Magnetics ASME Journal of Tribology, Proceedings of the Institute Mechanical Engineering Part J: Journal Engineering Tribolology STLE Tribology Letters, Tribology Transactions, Microsystem Technologies, and Journal of Applied Physics. Other, more specific details can be found in very specialized physics and chemistry journals. To be at the leading edge, it is only possible working for the R&D departments of the HDD companies (Seagate, Hitachi, Western Digital, Maxtor, Samsung, Fujitsu and Toshiba, Iomega), head and disk manufacturers (SAE, ALPS, Komag, Showa Denko), or working with one of the few universities that actively support research in this area (i.e., in USA, CML at the University of California, Berkeley; the Center for Magnetic Recording Research [CMRR] at the University of California San Diego, and Data Storage System Center at Carnegie Mellon University, and Nanotribology Laboratory for Information Storage and MEMS/NEMS [NLIM] at The Ohio State University). Obviously, if there are new findings or developments that can give an edge to one company over its competitors, these are not published until they are no longer a secret (i.e., the developments are implemented and the product is on the market). Other countries where research is strong are Japan, Singapore, and Korea because of their relationship with the hard disk drive industry.
Acknowledgments José Castillo would like to express his gratitude to Iomega for their support, to Chris Hahm for innumerable discussion and for providing the AFM and many of the photos presented in this paper, to Michele and Giordano Zett, friends who reviewed the entire document, and especially to my wife Carolina, for being patient and understanding and for her numerous contribution especially in most of the drawings. 10 The author believes that more than 80 to 90% of the journal published papers are published in these publications.
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Nomenclature and Terminology — Hard Disk Drive Industry Gbpsi Tera Penta Exa ABS a-C a-C : H or CHx a-C : N or CNx CSS DLC ECR-CVD ESD FCA FDB FHM FHt Gb GB GMR HDD IBC IBD IDEMA L/UL LZT MR PCVD PFPE RF-PECVD Stiction ULF vdW
Gbit/in.2 = Giga bit per square inch = 109 bits/in.2 1012 1015 1018 air bearing surface amorphous carbon (DLC) hydrogenated carbon (DLC) nitrogenated carbon (DLC) contact start–stop DC/RF sputtering diamond-like carbon electron–cyclotron resonance chemical vapor deposition electrostatic discharge filtered cathodic arc fluid dynamic bearing flying height modulation flying height giga bit giga byte giant magneto-resistance hard disk drive ion-beam carbon direct ion-beam deposition international Disk Drive Equipment and Materials Association load/unload laser zone texturing magneto-resistance plasma chemical vapor deposition perfluoropolyethers RF-plasma enhanced chemical vapor deposition high static friction that cause failure ultra low fly van der Waals
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[8] Suk, M., Dennig, P., and Gillis, D., Magnetic erasures due to impact induced interfacial heating and magnetostriction, ASME J. Tribol., 122, 264, 2000. [9] Bhushan, B., Methodology for roughness measurement and contact analysis for optimization of interface roughness, IEEE Trans. Magn., 32, 1819, 1996. [10] Bhushan, B. and Gupta, B.K., Handbook of Tribology: Materials, Coatings and Surface Treatments, McGraw Hill, New York (1991). Reprint, Krieger Publishing Co., Malabar, FL, 1997. [11] Bhushan, B., Chemical, mechanical and tribological characterization of ultra-thin and hard amorphous carbon coatings as thin as 3.5 nm: recent developments, Diamond Relat. Mater., 8, 1985, 1999. [12] Cho, N.H. et al., Chemical structure and physical properties of diamond-like amorphous carbon films prepared by magneton sputtering, J. Mater. Res., 5, 2543, 1990. [13] Bhushan, B. et al., Characterization of chemical bonding and physical characteristics of diamondlike amorphous carbon and diamond films, J. Mater. Res., 7, 404, 1992. [14] Robertson, J., Amorphous carbon, Adv. Phys., 35, 317, 1986. [15] Scharf, T.W. et al., Mechanical strength and wear resistance of 5 nm IBD carbon overcoats for magnetic disks, IEEE Trans. Magn., 37, 1792, 2001. [16] Yen, B.K. et al., Effect of N doping on structure and properties of DLC films produced by plasma beam deposition, IEEE Trans. Magn., 37, 1786, 2001. [17] Yamamoto, T., Toyoguchi, T., and Honda, F., Ultrathin amorphous C : H overcoats by pCVD on thin film media, IEEE Trans. Magn., 36, 115, 2000. [18] Yamamoto, T. et al., Ultrathin amorphous carbon overcoats by filtered cathodic arc deposition, IEEE Trans. Magn., 39, 2201, 2003. [19] Hyodo, H., Yamamoto, T., and Toyoguchi, T., Properties of tetrahedral amorphous carbon film by filtered cathodic arc deposition for disk overcoat, IEEE Trans. Magn., 37, 1789, 2001. [20] Wang, S., Chu, X., and Liu, J.J., A variable-speed drag test for characterizing wear of ultra-thin carbon films, IEEE Trans. Magn., 37, 1795, 2001. [21] Li, D.J. and Chung, Y.W., Ultrasmooth CNx overcoats for next-generation hard disks, IEEE Trans. Magn., 39, 765, 2003. [22] Pirzada, S.A. et al., Functionalized ion-beam carbon for magnetic disks, IEEE Trans. Magn., 38, 2117, 2002. [23] Chu, M.-Y., Bhushan, B., and De Jonghe, L., Wear behavior of ceramic sliders in sliding contact with rigid magnetic thin-film disks, STLE Tribol. Trans., 35, 603, 1992. [24] Klaus, E.E. and Bhushan, B., Lubricants in magnetic media — a review, in Tribology and Mechanics of Magnetic Storage Systems, Bhushan, B. and Eiss, N.S. (Eds), SP-19, STLE, Park Ridge, Illinois, p. 7, 1985. [25] Bhushan, B. (Ed.), Macro- and microtribology of magnetic storage devices, in Modern Tribology Handbook, CRC Press, Boca Raton, FL, p. 1413, 2001. [26] Zhao, Z. and Bhushan, B., Effect of bonded-lubricant films on the tribological performance of magnetic thin-film rigid disks, Wear, 202, 50, 1996. [27] Bhushan, B. and Zhao, Z., Macro- and microscale studies of molecularly-thick boundary layers of perfluoropolyether lubricants for magnetic thin-film rigid disks, J. Info. Storage Proc. Syst., 1, 1, 1999. [28] Zhao, Z., Bhushan, B., and Kajdas, C., Effect of thermal treatment and sliding on chemical bonding of PFPE lubricant films with DLC surfaces, J. Info. Storage Proc. Syst., 1, 259, 1999. [29] Yanagisawa, M., Adsorption of perfluoro-polyethers on carbon surfaces, in Tribology and Mechanics of Magnetic Storage Systems, Vol. 9, SP-36, STLE, Park Ridge, Illinois, p. 25, 1994. [30] Yang, M. et al., Environmental effects on phosphazene lubricated thin-film disks, IEEE Trans. Magn., 30, 4143, 1994. [31] Chen, C.-Y. et al., The decomposition mechanisms and thermal stability of ZDOL lubricant on hydrogenated carbon overcoats, ASME J. Tribol., 122, 458, 2000.
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[32] Tagawa, N. et al., Spreading characteristics of cyclotriphosphazine-terminated perfluoropolyether films on carbon surfaces, IEEE Trans. Magn., 39, 2441, 2003. [33] Tagawa, N. et al., Spreading of novel cyclotriphosphazine-terminated PFPE films on carbon surfaces, ASME J. Tribol., 126, 754, 2004. [34] Gao, C. and Bhushan, B., Tribological performance of magnetic thin-film glass disks: its relation to surface roughness and lubricant structure and its thickness, Wear, 190, 60, 1995. [35] Zhao, Z. and Bhushan, B., Effect of lubricant thickness, viscosity and rest time on long-term stiction in magnetic thin-film rigid disks, IEEE Trans. Magn., 34, 1708, 1998. [36] Bhushan, B. and Zhao, Z., Friction/stiction and wear studies of magnetic thin-film disks with two polar perfluoropolyether lubricants, IEEE Trans. Magn., 33, 918, 1997. [37] Zhao, Z. and Bhushan, B., Effect of environment on the friction/stiction and durability of lubricated magnetic thin-film disks, Proc. Inst. Mech. Eng., Part J: J. Eng. Tribol., 211, 295, 1997. [38] Hahm, C.D. and Bhushan, B., Lubricant film thickness mapping using a capacitance technique on magnetic thin-film rigid disks, Rev. Sci. Instrum., 69, 3339, 1998. [39] Xu, J. and Bhushan, B., Friction and durability of ceramic slider materials in contact with lubricated thin-film rigid disks, Proc. Inst. Mech. Eng., Part J: J. Eng. Tribol., 211, 303, 1997. [40] Kasai, P., Tang, W.T., and Wheeler, P., Degradation of perfluoropolyethers catalyzed by aluminum, Appl. Surf. Sci., 51, 201, 1991. [41] Kasai, P., Perfluoroethers: intramolecular disproportionation, Macromolecules, 25, 6791, 1992. [42] Zhao, Z. and Bhushan, B., Tribological performance of PFPE and X-1P lubricants at head–disk interface — part I. Experimental results, Tribol. Lett., 6, 129, 1999. [43] Zhao, X. and Bhushan, B., Lubrication studies of head–disk interfaces in a controlled environment part 1: effects of disk texture, lubricant thermal treatment and lubricant additive on durability of head–disk interface, Proc. Inst. Mech. Eng., Part J: J. Eng. Tribol., 214, 535, 2000. [44] Vurens, G., Zehringer, R., and Saperstein, D., The decomposition mechanism of perfluoropolyether lubricants during wear, in Surf. Sci. Inv. Tribol., Chung, Y.W., Homola, A.M., and Street, G.B. (Eds), ACS, Washington, DC, 169, 1992. [45] Zhao, X., Bhushan, B., and Kajdas, C., Lubrication studies of head–disk interfaces in a controlled environment part 2: degradation mechanisms of perfluoropolyether lubricants, Proc. Inst. Mech. Eng., Part J: J. Eng. Tribol., 214, 547, 2000. [46] Zhao, X. and Bhushan, B., Studies on degradation mechanisms of lubricants for magnetic thin-film rigid disks, Proc. Inst. Mech. Eng., Part J: J. Eng. Tribol., 215, 173, 2001. [47] Murayama, H., Sano, K., and Yokoyama, E., Lubricant bonding via terminal bond network, IEEE Trans. Magn., 31, 2922, 1995. [48] Choi, J., Kawaguchi, M., and Kato, T., Possibility of organic monolayer films as lubricants for disk drives: comparative study of PFPE and organosilane, ASME J. Tribol., 125, 850, 2003. [49] Choi, J., Kawaguchi, M., and Kato, T., Nanoscale lubricant with strongly bonded phase and mobile phase, Tribol. Lett., 15, 353, 2003. [50] Kondo, K. et al., Design and construction of magnetic storage devices, in Handbook of Micro/Nano Tribology, 2nd ed., Bhushan, B. (Ed.), CRC Press, Boca Raton, FL, p. 597, 1999. [51] Ramirez, A.G. et al., The effects of slider material on the gasification of carbon, ASME J. Tribol., 124, 771, 2002. [52] Zhao, Z. and Bhushan, B., Effect of head slider DLC overcoats produced by various deposition techniques on the interface failure, IEEE Trans. Magn., 36, 2663, 2000. [53] Song, D., Schnur, D., and Boutaghou, Z.E., Discharge mechanism for electrostatic fly control, IEEE Trans. Magn., 40, 3162, 2004. [54] Khonsari, M. and Booser, E., Applied Tribology, John Wiley & Sons, New York, 2001. [55] Fukui, S. and Kaneko, R., Analysis of ultra-thin gas film lubrication based on linearized Boltzmann equation: first report — derivation of a generalized lubrication equation including thermal creep flow, ASME J. Tribol., 110, 253, 1988.
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[56] Huang, W. and Bogy, D.B., The effect of the accommodation coefficient on slider air bearing simulation, ASME J. Tribol., 122, 427, 2000. [57] Garnier, M.F., Tang, T., and White, J.W., U.S. Patent No. 3,855,625, Magnetic head slider assembly, Dec. 17th, 1974. [58] Stiction, Webopedia — online dictionary for words, phrases and abbreviations related to computer and Internet technology (http://www.webopedia.com/TERM/s/stiction.html), Nov. 25, 2004. [59] Chilamakuri, S.K. and Bhushan, B., A comprehensive kinetic meniscus model for prediction of long-term static friction, J. Appl. Phys., 86, 4649, 1999. [60] Chilamakuri, S.K., Zhao, X., and Bhushan, B., Failure analysis of laser-textured surfaces, Proc. Inst. Mech. Eng., Part J: J. Eng. Tribol., 214, 471, 2000. [61] Yamada, T. and Bogy, D.B., Load–unload slider dynamics in magnetic disk drives, IEEE Trans. Magn., 24, 2742, 1988. [62] Bhushan, B. and Tambe, N.S., Role of particulate contamination on friction and wear and durability of load/unload and padded picosliders, IEEE Trans. Magn., 39, 857, 2003. [63] Hiller, B., Yaeger, J.R.D., and Sonnenfeld, R.G., Ramp load/unload friction dependence on temperature, velocity and ramp material, IEEE Trans. Magn., 37, 1852, 2001. [64] Suk, M. and Jen, D., Potential data loss due to head/disk contacts during dynamic load/unload, IEEE Trans. Magn., 34, 1711, 1998. [65] Tanaka, H., Kohira, H., and Matsumoto, M., Effect of air-bearing design on slider dynamics during unloading process, IEEE Trans. Magn., 37, 1818, 2001. [66] Zeng, Q.H., Chapin, M., and Bogy, D.B., Dynamics of the unload process for negative pressure sliders, IEEE Trans. Magn., 35, 916, 1999. [67] University of California, Berkeley, Load–Unload simulation software. Computer Mechanical Laboratory (CML) (http://cml.me.berkeley.edu/). [68] Zeng, Q.H. and Bogy, D.B., A Simplified 4-DOF suspension model for dynamic load/unload simulation and its application, J. Tribol., 122, 274, 2000. [69] Weissner, S., Zander, U., and Talke, F.E., A new finite-element based suspension model including displacement limiters for load/unload simulations, ASME J. Tribol., 125, 162, 2003. [70] Lee, S.-C. and Polycarpou, A.A., Adhesion forces for sub-10 nm flying-height magnetic storage head–disk interfaces, ASME J. Tribol., 126, 334, 2004. [71] Zeng, Q.H. et al., Flyability and flying height modulation measurement of sliders with sub-10 nm flying heights, IEEE Trans. Magn., 37, 894, 2001. [72] Thornton, B.H. and Bogy, D.B., Head–disk interface dynamic instability due to intermolecular forces, IEEE Trans. Magn., 39, 2420, 2003. [73] Ono, K. and Ohara, S., Experimental identification of elastic force, damping force and adhesion force in collision of a spherical slider with a magnetic disk (INVITED), ASME Tribology Symposium, Santa Clara University, CA, USA, June 16th, 2004. [74] Kato, T., Watanabe, S., and Matsuoka, H., Dynamic characteristics of an in-contact headslider considering meniscus force: Part 1 — formulation and application to the disk with sinusoidal undulation, ASME J. Tribol., 122, 633, 2000. [75] Feng, Z., Cha, E., and Zhang, X., A study of electrical charge at head–disk interface, ASME Information Storage and Processing Systems Conference 2004, Santa Clara University, CA, USA, June 15th, 2004. [76] Knigge, B.E. et al., Influence of contact potential on slider-disk spacing: simulation and experiment, IEEE Trans. Magn., 40, 3165, 2004. [77] Li, J. et al., Effects of intermolecular forces on deep sub-10 nm spaced sliders, IEEE Trans. Magn., 38, 2141, 2002. [78] Thornton, B.H. and Bogy, D.B., A parametric study of head–disk interface instability due to intermolecular forces, IEEE Trans. Magn., 40, 337, 2004. [79] Hamilton, H., Anderson, R., and Goodson, K., Contact perpendicular recording on rigid media, IEEE Trans. Magn., 27, 4921, 1991.
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[80] Yeack-Scranton, C.E. et al., An active slider for practical contact recording, IEEE Trans. Magn., 26, 2478, 1990. [81] Bogy, D.B. et al., Some critical tribological issues in contact and near-contact recording, IEEE Trans. Magn., 29, 230, 1993. [82] Leung, C.M. and Gitis, N.V., Tri-pad airbearing technology, IEEE Trans. Magn., 32, 3657, 1996. [83] Wang, R.-H. et al., Challenges of the head–disk interface for near contact and contact recording, IEEE Trans. Magn., 35, 2466, 1999. [84] Chao, J.L. and Russak, M.A., Media tribology optimization for proximity recording, IEEE Trans. Magn., 33, 897, 1997. [85] Itoh, J. et al., An experimental investigation for continuous contact recording technology, IEEE Trans. Magn., 37, 1806, 2001. [86] Kawakubo, Y. et al., Wear life prediction of contact recording head, IEEE Trans. Magn., 39, 888, 2003. [87] Singh, G.P. et al., A novel wear-in-pad approach to minimizing spacing at the head/disk interface, IEEE Trans. Magn., 40, 3148, 2004. [88] Mate, C.M. et al., Dynamics of contacting head–disk interfaces, IEEE Trans. Magn., 40, 3156, 2004. [89] Liu, B., Yuan, Z.-M., and Man, Y.-J., Tribo-magnetics and nanometer spaced head–disk systems, IEEE Trans. Magn., 37, 918, 2001. [90] Weller, D. and Moser, A., Thermal effect limits in ultrahigh-density magnetic recording, IEEE Trans. Magn., 35, 4423, 1999.
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17 Biotribology: Material Design, Lubrication, and Wear in Artificial Hip Joints 17.1 Biotribology . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
17-1
Definition • General Methodology of Biotribology Studies
17.2 Hip Joints and Artificial Replacements . . . . . . . . . . . . . . .
17-2
Natural Hip Joints and Artificial Hip Replacements • Bearing Systems and Operating Conditions
Z.M. Jin and J. Fisher School of Mechanical Engineering, Institute of Medical and Biological Engineering, University of Leeds
E. Ingham School of Biochemistry and Microbiology, Institute of Medical and Biological Engineering, University of Leeds
17.3 Material Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
17-5
Biocompatibility • Biomaterials • Bearing Couples • Fixation • Biological Responses to Wear Particles
17.4 Contact Mechanics, Lubrication, and Wear in Artificial Hip Joints . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
17-7
Introduction • UHMWPE-on-Metal • Metal-on-Metal • Ceramic-on-Ceramic • Other Bearing Materials and Combinations
17.5 Summary . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
17-15 17-17
17.1 Biotribology 17.1.1 Definition Biotribology is generally involved with all aspects of tribology related to biological systems [1]. Typical examples of tribology applied to biology include: • • • •
Wear of denture [2,3] Friction of skin and comfort of clothes, socks, and shoes [4,5], and slipperiness [6,7] Tribology of contact lenses and ocular tribology [8] Tribology at microlevels — inside cells, vessels, and capillaries such as lubrication by plasma of red blood cells in narrow capillaries [9] • Wear of replacement heart valves [10] • Lubrication of the pump in total artificial hearts [11] 17-1
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• Wear of screws and plates in bone fracture repair [12] • Lubrication in pericardium and pleural surfaces [13] • Tribology of natural synovial joints and artificial replacements [14,15] The details of each of the topics listed above can be found in the various references cited. The main focus of this chapter is to review studies of contact mechanics, friction, lubrication, and wear in artificial hip joints.
17.1.2 General Methodology of Biotribology Studies The general approach adopted for biotribological studies involves the following steps: • Obtain input conditions from anatomical and physiological considerations • Select appropriate tribological models of friction, wear, and lubrication • Discuss importance of tribology in the context of biology Although mainly built on the fundamentals of engineering tribology, biotribological studies have extended well beyond these engineering disciplines. For example, the unique load support from the pressurized fluid within articular cartilage provides further understanding of the mixed lubrication regime in synovial joints [16]. Unique lubrication mechanisms associated with biological surfaces and lubricants such as brush-like structures may help to design alternative engineering solutions [17,18]. Tribological inputs and biological responses have been shown to be particularly important in the design and optimization of the bearing surfaces for artificial joints [19–21].
17.2 Hip Joints and Artificial Replacements 17.2.1 Natural Hip Joints and Artificial Hip Replacements The natural hip joint, as shown schematically in Figure 17.1(a) for its major bearing components, is a remarkable bearing. It is covered by articular cartilage, which is an extremely complex, structured biological material mainly consisting of solid and fluid phases (biphasic). The lubrication mechanism in the natural synovial joint is still unclear, although a number of suggestions have been made. The low elastic modulus associated with articular cartilage promotes surface deformation and hence elastohydrodynamic lubrication at both macroscopic and microscopic scales during steady-state walking cycles [22]. However, the lubrication mechanism in synovial joints remains effective under adverse conditions such as prolonged loading, start-up, and stopping. One of the proposed alternative lubrication mechanisms is the load support due to the fluid phase in articular cartilage [16,23,24]. Therefore, the prediction of the load support due to the fluid phase of the articular cartilage may provide further understanding of the lubrication mechanism in natural synovial joints. Natural synovial joints are expected to function in the human body for a lifetime, while transmitting large dynamic loads and yet accommodating a wide range of movements. However, diseases such as osteoarthritis, rheumatoid arthritis, and trauma often require these natural bearings to be replaced by artificial ones. Total joint replacement has been the most successful surgical treatment for hip joint diseases in the last 50 years. Currently, over one million hip joint replacements are carried out worldwide every year. The majority of these devices utilize a material combination of ultra high molecular weight polyethylene (UHMWPE) as the acetabular cup articulating against either a ceramic (alumina) or a metallic (stainless steel, cobalt-based alloy) femoral head as shown in Figure 17.1(b). These man made bearings can sometimes last more than 20 years in the body without failure. However, osteolysis and subsequent loosening of the prosthetic components has recently been identified as the major clinical limiting factor, often leading to revision after only about 10 years. This complication has become more significant recently because of the increasing use of these devices in younger, more active, and demanding patients with life expectancies after surgery in excess of 25 years. It is generally accepted that osteolysis and
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(a)
(b) Articular cartilage
Bone
Synovial fluid
Bone Synovial membrane
(c)
(a)
(b)
(c)
(d)
(e)
FIGURE 17.1 (a) A schematic diagram of the synovial hip joint showing articular cartilage and other bearing components. (b) A typical artificial hip joint consisting of a metallic femoral head and an UHMWPE acetabular cup. (c) Five examples of hip implants (a) Charnley UHMWPE-on-Metal, (b) Muller UHMWPE-on-Metal, (c) UHMWPEon-Ceramic, (d) Ceramic-on-Ceramic, and (e) Metal-on-Metal.
loosening is most frequently caused by adverse tissue reactions to UHMWPE wear particles. Consequently, minimizing UHMWPE wear and wear particle generation is an important strategy in prolonging the service life of artificial joint replacements, and a number of bearing material combinations have been introduced recently to reduce wear and wear particle load as shown in Figure 17.1(c) and discussed in Section 17.3. Tribological studies of artificial hip joints play a major role in the long-term success of these medical devices. Wear particles are mainly generated by direct surface contact, causing adhesive and abrasive wear, as well as subsurface stresses and fatigue. Much research work in this field is now focusing on wear and wear debris induced osteolysis in currently used artificial hip joints, through various tribological studies of contact mechanics, friction, lubrication, and wear using both experimental and theoretical tools.
17.2.2 Bearing Systems and Operating Conditions The bearing surfaces of natural hip joints are articular cartilage. Diseased or damaged cartilage can be repaired or replaced with the aim to retain some or all of the normal functions using minimally invasive and conservative bone preserving surgeries. For example, if the cartilage surface is damaged only at one site, it can be repaired locally using tissue engineering approaches or replaced either locally (partial hemi-arthroplasty [25,26]) or totally (hemi-arthroplasty) depending on the severity of the damage. More recently, an intraarticular metallic spacer technology has been refined and reintroduced between the two cartilage surfaces [27]. However if articular cartilages are damaged at both sites but the bone quality is still maintained as found in some younger patients, hip resurfacing procedures are usually performed to replace the damaged cartilage surfaces. Treatment of end stage arthritis currently involves major
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surgical intervention and total joint replacement. Therefore, various combinations of bearing surfaces are encountered in treating hip diseases, including cartilage/tissue engineered cartilage, cartilage/biomaterial, and biomaterial/biomaterial. The potential candidate biomaterials are discussed in Section 17.3. Hip implants in vivo are subjected to intermittent motion and it is important to remember that the tribological conditions can frequently change between steady-state walking and start-up/stopping. Steadystate walking cycles consist of both stance phase and swing phase [28]. During the stance phase, the load can reach up to five times bodyweight while the surface velocities are relatively low. During the swing phase, the load is greatly reduced and the sliding velocity increases. In addition, the directions of both the total load and the velocity vectors change during walking, although the major load component is in the vertical direction and the major velocity component is in the flexion/extension direction. Figures 17.2(a) and 17.2(b) show the transient variation of the load in the vertical direction and the resultant angular velocity (mainly in the flexion/extension direction) in hip joints experienced during cyclic walking which has often been used for theoretical lubrication analyses [29]. The load and speed employed in hip simulators range from a three-dimensional physiological pattern to simplified versions of specified components, generally consisting of a vertical load and a flexion/extension motion, with an internal–external rotation often added [31]. An average load of between 1200 N (averaged over a complete cycle) and 2500 N (averaged over the stance phase), equivalent to about 1.6 to 3 times bodyweight of 750 N and an average resultant angular velocity of about 1.5 rad/sec (mainly in the flexion/extension direction) have been suggested for steady-state lubrication analysis under in vivo conditions [29,32]. The lubricant in healthy natural joints is called synovial fluid and the lubricant in total joint replacements is periprosthetic synovial fluid, similar to the synovial fluid obtained from patients with osteoarthritis [33,34]. The lubricant used for simulator testing is usually bovine serum, diluted to various concentrations. Laboratory measurements of viscosity at different shear rates for various types of synovial fluids from both normal and diseased or replaced joints and bovine serum have all shown powerful non-Newtonian shear thinning characteristics under relatively low shear rates up to 103 l/sec but this shear thinning declines significantly at higher shear rates [35]. It has been estimated by Jin et al. [32] that the realistic shear rates experienced in both natural and replaced hip joints are of the order of 105 to 107 l/sec under steady walking conditions. Thus, it is often possible to approximate the non-Newtonian synovial fluid with reasonable accuracy as a Newtonian fluid for the lubrication analysis. Furthermore, the major piezo-viscous effect observed for mineral oils does not appear to be effective in synovial fluid from osteoarthritic patients for pressures up to 100 mPa [35] and this behavior is unlikely to change for synovial fluid in arthroplasty patients. This pressure level is also unlikely to be exceeded in fluid-films encountered in UHMWPEon-metal [36] or even in more rigid metal-on-metal [37] or ceramic-on-ceramic hip implants [38]. Therefore, the synovial fluid found in arthroplasty patients is not likely to be influenced by film pressures
(b)
(a) 4000
5
Paul [28]
Johnston and Smidt [30] 4 v (rad/sec)
w (N)
3000 2000 1000
3 2 1
0
0 0
0.2
0.4 t (sec)
0.6
0.8
1
0
0.2
0.4
0.6
0.8
1
t (sec)
FIGURE 17.2 (a) Transient load (w) variation in the hip joint as function of time (t ) during walking for a body weight of 750 N (redrawn from Medley et al., World Tribology Forum in Arthroplasty, C. Rieker et al., Eds, Hans Huber, 2001, 125.) (b) Transient resultant angular velocity (ω) in the hip joint as a function of time during walking (redrawn from Medley et al., World Tribology Forum in Arthroplasty, C. Rieker et al., Eds, Hans Huber, 2001, 125.)
© 2006 by Taylor & Francis Group, LLC
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or by non-Newtonian shear thinning. Representative viscosities of 2.5 and 0.9 mPa have been measured for periprosthetic synovial fluid and 25% bovine serum respectively [39].
17.3 Material Design 17.3.1 Biocompatibility The materials used for artificial hip joint replacements must be biocompatible. Both effects of the implant on the host tissues and effects of the host environment on the implant must be addressed. On one hand, implanted biomaterials may cause infection, toxicity, hypersensitivity, and so forth, while on the other they may be subjected to corrosion, degradation, and so on. Further details on the implant–tissue interactions can be found elsewhere [40]. However, it should be noted that there are different biological effects between bulk and particulate forms of any materials as reviewed in Section 17.3.5.
17.3.2 Biomaterials The major biomaterials used to repair or replace damaged articular cartilage include: • Polymers: polyurethane, polyvinyl alcohol hydrogel, UHMWPE, and so on. • Metals: stainless steel, cobalt chromium alloy, and so on. • Ceramics: alumina, zirconia, oxinium, and so on. Various bearing couples can be constructed and this will be discussed in Section 17.3.3.
17.3.3 Bearing Couples The bearing couple for artificial hip joints can be generally classified into two categories: • Soft-on-hard: UHMWPE-on-metal or ceramic • Hard-on-hard: Metal-on-metal, Ceramic-on-ceramic Table 17.1 gives the detailed combinations currently in clinical use or being researched. Two important parameters characterizing the mechanical properties of the biomaterials used to construct artificial hip joints include elastic modulus and Poisson’s ratio as summarized in Table 17.2. In addition to these mechanical parameters, the geometry of the bearing couples is also important. For a ball-in-socket configuration, the main geometric parameters are the radii of the femoral head (Rhead ) and the acetabular cup (Rcup ) or the radial clearance (c = Rcup − Rhead ) between the femoral head and the acetabular cup. These can be combined as an equivalent radius defined as: R=
Rcup Rhead Rhead (Rhead + c ) = Rcup − Rhead c
(17.1)
Typical equivalent radii used in current artificial hip joints are summarized in Table 17.3. TABLE 17.1
Bearing Couples Available for Hip Implants in Clinical Use or Being Researched Acetabular cup
Femoral head Stainless steel CoCr Alumina Zirconia Alumina composite
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UHMWPE
Cross-linked UHMWPE
CoCr
Alumina
Polyurethane
Alumina Composite
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Handbook of Lubrication and Tribology TABLE 17.2 Elastic Modulus (E) and Poisson’s Ratio (ν) for Various Bearing Materials Used in Hip Implants Bearing/structural materials Polyurethane UHMWPE Cross-linked UHMWPE Bone cement Bone Stainless steel CoCr Zirconia Alumina
E (GPa)
ν
0.02 0.5–1 0.8–2
0.4–0.5 0.4 0.4
2.5 0.8–17 210 230 210 380
0.25 0.3 0.3 0.3 0.26 0.26
TABLE 17.3 Typical Values (Range) of Femoral Head Radius, Radial Clearance, and Effective Radius for Hip Implants with Various Bearing Couples Bearing couples Polyurethane-on-metal UHMWPE-on-metal Cross-linked UHMWPE-on-metal Metal-on-metal Ceramic-on-ceramic
Femoral head radius (mm)
Radial clearance (µm)
Equivalent radius (m)
Ref.
16 (16–25) 14 (11–16) 14 (14–22)
∼250 150 (80–180) 150 (100–200)
1.04 (1–2) 1.32 (0.5–2.5) 1.32 (1–5)
[41,42] [36] [43]
14 (14–30) 14 (14–16)
30 (30–150) 40 (10–40)
∼5 ∼5
[38]
17.3.4 Fixation The tribology performance of artificial hip joints mainly depends on the bearing surfaces, but fixation of the implant to the bone is equally important. Successful tribological functions cannot be maintained if fixation fails. Furthermore, fixation may also influence tribological performance. For example, the status of the fixation interface may change both the contact mechanics and lubrication at the bearing surfaces, and equally the friction at the bearing surfaces may impose additional stresses at the fixation interface. This is particularly important in some recent forms of artificial hip joints such as the hip resurfacing prosthesis. Another important consideration for implant fixation includes optimization of the stem to minimize the problem of stress shielding [44]. There are basically two fixation methods; cemented and cementless. Cemented fixation uses PMMA bone cement (polymethyl methacrylate) and depends on the mechanical interlock between the bone and the cement. Cementless fixation generally relies on the primary mechanical stability achieved through interference fit and anchorage, often supplemented by secondary long-term biological bone in-growth.
17.3.5 Biological Responses to Wear Particles Current designs of hip prostheses do not wear out, rather they fail due to late aseptic loosening as a result of bone loss (osteolysis) which is caused by an adverse biological response to particulate wear debris. Particles of UHMWPE released from the articulating interface enter the periprosthetic tissue where they are taken up by macrophages. In their frustrated attempts to eliminate the bio-inert particles, the macrophages release a range of mediators of inflammation including cytokines which have osteolytic effects. Studies have shown that particles of a critical size (0.1 to 1 µm) are the most biologically reactive [45]. The result is the development of an inflamed granulomatous tissue rich in macrophages, multinucleated giant cells,
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and UHMWPE wear debris. Cytokines produced by the particle activated macrophages are concerned in the recruitment, differentiation, and activation of osteoclasts, the cells responsible for bone resorption. The concentration of wear debris particles from periprosthetic tissues is directly related to the duration of implantation [46] and there are billions of submicrometer particles generated per gram of tissue. It has been reported that osteolysis is likely to occur when the threshold of particles exceeds 1 × 1010 /g of tissue [47]. Each milligram of polyethylene wear has been estimated to generate 1.3 × 1010 particles [48]. The biological events have been recently reviewed by Ingham and Fisher [45].
17.4 Contact Mechanics, Lubrication, and Wear in Artificial Hip Joints 17.4.1 Introduction Contact mechanics is the study of load transfer between two contacting solids. The main parameters determined from contact mechanics analysis include contact area and stresses. Contact stresses are generally related to structural failure and fatigue-related wear mechanisms. A large contact area is often required to produce a low contact stress under a given load. However, if the contact area is too large, the contact may well be extended to the equatorial region and the edge of the cup, not only leading to stress concentrations — limiting the normal movement of the hip joints, but also blocking the lubricant entry and causing lubricant starvation and depletion. In addition to the tribological studies at the bearing surfaces, contact mechanics can also be used to investigate the implantation simulation of the prosthetic components such as press-fit surgical procedure of acetabular cups to examine the stresses in the bone and the deformation of the prosthetic components. Furthermore, contact mechanics at microscopic scales can provide important information for the study of the surface damage which can lead to a significant increase in wear. Contact mechanics studies of hip implants can be achieved through both theoretical modeling and experimental measurements. Experimental methods usually use a pressure-sensitive film [49] or sensors to measure the contact area and the contact pressure experienced at the bearing surfaces [50,51]. Theoretical predictions are usually carried out numerically using the finite element or finite difference method. Sometimes, approximate models based on the Hertz contact theory or simple analytical solutions based on the concept of column model can provide a preliminary estimation [52]. Friction plays an important role in the stresses generated at the bearing surfaces as well as the load transmission from the bearing surfaces to the fixation interface. However, friction measurement is often employed to indicate the lubrication regime based on the Stribeck curve used in engineering [53]. The measurement of friction experienced in hip implants is usually achieved by using pendulum simulators with the cup fixed to a lower stage floating on hydrostatic bearings so that virtually all of the frictional torque comes from the implant. It is particularly important to minimize experimental inaccuracies due to mechanical vibrations and slight misalignments between the centers of the rotation and the femoral ball in order to measure low friction, particularly in bearing couples which may enjoy fluid-film lubrication. It is often necessary to simplify the three-dimensional loading and motion patterns discussed in Section 17.2.2 in order to obtain accurate friction measurements, a typical example being a flexion/extension motion under a vertical load. However, it is important to employ a dynamic load representative of walking cycles in order to investigate the potential squeeze-film effect [54]. The friction magnitude alone often indicates the mode of lubrication, since each mechanism is associated with broad ranges of the coefficient of friction [15]. This can be further extended to investigate the variation of the friction coefficient against the Sommerfeld number, that is, the Stribeck curve, in order to assess lubrication regimes. If the measured friction factors remain constant, fall, or increase as the Sommerfeld number is increased, the associated modes of lubrication are boundary, mixed, or fluid-film respectively. Lubrication plays an important role in the tribological and clinical performance of artificial hip joints. Since a synovial fluid type of lubricant forms in the joint capsule after total hip arthroplasty as discussed
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in Section 17.2.2, it may be possible to promote maximum lubricant protection of the implant surfaces. The inherent boundary lubricating ability of the synovial fluid can always act to reduce the severity of direct asperity contacts. It is, perhaps, more effective to develop bearings that promote fluid-film lubrication (while avoiding fatigue from subsurface stresses), thus separating the surfaces fully or partially and reducing the proportion of the total load carried by asperity contacts. Conventional engineering methods of assessing lubrication regimes have been applied to hip implants. These can be broadly classified into two categories, experimental measurements and theoretical predictions. The experimental methods usually have involved either friction measurements as discussed above, or the detection of separation between the two bearing surfaces using a simple resistivity technique. Separation of the head and cup surfaces by a lubricant can be detected directly by applying a small electrical potential (voltage) across the head/cup contact. It is necessary to have conducting bearing surfaces that are insulated from each other, except when the tips of the asperities are in contact, thus making the contact itself a resistance element in the circuit. The voltage causes a correspondingly small current flow through asperity contacts. If the surfaces are separated, the film resistance is large enough to prevent any detectable current flow [55]. The theoretical predictions have usually been based on the λ ratio defined as follows [29,32]: λ=
2 [Ra_head
hmin 2 + Ra_cup ]1/2
or
λ=
2 [σhead
hcen 2 ]1/2 + σcup
(17.2)
The key to the theoretical assessment of the potential for fluid-film lubrication is the accurate prediction of a representative film thickness, usually the minimum (hmin ) or the average (hcen ), for the bearing and accurate measurements of average surface roughness (Ra ) or root mean square roughness (σ ). The theoretical prediction of the lubricant film thickness is generally involved with the simultaneous solution to both Reynolds and the elasticity equations in appropriate spherical ball-in-socket coordinates and under both anatomical and physiological conditions. In general, the finite element method is used to solve the elasticity equation due to the complex spherical geometry and underlying supports to the prosthetic components [37]. The application of the fast Fourier transform (FFT) technique has resulted in a significant reduction in the computing time required to calculate the elastic deformation [56]. Consequently, it is possible to adopt a large number of grid points in the numerical solution to the lubrication problem and more realistic conditions such as the low viscosity associated with the synovial fluid and transient walking conditions can be considered [57,58]. The analysis of the mixed lubrication regime, which consists of both fluid-film and boundary lubricated regions, has also been attempted for artificial hip joints [59]. The most important tribological parameter which affects the clinical outcome of artificial hip joints is wear and particularly wear debris. Compared with contact mechanics, friction, and lubrication, wear has been the most difficult to study, particularly wear mechanisms. Consequently extensive experimental studies have been carried out using a wide range of laboratory equipment, test methods, and measuring systems in total replacement hip joints. The three major forms of equipment have been: • Pin-on-disc machines • Pin-on-plate machines • Joint simulators The pin-on-disc machine has been widely used in tribology and is particularly useful in the evaluation of the nature, wear, and friction of material pairs under well-controlled, steady-state conditions of load, sliding speed, and environment. The pin-on-plate machine sacrifices the steady sliding speed between specimens, but partially simulates the reciprocating action broadly associated with the hip joint during walking. It is sometimes necessary to increase the complexity of these simple devices to introduce a rotation (multidirectional motion) [60]. Nevertheless, the main purpose of these simple devices is to screen potential material combinations. As interest in the comparative performance of hip joints of different designs and bearing couples developed, it has became necessary to design and build joint simulators for
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laboratory studies. These machines simulate to a greater or lesser extent the three-dimensional loading and motion patterns experienced by hip joints, while immersing the test joints in a lubricant deemed to be physically and chemically similar to synovial fluid. Wear can be evaluated by dimensional or gravimetric means. An early form of versatile hip joint simulator developed at Leeds in the 1980s [61] has formed the basis for a number of subsequent designs. Hip joint simulators have since developed enormously, in both sophistication and number. This is a clear indication not only of the value of such devices, but also of the international recognition that careful laboratory evaluation of new products should precede clinical trials. It should be pointed out that although contact mechanics, friction, lubrication, and wear have been discussed separately above, it is important to integrate and couple these studies in order to provide an overall tribological evaluation. For example, simulators have been used mainly for wear studies, which themselves reveal much about the prevalent modes of lubrication, and they have also been used more recently for direct studies of lubrication [55]. Contact mechanics, coupled with the kinematics in hip joints, has been used to predict the wear based on the classical Archard equation [62]. A better understanding of the mixed lubrication mechanism in different forms of artificial hip joints may provide more realistic wear predictions.
17.4.2 UHMWPE-on-Metal 17.4.2.1 Contact Mechanics The first contact mechanics analysis for artificial hip joints employing UHMWPE-on-metal material combinations was performed by Bartel et al. [63] using a simple elasticity equation based on the concept of the constrained column model. The use of the simple constrained column model was also validated by the finite element method. The importance of different design parameters such as the femoral head radius, the clearance, and the UHMWPE cup wall thickness was shown. Subsequently, a parametric analysis was carried by Jin et al. [52] who identified the regions of low contact pressures for different combinations of the main design parameters. Other factors such as the metal backing, cement, or bone have been shown to have small effects on the predicted contact parameters [36,64]. Since then, more realistic constitutive models have been proposed for UHMWPE [65,66] and have yet to be incorporated into the contact mechanics analysis for UHMWPE-on-metal hip implants. Contact mechanics has also been integrated with wear prediction in a fully coupled analysis during the gait cycle [62]. The effect of the surface topography of the UHMWPE bearing surface has been considered by Wang et al. [67]. It was shown by these authors that the real contact area was only a small fraction of the nominal contact area and consequently the real contact pressure was significantly higher than the nominal contact pressure. 17.4.2.2 Friction The experimental studies of UHMWPE-on-metal hip implants by Unsworth et al. [68] clearly indicated boundary or mixed lubrication under steady load, although a suddenly applied load was found to invoke squeeze-film lubrication action. Broadly similar conclusions were reached by O’Kelly et al. [69] under dynamic loading conditions. A more recent study by Scholes and Unsworth [54], in which a simple harmonic oscillatory motion and dynamic loading were applied in a hip function simulator, indicated that the friction factors were in the range 0.02 to 0.06 for 28 mm diameter metal heads and UHMWPE cups, broadly representative of mixed lubrication. 17.4.2.3 Lubrication Fluid-film lubrication studies in UHMWPE-on-metal hip implants have been largely based on the elastohydrodynamic mechanism associated with the relatively compliant polyethylene material. Initial attempts to predict the lubricating film thickness were made by using an equivalent ball-on-plane configuration with the equivalent radius defined in Equation (17.1), a quasi-static condition and a semiinfinite solid model, and the Hamrock and Dowson [70] film thickness formulae applied accordingly [32]. The effects of the ball-in-socket configuration and the finite thickness of the acetabular cup were subsequently examined,
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initially under quasi-static conditions [71,72] and later extended to transient walking conditions [73]. It was shown by these authors that the average of the predicted transient lubricating film thickness over one normal walking cycle was remarkably close to that estimated under the quasi-static condition based on the average load and speed. The predicted average lubricating film thickness was in the range between 0.1 and 0.2 µm (as summarized in Table 17.4 and in Section 17.4.5). Therefore, a mixed lubrication regime was predicted, since typical average surface roughness between 0.1 and 1 µm has been reported for UHMWPE bearing surfaces [74], and this is consistent with the friction measurement discussed in Section 17.4.2.2. This also demonstrates the importance of the mixed lubrication analysis of UHMWPE-on-metal hip implants [59]. 17.4.2.4 Wear The majority of simulator studies reported in the literature have focused on UHMWPE-on-metal hip implants. Although different simulators with different kinematic and loading input conditions have been adopted, the wear results obtained are relatively consistent, provided that a multidirectional motion path is produced on the UHMWPE bearing surface [31,75,76]. This finding is consistent with the mixed lubrication regime experienced in UHMWPE-on-metal hip implants deduced from the friction measurement and the theoretical predictions in Sections 17.4.2.2 and 17.4.2.3 respectively. Consequently, the wear in these joints is mainly governed by the adhesive, abrasive, and surface fatigue mechanisms associated with boundary/mixed lubrication regimes. Increases in the femoral head radius lead to almost linear increases in volumetric wear, as a result of increases in sliding distance in the boundary/mixed lubrication regime. Important boundary lubricating constituents from bovine serum and synovial fluid include lipids and proteins [77,78]. A slight increase in the phospholipid concentration has been shown to lead to a significant reduction in wear rates [79]. The effect of protein concentration has been investigated by Liao et al. [80]. The implication of these findings lies in the importance of the concentration of bovine serum used for simulator testing [81], which can range from 25 to 100%, the lower limit being recommended by ISO/FDIS/14242-1. The importance of the clearance between the femoral head and the cup has been investigated by Wang et al. [82]. It was shown by these authors that an increase in the radial clearance resulted in a decrease in the initial wear rate, opposite to the general expectation that an increase in the radial clearance would increase the contact stress and hence wear. A similar finding was reported by Barbour et al. [83] using a simple pin-on-plate wear tester that an increase in the nominal contact stress could result in a decrease in the wear factor. This highlights the complexity between wear and contact pressure for UHMWPE-on-metal hip implants. It is plausible to explain these results on the basis of microscopic asperity contacts between relatively rough UHMWPE and smooth counterface bearing surfaces [67]. A decrease in the nominal contact stress, due to an increase in the nominal contact area, can result in an increase in the real area of contact and potential wear. Recently, there has been significant interest in the study of the wear of hip implants employing crosslinked UHMWPE cups which showed a remarkable reduction in wear volume from simulator testing [84]. However, there is some controversy regarding the amount of wear reduction, depending on cross-linking, kinematics, counterface roughness, and bovine serum concentration [85]. Currently, there is still a lack of understanding of wear and lubrication mechanisms experienced with cross-linked UHMWPE materials. The early retrieved components showed significant creep and plastic deformation of one type of crosslinked UHMWPE, but the wear still seemed to be much lower than that of conventional UHMWPE [86]. Surprisingly, such surface characteristics had not been found in extensive simulator studies of the same cross-linked UHMWPE [84]. In addition to the linear and volumetric wear discussed above, extensive studies have been conducted on the morphology of the UHMWPE wear particles generated by different types of UHMWPE. There have been extensive studies of the UHMWPE particles isolated from tissues retrieved from failed total hip replacements. These studies have shown large platelet-like particles, up to 250 µm in length, fibrils, shreds and submicrometer globule-shaped spheroids between 0.1 and 0.5 µm in diameter [87–90]. The corresponding mode of the frequency distribution of particles in vivo from these studies is 0.1–0.5 µm, although the larger particles may account for a high proportion of the total volume of particles within
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TABLE 17.4 Typical Values (Range) of Femoral Head Radius, Diametral Clearance and Corresponding Predictions of Maximum Contact Pressure, Total Angle of the Contact Area, and Average Lubricating Film Thickness for Hip Implants with Various Bearing Couples Bearing couples UHMWPE-on-metal Metal-on-metal
Ceramic-on-ceramic
Specific features
Head radius (mm)
Diametral clearance (mm)
Max. contact pressure (MPa)
Total contact angle (◦ )
Average film thickness (µm)
Metal-backed cups Thick walled cup Taper-connected cup Metasul McKee–Farrar Resurfacing Resurfacing Ideal condition Microseparation
11–16 14 14 14 17.5 25 25 14 14
0.15–0.36 0.06 0.06 0.12 0.16 0.3 0.10 0.08 0.08
10–23 50 32–37 45 20 57 19.4 80 300–400
80–100 40 50 40 50 24 38 30
0.1–0.2 0.03 0.06 0.03 0.06 0.03
[72] [37] [97] [57,99] [106,143] [105,144] [145] [38] [132]
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the tissues. Analysis of wear particulates using frequency distributions as a function of size is not adequate to differentiate between the particles generated under different conditions [48,91,92]. Analysis of the mass distribution as a function of size has been shown to discriminate between samples [48,91,92]. Studies of UHMWPE wear particles generated in vitro in hip joint simulators have shown that there is a larger proportion of the mass of particles generated in the 0.1–1 µm sized range than that isolated from periprosthetic tissues [93,94]. This may indicate that in vivo, the smaller particles are disseminated more widely away from the implant site. Recently, improvements to particle imaging techniques have shown nanometer sized polyethylene particles in in vitro wear simulations. These nanometer sized particles account for the greatest number of particles generated, but they account for a very small proportion of the total volume [95]. There is evidence that some cross-linked polyethylenes may give rise to an increased proportion, per unit volume of wear, of wear particles in the 0.1–1.0 µm size range, which as noted in Section 17.3.5 is the critical size range for macrophage activation [96]. However, the reduction in total wear volume achieved with cross-linked polyethylenes may lead to a reduction in the volumetric concentration particles in the 0.1–1 µm size range generated per unit time in vivo.
17.4.3 Metal-on-Metal 17.4.3.1 Contact Mechanics Finite element method is usually adopted to study the contact mechanics in metal-on-metal hip implants. Unlike hip implants with UHMWPE-on-metal material combinations, the deformation of both the femoral head and acetabular cup have to be considered as well as cement, bone, and other support materials. It has been shown that the design parameters such as the clearance and the head diameter have a large effect on the predicted contact parameters [32]. Furthermore, the wall thickness and structure of the metallic acetabular cup such as taper-connected and sandwiched UHMWPE can also influence the contact mechanics at the bearing surfaces [97–99]. Despite a wide range of parameters and conditions, the maximum contact pressure in some metal-on-metal hip implants are in the range between 20 and 60 MPa as summarized in Table 17.4. 17.4.3.2 Friction For metal-on-metal hip implants tested in a friction hip simulator, a mixed lubrication regime has been shown to be dominant by Scholes and Unsworth [54] and Williams et al. [100]. A similar conclusion was found more recently when a number of large-sized metal-on-metal hip resurfacing prostheses with a diameter of 50 mm and various radial clearances between 20 and 190 µm were tested [101,102]. However, the friction testing of the worn components after the running-in period has shown the potential for fluid-film lubrication by Unsworth et al. [103]. Although these authors attributed this observation to the microscopic modification of the metallic bearing surface, the macroscopic change of the bearing geometry and effectively reduced radial clearance may be more important [104]. 17.4.3.3 Lubrication The modes of lubrication in metal-on-metal hip implants have been found to be complex and quite sensitive to design and manufacturing parameters, and kinematic and loading conditions. As a result, various lubrication regimes are possible, ranging from boundary, mixed, to fluid-film. The simple resistivity technique discussed in Section 17.4.1 has proved particularly useful for metal-onmetal hip implants. The direct measurement of the separation between the two metallic bearing surfaces tested in a hip simulator under normal walking conditions was carried out by Dowson et al. [55]. It was shown by these authors that the mode of lubrication was mixed, with some periods of very effective surface separation and others of metal-to-metal contact in each simulated walking cycle. However, this valuable experimental technique does not, with any certainty indicate whether the lubricating film is hydrodynamic or boundary in nature due to protein films. Furthermore, it should be noted that under no conditions was the joint fully separated throughout the whole cycle and contact always occurred; hence wear was inevitable.
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A general numerical methodology for the steady-state elastohydrodynamic lubrication (EHL) analysis of hip implants based on the combination of the finite element and finite different method was developed by Jagatia and Jin [37]. The general methodology has since been applied to examine EHL in various metal-on-metal hip implants, including a thick walled cup [37]; resurfacing [105]; Mckee–Farrar [106]; and sandwiched cup [57]. The major findings from these studies have confirmed the importance of clearance and head size, and also pointed out the importance of the cup wall thickness and structural support. The effect of transient loading and motion associated with walking on EHL of metal-on-metal hip implants was investigated by Chan et al. [107] based on a simple linear superposition technique. It was concluded by these authors that the variation in the transient lubricating film thickness was small, the average value being similar to that estimated from steady-state analysis using the average load and speed. However, a more complete transient EHL analysis by Williams et al. [100] revealed a much larger variation in the predicted transient lubricating film thickness, particularly in the stance phase when the swing phase load was reduced. Further transient EHL analysis was performed by Jalali-Vahid et al. [108] under start-up and stopping conditions. It was shown by these authors that the time taken to establish a steady cyclic solution after start-up was rather short, of the order of a few seconds, while the time taken for the lubricating film thickness to reduce to the magnitude of typical metallic surface roughness after stopping was much longer. 17.4.3.4 Wear The complex lubrication mechanism in metal-on-metal hip implants discussed in Section 17.4.3.2 leads to complex wear. The importance of the lubrication regime on wear volume was first demonstrated by Chan et al. [109]. The volumetric wear was shown to correlate fairly well with the λ ratio defined in Equation (17.2). An increase in the λ ratio led to an improvement in the lubrication regime and hence a significant reduction in the wear volume. However, even for a given λ ratio, particularly in the mixed lubrication regime, there is usually a large scatter in the wear rates, highlighting the complexity of the lubrication and wear mechanisms involved. Furthermore, the use of the average surface roughness in the calculation of the λ ratio is questionable, particularly for the topography of the metallic bearing surface which predominantly consists of valleys [110]. Nevertheless, there appears to be an optimum range of the clearance, depending on the bearing system. If the clearance is too tight, the equatorial and edge contact between the two bearing surfaces may occur, not only elevating the contact stress but also restricting the lubricant entry and causing lubricant starvation. If the clearance is too large, the contact stress may be significantly increased and the mode of lubrication may be shifted toward the boundary lubrication regime. In both cases, wear can increase significantly [111]. The optimum radial clearance was found to be about 30 µm for the metal-on-metal hip implant with 28 mm diameter heads (Ultima®, made by DePuy Orthopaedics) considered by Farrar and Schmidt [111]. However, Muller [112] considered the optimal radial clearance to be around 75 µm for Metasul® metal-on-metal hip implants of 28 mm diameter heads made originally by Sulzer. This is consistent with the contact mechanics study of the Metasul as discussed in Section 17.4.3.1 that the contact pressure experienced at the bearing surfaces was reduced significantly by the UHMWPE backing. Therefore, the clearance should be optimized for a given bearing system. Furthermore, manufacturing errors such as nonsphericity may compound the effect of the clearance. Scholes et al. [113] showed that a reduction in the radial clearance from 40 to 22 µm for a fixed femoral head radius of 14 mm resulted in a smaller wear rate, but the difference was not found to be significant. It should be noted that the wear in metal-on-metal hip implants has been shown to exhibit biphasic characteristics, a rapid initial running-in phase followed by a gentle steady-state phase [114]. Consequently, the effective clearance after the running-in period and the corresponding lubrication may be quite different [115]. The femoral head radius is generally expected to have a large effect on lubrication and wear, since it is not only related to the equivalent radius as defined in Equation (17.1) and entrainment velocity associated with fluid-film lubrication, but also the sliding distance. This has been confirmed by Smith et al. [116,117] who showed that an increase in the femoral head diameter from 16 to 22.225 mm
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resulted in an increase in the wear rate, consistent with the boundary lubrication regime. However, a further increase in the femoral head radius beyond 22.225 mm led to significant wear reduction, presumably due to improved fluid-film lubrication and reduced asperity contacts [118]. The benefit of large femoral heads in wear reduction has been further demonstrated in a study of metal-on-metal hip resurfacing prostheses [119]. However, it should be pointed out that it is the combination of both the clearance and the head diameter that is required to achieve an equivalent radius as defined in Equation (17.1) and an effective fluid-film lubrication. A too large clearance may compromise lubrication, and adversely increase runningin wear due to the increased sliding distance associated with the large femoral head size as demonstrated recently by Rieker et al. [120]. In addition to the design parameters of the bearing surfaces, the effect of kinematic and loading conditions can also influence lubrication and wear of metal-on-metal hip implants. Firkins et al. [121] have shown that in the simulator with two input motions which produced an open elliptical wear path with greater eccentricity, the wear rate was at least ten times higher than that in the simulator with three independent input motions which produced a low level of eccentricity. However, the detailed variation of the motion cycles has been found to have a small effect on wear [122]. A decrease in the swing phase load has been found to lead to a large wear reduction [123], consistent with the fluid-film lubrication analysis discussed in Section 17.4.3.2. A protocol incorporating start-up and stopping has been proposed by Roter et al. [124]. Other adverse conditions such as microseparation [125] and stumbling [119] have also been shown to increase wear. The metallic wear particles generated at the bearing surfaces are invariably in the nanometer size range, from 20 to 110 nm. They are uniformly round to oval in size [126]. Studies of the cellular response to these nanometer size metal particles have shown that they have low osteolytic potential compared with UHMWPE particles [127]; however, they are toxic to cells at relatively low volumetric concentrations [128] and there are concerns regarding the long-term effects of these metal particles in patients, since due to their small size they will disseminate from the implant site [129,130].
17.4.4 Ceramic-on-Ceramic 17.4.4.1 Contact Mechanics Limited contact mechanics studies have been conducted for ceramic-on-ceramic hip implants, since the major focus has been on the taper connection between the ceramic head and the metallic stem [131]. There are basically two forms of current ceramic cup designs; sandwiched or taper connected. Since the wall of the ceramic cup should be sufficiently thick in order to minimize potential tensile stresses due to bending, Hertz contact theory has been found to be adequate for the prediction of the contact parameters at the bearing surfaces [38]. However, under microseparation conditions, the edge of the cup comes into contact with the head and a significant increase in the contact pressure has been predicted [132]. 17.4.4.2 Friction For ceramic-on-ceramic hip implants, a very small friction factor can be achieved with a synthetic lubricant of similar viscosity to that of synovial fluid, indicative of fluid-film lubrication [54]. However, the friction factor can be significantly affected by the proteins contained within biological lubricants such as bovine serum, indicating that a boundary lubrication mechanism is dominating. 17.4.4.3 Lubrication The resistivity technique applied for metal-on-metal hip implants has also been used to measure the separation in ceramic-on-ceramic hip implants with appropriate conducting coatings such as titanium nitride [133]. Surface separation of the femoral head and acetabular cup by a lubricating film has been demonstrated by these authors. But once again, it is not clear whether such a separation is due to the hydrodynamic fluid-film or the boundary proteins. Under identical operating conditions and for the same geometrical parameters, the predicted lubricating film thickness in ceramic-on-ceramic hip implants should be slightly reduced compared with metal-on-metal bearings due to a higher elastic modulus.
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Coupled with the recent improvement in the surface finish of metallic bearing surfaces, the predicted lubrication regime in ceramic-on-ceramic hip implants may not be much better than that in metal-onmetal bearings. This is particularly true since the clearance used in the majority of current ceramic-onceramic hip implants is not much different from that in metal-on-metal hip implants [38]. Full fluid-film lubrication appears to be possible in ceramic-on-ceramic hip implants under most conditions, but only for small clearances and smooth bearing surfaces. 17.4.4.4 Wear The biphasic wear observed in metal-on-metal bearings has also been found in ceramic-on-ceramic hip implants [134]. This therefore suggests that different lubrication status exist between the running-in and steady-state periods. Under normal conditions, extremely low wear rates have been found in ceramic-onceramic hip implants [135,136]. However, this is not always reflected in the clinical wear rates and the pattern observed on the retrieved components. Microseparation has been introduced into simulator testing and a significant increase in wear has been demonstrated [137]. Thus improved material microstructures are still important [138]. Ceramic wear debris produced in vivo has been shown to have a bimodal size distribution [139]. Particles in the nanometer size range (10–20 nm) are the predominant type found, however, larger particles in the 0.1–10 µm size range are generated [139]. Studies of the particles generated in hip simulations have shown that, under standard wear conditions, only the nanometer sized particles are generated. When microseparation is introduced into hip simulations, the larger-sized particles are also produced [140]. Thus microseparation not only reproduces clinical wear rates for ceramic-on-ceramic bearings but also reproduces the clinically relevant particle morphologies [140]. Studies of the biological response to ceramic wear particles, produced under microseparation conditions have shown that the particles are capable of inducing the release of osteolytic cytokines by macrophages in vitro, but only at volumetric concentrations that are unlikely to be generated clinically in the absence of catastrophic wear [141]. Moreover, ceramic particles are relatively bio-inert and, in contrast to metal particles, they have very low cytotoxic potential [128]. Tables 17.4 to 17.7 summarize the maximum contact pressure, typical friction coefficient, lubricating film thickness, wear, and biological responses in various bearing couples for hip implants discussed in Section 17.4.
17.4.5 Other Bearing Materials and Combinations In addition to the hip implants discussed in Sections 17.4.2 to 17.4.4 which are in current clinical use, other bearing materials and couples are also being researched extensively. Cushion form bearings using compliant materials, such as polyurethane and hydrogel, have been investigated to improve fluid-film lubrication in artificial hip joints [41,42,146–150]. The wear rate in ceramic-on-metal combination has shown to be even lower than that in metal-on-metal, due to the differential hardness between the metallic and ceramic materials [151]. Surface engineered coatings have been introduced into metal-on-metal bearings, with a significant wear reduction [152]. New ceramic materials have been developed, by combining both alumina and zirconia, in order to improve both toughness and wear performance [153].
17.5 Summary It is clear from tribological studies of contact mechanics, friction, wear, and lubrication discussed in this chapter that different lubrication regimes exist in different forms of hip implants, which can have a strong influence on the wear and the generation of wear particles from the bearing surfaces. Wear debris is now widely recognized as a major factor in clinical problems such as osteolysis and the loosening of implants. Fluid-film lubrication is the most effective method of reducing wear and wear particle generation in most hip bearings. It should be noted, however, that even if a continuous fluid-film is present, wear can still take place due to erosion and fatigue. More importantly, adverse conditions such
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Handbook of Lubrication and Tribology TABLE 17.5 Measurement of Friction Factors in a Pendulum Simulator and Corresponding Assessment of Lubrication Regimes [54,146]
Bearing couples
Friction factor
Cushion form bearing UHMWPE-on-metal Metal-on-metal Ceramic-on-ceramic
0.002–0.01 0.06–0.08 0.22–0.27 0.002–0.2
Variation of friction factor against increasing Sommerfeld number
Indicated lubrication regimes
Increasing Constant/decreasing Decreasing Increasing
Fluid-film Boundary/mixed Mixed Fluid-film/mixed
TABLE 17.6 Typical Volumetric and Linear Wear Rates for Various Hip Implants Volumetric wear rate (mm3 /yr)
Bearing couples UHMWPE-on-metal UHMWPE-on-ceramic Metal-on-metal Ceramic-on-ceramic
30–100 15–50 0.1–1 0.05–1
Linear wear rate (µm/yr) 100–300 50–150 2–20 1–20
One year was assumed to be equal to 1 million cycles.
TABLE 17.7 Implants
Typical Wear Debris Sizes and Biological Responses for Various Hip
Bearing couples UHMWPE-on-metal/ceramic Metal-on-metal Ceramic-on-ceramic
Dominant particle diameters (µm) UHMWPE, 0.1 to 1 Metallic, 0.02 to 0.1 Ceramic,a 0.01 to 0.02 Ceramic,b 0.1 to 10
Biological responses Macrophages/osteoclasts/osteolysis Low osteolysis, toxic Bio-inert, low cytotoxic Macrophages/osteoclasts/osteolysis
a Under standard wear testing condition. b Under microseparation condition.
as start-up, stopping, and microseparation can lead to a significant increase in wear, particularly for hip implants with hard-on-hard bearing couples. The inherent wear resistance of the bearing couples associated with optimal microstructures and surface coatings may play a more important role in implant performance. For UHMWPE-on-metal/ceramic hip implants, the mode of lubrication is predominantly boundary, or perhaps mixed under favorable conditions. Changes in the design parameters such as the radial clearance and the femoral head radius are unlikely to change the mode of lubrication and thus the only remaining lubrication mechanism might involve boundary additives. Further understanding of the wear mechanism associated with the cross-linked UHMWPE should play an important role in its widespread use in different joints and forms. The wear and wear debris generation in metal-on-metal hip implants are highly lubrication-dependent. The mode of lubrication ranges from boundary to fluid-film including the mixed regime, and the lubrication regime experienced in a particular bearing system depends critically on the design parameters such as the radial clearance, the femoral head radius, and the bearing structure, as well as the manufacturing of the bearing surfaces in terms of both the sphericity and the smoothness. Design optimization of these
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parameters in terms of fluid-film lubrication may offer further opportunities to reduce wear and the generation of wear particles, particularly during the running-in period. Ceramic-on-ceramic hip implants appear to have the best chance to benefit from fluid-film lubrication, but only with small clearances. Surface parameters such as hardness, microstructure, and roughness are still important in order to reduce wear under adverse conditions such as microseparation. Further improvement of ceramic-on-ceramic bearings in terms of reducing occasional shock loading within the material or through introducing a soft-mounting system might prove beneficial. The topography of the bearing surfaces plays an important role in lubrication and wear in most types of hip implant. Lubrication analysis that directly includes surface roughness will be important for both conventional and cross-linked UHMWPE cups against either metallic or ceramic femoral head and metal-on-metal hip implants. In particular for metal-on-metal hip implants, consideration of the surface topography (particularly the valleys) may suggest possibilities for nano-sculptured surfaces to improve lubrication and reduce wear. Further structural optimization of the bearing surface mounting in acetabular cups to achieve continuous fluid-film lubrication may offer opportunities to reduce wear to even lower levels in hard-on-hard bearings. A full transient EHL analysis for a ball-in-socket configuration will be required for this purpose. It should be pointed out that the ultimate goal of contact mechanics and fluid-film lubrication analyses or friction measurements in hip implants is to provide design guidelines in order to reduce wear and wear particle generation. Further studies should aim at integrating both the analysis of fluid-film and boundary lubrication with contact mechanics to ultimately develop theoretical wear models. Such theoretical models would offer, in conjunction with simulator testing and other experimental approaches, a screening tool for future development of new bearing couples, thus reducing the time for preclinical testing along with costs and improving patient safety [154]. In addition to the development of methodologies for tribological analyses of artificial hip joints, new forms of bearing couples discussed in Section 17.4.5 and biological surface replacements using minimum invasive approaches may become available. This may introduce new tribological challenges as well as the interaction between the tribology of the bearing surfaces and the fixation of the prosthetic components may become important. Tribology will continue to play an important role in optimizing the bearing couple and ensuring the long-term clinical success of artificial hip joints.
References [1] Dowson, D. and Wright, V., Bio-tribology, in Proceedings of the Conference on The Rheology of Lubrication, The Institute of Petroleum, The Institution of Mechanical Engineers and the British Society of Rheology, 1973, 81–88. [2] Litonjua, L.A. et al., Tooth wear: attrition, erosion, and abrasion, Quintessence Int., 34, 435, 2003. [3] Turssi, C.P. et al., Wear of dental resin composites: insights into underlying processes and assessment methods — a review, J. Biomed. Mater. Res., 65B, 280, 2003. [4] Dowson, D., Tribology and the skin structure, in Bioengineering of Skin: Methods and Instrumentation Part III: General Aspects, E. Berardesca et al., Eds, CRC Press, Boca Raton, FL, 1995, 159. [5] Sivamani, R.K. et al., Coefficient of friction: tribological studies in man — an overview, Skin Res. Technol., 9, 227, 2003. [6] Gronqvist, R. et al., Measurement of slipperiness: fundamental concepts and definitions, Ergonomics, 44, 1102, 2001. [7] Maynard, W.S., Tribology: preventing slips and falls in the workplace, Occup. Health Safety, 71, 134, 2002. [8] Holly, F.J. and Holly, T.F., Advances in ocular tribology, Adv. Exp. Med. Biol., 350, 275, 1994. [9] Secomb, T.W., Hsu, R., and Pries, A.R., Blood flow and red blood cell deformation in nonuniform capillaries: effects of the endothelial surface layer, Microcirculation, 9, 189, 2002.
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[33] Saarri, H. et al., Hyaluronate in total hip replacement, J. Rheumatol., 20, 87, 1993. [34] Delecrin, J. et al., Changes in joint fluid after total arthroplasty, Clin. Orthopaed. Relat. Res., 307, 240, 1994. [35] Cooke, A.V., Dowson, D., and Wright, V., The rheology of synovial fluid and some potential synthetic lubricants for degenerate synovial joints, Eng. Med., 7, 66, 1978. [36] Jin, Z.M. et al., An axisymmetric contact model of ultra high molecular weight polyethylene cups against metallic femoral heads for artificial hip joint replacements, Proc. Instn. Mech. Engrs., J. Eng. Med., 213, 317, 1999. [37] Jagatia, M. and Jin, Z.M., Elastohydrodynamic lubrication of metal-on-metal hip prosthesis under steady-state entraining motion, Proc. Instn. Mech. Engrs., J. Eng. Med., 215, 531, 2001. [38] Mak, M.M. and Jin, Z.M., Analysis of contact mechanics in ceramic-on-ceramic hip joint replacements, Proc. Instn. Mech. Engrs., J. Eng. Med., 216, 231, 2002. [39] Yao, J.Q. et al., The influence of lubricant and material on polymer/CoCr sliding friction, Wear, 255, 780, 2003. [40] Ratner, B.D. et al., Biomaterials Science, Academic Press, London 1996. [41] Dowson, D. et al., Design consideration of cushion form bearings in artificial hip joints, Proc. Instn. Mech. Engrs., J. Eng. Med., 205, 59, 1991. [42] Jennings, L.M. and Fisher, J., A biomechanical and tribological investigation of a novel compliant all polyurethane acetabular resurfacing system, in Proceedings of the International Conference of Engineers and Surgeons Joined at the Hip, IMechE, 2002, C601/032. [43] Muratoglu, O.K. et al., Larger diameter femoral heads used in conjunction with a highly cross-linked ultra-high molecular weight polyethylene — a new concept, J. Arthroplasty, 16, 24, 2001. [44] Huiskes, R., Stress shielding and bone resorption in THA: clinical versus computer-simulation studies, Acta Orthop. Belg., 59, 118, 1993. [45] Ingham, E. and Fisher, J., The role of macrophages in osteolysis of total joint replacements, Biomaterials, 26, 1271–1286, 2005. [46] Hirakawa, K. et al., Comparison and quantitation of wear debris of failed total hip and knee arthroplasty, J. Biomed. Mater. Res., 31, 257, 1998. [47] Revell, P.A., Biological reaction to debris in relation to joint prostheses, Proc. Instn. Mech. Engrs., J. Eng. Med., 211, 187, 1997. [48] Tipper, J.L. et al., Quantitative comparison of polyethylene wear debris, wear rate and head damage in retrieved Charnley hip prostheses, J. Mater. Sci. Mater. Med., 11, 117, 2000. [49] Koh, J.L. et al., The effect of graft height mismatch on contact pressure following osteochondral grafting: a biomechanical study, Am. J. Sports Med., 32, 317, 2004. [50] Wilson, D.R. et al., Accuracy and repeatability of a pressure measurement system in the patellofemoral joint, J. Biomech., 36, 1909, 2003. [51] Müller, O. et al., Three-dimensional measurements of the pressure distribution in artificial joints with a capacitive sensor array, J. Biomech., 37, 1623, 2004. [52] Jin, Z.M., Dowson, D., and Fisher, J., A parametric analysis of the contact stress in UHMWPE acetabular cups, Med. Eng. Phys., 16, 398, 1994. [53] Hall, R.M. and Unsworth, A., Friction in hip prostheses, Biomaterials, 18, 1017, 1997. [54] Scholes, S.C. and Unsworth, A., Comparison of friction and lubrication of different hip prostheses, Proc. Instn. Mech. Engrs., J. Eng. Med., 214, 49–57, 2000. [55] Dowson, D., McNie, C.M., and Goldsmith, A.A.J., Direct experimental evidence of lubrication in a metal-on-metal total hip replacement tested in a joint simulator, Proc. Instn. Mech. Engrs., J. Mech. Eng. Sci., 214, 75, 2000. [56] Wang, F.C. and Jin, Z.M., Prediction of elastic deformation of acetabular cup and femoral head for lubrication analysis of artificial hip joints, Proc. Instn. Mech. Engrs., J. Eng. Tribol., 218, 201, 2004.
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[57] Liu, F. et al., Elastohydrodynamic lubrication analysis of a metal-on-metal hip implant employing a metallic cup with an UHMWPE backing under steady-state conditions, Proc. Instn. Mech. Engrs., J. Eng. Med., 218, 261, 2004. [58] Liu, F. et al., Transient elastohydrodynamic lubrication analysis of a metal-on-metal hip implant under simulator-tested conditions, in Computational Mechanics, WCCM VI in conjunction with APCOM ’04, Sept. 5–10, 2004, Beijing, China, Tsinghua University Press & Springer-Verlag. [59] Wang, F.C. and Jin, Z.M., Lubrication modelling of artificial hip joints: from fluid film to boundary lubrication regimes, in Proceedings of ESDA, 7th Biennial Conference on Engineering Systems Design and Analysis, Manchester, UK, 2004. [60] Saikko, V., A multidirectional motion pin-on-disk wear test method for prosthetic joint materials, J. Biomed. Mater. Res., 41, 58, 1998. [61] Dowson, D. and Jobbins, B.J., Design and development of a versatile hip joint simulator and a preliminary assessment of wear and creep in Charnley total replacement hip joints, Eng. Med., 17, 111, 1988. [62] Maxian, T.A. et al., The Frank Stinchfield Award. 3-Dimensional sliding/contact computational simulation of total hip wear, Clin. Orthopaed. Relat. Res., 333, 41, 1996. [63] Bartel, D.L. et al., The effect of conformity and plastic thickness on contact stresses in metal-backed plastic implants, J. Biomech. Eng., 107, 193, 1985. [64] Kurtz, S.M., Edidin, A.A., and Bartel, D.L., The role of backside polishing, cup angle, and polyethylene thickness on the contact stresses in metal-backed acetabular components, J. Biomech., 30, 639, 1997. [65] Lee, K.Y., Pienkowski, D., and Lee, S., Dynamic compressive creep of extruded ultra-high molecular weight polyethylene, KSME Int. J., 17, 1332, 2003. [66] Bergstrom, J.S., Rimnac, C.M., and Kurtz, S.M., An augmented hybrid constitutive model for simulation of unloading and cyclic loading behavior of conventional and highly crosslinked UHMWPE, Biomaterials, 25, 2171, 2004. [67] Wang, F.C. et al., Microscopic asperity contact and deformation of UHMWPE bearing surfaces, Proc. Instn. Mech. Engrs., J. Eng. Med., 217, 477, 2003. [68] Unsworth, A. et al., The frictional behaviour of human synovial joints — part II. Artificial joints, Trans. ASME, J. Lubr. Technol., 97, 377, 1975. [69] O’Kelly, J. et al., Pendulum and simulator for studies of friction in hip joints, in Evaluation of Artificial Joints, D. Dowson and V. Wright, Eds, The Biological Engineering Society, 19, 1977. [70] Hamrock, B.J. and Dowson, D., Elastohydrodynamic lubrication of elliptical contacts for materials of low elastic modulus. I: fully flooded conjunction, Trans. ASME, J. Lubr. Technol., 100, 236, 1978. [71] Jalali-Vahid, D. et al., Elastohydrodynamic lubrication analysis of UHMWPE hip joint replacements, in Proceedings of 26th Leeds-Lyon Symposium on Tribology, D. Dowson et al., Eds, Leeds, 2000, 329. [72] Jalali-Vahid, D. et al., Prediction of lubrication film thickness in UHMWPE hip joint replacements, J. Biomech., 34, 261, 2001. [73] Jalali-Vahid, D. and Jin, Z.M., Transient elastohydrodynamic lubrication analysis of UHMWPE hip joint replacements, Proc. Instn. Mech. Engrs., J. Mech. Eng. Sci., 216, 409, 2002. [74] Elfick, A.P. et al., Surface topography of retrieved PCA acetabular liners: proposal for a novel wear mechanism, J. Mater. Sci. Lett., 17, 1085, 1998. [75] Smith, S.L. and Unsworth, A., Simplified motion and loading compared to physiological motion and loading in a hip joint simulator, Proc. Instn. Mech. Engrs., J. Eng. Med., 214, 233, 2000. [76] Kaddick, C. and Wimmer, M.A., Hip simulator wear testing according to the newly introduced standard ISO 14242, Proc. Instn. Mech. Engrs., J. Eng. Med., 215, 429, 2001. [77] Hills, B.A., Boundary lubrication in vivo, Proc. Instn. Mech. Engrs., J. Eng. Med., 214, 83, 2000. [78] Purbach, B., Hills, B.A., and Wroblewski, B.M., Surface-active phospholipid in total hip arthroplasty, Clin. Orthopaed. Relat. Res., 396, 115, 2002.
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Biotribology
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[79] Bell, J. et al., The influence of phospholipid concentration in protein-containing lubricants on the wear of ultra-high molecular weight polyethylene in artificial hip joints, Proc. Instn. Mech. Engrs., J. Eng. Med., 215, 259, 2001. [80] Liao, Y.S., Benya, P.D., and McKellop, H.A., Effect of protein lubrication on the wear properties of materials for prosthetic joints, J. Biomed. Mater. Res. (Appl. Biomater.), 48, 465, 1999. [81] Wang, A., Essner, A. and Schmidig, G., The effects of lubricant composition on in vitro wear testing of polymeric acetabular components, J. Biomed. Mater. Res. Part B — Appl. Biomater., 68B, 45, 2004. [82] Wang, A., Essner, A. and Klein, R., Effect of contact stress on friction and wear of ultra-high molecular weight polyethylene in total hip replacement, Proc. Instn. Mech. Engrs., J. Eng. Med., 215, 133–139, 2001. [83] Barbour, P.S.M., Barton, D.C., and Fisher, J., The influence of contact stress on the wear of UHMWPE for total replacement hip prostheses, Wear, 181, 250, 1995. [84] Muratoglu, O.K. et al., Unified wear model for highly crosslinked ultra-high molecular weight polyethylene (UHMWPE), J. Biomater., 20, 1463, 1999. [85] Galvin, A. et al., Reduction in wear of crosslinked polyethylene under different tribological conditions, in Proceedings of the International Conference of Engineers and Surgeons Joined at the Hip, IMechE, 2002, C601/005. [86] Muratoglu, O.K. et al., Surface analysis of early retrieved acetabular polyethylene liners: a comparison of standard and highly crossslinked polyethylenes, in Proceedings of Transactions of the 48th Annual Meeting of the ORS, 2002, 1029. [87] Campbell, P. et al., Isolation of predominantly sub-micron sized UHMWPE wear particles from periprosthetic tissues, J. Biomed. Mater. Res., 29, 127, 1995. [88] Maloney, W.J. et al., Isolation and characterisation of wear debris generated in patients who have had failure of a hip arthroplasty without cement. J. Bone Joint Surg., 77A, 1301, 1994. [89] Margevicius, K.T. et al., Isolation and characterisation of debris from around total joint prostheses, J. Bone Joint Surg., 76A, 1664, 1994. [90] Shanbhag, A.S. et al., Composition and morphology of wear debris in failed uncemented total hip replacement, J. Bone Joint Surg., 76B, 60, 1994. [91] Howling, G.I. et al., Quantitative characterisation of polyethylene debris isolated from periprosthetic tissue in early failure knee implants and early and late failure Charnley hip implants, J. Biomed. Mater. Res. Appl. Biomater., 58, 415, 2001. [92] Bell, J. et al., Quantitative analysis of UHMWPE wear debris isolated from the periprosthetic femoral tissues from a series of Charnley total hip arthroplasties. Bio-Med. Mater. Eng., 12, 189, 2002. [93] Endo, M. et al., Comparison of wear, wear debris and functional biological activity of moderately crosslinked and non-crosslinked polyethylenes in hip prostheses, Proc. Instn. Mech. Engrs., J. Eng. Med., 216, 111, 2002. [94] Tipper, J.L. et al., Characterisation of wear debris from UHMWPE on zirconia ceramic, metalon-metal and alumina ceramic-on-ceramic hip prostheses generated in a physiological anatomical hip joint simulator, Wear, 250, 120, 2001. [95] Galvin, A.L. et al., Nanometre size polyethylene particles from hip and knee simulator studies, in Proceedings of the 50th Annual Meeting of the Orthopaedic Research Society, San Francisco, USA, 2004, 1502. [96] Ingram, J.H. et al., The influence of molecular weight, crosslinking and counterface roughness on TNF-alpha production by macrophages in response to ultra high molecular weight polyethylene particles, Biomaterials, 17, 3511, 2004. [97] Besong, A.A. et al., Contact mechanics of a novel metal-on-metal total hip replacement, Proc. Instn. Mech. Engrs., J. Eng. Med., 215, 543, 2001. [98] Verdonschot, N. et al., Effects of metal-inlay thickness in polyethylene cups with metal-on-metal bearings, Clin. Orthop., 404, 353, 2002.
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[99] Liu, F. et al., Contact mechanics of metal-on-metal hip implants employing a metallic cup with an UHMWPE backing, Proc. Instn. Mech. Engrs., J. Eng. Med., 217, 207, 2003. [100] Williams, S. et al., Effect of swing phase load conditions on the wear, friction and lubrication in metal-on-metal hips, in Proceedings of the 50th Annual Meeting of the Orthopaedic Research Society, 2004, 1504. [101] Kampen, M., Scholes, S.C., and Unsworth, A., The lubrication regime in a metal-on-metal total hip replacement, in Proceedings of the International Conference of Engineers and Surgeons Joined at the Hip, IMechE, 2002, C601/035. [102] Udofia, I.J., Tribology of Metal-On-Metal Hip Resurfacing Prostheses, PhD Dissertation, University of Bradford, 2003. [103] Unsworth, A. et al., Fluid film lubrication of metal-on-metal hip joints — fact or fiction, in Proceedings of the 16th Annual Symposium of the International Society for Technology in Arthroplasty, (ISTA 2003), San Francisco, 2003, 150. [104] Hu, X.Q., Isaac, G.H., and Fisher, J., Changes in the contact area during the bedding-in wear of different sizes of metal on metal hip prostheses, Bio-Med. Mater. Eng., 14, 145, 2004. [105] Udofia, I.J. and Jin, Z.M., Elastohydrodynamic lubrication analysis of metal-on-metal hip resurfacing prostheses, J. Biomech., 36, 537, 2003. [106] Yew, A. et al., Analysis of elastohydrodynamic lubrication in Mckee-Farrar metal-on-metal hip joint replacements, Proc. Instn. Mech. Engrs., J. Eng. Med., 218, 27, 2004. [107] Chan, F.W. et al., Numerical analysis of time-varying fluid film thickness in metal-metal hip implants in simulator tests, in Alternative Bearing Surfaces in Total Joint Replacement, J.J. Jacobs and T.L. Craig, Eds, ASTM STP 1346, ASTM, West Conshohocken, PA, USA, 1998, 111. [108] Jalali-Vahid, D., Jin, Z.M., and Dowson, D., Elastohydrodynamic lubrication analysis of metalon-metal hip implants under start-up and stopping conditions, in Transient Processes in Tribology, G. Dalmaz et al., Eds, Elsevier, Amsterdam, 2004, 751. [109] Chan, F.W. et al., The Otto Aufranc Award — wear and lubrication of metal-on-metal hip implants, Clin. Orthopaed. Relat. Res., 369, 10, 1999. [110] Dowson, D., Lubrication regimes in total hip replacements, in Proceedings of the International Conference of Engineers and Surgeons Joined at the Hip, IMechE, 2002, C601/051. [111] Farrar, R. and Schmidt, M.B., The effect of diametral clearance on wear between head and cup for metal-on-metal articulations, in Proceedings of the 43rd Orthopaedic Research Society, San Francisco, 1997, 71. [112] Muller, M.E., The benefits of metal-on-metal total hip replacements, Clin. Orthopaed. Relat. Res., 311, 54, 1995. [113] Scholes, S.C., Green, S.M., and Unsworth, A., The wear of metal-on-metal total hip prostheses measured in a hip simulator, Proc. Instn. Mech. Engrs., J. Eng. Med., 215, 523, 2001. [114] Rieker, C., Konrad, R., and Schon, R., In vitro comparison of the two hard–hard articulations for total hip replacements, Proc. Instn. Mech. Engrs., J. Eng. Med., 215, 153, 2001. [115] Liu, F. et al., Running-in wear and lubrication of metal-on-metal hip implants, in Proceedings of BORS, Spring 2004, Bristol. [116] Smith, S.L., Dowson, D., and Goldsmith, A.A.J., The lubrication of metal-on-metal total hip joints: a slide down the Stribeck curve, Proc. Instn. Mech. Engrs., J. Eng. Tribology, 215, 483, 2001. [117] Smith, S.L., Dowson, D., and Goldsmith, A.A.J., The effect of femoral head diameter upon lubrication and wear of metal-on-metal total hip replacements, Proc. Instn. Mech. Engrs., J. Eng. Med., 215, 161, 2001. [118] Jin, Z.M., Analysis of mixed lubrication in metal-on-metal hip joint replacements, Proc. Instn. Mech. Engrs., J. Eng. Med., 216, 85, 2002. [119] Bowsher, J.G. et al., Hip simulator testing — the next generation? in Proceedings of the International Conference of Engineers and Surgeons Joined at the Hip, IMechE, 2002, C601/021.
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Biotribology
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[120] Rieker, C.B. et al., In vitro tribology of large metal-on-metal implants, in Proceedings of the 50th Transactions of Orthopaedic Research Society, 2004, 0123. [121] Firkins, P.J. et al., Influence of simulator kinematics on the wear of metal-on-metal hip prostheses, Proc. Instn. Mech. Engrs., J. Eng. Med., 215, 119, 2001. [122] Smith, S.L., Dowson, D., and Goldsmith, A.A.J., The effect of diametral clearance, motion and loading cycles upon lubrication of metal-on-metal total hip replacements, Proc. Instn. Mech. Engrs., J. Mech. Eng. Sci., 215, 1, 2001. [123] Williams, S. et al., Metal-on-metal bearing wear with different swing phase loads, J. Biomed. Mater. Res., 70B, 233, 2004. [124] Roter, G.E. et al., Intermittent motion: a novel simulator protocol for the wear of metal-metal hip implants, in Tribology Series 40, D. Dowson et al., Eds, Elsevier, Amsterdam, 2002, 367. [125] Fisher, J. et al., Functional biological activity and osteolytic potential of wear debris generated in artificial hip joints, in Proceedings of the International Conference of Engineers and Surgeons Joined at the Hip, IMechE, 2002, C601/011. [126] Brown, C. et al., Use of high resolution microscopy to characterise wear debris produced by metalon-metal hip simulations, in Proceedings of the 50th Annual Meeting of the Orthopaedic Research Society, San Francisco, USA, 2004, 1512. [127] Ingham, E. and Fisher, J., Can metal particles (theoretically) cause osteolysis? in Proceedings of Second International Conference on Metal–Metal Hip Prostheses: Past Performance and Future Directions, Montreal, Canada, 2003. [128] Germain, M.A. et al., Comparison of the cytotoxicity of clinically relevant cobalt–chromium and alumina ceramic wear particles in vitro, Biomaterials, 24, 469, 2003. [129] Case, C.P. et al., Widespread dissemination of metal debris from implants, J. Bone Joint Surg., 76B, 701, 1994. [130] Case, C.P. et al., Preliminary observations on possible premalignant changes in bone marrow adjacent to worn total hip arthroplasty implants, Clin. Orthopaed. Relat. Res., 329S, S269, 1996. [131] Weisse, B. et al., Improvement of the reliability of ceramic hip joint implants, J. Biomech., 36, 1633, 2003. [132] Mak, M.M. et al., Effect of micro-separation on contact mechanics in ceramic-on-ceramic hip joint replacements, Proc. Instn. Mech. Engrs., J. Eng. Med., 216, 403, 2002. [133] Smith, S.L. et al., Direct evidence of lubrication in ceramic-on-ceramic total hip replacements, Proc. Instn. Mech. Engrs., J. Mech. Eng. Sci., 215, 265, 2001. [134] Oonishi, H. et al., Alumina hip joints characterized by run-in wear and steady-state wear to 14 million cycles in hip-simulator model, J. Biomed. Mater. Res., 70A, 523, 2004. [135] Nevelos, J.E. et al., Wear of HIPed and non-HIPed alumina–alumina hip joints under standard and severe simulator testing conditions, Biomaterials, 22, 2191, 2001. [136] Cheng, N.Y. et al., Lubrication and wear of alumina–alumina hip bearings, in Tribology Series 40, D. Dowson et al., Eds, Elsevier, Amsterdam, 377, 2001. [137] Nevelos, J.E. et al., Micro-separation of the centres of alumina–alumina artificial hip joints during simulator testing produces clinically relevant wear rates and patterns, J. Arthroplasty, 15, 793, 2000. [138] Stewart, T. et al., The performance of new ceramic articulations in hip simulator studies with microseparation, in Proceedings of the International Conference of Engineers and Surgeons Joined at the Hip, IMechE, 2002, C601/038. [139] Hatton, A. et al., Alumina–alumina artificial hip joints — part I: a histological analysis and characterisation of wear debris by laser capture microdissection of tissues retrieved at revision, Biomaterials, 23, 3429, 2002. [140] Tipper, J.L. et al., Alumina–alumina artificial hip joints — part II: characterisation of the wear debris from in vitro hip joint simulations, Biomaterials, 23, 3441, 2002. [141] Hatton, A. et al., Effects of clinically relevant alumina ceramic wear particles on TNF-α production by human peripheral blood mononuclear cells, Biomaterials, 24, 1193, 2003.
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[142] Jagatia, M. and Jin, Z.M., Analysis of elastohydrodynamic lubrication in a novel metal-on-metal hip joint replacement, Proc. Instn. Mech. Engrs., J. Eng. Med., 216, 185, 2002. [143] Yew, A. et al., Analysis of contact mechanics in McKee-Farrar metal-on-metal hip implants, Proc. Instn. Mech. Engrs., J. Eng. Med., 217, 333, 2003. [144] Udofia, I.J., Yew, A., and Jin, Z.M., Contact mechanics analysis of metal-on-metal hip resurfacing prostheses, Proc. Instn. Mech. Engrs., J. Eng. Med., 218, 293–305, 2005. [145] Liu, F. et al., Comparison of contact mechanics between a total hip replacement and a hip resurfacing with a metal-on-metal articulation, in Proceedings of BORS, Spring 2004, Bristol. [146] Auger, D.D. et al., Friction and lubrication in cushion form bearings for artificial hip joints. Engineering in medicine, Proc. Instn. Mech. Engrs., J. Eng. Med., 207, 25, 1993. [147] Medley, J.B. et al., Hydrophyllic polyurethane elastomers for hemiarthroplasty, Eng. Med., 9, 59, 1980. [148] Bigsby, R.J.A. et al., A comparative tribological study of the wear of composite cushion cups in a physiological hip joint simulator, J. Biomech., 31, 363, 1998. [149] Murakami, T. et al., Adaptive multimode lubrication in natural synovial joints and artificial joints, Proc. Instn. Mech. Engrs., J. Eng. Med., 212, 23, 1998. [150] Smith, S.L. et al., A tribological study of UHMWPE acetabular cups and polyurethane compliant layer acetabular cups, J. Biomed. Mater. Res., 53, 710, 2000. [151] Firkins, P.J. et al., A novel low wearing differential hardness, ceramic-on-metal hip joint prosthesis, J. Biomech., 34, 1291, 2001. [152] Fisher, J. et al., An in vitro study of the reduction in wear of metal-on-metal hip prostheses using surface-engineered femoral heads, Proc. Instn. Mech. Engrs., J. Eng. Med., 216, 219, 2002. [153] Morita, Y. et al., Wear properties of alumina/zirconia composite ceramics for joint prostheses measured with an end-face apparatus, Biomed. Mater. Eng., 14, 263, 2004. [154] Jin, Z.M., Medley, J.B., and Dowson, D., Fluid film lubrication in artificial hip joints, in Proceedings of 29th Leeds-Lyon Symposium on Tribology, D. Dowson et al., Eds, 2003, 237.
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II Industrial Lubrication Practices
© 2006 by Taylor & Francis Group, LLC
18 Steel Industry 18.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.2 Lubrication . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
18-3 18-5
Types of Lubricants
18.3 Problem Areas . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
18-10
High Temperatures • Impact Forces • Water Entry • Pumping Greases • Increase in Speeds
18.4 Lubricant Guidelines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.5 Lubricant Application Methods . . . . . . . . . . . . . . . . . . . . . .
18-11 18-11
Circulating Oil Systems • Centralized Grease Systems • Oil Mist Systems • Air–Oil System
18.6 Plant Equipment — Auxiliary Services. . . . . . . . . . . . . . .
18-20
Electrical Machinery • Utilities Systems • Mobile Equipment • Maintenance Shops • Hydraulics • Overhead Cranes
18.7 Iron Making Equipment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
18-25
Coke Manufacture • Limestone and Iron Ore • Air • Blast Furnace
18.8 Steel Making Equipment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.9 Steel Shaping — Casting . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.10 Rolling Processes and Components . . . . . . . . . . . . . . . . . .
18-30 18-35 18-38
General — The Mills
18.11 Mill Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
18-42
Reheat Furnaces • Mill Tables • Manipulators, Entry, and Side Guides • Cooling Beds — Transfer Tables • Mill Stand Complex
Rick Schrama Dofasco Inc.
18.12 Steel Finishing Facilities . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
18-58 18-58
18.1 Introduction The steel industry can be broken down into several distinct types of producing groups. These are: fully integrated plants, mini-mills, and finishing operations. A general product flow of the steel industry process is shown in Figure 18.1. An integrated steel plant takes iron ore, coal, limestone, scrap, and ferro alloys to produce steel. The plants have coal handling facilities, coke oven batteries, blast furnaces, melt shops with basic oxygen furnaces and electric arc furnaces, ladle metallurgy, continuous casting and ingot pouring shops, hot rolling mills, cold processing facilities and various utilities, material handling, and
18-3
© 2006 by Taylor & Francis Group, LLC
18-4
© 2006 by Taylor & Francis Group, LLC
Handbook of Lubrication and Tribology
FIGURE 18.1 Flow diagram showing the principal process steps involved in converting raw materials into the major product forms excluding coated products. (Printed with permission, Association of Iron and Steel Technology, The Making, Shaping and Heat Treating of Steel, 10th Edition.)
Steel Industry
18-5
environmental equipment. The mini-mills take scrap steel and ferro alloys and process them through electric arc furnaces, ladle metallurgy, and hot rolling facilities with some cold processing in some cases. The finishing operations plants are stand alone cold finishing operations that produce wire, pipe, galvanized, tin plate, painted sheet, and various rod products without any steel making or hot rolling divisions. Within the fully integrated type of plant, there are five main categories of area in the plant, which includes the following: 1. The first group consisting of the utilities and transportation groups that include electrical equipment, that is, motors and generators, steam, compressed air and gas generating equipment such as turbines, compressors, and valves and overhead cranes, and all mobile equipment such as transporters, railroad stock, and trucks. 2. The second major group is the primary end of the integrated steel plant. This includes all of the machinery used in the handling and preparation of raw materials such as limestone, coal, iron ore, sinter and pellets used to manufacture coke, and sinter and molten iron. In the sinter plants, coke ovens, blast furnaces, and direct reduction plants. 3. The third group is the steel making area of the steel plant. Today, steel is produced using variations of the electric arc furnace technology and the basic oxygen furnace technology. In the former type of furnace, steel scrap is melted with carbon electrodes in a water-cooled furnace, which tips on an arc to pour the molten steel into a ladle. The basic oxygen furnace (BOF) takes molten iron, scrap steel, and ferro alloys in a refractory lined steel vessel. The furnace is held in place using a trunnion ring that is supported by a pair of pillow blocks. A pinion and bull gear arrangement rotates the 20 to 40 t capacity vessels. 4. The fourth group consists of machinery and equipment used to shape the steel slab, bloom, ingot, or billet. It starts with the continuous casting machines with their ladle turrets, tundishes, molds and oscillating machinery, segments, and then the cutoff equipment. From there the steel goes to the reheat furnaces and then onto the hot rolling end of the steel mill. The rolling mills and accessories shape the steel into round, square and hex bars, strip, plate, merchant bars and structural shapes in what is termed a “semifinished” product. The steel could also go from an ingot and then be forged into a shape followed by heat treating and machining processes. These areas of the steel mill use large quantities of lubricants to control friction and wear problems. 5. The fifth and final group consists of the finishing facilities in a steel mill. This is where the steel is pickled and cold worked into wire, tubes, pipe, cold rolled strip, and coated steel in the forms of galvanized, tin plate, chrome free steel, plated, and heat treated. This group also includes roll forming, shearing, trimming, leveling, straightening, and packaging types of operations. These operations can also consume large amounts of lubricants depending on the types of drive systems, fans, and bearings that are selected by the equipment builders.
18.2 Lubrication 18.2.1 Types of Lubricants The steel industry employs four main categories of lubricants in its facilities (1) process lubricants, (2) hydraulic fluids, (3) metal working lubricants, and (4) machinery lubricants. Table 18.1 shows a further breakdown of the fluids that are used in the steel plants. 18.2.1.1 Process Lubricants The process type lubricants come in direct contact with the steel surfaces during the various pickling, rolling, shaping, plating, and heat-treating operations. Rolling oils used in cold rolling and temper mills constitute the largest volume in this group of fluids. The roll coolants, roll oils, and pickler oils applied to the strip and sheet in cold rolled tin plate, cold rolled sheet, and hot rolled strip rolling operations keep the
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Handbook of Lubrication and Tribology
rolls cool, extend the roll life, control roll bite friction, clean the strip, and influence the strip shape. The volume and expenditures for process rolling oils in cold rolling operations far exceed the lubricants used for machinery maintenance and reliability. Palm oil at one time was the original vegetable oil used for cold rolling. Today, most rolling oils are made from the tallows of beef and pork with surfactants and biocides added to the oil to extend the life of the oil. Some synthetic products from esters have been developed to compete with the tallow based oils. The oil is mixed with water in a 2 to 4% solution. It is kept heated to 160◦ F and then sprayed through nozzle systems onto the rolls and strip. The ability of the oil in an emulsion to preferentially wet out on the roll and steel strip at the roll bite is a basic critical property. This is termed “Saponization.” There are once-through or “direct application” systems and recirculating or “solution” systems used by the steel industry. The second major process fluid or lubricant is the drawing lubricant. This material is used on the successive drawing dies to reduce friction and heating as steel rod is drawn in a cold state through the dies to produce wire or for producing cold drawn tube from hot rolled tubes.The processes can be done wet or dry. In the wet process, two basic lubricant types are employed (1) solutions of high fat, low soap for carbon steels and (2) 100 to 300 SSU (20 to 65 cSt) at 100◦ F (40◦ C) oil compounded with sulfur, chlorine, and fat. The dry process uses sodium, calcium, or aluminum soaps, or soap-fat compounds. Sodium soap requires its removal from the surface prior to the next operation. The aluminum soap has been used in cold heading. The soap-fat compound was used when extra clean brightware was required. In the wet process, there are two basic lubricant types (1) solutions of high fat, low soap for carbon steels and (2) 100 to 300 SSU (20 to 65 cSt) at 100◦ F (40◦ C) oil compounded with sulfur, chlorine, and fat. 18.2.1.2 Hydraulic Fluids Hydraulic fluids are used in the hydraulic or fluid power systems throughout the plant. Maintenance departments along with the resident hydraulic specialists or fluid power technicians and engineers traditionally maintain these fluids. The industry uses large volumes of these fluids if the systems are not well maintained and are not kept tight or free from leaks and damage. The fluids include petroleum oils (A.W. or antiwear type), high water base fluids (soluble and synthetic), invert emulsions, water glycols, polyolesters, polyolethers, phosphate esters, polyalphaolefins (PAOs), and polyalkylene glycols (PAGS). A majority of the hydraulic fluids in a steel plant is of petroleum base except where heat or possible contact with hot metal is involved. In these situations the water glycols and ester based fire resistant fluids are used. 18.2.1.3 Metal Working Lubricants These fluids are used primarily in the maintenance departments that have machine shops and the manufacturing shops or roll shops. They consist of the cutting and grinding fluids used to remove metal from mill rolls, fabrications, and steel mill components. 18.2.1.4 Machinery Lubricants The lubricants in this category consist of the greases, gear oils, turbine oils, circulating oils, pastes, and solid lubricants used to lubricant the steel mill machinery such as bearings, ways, liner surfaces, gears, slides, chain, and sprockets. The range of products used by the maintenance department in a typical integrated steel plant are shown in Table 18.1. Maintenance groups in a modern steel mill try to consolidate the products wherever possible to reduce costs. Most plants now have some form of computerized maintenance management system (CMMS) that looks after the warehousing of the products, consumption of the lubricants, purchasing and the work order planning, and execution. Wherever it is practical, lubricants are being brought into the plant in bulk form for flo bins or storage tanks. This eliminates the storage and handling of drums. Most lubricants supplied to the steel mill come with a product code that is supplied by the end-user.
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Steel Industry TABLE 18.1
18-7 Typical Products by Type
Type of product Machinery lubricants Black oils Spindle oils Way lubricants Cylinder oils
Straight oils Refined circulating oil Turbine oils Mist oils Motor oils Gear oils Synthetic oils Miscellaneous oils Open gear compounds Block greases Cup grease Mill grease
Grease with fillers Miscellaneous greases Hydraulic fluids Antiwear hydraulic oils Fire resistant fluids Soluble oils Metalworking lubricants Cutting oils, etc. Process lubricants Slushing oils Roll oils Pickler oils Wire drawing compounds
© 2006 by Taylor & Francis Group, LLC
Pertinent information — principal uses Residual, heavy, tack, dark oils — for use on open gears, skids, chains Low viscosity — turbine quality — machine tool bearings Good stick-slip properties — heavy duty EP lubricant with oiliness additive — machine tool ways High viscosity, both compounded, and noncompounded with degras or acidless tallow — worm gears, steam cylinders Noninhibited oil for once-through applications of high loss circulating systems — bearings, gears, etc. Inhibited with good demulsibility and antirust — backup roll bearings, primarily Inhibited and highly refined, for turbines and blowers Good misting properties, low stray mist, no gumming properties — rolling element bearings API service classifications for diesel and gasoline engines EP compounded oils — mill drives and pinions Silicones, diesters, polyalphaolefins — for special high temperature needs Aerosols and other specialties — miscellaneous High viscosity, tacky, residual oils, or asphaltic compounds — EP additive — open gears NLGI 4, 5, and 6 grade EP greases — old design mill bearings Calcium soap grease with light viscosity oil — miscellaneous centralized systems on mills NLGI 1 and 2 grades with EP additives and high viscosity oil, water resistant, Ca, Li, and Al soap thickeners Grease with MoS2 , graphite, or Teflon , etc. as filler — high-load boundary lubrication applications Micro-gel, clay thickened, etc. — high temperature
% Consumption (9 plant sampling) 1.6 Trace 0.3 0.9
6.5 13.75 2.5 0.72 2.9 13.0 Trace 0.75 0.8
0.25 8.0
0.4 1.0
Low viscosity refined oils with friction modifier — for fluid power systems Invert emulsions, water glycols, and organic esters — for fluid power Lubricating oil with emulsifier for use in hydraulics or light machine tool operations
11.0
Chemical and petroleum base products for use in cutting, grinding, drawing, etc.
2.0
Antirust, inhibited, light oil — for use as sheet protection Fatty lubricants used in the rolling process — hot and cold rolling Light mineral oil with rust inhibitor — sheet protection after pickling Soaps with fillers for wire drawing
7.8
4.0 2.0
15.2 3.8 0.7
18-8 TABLE 18.2
Handbook of Lubrication and Tribology Mill Bearing Grease Comparison Thickener
Test method Water spray-off (%)a Water emulsion testa Water absorbed (ml) Penetration change Adhesion change (psi) Rust protection change Mobility at 20◦ F (−6.6◦ C) (g/min) Mobility at 0◦ F (−17.8◦ C) (g/min) Dropping point, ◦ F (◦ C) Water washout, 175◦ F (80◦ C) (g)a Roll shear stability (% change) Roll shear stability with 200◦ F (93◦ C) H2 O (% change)a Mechanical shear stability (% change)
Ca
Ca–Pb
Li–12 OH
Black Li–12 OH
Al complex 5/10
Exp. Li–12 OH
ASTM (tent.)
13
28.7
98.2
49
BSCo. D217 BSCo. D1743 S.O.D.
50 272/289 11.5/9.7 1/3 0.9
92 335/352 12.7/15.4 1/1 11.2
35 290/310 8.8/12.6 1/1 105.00
108 308/347 14.4/12.8 3/1 14.0
S.O.D.
0.03
0.62
21.2
0.2
D2265 D1264
209 (98) 3.9
272 (133) 12.9
380 (193) 15.7
380 (193) 8.9
D1831
—
5.4
5.5
1.0
2.8
2.2
D1831 Mod.
—
—
—
—
19.8
5.9
D217
10.4
3.4
7.0
4.4
2.1
0.6
40 300/330 9.6/8.5 1/1 15.5 4.0 500 (260) 5.3
5 55 282/326 8.6/8.6 1/1 10 0.9 355 (179) 1.5
Note: Properties of mill greases used in roll neck bearings illustrating the results of test methods involving the effects of water contamination. a Tests devised to simulate water effects on grease. Source: Reprinted from Cichelli, A.E., NLGI (Nad. Lubr. Grease Inst.) Spokesman, April 1980. With permission.
Refined circulating oils are extremely important to the circulating oil systems in the plant. These systems provide oil for the oil film bearings used in the rolling mill backup roll assemblies, turbine and compressor oil systems, and large pinion stand gearboxes. Oils are subjected to water, rolling oil solutions, dirt, and high temperatures in their working environments and are fully inhibited for corrosion, rust, and oxidation. Additionally, the oils have excellent demulsibility characteristics to handle water. Gear oils used in the steel plant are traditionally compounded with extreme pressure (EP) additives of sulfur/phosphorus and zinc dialkyldithiophosphate (ZDDP). The oil viscosities range from ISO 68 to ISO 680 (350 to 3600 SSU) at 100◦ F. Synthetic gear oils made from PAO and PAG base stocks are also in the plant. These lubricants are used in applications where there are very wide temperature ranges and a high viscosity index (VI) is required. Mill greases make up an important role in the steel plant. Every mill has their own grease type that they prefer. Soap types include: aluminum complex, lithium complex, calcium complex, calcium sulfonate, bentone, and polyurea. Base oils can also vary from a straight mineral oil to PAO and PAG base stocks in combination with a full additive package that may include EP additives. The greases must be able to lubricate under many of the following extreme conditions: water impingement, scale, wide temperature extremes, high-applied loads, high-shear rates, etc. Typical applications include both plain journal bearings and rolling element bearings, pins and bushings in oscillatory motions, sliding ways, and wear liner surfaces. Table 18.2 shows some of the grease soaps and their properties. Lubricant test methods employed by the industry to test both used and new lubricants have been developed for the special needs in the steel industry over the years. The principle test methods used are shown in Table 18.3. ASTM (American Society for Testing Methods) is the main source of all methods.
© 2006 by Taylor & Francis Group, LLC
Steel Industry TABLE 18.3
18-9
Principal Lubricant Test Methods Used in Steel Industry
Tests and properties Oils Rust prevention (oils)
ASTM No. D665
Vapor space corrosion
D2270
Demulsibility
D2711
EP oil oxidation Static heat
D2893 Cine, milling
Panel coking
Federal 3462-T
Misting properties Timken EP FZG gear oil tests
D3705 D3323 —
Dissolved gases
USS
Particulate contamination
D2390
Oil spot test Quenchometer Stacked sheet corrosion
USS — —
Environmental cabinet
—
Hydraulic pump wear
D2882
Accelerated hyd pump wear
USS BSC
Freeze-thaw cycling
—
Greases Rust prevention Emulsion characteristics
D1743 BSCo.
Roll stability
D1831
Water spray-off
GM/BSC
Timken retention
USS
Cohesion/adhesion
BSC
Grease mobility
—
Bleeding
—
Plastic plate abrasion Static heat
D1404 USS
Pressure oil separation Timken EP
USS D3323
Function of test Three pan method measuring ability of an oil to prevent erosion of immersed polished steel Measure of effectiveness of vapor phase corrosion inhibitor, particularly in hydraulic fluids An indicator of either a new or used oil’s ability to separate from water Measures oxidation resistance of nonoxidation inhibited EP oils Screening test for determining resistance of oils to oxidize at elevated temperatures Determines temperature at which decomposition products form and deposit on heated metal panel Determines oil’s ability to function in mist systems Measure of level of EP properties (of an oil or grease) Gear rig measuring EP characteristics of gear oil under varying loads and sliding velocities Vacuum test for indicating presence of dissolved gases in hydraulic fluids — hence bulk density A measure of paniculate matter — principally used for hydraulic fluids in servo systems Quick blotter test for indication of motor oil condition Measures relative quenching speeds of oils Actual field tests evaluating performance of slushing oils for protecting cold rolled sheets from corrosion Laboratory simulated environmental cabinet for evaluating rust preventives Simulated field condition — measures ability of hydraulic fluids to prevent vane pump wear — up to 1000 psi High pressure extended life test on pump stands — evaluate ability to prevent wear Used on invert emulsions to measure ability to resist oil separation at low temperature Measures ability to protect polished steel surfaces from rust Measures ability of grease to resist emulsification and the effect of emulsification on properties A simulated roll and plain bearing to measure resistance to shear breakdown (modification includes adding 50% water at 200◦ F) Measures ability to resist from being forced off a flat surface by water jet sprays Measures ability of EP lubricant to retain EP properties under steady load without replenishment A technique for measuring a lubricant’s tenacity to metal and also within itself; useful for wire rope lubricants A measure of the ability of grease to be pumped at low-ambient temperatures (Std. Oil Development Mobilometer apparatus) Indication of a grease’s ability to retain oil in a soap matrix — in storage considerations Screening test for presence of abrasive dirt Measure of hardening tendency at elevated temperatures — important in grease systems Measures ability to avoid caking in grease systems under pressure See under oils above
Note: Supplementary to the multi-industry methods covered in the chapter on Lubricant Properties and Test Methods in Volume I.
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Handbook of Lubrication and Tribology
18.3 Problem Areas The steel industry has some problem areas associated with lubrication. Most of these while centered on the rolling processes can also show up in other parts of the steel making process. The major areas of concern are: 1. 2. 3. 4. 5. 6. 7. 8. 9.
high temperatures experienced in the hot mills high impact loads large volumes of cooling water and scale interfering with lubrication difficulties in pumping grease long distances to remote locations under cold conditions ever increasing speeds and loads to improve production rates the need to reduce the amount of spent lubricant getting into the water systems and the environment keeping the lubricants out of the rolling oil solutions in the cold rolling and tempering operations hydraulic fluids that are fire resistant and able to handle high pressures lubricants for cold climates
18.3.1 High Temperatures The rolling mill areas, especially in the hot rolling end are not beyond the capabilities of mineral oil soap based greases. Principle soap types used in the industry are: lithium soap, lithium complex, aluminum complex, and calcium sulfonate soap thickeners. Some soaps use synthetic oils such as PAO’s around the furnace areas. Bentone or clay base greases are also used in many applications.
18.3.2 Impact Forces The impact forces seen in roller bearings in the hot rolling area are best accommodated with the use of high viscosity oils, whether this is for a pinion gearbox or the base oil in the bearing greases. The minimum recommended is a 220 cSt oil with many applications using a 460 or 680 cSt oil.
18.3.3 Water Entry Water entry into oil-lubricated bearings is one of the most complex in a rolling mill problems especially so for the oil film babbitted bearings used on the backup roll assemblies. Water is kept out of the oil inside the bearing by enhancing the sealing practices, use of water deflectors, and the use of oils with excellent demulsibility properties. At the same time, the high pressure roll cooling water and descale sprays used in the hot mills require mill greases to be resistant to water washout and water sprayoff in addition to being able to lubricate the bearings or slideway surfaces. Grease is introduced into the bearings on a regular interval, the frequency dependent on the speed, temperatures, and mill operating conditions. Table 18.2 shows typical at the properties of various steel mill greases.
18.3.4 Pumping Greases Many steel mills are located in geographical locations where low temperature can be a factor in the transport and delivery of grease to the lubrication point. Low temperature pumpability of the grease over long distances is an issue in many companies, especially if a high-viscosity base oil is required.
18.3.5 Increase in Speeds Rolling element bearings require good application techniques to get the right amount of lubricant to the right location at high speeds. Oil mist and air–oil systems using high-viscosity oils are used in many work roll application in the cold rolling.
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Steel Industry
18-11
TABLE 18.4
Turbine Oil
Property Viscosity at 100◦ F Viscosity index API gravity Pour point COC flash point Rust prevention Oxidation test Emulsion characteristics at 130◦ F/54◦ C
ASTM method D-445 D-2270 D-287 D-97 D-92 D-665 D-943 D-1401
Value 150, 315, 465, 700, 1.000 SUS >80 >30 <20◦ F >375◦ F/190◦ C Pass <1.0 neutralization number after 1000-h ml oil: 40; ml water: 37; ml emulsion: 3; min: <20
General properties for the oil include: high oxidation and corrosion resistance, very good water separation for steam turbines, machine tools, and gear turbines.
TABLE 18.5
Hydraulic Oil (Extra-Duty, Antiwear)
Property
ASTM method
Viscosity at 100◦ F Viscosity index Hydraulic pump test
D-445 D-2270 D-2882
4-Ball wear test (40 kg. 1800 rpm. 130◦ F, 1 h) Oxidation test Rust prevention Winer emulsion test at 130◦ F/54◦ C
D-2266
Application dependent >90 Pass 1900 psi, 100 h at 150◦ F on a 10-gal reservoir with a Vickers V-I04-C-I0 vane-type pump, <0.010% total wear <0.50 mm scar diameter
D-943 D-665 D-1401
<1.0 neutralization number after 1000 h Pass ml oil: 40; ml water: 37; ml emulsion: 3, after 30 min
Value
General properties for the oil include: high wear protection, high oxidation, corrosion resistance, and good demulsibility for fluid power systems subjected to shock loads, variable speeds, and continuous operation.
18.4 Lubricant Guidelines Tables 18.4 to 18.34 provide some guidelines for the maintenance and lubrication personnel in the steel mills and the lubricant manufacturers of the various lubricants that can be used in the facilities. The typical critical properties that might be required for the grease or oil to perform under certain applications are given in order to supply minimum coverage. Premium products with properties that exceed those shown in the guidelines may be required for an application so as to meet performance requirements, health and safety, disposal, and federal or state regulations. These guidelines were also published in the CRC Tribology Data Handbook.
18.5 Lubricant Application Methods In a steel mill, manpower issues are always present. In order to keep manpower costs low, most plants tend to use centralized delivery systems for their lubricants. Grease lines in cold climates are generally heat traced to ensure delivery to the right location. Circulating oil systems can be small reservoirs are large units that handle a complete 5 stand cold mill or 7 stand finishing mill in the hot mill.
18.5.1 Circulating Oil Systems These systems supply oil to the mill backup roll bearings or pinion gearboxes from a central reservoir. Most systems consist of two tanks, one which is operating and a second tank that is off-line to allow the
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Handbook of Lubrication and Tribology TABLE 18.6
Circulating Oil
Property Viscosity at 100◦ F
ASTM method
Viscosity index COC flash point
D-445 D-567 D-92
Neut. number Pour point Foaming characteristics Emulsion characteristic at 180◦ F Oxidation test by static oxygen 4-Ball wear at 5 kg, 1800 rpm, 130◦ F, 1 h Coefficient of friction
D-974 D-97 D-892 D-1401 D-2272 D-2266 D-2266
Value 700, 1000, 1500, 2150, 3150 SUS >80 >350◦ F for oils under 1000 SUS >450◦ F for oils over 1000 SUS <0.10 <20◦ F Nonfoaming ml oil: 40; ml water: 37; ml emulsion: 3, after 40 min >24 h for 60 mm pressure drop <0.60 mm scar diameter <0.1 under wear test conditions
General properties required include moderate oxidation stability and rapid water separation for plain bearings, machine tools, low-medium load gears, and backup roll bearings where there is water, dirt, and heat.
TABLE 18.7
Extra-Duty Circulating Oil
Property Viscosity at 100◦ F Viscosity index COC flash point Pour point Foaming characteristics Rust protection Water separation DDT at 180◦ F/82◦ C Demulsibility characteristics Water separation test Rotary bomb oxidation FZG test 4-Ball EP 4-Ball wear at 20 kg, 1800 rpm, 130◦ F/54◦ C, 1 h Field test
ASTM method D-445 D-2270 D-92 D-97 D-892 D-665 D-2711 D-1401 D-2272 DIN A/8.3/90 D-2783 D-2266
Value 315, 465, 700, 1000, 1500, 2150, 3150, SUS >90 >350◦ F/176◦ C for oils under 1000 SUS >450◦ F/232◦ C for oils over 1000 SUS <10◦ F Zero break at 600 sec at 100◦ F Pass <5% top. >95% bottom >36 ml ml oil: 40; ml water: 37; ml emulsion:3, after 40 min >120 min Pass stage 9 >150 kg, >30 LWI <.50 mm Satisfactory performance for the application intended
General properties required include high oxidation stability, rapid water separation, and antiwear properties for reduction gear and bearings where loads are greater than can be lubricated by a straight mineral oil.
oil to rest so that water and contamination in the oil can be settled out or removed from the oil. This is normally done every 4 to 6 weeks. The oil film bearings in the backup roll chocks are supplied from circulating oil systems. Each system has pumps, a reservoir with tank heaters, a floating suction, magnetic separators and return line strainers, pressure regulating and flow control monitoring valves, an oil cooler, and a filtration system. Many systems also come with vacuum dehydrator equipment to remove the water. Many mills have sight flow gauges or drip legs in critical locations to detect the presence of water entry into the system. These systems generally have monthly samples taken and sent to laboratories to analyze for viscosity, water content, sediment, total acid number (TAN) and wear metals. Reservoirs in these systems are sized to allow oil retention time in the tank from 30 to 45 min. Thus, for a 10,000 gall tank, the system should be circulated beyond 330 gpm. Reservoirs are divided into dirty (return side) and clean side sections by plate weirs. Any sediment or wear metals returning from the bearings or gearing is caught by the dirty side and the overflow goes into the clean side. The tank suction
© 2006 by Taylor & Francis Group, LLC
Steel Industry TABLE 18.8
18-13 Ball and Roller Bearing Grease ASTM method
Property Penetration 10,000 Strokes Roll stability: 2 h Drop point Oxidation Pressure oil separation Wheel bearing test, 6 h at 235◦ F Water washout Grease mobility 4-Ball wear test (7.5 kg, 1800 rpm, 130◦ F, 1 h) Field test
D-217 D-217 D-1831 D-2265 D-942 D-1263 D-1264 US Steel method D-2266
Value Suitable for the application. NLGI Grades 0, 1, 2, 3 <10% change <25% change <350◦ F <5 psi drop <40 (Quarter scale penetration) <5% loss <5% loss at 100◦ F, <15% loss at 175◦ F >0.1 g flow per second at 0◦ F <0.60 mm sear diameter Satisfactory for the intended application
General properties required include: grease — provides high stability, works at low and high temperatures and under normal pressure for plain bearings, machine tools, low-medium load gears, and backup roll bearings where there are dirt, water, and heat.
TABLE 18.9
Circulating Turbine Oil ASTM method
Property Viscosity at 100◦ F API gravity Viscosity index Pour point COC flash point Rust prevention Oxidation test Emulsion characteristics at 130◦ F/54◦ C Field test
D-445 D-287 D-2270 D-97 D-92 D-665 D-943 D-1401
Value 150, 315, 465, 700, or 1000 SUS >20 >80 <20◦ F, lower, dependent upon application >375◦ F/190◦ C Pass Not to exceed 2.0 neutralization number after 1000 h ml oil: 40; ml water: 37; ml emulsion: 3, after 40 min Satisfactory for the application intended
General properties required include: turbine oil — for circulating systems provides good oxidation and corrosion resistance along with good water separation for steam turbine, machine tool, precision hydraulic machine, and ring oil bearing applications.
TABLE 18.10
Engine Oil
Property Viscosity at 100◦ F Viscosity index COC flash point Neut. no. Pour point 4-Ball wear test (5 kg, 1800 rpm, 130◦ F, 1 h) Coefficient of friction Field test
ASTM method
Value
D-445 D-2270 D-92 D-974 D-97 D-2266
215, 465, and 700 SUS >50 >340◦ F <0.10 <20◦ F or lower depending upon application <0.6 mm
D-2266
<0.1 Satisfactory for the application intended
General properties required include: oil —provides ordinary oxidation stability with the correct viscosity for plain bearings, slides, machine tool bearings, and low- or normal-load enclosed gears where extreme pressure or oxidation-inhibited qualities are not necessary.
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18-14
Handbook of Lubrication and Tribology TABLE 18.11
Circulating Engine Oil ASTM method
Property Viscosity at 100◦ F Viscosity index COC flash point Neut. no. Pour point Emulsion characteristics at 130◦ F/54◦ C 4-Ball wear text (5 kg, 1800 rpm. 130◦ F/54◦ C, 1 h) Coefficient of friction Field test
Value
D-445 D-2270 D-92 D-974 D-97 D-1401
215, 315, and 465 SUS >50 >340◦ F <0.10 <+20◦ F or lower depending on application ml oil: 40; ml water: 37; ml emulsion: 3, after 60 min
D-2266
<0.60 mm
D-2266
<0.1 Under conditions of wear test Satisfactory for the application intended
General properties required include: circulating oil —provides oxidation stability and water separation for plain bearings, slides, machine tools, and low- or normal-load enclosed gears where extreme pressure or oxidation-inhibited qualities are not necessary.
TABLE 18.12
Sendzimir Rolling Oil
Property Base oil type Viscosity at 100◦ F COC flash point Color Rust prevention Sulfur, % as S Corrosion Copper strip, 3 h at 212◦ F Steel strip, 3 h at 212◦ F Field test
ASTM method D-445 D-92 D-665 D-129 D-130
Value May range from naphthenic to paraffinic in type 75 and 105 SUS >300◦ F/149◦ C <4 Pass (synthetic sea water) <0.3 2 Max classification la Classification Satisfactory for the application intended
General properties required include: rolling oil —supplies low viscosity, rust resistance, and oxidation stability for circulating pumps from storage tanks. During each pass to the mill the lubricant is reconditioned by a combination of settling, centrifuging, and filtering. In the mill, the oil cascades over and through the backup bearings to lubricate them. In addition, the oil is sprayed into the roll-bite area.
utilizes a floating suction line. For high-viscosity oils, the settling time is important since the returning oil may have entrained air, water, and sediment in it. Systems are designed to have 20 to 50% extra pump capacity to allow for additional oil flow and also to take into account mill speed changes. Orifices, sight glasses, flow control devices, and turbine type metering units are located in front of the bearings and gearing to control the flow into each lubrication point. Most large systems have some form of steam or electric heating coils in the reservoir to maintain tank temperatures especially during extended shutdowns. Additionally, the systems have tube and shell or plate coolers to remove excess heat from the oil before the oil goes back to the system. The tanks are also equipped with tank level gauges to control the amount of oil in the tank. Most systems have some form of inline filtration system to remove the debris and sediment from the oil. Traditionally this was done by duplex strainer baskets with 150 µm stainless steel cleanable baskets on the pressure side of the pumps. At the mill stand or the gearbox, there would be a single strainer to remove additional material before the oil went to the lubrication point. Modern systems are being supplied with off-line kidney loop systems to remove dirt and oxidation products down to 25 µm. These systems
© 2006 by Taylor & Francis Group, LLC
Steel Industry TABLE 18.13
18-15 Inhibited Hydraulic Oil
Property
ASTM method
Viscosity Viscosity index Hydraulic pump test
D-445 D-2270 D-2882
4-Ball wear test (40 kg, 1800 rpm, 1–30◦ F/54◦ C, 1 h) Oxidation test Water emulsion test at 130◦ F/54◦ C Rust prevention test
D-2266
Suitable for the specified application >80 Must pass modified 100-h, 1900 psi test at 150◦ F (Vickers V-105-C-10 vane-type pump) with 10-gal reservoir fill; <0.05% total wear. Must maintain 1900 psi throughout the test. <0.80 mm scar diameter
D-943 D-1401 D-665
<1.0 neutralization number after 1000 h ml oil: 40; ml water: 37; ml emulsion: 3, after 30 min Must pass
Value
General properties required include: hydraulic oil — provides moderate wear protection, high oxidation and corrosion resistance, and good demulsibility for precision oil hydraulic systems.
TABLE 18.14
Open Gear Lubricants — EP Type ASTM method
Property Drum viscosity at 210◦ F Copper strip corrosion 4-Ball EP test
D-445 D-130 D-2783
4-Ball wear test
D-2266
Value 500, 750, 1000, 1500, 2000, and 3000 SUS <2b. classification >200 kg Weld point >30.0 Load-wear index <0.8 mm scar diameter, 20 kg, 1 h, 130◦ F, 1800 rpm
General properties required include: lubricant — provides adhesive and extreme pressure properties when applied by spray or brush for protection of gears and other mechanical parts that are difficult to lubricate by pressure system or bath. The protective coating must adhere to the surface to be lubricated without peeling, scaling, or excessive throw-off.
TABLE 18.15
Open Gear Lubricants — Non-EP Type
Property Drum viscosity at 210◦ F Copper strip corrosion 4-Ball EP test 4-Ball wear test
ASTM method D-445 D-130 D-2783 D-2266
Value 500, 750, 1000, 1500, 2000 and 3000 SUS
General properties required include: lubricant — provides adhesive and non-EP properties when applied by spray or brush for protection of gears and other mechanical parts that are difficult to lubricate by pressure system or bath. The protective coating must adhere to the surface to be lubricated without peeling, scaling, or excessive throw-off.
have their own pumps, coolers and heater and normally pump about one fifth of the tank’s capacity. New systems are also being supplied with in-line pressure filters with filter elements that are set at 100 to 75 µm.
18.5.2 Centralized Grease Systems These systems are the backbone of a rolling or process line in the steel industry. Systems vary from the single line progressive, single line injectors to the dual line design. The size of the system and number of
© 2006 by Taylor & Francis Group, LLC
18-16
Handbook of Lubrication and Tribology TABLE 18.16
Inverted Emulsions
Property
ASTM method
Viscosity Hydraulic pump test
D-445 D-2882
4-Ball wear test (40 kg, 1800 rpm, 130◦ F/54◦ C, 1 h) Thermal stability test at 200◦ F (U.S. Steel Method) Fire-resistance
D-2266
Rust prevention test
D-665A
Value Suitable for the specified application Must pass modified 100-h, 1900 psi test at 150◦ F (Vickers V-104-C-10 vane-type pump) with 10-gal reservoir fill: <0.10% total wear, must maintain 1900 psi throughout the test; no fluid or vapor-phase corrosion permitted <1.00 mm scar diameter No separation after 1 h in a water bath Must pass fire resistance tests as described in Schedule 30 of the U.S. Bureau of Mines Must pass
General properties required include: oil —supplies emulsion stability, fire resistance, moderate wear resistance, and correct viscosity for fluid power systems that are subject to fire hazards in mines and mills.
TABLE 18.17
Pale Paraffin Slushing Oil
Property Base Viscosity at 100◦ F Viscosity index Neut. no. COC flash point Pour point Field test
ASTM method D-445 D-2270 D-974 D-92 D-97
Value Midcontinent Paraffin-Base Oil preferred; naphthenic acceptable 60, 75, and 105 SUS >70 paraffin – >30 naphthenic <0.05 >270◦ F/132◦ C naphthenic–>300◦ F/149◦ C paraffin <+35◦ F/2◦ C Suitable for the application intended
General properties required include: this straight mineral pale oil —provides low viscosity at normal temperatures in a paraffinic base for short-term slushing or coating oil, process oil, rolling oil, cutback oil for heavier base oils and compounds.
lubrication vary from 4 to 6 points to several hundred. The systems have a reservoir, air operated pump, filtration from the reservoir, and some form of distribution or divider block to provide proper grease quantities to each bearing, gear, slideway or pin, and bushing. Many other pieces of machinery such as motor bearings, conveyor rollers, couplings, and pump bearings are lubricated manually according to a prescribed relubrication interval. The system may be as small as a 1 l grease-dispensing container to a large system with a 40,000 lb capacity reservoir. Bulk system, which the latter represents, is a growing trend in the industry instead of using drums or flo bins. The bulk tanks are used to provide grease to smaller 3,000 to 4,000 lb tanks throughout the system. Each delivery of grease from the lubricant manufacturer is sampled and tested to ensure that it meets the mill’s requirements. Many systems have condition monitoring instrumentation to monitor that grease is being dispensed through the main lines and the distribution blocks. In recent years, the roller bearing industry has brought a new roller bearing design into the marketplace for rolling mill work roll bearings and caster segment bearings. This is the realm of the sealed bearing where contact lip seals using elastomers or metal shields are used to keep the lubricant in and the dirt or water out of the bearing assembly. The bearings are pregreased with a high-viscosity grease that is water
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Steel Industry
18-17
TABLE 18.18
Noninhibited Hydraulic Oil
Property
ASTM method
Viscosity Viscosity index Neut. no. Hydraulic pump test
D-445 D-2270 D-974 D-2882
Demulsibility at 130◦ F/54◦ C 4-Ball wear test (40 kg, 1800 rpm, 130◦ F, 1 h)
D-I40I D-2266
Value Suitable for the specified application >50 <0.10 Must pass standard 1000-h, 1000 psi test at 175◦ F/80◦ C (Vickers V-104-E vane-type pump) with 3 12 -gal reservoir fill; <0.10% total wear; must maintain 1000 psi throughout the test ml oil: 40; ml water: 37; ml emulsion: 3, after 60 min <1.50 mm scar diameter
General properties required include: hydraulic oil-provides wear and demulsibility protection for mill and mine hydraulic systems that are prone to continued high leakage.
TABLE 18.19
Roll Neck Spray or Gear Oil ASTM method
Property Viscosity, SUS
D-445
Copper strip corrosion Grease mobility 150 psi. grams flow per second Evaporation loss 4-Ball EP test
D-130 U.S. Steel method
4-Ball wear test (20 kg, 1800 rpm, 130◦ F/54◦ C, 1 h) Field test
D-2595 D-2783 D-2266
Value Suitable for the application intended. 150, 215, 465, 700, and 1000 at 210◦ F suggested 2b Classification For viscosities of 300 SUS or above: >0.5 g/s at 60◦ F, for viscosities of 300 SUS or lower: >0.06 g/s at 20◦ F >1% (nonsolvent compounding) >200-kg Weld point >30.0 Load-wear index <0.8 mm scar diameter Satisfactory for the application intended
General properties required include: lubricant —supplies adhesive properties under water and extreme pressure when applied using spray, bath, or pump methods for applications in blooming mill, billet mill, slabbing mill, bar mill, plate mill, roll neck bearings of the fabric, bronze, babbitt, or combination and segmented types. These bearings are subjected to large quantities of water and boundary lubrication.
resistant with a very low oil bleed rate. The bearings are installed for a 6 to 8 month campaign. At the end of the campaign, the bearings are washed and fresh grease is put into them. Four-row tapered roller bearing and spherical roller bearings are the most popular. Sealed or shielded ball bearings are used for pumps, compressors, and motor applications. Another development in the past few years is the use of single point lubricators for bearings used in locations that are difficult to get to for relubrication. These units have a small reservoir (up to 260 ml in size) with a timer to set the frequency of lubrication and a method to force the grease into the bearing whether it is a motor and screw or a gas filled bladder.
18.5.3 Oil Mist Systems This method of lubrication takes oil and atomizes it into an air stream through the use of a venturi. It is reclassified or condensed by a nozzle at the point of lubrication. Oil mist has the following advantages (1) low oil consumption, (2) low power loss in rolling bearing applications, (3) low initial cost for system and installation, and (4) lower temperatures in bearings. The disadvantages of oil mist are (1) oil vapor
© 2006 by Taylor & Francis Group, LLC
18-18 TABLE 18.20
Handbook of Lubrication and Tribology Mist EP Gear Oil ASTM method
Property Viscosity, SUS
D-445
Viscosity index Copper strip corrosion at 212◦ F (100◦ C) COC, flash point Pour point
D-2270 D-130 D-92 D-97
Dynamic demulsibility at l80◦ F/82◦ C Extreme pressure gear oil oxidation 312 h, 203◦ F/95◦ C, 10 l of dry air per h Timken load arm test 4-Ball EP test
D-2711 D-2893
FZG test Field lest
D-2782 D-2783 DIN A/8.3/90
Value Suitable for the application intended: 50-70-90-110-130150-250-500 at 210◦ F/100◦ C >85 1a Classification >400◦ F/204◦ C <20◦ F for oils up to 130 SUS at 210◦ F/(100◦ C); lower pour point may be required for certain applications <7% increase in viscosity (SUS at 210◦ /100◦ C) <7% increase in viscosity (SUS at 210◦ F/100◦ C) >60 lbs O.K. >250-kg Weld point >40.0 Load wear index >Ninth stage Satisfactory for the application intended
General properties required include: oil —has extra high load-carrying capacity, oxidation stability, and water separation for lubrication of bearings and gear drives in extra heavy duty service. This service is subjected to dirt, scale, heat, and boundary lubrication.
TABLE 18.21
Extra-Duty Gear Oil
Property
ASTM method
Viscosity, SUS
D-445
Viscosity index Copper strip corrosion at 212◦ F (100◦ C) COC flush point Pour point
D-2270 D-130 D-92 D-97
Extreme pressure gear oil oxidation, 312 h, 203◦ F/95◦ C, 10 l of dry air per h Timken load arm test 4-Ball EP test
D-2893
FZG test Field test
D-2782 D-2783 DIN A/8.3/90
Value Suitable for the application intended: 50-70-90-110-130150-250-500 at 210◦ F/100◦ C >85 1a Classification >400◦ F/204◦ C <20◦ F for oils up to 130 SUS at 210◦ F (100◦ C); lower pour point may be required for certain applications <7% increase in viscosity (SUS at 210◦ F/100◦ C) >60 lbs O.K. >250-kg Weld point >40.0 Load-wear index >Ninth stage Satisfactory for the application intended
General properties required include: gear oil —provides extra high load-carrying capacity, oxidation stability, and water separation for lubrication of bearings and gear drives in extra-heavy-duty service. This service is subjected to dirt, scale, heat, and boundary lubrication.
discharged into the atmosphere, (2) inability to cool high speed roller bearings, (3) potential sudden failure due to small oil quantities, and (4) limitations to the types of oils that can be misted. Oil mist is used in the following applications: shaker screen bearings, electrolytic tinning lines, runout table bearings, spindle lubrication, table gears, chains, gearboxes, wire rope, work roll bearings, roller bearings used for backup roll assemblies in flat rolled mills, and rod and wire mills. Oils used for mist applications contain polymers to suppress stray mist as well as resisting oxidation at high temperatures. The lines can be monitored for pressure to ensure that any carbonaceous deposits will not clog nozzle holes. An increase in differential pressure in the system indicates a blockage somewhere.
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Steel Industry
18-19
TABLE 18.22
Hypoid Gear Oil ASTM method
Property Viscosity COC flash point Copper strip corrosion Timken EP test 4-Ball EP test 4-Ball wear lest (20 kg, 1800 rpm, 130◦ F/54◦ C, 1 h) Falex lubricant test at 500 lbs. — 1 h Field test
Value
D-445 D-92 D-130 D-2782 D-2783 D-2266
Suitable for the application intended: SAE #90, #140 >350◦ F/177◦ C 1b Classification >50 lb O.K. >250-kg Weld point >45 Load-wear index <0.8 mm scar diameter
D-2670
<25 Teeth of wear Satisfactory for the application intended
General properties required include: oil —has high load-carrying capacity, high antiweld and oxidation stability for lubrication of bearings and hypoid gears in heavy duty service.
TABLE 18.23
Combination Synthetic-Petroleum Fluids
Property
ASTM method
Viscosity Hydraulic pump test
D-445 D-2882
4-Ball wear test (40 kg. 1800 rpm, 130◦ F/54◦ C, 1 h) Fire-resistance
D-2266
Suitable for the specified application Must pass modified 100-h, 1900 psi test at 150◦ F (Vickers V-105-C-10 vane-type pump) with 10 gal. reservoir fill; must show <0.010% total wear; must maintain 1900 psi throughout the test; must show no liquid or vapor-phase corrosion <0.80 mm scar diameter
Rust prevention test
D-665
Must pass the spray flammability test as described in Schedule 30 of the U.S. Bureau of Mines Pass
Value
General properties required include: fluids —provide fire resistance and excellent wear protection for hydraulic systems that are subject to fire hazards. Suitable for systems operating at high pressures (to 5000 psi). They are limited to a high temperature of 180◦ F and a low temperature of 30◦ F. They must be compatible with seals and packings of Viton, Teflon, and Nylon.
TABLE 18.24 Insulating Oil for Transformers and Oil Switches Property Viscosity at 100◦ F API gravity COC flash point Neutralization no. Pour point AC dielectric strength Interfacial tension
ASTM method D-445 D-287 D-92 D-974 D-97 D-117 D-971
Value 60 SUS >API 26 >275◦ F/135◦ C >0.05 <−40◦ F >30 kv >30
General properties required include: oil —has a wellrefined high-dielectric strength for use as an electric insulating oil on transformers of the open type, conservator type, and inert gas blanket type. Oil switches.
© 2006 by Taylor & Francis Group, LLC
18-20
Handbook of Lubrication and Tribology
TABLE 18.25
Extreme Pressure Oil ASTM method
Property Viscosity at 210◦ F/100◦ C
D-445 D-92 D-130 D-97
COC flash point Copper strip corrosion Pour point Extreme pressure gear oil oxidation at 312 h, 203◦ F, 10 l/h air dry Timken EP test 4-Ball EP test 4-Ball wear test (20 kg, 1800 rpm. 1 h, 130◦ F) Coefficient of friction Field test
Value
D-2893
Suitable for the application intended; 50-70-90-110-130-150-250-500 SUS >400◦ F/204◦ C la Classification <20◦ F for oils up to 130 SUS at 210◦ F/100◦ C, lower pour point may be required for certain applications <15% increase in viscosity at 210◦ F/100◦ C
D-2782 D-2783 D-2266
>40 lb O.K. >200-kg Weld point >30.0 Load-wear index <0.8 mm scar diameter
D-2266
<0.1 Under conditions of the wear test Satisfactory for the application intended
General properties required include: oil —has moderate load-carrying capacity, oxidation stability, and water separation for the lubrication of bearings and gear drives in heavy-duty service. The service is subjected to large quantities of water, dirt, scale, heat, and boundary lubrication.
TABLE 18.26
Heavy-Duty Brake Fluid — SAE-J-1703
Property Viscosity at −40◦ F at 130◦ F COC flash point, ◦ F Boiling point, ◦ F Evaporation at 210◦ F/98.8◦ C Compatibility
ASTM method D-445 D-92 D-5307 D-5800
Value <8000 >39 >180 >375 <80% Shall conform to compatibility test as outlined in SAE Standard Procedure and must be completely miscible with similar named products
General properties required include: a nonmineral oil-type suitable for use in hydraulic brake equipment in industrial, automotive, and electric overhead traveling (EOT) crane equipment.
18.5.4 Air–Oil System This system is similar to the oil mist system except for the following factors (1) the oil is not atomized, it remains as oil and is propelled in the tubing by air, (2) the air is compressed air at 30 to 40 psi, (3) the system can be monitored up to the lubrication point for pressure and flow changes, and (4) the lubrication point is pressurized in the bearing-housing cavity to keep contamination out. These systems are being used on casters for the segment bearings and in rolling mills for the work roll and backup roll roller bearings. They have also been used to lubricate the steady rest babbitts on large roll grinders and to lubricate roller bearings on large fume extraction fans.
18.6 Plant Equipment — Auxiliary Services The auxiliary equipment in the steel mill includes: mobile equipment such as slab haulers, transporter, coil tractors, and coal scrapers; electrical rotating machinery; utilities systems for compressed air, water, steam, and gases; and material handling equipment such as overhead cranes, gantry cranes, and ore bridges.
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Steel Industry
18-21
TABLE 18.27 Water–Glycol Fluids Property
ASTM method
Viscosity Hydraulic pump test
D-445 D-2882
4-Ball wear test (40 kg. 1800 rpm. 130◦ F/54◦ C. 1 h) Fire resistance
D-2266
Suitable for the specified application Must pass the modified 100-h, 1900-psi test at 150◦ F (Vickers V-105-C-10 vane-type pump) with 10 gal. reservoir fill; <0.10% total wear; must maintain 1900 psi throughout test; no fluid or vapor-phase corrosion permitted <1.00 mm scar diameter
Rust prevention test
D-665
Must pass the spray flammability test as described in Schedule 30 of the U.S. Bureau of Mines Must pass
Value
General properties required include: these fluids —provide fire resistance and moderate wear protection. They must be compatible (to 200◦ F) with seals and packings of homogeneous Buna-N, fabricated Neoprene, Viton, and Teflon.
TABLE 18.28
Phosphate Ester Fluids
Property
ASTM method
Viscosity Hydraulic pump test
D-445 D-2882
4-Ball wear test (40 kg, 1800 rpm, 130◦ F/54◦ C, 1 h) Fire-resistance
D-2266
Suitable for the specified application Must pass modified 100-h, 1900 psi test at 150◦ F (Vickers V-105 C-10 vane-type pump) with 10 gal. reservoir fill; <0.010% total wear: must maintain 1900 psi throughout the test; no liquid or vapor-phase corrosion permitted <0.50 mm scar diameter
Rust prevention test
D-665A
Must pass the spray flammability test as described in Schedule 30 of the U.S. Bureau of Mines Must pass
Value
General properties required include: these fluids —provide fire resistance and excellent wear protection on hydraulic systems that are subject to fire hazards. Suitable for systems operating at high pressures (to 5000 psi) and high temperatures (to 225◦ F/107◦ C). Must be compatible (to 200◦ F) with seals and packings of Viton, Teflon, and homogeneous or fabricated Butyl.
18.6.1 Electrical Machinery Motors and generators are very low consumers of bearings whether it be oil film or roller bearings. Smaller motors have roller bearings that are grease lubricated or sealed with grease. Over lubrication, contamination, heat generation, and moisture are common problems. Lithium complex and polyurea soap technologies are the most common. Main drive motors of high horse power (hp) used in the rolling mills are oil lubricated with a circulating oil system and a hydrostatic lift arrangement to ensure that the babbitted bearings receive proper amounts of lubricant.
18.6.2 Utilities Systems The steel industry uses very large quantities of electricity of which about 20% is generated internally in the plants using steam or gas. Steam boilers, turbine generators, and gas driven generators are lubricated the same as the public utilities generating equipment. Blowers and compressors supply air to the plants for two main uses: cold air blast or wind for use in the blast furnaces to produce iron and compressed air for pneumatic devices and controls in the plant. The blowers are a form of turbine that can produce air volumes up to 250,000 cfm (117.9 m3 /sec) and 60 psi to deliver air to the blast furnace tuyere in the form of carbon monoxide for the iron ore to be converted
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18-22
Handbook of Lubrication and Tribology TABLE 18.29
Mill Utility Grease ASTM method
Property
Value
Penetration Roll stability — 2 h Dropping point Oxidation 100 h at 210◦ F Wheel bearing test Water washout test
D-217 D-1831 D-566 or D-2265 D-942 D-1263 D-1264
Grease mobility 4-Ball EP test 4-Ball wear test (20 kg, 1800 rpm, 130◦ F, 1 h) Timken EP test Field test
U.S. Steel method D-2596 D-2266
Suitable for the application; NLGI 0, 1, 2, 3 <25% >350◦ F <20 psi drop >5% loss at 235◦ F <5% loss at 100◦ F <10% loss at 175◦ F >0.01 g/sec or 0.10 g if pumpability is critical Load — wear index >30.0 Weld point >200 kg <0.60 mm scar diameter >40 lb Satisfactory for the intended application
D-2509
General properties required include: grease —has high-temperature, low-temperature, water-resistant, and extreme pressure properties for use in centralized pressure systems supplying lubricants for the operating conditions found on ball, roller, and plain bearings, including roll necks, and general purpose lubrication in the presence of large quantities of water where adherence to metal is most important. Also requires dispensing through bulk systems for plantwide use.
TABLE 18.30
High Temperature Grease
Property
ASTM method
Penetration Grease worker Roll stability — 2 h Dropping point Oxidation by oxygen bomb Wheel bearing test, 6 h at 235◦ F Grease mobility
D-2177 D-217 D-1831 D-566 or D-2265 D-942 D-1263 U.S. Steel method
4-Ball wear test (7.5 kg, 1800 rpm, 130◦ F, 1 h) Field test
D-2266
Value 0, 1, 2, 3 <10% change in 10,000 strokes <15% change >500◦ F <10 psi drop <2% loss >0.1 g flow per second at 0◦ F if pumpability is critical >0.6 mm scar diameter Satisfactory for the intended application
General properties required include: grease —has high-temperature stability where bearings operate above 250◦ F requiring lubricants with dropping points above 500◦ F. High-temperature grease is used for ball and roller bearing lubrication covering a wide range of conditions, such as exposure to water, high temperature, low temperature, shearing, oxidation, and rust. These conditions are illustrated by fan motor bearings, furnace-car wheel bearings, annealing furnaces, drying ovens, sintering plants, and soaking pits where the lubricant must serve for long periods under severe conditions without being replenished.
into pure iron. Steam turbine driven turbo blowers have the same type of circulating oil system as the steam turbines and generators. They come with bearing, pressure, and temperature monitoring devices. Compressed air generation is done by reciprocating piston units, screw type, and turbine wheel design units. Air compressor lubrication is always a point of concern with regard to fires from overlubrication and the use of oils that deposit carbon on valves. Naphthenic oils as well as synthetic oil in the form of PAOs, silicones, and diester are used to extend drain intervals and improve fire resistance. There are many gas compressors used to move argon, nitrogen, and oxygen in the plant. These units for the most part are
© 2006 by Taylor & Francis Group, LLC
Steel Industry TABLE 18.31
18-23 High Temperature EP Greases — Complex Soaps and Nonsoaps ASTM method
Property Penetration Motormatic grease worker 10,000 strokes Roll stability — 2 h Dropping point Oxidation by oxygen bomb Wheel bearing test, 6 h at 235◦ F Grease mobility 4-Ball wear test (7.5 kg, 1800 rpm, 130◦ F, 1 h) Field test
D-217 D-217 D-1831 D-566 or D-2265 D-942 D-1263 U.S. Steel method D-2266
Value Suitable for the application. NIG <10% change <15% change >500◦ F <10 psi drop <2% loss >0.l g flow per second pumpability is critical >0.60 mm scar diameter Satisfactory for the intended application
General properties required include: greases —have high-temperature and pressure properties for bearings which operate above required EP lubricants with dropping points above 500◦ F. General applications include: ball and roll lubrication covering a wide range of conditions, such as exposure to water, extreme pressure, high temperature, low temperature, shearing, oxidation, and rust.
TABLE 18.32
Molybdenum Disulfide Grease ASTM method
Property Penetration Filler Falex wear test 500 lb load — 1 h Rust prevention Water washout test at 175◦ F Plastic plate abrasion Field test
D-217 D-2670 D-1743 D-1264 D-1404
Value Suitable for the application. NLGI 0, 1, 2 >2% Molybdenum disulfide (commercial technical fine grade) <20 teeth wear Pass 24 h <15% Loss <50 Scratches Satisfactory for the intended application
General properties required include: grease —reduces friction; adheres to metal surfaces; has limited loadcarrying capacity; and resists heat in slow-speed plain-bearing and sliding-surface applications operated under marginal or boundary lubrication conditions. Not generally recommended for antifriction bearings.
TABLE 18.33
Roll Neck Grease
Property
ASTM method
Value
Oil viscosity
D-445
>75 SUS at 210◦ F; up to 150 SUS at 210◦ F preferred when
Penetration Drop point Timken test 4 Ball EP test 4-Ball wear test (5 kg, 1800 rpm, 130◦ F, 1 h) Grease mobility test Field test
D-217 D-566 or D-2265 D-2509 D-2596 D-2266
pumpability is not critical NLGI Grade 0, 1, 2 >200◦ F >20 lb LWI >20.0 load-wear index Weld point >150 kg >0.60 mm scar
U.S. Steel method
>0.1 g flow per second at 0◦ F where pumpability is critical Satisfactory for the application intended
These greases require adhesive film, water resistance, mild EP, and pumpability properties for applications in blooming-mill, billet-mill, slabbing-mill, bar-mill, plate-mill roll-neck bearings of the fabric, bronze, babbitt, or combination and segmented types. These bearings are subjected to large quantities of water and boundary lubrication. The method of application is by centralized lubricating systems. In several cases these systems serve mill screws and nuts, table roll bearings, spindle bearings, etc.
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18-24
Handbook of Lubrication and Tribology TABLE 18.34
Journal Roller Bearing Grease — AAR Specification M-942-92
Property Corrosion Penetration Dropping point Oxidation stability psi drop in 200 h, max psi drop in 500 h, max Structure stability Max. increase in penetration at 77◦ F after 100,000 double strokes in standard grease worker Moisture
ASTM method D-1743 D-217 D-566 D-942
Value 1 Acceptable 2 Rejectable 290–320 325◦ F (min) 10 25
D-217 +25, −25 D-128
0.10% (max)
General properties required include: grease —requires adhesive film, water resistance, mild EP, and pumpability properties for the following applications: a. For Association of American Railroads (AAR) approval, the grease shall satisfactorily lubricate each of the freight car roller bearing designs during an 8-week simulated service test at the AAR Central Research Laboratory. b. The grease must maintain stable consistency in the bearing assemblies of >270 or <340 penetration at 77◦ F during the 8-week accelerated test. c. In addition, when 50% of the test grease is mixed with the remaining 50% composed of equal proportions of all other AAR types of approved journal roller bearing greases as identified by the metal base, it shall meet the requirements set forth in Sections 3a and 3h during an 8-week accelerated service test. Please note that Bentones are prohibited for use as a thickener, viscosity improvers are also prohibited, while extreme pressure additives can be utilized in the grease formulation.
reciprocating designs that require a lubricant that is fire resistant and may use a grease that does not have any petroleum products in it. The utilities in the steel plants include water systems and the treatment of water prior to going to the mill and upon return to the treatment center to remove contaminants such as scale and heat. Most steel plants are equipped with clarifiers and cooling towers. The large submersible pumps and centrifugal pumps with flow rates varying up to 40,000 gpm are lubricated as they would be in any facility, either by a circulating oil system or a grease system. Many water systems use standard API centrifugal pumps that are grease or oil lubricated. In hot rolling areas, water-cooling systems for the strip and rolls are supplied water from large single stage centrifugal pumps. High pressure descale water systems (up to 3,500 psi operating pressure) in the hot mills use multi-stage centrifugal or in some locations barrel pumps to provide water to the descale headers and nozzles on the rolling mills. These pumps can have turbine or electric motor drivers with horsepowers into the 2,000 to 5,000 hp ranges. The pumps may have oil film bearings or roller bearings that are grease or oil lubricated depending of the design.
18.6.3 Mobile Equipment The trucks, transporters, off road equipment, slab and coil haulers, coal scrapers and bulldozers, and other earthmovers as wells as coil tractors and rail equipment such as diesel engines and freight cars are lubricated according to Society of Automotive Engineers (SAE), API, and Association of American Railboards (AAR) guidelines and standards.
18.6.4 Maintenance Shops Although the recent trend is to do away with in-house maintenance facilities such as machine shops and weld/fabricating shops, many steel mills still keep shops with machine tools, grinders, and other machinery to service the plant. This machinery is lubricated according to the vendors’ requirements.
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18.6.5 Hydraulics Fluid power is used throughout all processes in the steel plant with systems generally in the 1800 to 2000 psi pressure range. Some systems such as for AGC (automatic gauge control) cylinders are in the 5000 psi operating range. In any location where heat and possible fire, fire-resistant fluids are used. These include polyol esters, polyol ethers, 95/5 high water base fluids (HWBF), water glycols, and phosphate esters. Fluid cleanliness and resistance to oxidation and varnishing are important properties in the oils.
18.6.6 Overhead Cranes Steel plants have hundreds of overhead electric traveling cranes. Their capacities range from 1 t maintenance units to 600 t capacity cranes used in the melt shop to move molten steel in ladles. The cranes in the melt shop and caster areas will have a primary hoist arrangement with a spreader beam and hooks to lift the ladles and scrap boxes. These multi-rope hoists can lift up to 600 t. On a separate trolley above this main hoist is a second hoisting arrangement, typically with a 100 to 150 hook block and a 30 to 50 t capacity secondary block for maintenance work. All of the hoists are controlled from one cab. Centralized grease systems are installed in the cranes as much as possible. AP bearings are being used for the crane wheel assemblies. These bearings are sealed and come prepacked with grease for the life of the bearing or the wheel assembly. Wire rope lubrication is an important aspect of the crane. Ropes are manufactured now with the lubricant in the rope structure so as to prevent any mess on the floor or fires. Double and triple reduction helical and spur gear gearboxes driven by electric motors and brake arrangements are used for the bridge and trolley drives. Theses type of gearboxes drive the hoisting drums. These boxes have their own lubrication systems, either by a built-in pump and spray arrangement or a bath set-up. EP gear oils with viscosities ranging from 220 up to 680 cSt are typically used for lubricants. In some mills where there are wide seasonal temperature swings and the duty cycle of the crane is such that they are required, synthetic EP gear oils from the PAO and PAG base stocks are being used.
18.7 Iron Making Equipment In the iron-making process of steel making, iron ore, sinter, limestone, coke, iron, pellets, air, and heavy fuel oil are combined in a blast furnace to produce pure iron. As the burden of raw material descends the furnace height, it meets preheated air from the stoves in the tuyeres. Molten iron and slag form in the furnace hearth and are tapped after every specified number of hours.
18.7.1 Coke Manufacture The first step in iron making is to produce coke from metallurgical coal. The bituminous coal comes via barge, ship, and rail to the plant. It is deposited into storage piles with car dumpers, conveyor systems, coal/ore bridges, and stacker/reclaimers. Coal is then crushed to size using coal crushers and hammer mills. It is discharged via a larry car into the ovens for the coking process to begin. The coal is then heated to 1800◦ F for a cycle time of 18 h. After that it is pushed into a self-propelled quench car by a pusher machine after the oven doors are opened with a door machine. The water-quenched coke is dumped onto a coke wharf that feeds a conveyor system to move the coke to the blast furnace. The coking process is shown in Figure 18.2. The coke oven gas and by-products from the coking process are gathered by turbo-exhausters and sent to the by-products plant for cleaning and separation. The principle products are naphthalene, benzene, xylene, toluene, CO, H2 , CH4 , H2 S, and NH3 . The by-products plant is piping systems, turbo-gas exhausters, gas compressors, gas cleaning and separating systems, and chemical pumps of API design. The process is shown in Figure 18.3 to give an indication of the types of chemicals that result from the coking process.
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FIGURE 18.2 Schematic representation of the sequence of operations involved in charging, leveling, and pushing in one coking cycle of a by-product coke oven. (Printed with permission, Association of Iron and Steel Technologies, The Making, Shaping and Heat Treating of Steel, 10th edition, p. 157.)
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FIGURE 18.3 Flow sheet showing the major steps involved in the carbonization of coal by the by-product process and the subsequent recovery of coal from the gases generated at the ovens. (Printed with permission, Association of Iron and Steel Technologies, The Making, Shaping and Heat Treatment of Steel, 10th Edition, p. 231.)
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FIGURE 18.4 View of a steam-turbine-driven axil-flow blast-furnace blower, with upper-half of casing removed. Vaned rotor of air compressor is at left, steam turbine rotor at right. (Courtesy, Ingersoll-Rand Comp.)
Fine particles of coal and coke breeze are very abrasive. The heat of the process and the corrosive nature of the coal chemistry by-products means that special attention has to be paid to the lubricants used and their properties. Larry cars and the door machines and pusher machines all have hydraulic systems that use fire resistant fluids.
18.7.2 Limestone and Iron Ore Car dumpers and ore bridges unload railroad cars, trucks, and ships. Stacker/reclaimers load conveyor systems to take the materials to the blast furnace. Most reclaimers contain large slewing ring bearings. The mechanical systems are generally lubricated with grease that has low temperature pumpability properties for centralized grease systems. Gearboxes generally have pumping systems built into the design of the box. Because of the abrasive dust, water from being outside, and other environmental conditions, wire rope lubrication is critical. It is usually built into the design of the strand structure so as to minimize contamination. Many plants have sinter machines that take fine ore dust and agglomerate it into a sinter using limestone as the binder. Sinter dust in gearboxes and pallet wheel bearing being subjected to high temperatures of 300 to 500◦ F are common problems. Greases with a high temperature dropping point are required. Pellet plants that form briquets from the sintered ore at the ore mine site have similar problems.
18.7.3 Air The air required to produce iron is moved using turbo blowers that are driven by steam turbines, electric motors and gas engines. It takes 35,000 to 60,000 ft3 of preheated air from the stoves to make 1 t of iron. The turbo-blowers (Figure 18.4) can supply up to 250,000 ft3 /min of air to the furnace. The bearings used on these machines can be exposed to temperatures of up to 250◦ F. The bearings are generally oil film bearings with Babbitt inserts and Kingsbury type thrust bearing. The bearings are lubricated with inhibited turbine oils in a circulating oil system. The units have G1 type dynamic balancing done to the rotor assemblies and have continuous vibration and temperature monitoring instrumentation. The hydraulic control systems use phosphate ester fluids usually for the fire resistance.
18.7.4 Blast Furnace An overall view of the blast furnace iron-making process is shown in Figure 18.5. The blast furnace is a truncated cone with hoppers at the top for receiving ore, coke, limestone, pellets, sinter, and with series
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FIGURE 18.5 Flow diagram depicting the principal units and auxiliaries in a modern blast-furnace plant, and showing the steps in the manufacture of pig iron from receipt of raw materials to disposal of pig iron and slag, as well as the methods for utilizing the furnace gases. (Printed with permission, Association of Iron and Steel Technology, The Making, Shaping and Heat Treating of Steel, 10th Edition.)
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of air intakes or tuyeres around the circumference at the bottom to introduce air that will turn into CO in the furnace. The iron ore is reduced as it makes its way down the furnace stack. Some iron is produced using a direct reduction method using natural gas and coal to reduce the iron ore to pure iron. This process produces smaller quantities of iron per annum with a greatly reduced capital investment in machinery. Today’s furnaces can produce up to 10,000 t of molten iron per day. The new furnaces use conveyor belt systems to move the raw materials to the top and into a rotating hopper that distributes the materials into the furnace burden. The older furnaces used a skip car and hoist arrangement to move carloads of materials up to the top and into the large bell- and small bell hoppers. The skip cars and bells were moved using an array of wire rope cables that were fed from a hoisting room using the same technology as found in a mine hoist arrangement. A detailed layout of a typical blast furnace is shown in Figure 18.6. The furnace stack can be up to 150–200 ft high and 30–50 ft in diameter at the bottom where the hearth is located. The furnace is basically refractory lined steel shell that has water-cooling staves on the outside to reduce heat generation on the shell. Stoves heat the air being supplied to furnace using coke oven and blast furnace gas. The furnaces are equipped with automatic central grease systems to deliver grease to all of the bearings, sheaves, and valves using and NLGI 0 or 1 grade grease. The skip hoists are lubricated via a circulating oil system using 460 or 680 cSt EP gear oil. In some locations high viscosity open gear lubricants are used instead of gear oil. Fluid power systems are being used more and more to control the air and raw material systems in the blast furnace. This includes: blower controls, cold blast main butterfly valve movements, hot stove changes and snort valve, furnace top (bleeder valves, stock hoppers, and chutes), mud gun for plugging the tap hole, tap hole drills to tap into the furnace hearth (similar to a mining drill), gas tap pressure controls, and dust catcher system isolation valve. The systems use polyol ester, water glycol, or phosphate ester fluids at 1800 to 3000 psi operating pressures. Molten iron and slag after tapping flow through ceramic troughs into hot metal ladle cars or torpedo cars as they are known in the industry at ground level. Car have up to 32 wheels that run on standard gauge rail to hold up to 600 t of metal. Car bearings are generally AP type tapered roller bearings. The cars are equipped with a gearbox and drive to rotate the ladle in the melt shop. Plain bearings made from bronze are used to hold the ladle for rotation.
18.8 Steel Making Equipment Most melt shops for producing steel from the iron and scrap use some form of BOF or Basic Oxygen Furnace to produce up to 350 t of steel in a 45 min tap to tap time. Figure 18.7 shows a typical shop layout with two vessels. The furnace is a refractory lined steel shell that is fixed into a trunnion ring for rotation from the charging to tapping positions. An oxygen lance is brought down above the bath of molten iron and scrap to create a reaction using O2 gas. Some furnaces have a tuyere system in the bottom to also feed argon into the bath for bottom stirring. Molten iron is transferred from the torpedo cars into large iron ladles. These ladles are round steel fabricated shells with a pouring spout on them. The ladles are refractory lined. On the outsides of the ladle shell are trunnion pins with wear sleeves for the overhead crane to lift the ladle and pour the molten iron into the furnace. These ladles can hold up to 150 t of metal. Scrap up to 150 t, is then added by crane to the furnace from a scrap box. The trunnion ring and vessel are supported on large diameter bearings and pillow blocks (Figure 18.8). The bearings used are traditionally spherical roller bearings with tapered adapter sleeves. These bearings can be in the 2.5 to 3 ft diameter range. Main of the pillow blocks incorporate a ladder bearing (two hardened plates with rollers in-between the plates) under the block to take up thermal expansion (Figure 18.9). Hydrostatic oil film bearings and spherical plain bearings using Teflon pads (“hockey Pucks”) have also been used. The bearings are lubricated with lithium soap greases that may contain or not MoS2 or graphite solid lubricants. The vessel and ring are rotated through a 270◦ arc using a large diameter
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FIGURE 18.6 Idealized cross-section of a typical modern blast-furnace. Details may vary from plant to plant. (Printed with permission, Association of Iron and Steel Technologies, The Making, Shaping and Heat Treating of Steel, 10th Edition, p. 545.)
helical or herringbone type bull gear and four pinions that are each driven by a double to triple reduction gearbox. Typical installations use four 1000 or 1200 rpm, 150 hp electric motors to drive the gearboxes and pinions so that final rotation of the vessel is 1.5 rpm in its fully loaded condition. A circulating oil system using EP gear oils in the 680 to 1000 cSt viscosity range supplies lubrication to all of the gearing and bearings.
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FIGURE 18.7 Schematic elevation showing the principle operating units of the basic oxygen process steel making shop. The storage bins contain: (i) limestone, (ii) fluospar, (iii) ore, (iv) calcitic lime, and (v) slobomitic lime. (Printed with permission, Association of Iron and Steel Technologies, The Making, Shaping and Heat Treating of Steel, 10th Edition, p. 606.)
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FIGURE 18.8 BOF Bearings. (Printed with prmission, NSK, Large Bearing Catalogue, p. 26.)
Electric arc furnaces are being used more for producing steel directly from scrap. These furnaces consist of a steel shell that is refractory lined. It is water cooled. The furnace is loaded with scrap from the top and then a steel lid that is refractory lined and water cooled is put over top of the shell. Large carbon electrodes are brought down above the bath and high voltage electricity is used to produce steel. This method traditionally was only used to make stainless steels or tool steels. Nowadays, they are also being used for carbon steels. It takes about two hours to produce between 150 and 250 t of steel. The furnace is rotated using a rack and pinion or hydraulic actuators to a tapping position. The steel produced by the two methods is moved in large ladles by overhead cranes to ladle metallurgy stations. Here the steel is refined further using argon and nitrogen in a vacuum and the additives are properly added for the grade of steel required. All of the furnaces have water cooled hoods and large fume and gas extraction systems to collect and clean the dust and gases from the process. There are also baghouses and other environmental devices on the furnaces. These systems have large centrifugal fans that may be a single unit with a direct drive 6000
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FIGURE 18.9
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Laddor Bearings. Printed with permission, NSK, Large Bearing. Catalogue, p. .
to 8000 hp electric motor or two or three smaller fans with 2000 to 3000 hp drives. The motors have film bearings that are ring or disc lubricated with oil. The fan bearings may be oil film bearings that are ring lubricated with oil or fed oil from a circulating oil system. Spherical rollers, split cylindrical roller bearings, and the SKF CARB bearings are used for many of the smaller fan installations. The bearings are grease lubricated, circulating oil fed, oil mist, or air–oil lubricated. The furnaces also have extensive water systems with API type process pumps.
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Ladleman’s platform
Tundish pre-heat
Ladle turret
Ladle
Tundish car Control room
Dummy bar Casting floor Spray chamber Cooling bed
Mold
Cross transfer
Strand guide
Pre cut-off table Shear cut-off Discharge table
FIGURE 18.10 Billet casting machine anatomy. (Printed with permission, Association of Iron and Steel Technologies, The Making, Shaping and Heat Treating of Steel, Casting, p. 5.)
18.9 Steel Shaping — Casting The molten steel is poured into ingots and allowed to solidify prior to transfer to the hot mill for shaping in the rolling mill. Nowadays, most of the initial shaping is done by continuous casting machines. The continuous casting of billets, slabs, and some shapes is now the preferred process instead of ingots, pour floors, stripper cranes, and soaking pits. The casters receive the molten steel in ladles weighing 350 to 500 t. The ladle is positioned in a two-position ladle turret over the casting machine. Figure 18.10 and Figure 18.11 show typical caster shop layouts with the types of machinery in the shops for billet and slab casting. The ladle turret has a large diameter triple race slewing ring (Figure 18.12) to position the new and empty ladles. These bearings may have 15 to 25 ft diameters. The bearing has an internal spur ring gear and pinion to rotate it at 1 to 2 rpm. The driver can be a hydraulic motor or electric motor/gearbox arrangement. The bearing and gearing are grease lubricated with a heavy viscosity EP NLGI 1 to 2 grease with a 460 to 680 cSt base oil. The steel flows from slidegates on the bottom of the ladle into the refractory lined tundish that directs the steel into a mould beneath. The flow rate matches the rate of solidification of the steel leaving the mould. The steel is drawn through a series of segments with roller assemblies top and bottom to shape the steel into a slab or billet. This is usually curved. At the bottom of the curve, the slab goes onto a straight section of tables with roller assemblies. Through the segments, the solidifying steel is sprayed with water to enhance the cooling. The copper lined mould may be vibrated or stirred to ensure no impurities form in the steel. The slab is cut with torches on an automatic machine at lengths to be fed to the rolling mill. Most casters create slabs in the 8 in. thick range. Newer casting technology calls for thin strip casting where the caster produces a 2.5 in. thick slab that is fed directly to the rolling mill via a tunnel furnace to maintain the temperature. . The caster segments consist of roller assemblies with spherical roller bearings and split spherical or cylindrical roller bearings with water cooled housings. Depending on the location of the roll arrangement in the caster, the rolls may be water cooled in the center. The bearings have C3 or C4 internal clearances for the thermal expansion (Figures 18.13 to Figure 18.15). The bearings are lubricated with a NLGI
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Tundish pre-heat Tundish car
Tundish
Ladle turret
Ladle lid manipulator Ladle
Ladleman’s platform
Ladle shroud manipulator
Casting floor Spray chamber Segment removal guides Cut-off table Cast product storage area
Strand guide Discharge table
Segment manipulator
Strand guide segment Pre cut-off table Cut-off Dummy bar receiver
Dummy bar
FIGURE 18.11 Slab casting machine anatomy. (Printed with permission, Association of Iron and Steel Technologies, The Making, Shaping and Heat Treating of Steel, Casting.)
FIGURE 18.12
Slewing ring bearings — most common.
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Segment Type Roll Type
Roll Dia Roll Pitch Segment inch inch Profile
Number Off
FIGURE 18.13 Strand segmental arrangement. (Printed with permission, Association of Iron and Steel Technologies, The Making, Shaping and Heat Treating of Steel, Casting p. 68.)
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Solid roll, dead stub shaft, no internal cooling A
A
Rotating, cooled center shaft, separate sleeves B
Solid roll, center bore cooled, separate barrels D
C
B
D
Solid roll, center bore cooled, split center bearing
C
E E
FIGURE 18.14 Split roll types. (Printed with permission, Association of Iron and Steel Technologies, The Making, Shaping and Heat Treating of Steel, Casting p. 72.)
grade 1 or 2 grease with a 400 to 460 cSt base oil from a centralized greased dispensing system. Many casters are being converted to an air-oil system to reduce the amount of lubricant consumed and keep the water and caster cleaner. The oil used is a 460 to 680 cSt viscosity EP gear oil or paper machine oil. The bearings traditionally can be reused for two to three campaigns in an air–oil system when compared to one to maybe two on a grease system. The single stand slab caster may have up to 1300 bearings to be lubricated. Aluminum complex grease has been the preferred grease for many casters on grease lubrication. Fluid power systems are on each of the segments to control the applied pressure on the roller assemblies in the curved section of the machine. Large diameter cylinders are used to lift the turret and ladle into position. Fire resistant fluids are used in pressures up to 3000 psi. The level of filtration of the fluid is kept below 10 µm.
18.10 Rolling Processes and Components 18.10.1 General — The Mills After the molten steel has been poured into ingots or moulds made from cast iron that sit on stools on railway cars and allowed to solidify, the ingots are stripped from the moulds. The ingots are sent to soaking pits to reheat the steel to 2400◦ F prior to rolling. The soaking pit machinery such as the covers and lift mechanisms require good EP grease with high temperature properties. Most pit cover transfer mechanisms have centralized lubrication systems.
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FIGURE 18.15 Printed with permission NSK, Large Bearing Catalogue, p. c2
Once the ingot is heated to temperature, it is removed from the pit by crane and put onto a table of rollers for transfer to the primary breakdown mill, better known as a slabbing mill or blooming mill depending on the product rolled. This mill starts the hot rolling process to roll the 20 to 40 t rectangular ingot into 4 to 16 in. thick rectangular cross-section and widths between 36 and 100 in. The smaller the section, the longer the piece and the more rolling passes required to achieve the cross-section. The rolling
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process involves squeezing the hot plastic steel ingot between a pair of rolls. The rolls grip the ingot and by contact friction pull the section through in a reshaped section. After each pass, the tables and rolls reverse direction as the rolls are brought closer together. As the steel reduces in size, it elongates until the final thickness is achieved. A slabber has two vertical rolls or edger rolls to take side reductions and two horizontal rolls. It rolls all four sides simultaneously to produce a rectangular cross-section. Blooming mills roll blooms (square cross-sections). These mills have manipulating equipment to tilt and turn the section to achieve the square required. Slabs are further processed into plates, strip, sheet, tin plate, and galvanized products. Blooms are further worked into structural shapes, rails, skelp, and billets. Skelp and billets go on to become rod, small pipe, and wire. The processing of the steel in this stage uses a variety of rolling mill designs and auxiliary machinery. When the steel is continuously cast, this entire part of the process is made redundant. It is a major energy, manpower, and capital investment cost on spare parts, machinery, and maintenance savings to a steel plant. Plate mills receive the reheated slabs and roll plates up to 200 in. wide. There are various mill designs with some being a 4-HI reversing steckel type design or a 2-HI reversing rolling mill and then a 4-HI 1 finisher. The mills can put up to 10 to 12 million lb of force on the plate to produce plate down to 16 in. thick. Bar and billet mill are smaller and faster, taking the blooms and processing them through a series of 2-HI roll stands (up to 22 stands in a line) in both the horizontal and vertical direction into small sections. These mills may operate at up to 4000 fpm and produce coils or straight bundles. Rod mills are similar to bar mills but with smaller rolls that can achieve up to 22,000 fpm. The product ranges from 0.218 to 0.500 in. diameter. A series of mill stands with contoured roll bodies and equipped with roller guides that can operate at speeds up to 80,000 rpm produce coils. The Morgan No-Twist mill is a special design rod mill for high productivity rolling. These mills use oil film bearings and special roller ball bearings with circulating oil systems for lubrication. Seamless tube mills take solid round billets that are heated and passed through a piecing mill to produce hollow seamless tube. The tube is then run through a series of mills to elongate and reduce the outside and inside diameters. Pipe mills take the narrow plate skelp and using heating furnaces and a series of forming rolls and electric butt welding or friction welding techniques produce pipe. Structural and rail mills process the reheated blooms through a series of stands, both in the vertical and horizontal direction to create wide flanges, angles, channels, etc. Hot strip mills process about 50% of all the steel tonnage produced. Figure 18.16 shows the general layout for a typical strip mill set-up. The new thin strip mills will not have the number of breakdown stands or roughers. The 2.5 to 1 in. thick transfer bar comes directly from the caster through a tunnel furnace to maintain the temperature of the bar and then into the 6 or 7 stand rolling mill. The slabs from the caster or slabbing mill are reheated and then rolled through a series of rolling mill stands. Many older mills have a 2-Hi scale breaker, several 4-Hi roughers and a 5 tp 7 stand 4-Hi finishing mill. Figure 18.17 gives one the indication of the water, heat, and speed present in a 7-stand 4-HI hot strip mill with its interstand looper and side guide equipment as well as the hydraulic systems for bending, side shift, and automatic gauge control. Slabs up to 14 in. thick are reduced to 0.47 in. thick. Delivery mill speeds from the finishing stands can exceed 4000 fpm. The product is coiled over rotating expanding mandrels afterward to be cooled. Cold rolling tandem mills (Figure 18.18) take the hot bands as they are known from the hot mill after pickling and oiling to remove rust and oxidation products on the surface and cold roll the strip to 0.010 to 0.015 in. thick in sheet mills and 0.008 in. thick in tin mills. These 4 or 5 stand mills use 4-Hi (two backup rolls and 2 work rolls per stand) and 6-Hi (two backup rolls, two intermediate rolls, two work rolls per stand) stack technology at speeds of 4000 to 6000 fpm and rolling loads of 2 to 3 million pounds separating force per stand. Sendzimir mills are single stand, high-speed reversing mills using small diameter rolls that are stacked together and baked up with a cluster of
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Steel Industry FIGURE 18.16 Hot strip rolling mill. Typical reductions per pass in the finishing stands of a hot-strip rolling mill equipped with four roughing stands and six finishing stands. (Printed with permission, Association of Iron and Steel Technologies The Making, Shaping, and Heat Treating of Steel, 10th Edition, p. 1075.)
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FIGURE 18.17 Technologies.)
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Typical hot mill finishing stands. (Printed with permission, Association of Iron and Steel
backing rolls to produce very close tolerance stainless steels. Rolling oils are required on all cold mills to control the friction at the roll bite and cool the mill rolls. Twin Stand double reduction mills are used to reduce the product from the tinplate cold mills to an even thinner gauge — 0.003 in. The mills operate at 5000 to 6000 fpm using a 4-HI mill stack technology. The product is annealed prior to is this process to relieve the strip of stresses from cold rolling. Single stand and two-stand temper mills temper the surface of the steel to increase the hardness and elongation of the strip. These 4-HI rolling mills operate up to 7,000 fpm and use separating forces of 500,000 to 750,000 pounds force per stand.
18.11 Mill Components 18.11.1 Reheat Furnaces There are several designs of reheat furnace that take the slabs and using natural gas or coke oven gas heat the slab up to 2400◦ F. These furnaces are a continuous operation with cold slabs going in one side and heated units coming out the other. The furnaces have hydraulics systems to move the slabs. They also have combustion air fans to move the air in the furnace.
18.11.2 Mill Tables The tables are a series of rollers that support and direct the hot steel sections to each mill stand with such names as approach and runout or delivery depending on the position in the mill. The rollers are 10 to
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FIGURE 18.18 52 -wide, 6 strand Tandem cold mill (Printed with permission, Association of Iron and Steel Technologies, The Making, Shaping and Heat Treating of Steel, 10th Edition, p. 1106)
20 in. diameter and are supported by spherical roller bearings on the ends. AP bearings have been used in some mills. The rolls can be direct driven by electric motors. More common is the use of chain drives for a series of rolls or a motor, gear reducer and a line shaft, and bevel gears. Lubrication of the bearings and the gearing and chain can be severe. Roller chain is lubricated in a bath. Bevel gears are grease lubricated or air–oil lubricated. The bearings are grease lubricated or air–oil lubricated from a centralized dispensing system. The properties of the lubricants used and the amount of lubrication present are critical to the life of the components. Temperatures up to 2000◦ F in radiant heat is seen by the bearings and the mill stand if there is a delay with a bar on the table. Cooling water, deflected mill scale (ferrous oxide), and shock forces also affect the bearing and drive performance.
18.11.3 Manipulators, Entry, and Side Guides This machinery positions the steel between passes prior to entry into the roll bite. High impact forces, water, mill scale, heat, and vibration affects the operation of the bearings, pins, and bushings drives and any wear liners or slideways present. For the most part all components are lubricated with a grease suitable for the mill environment from a central dispensing system.
18.11.4 Cooling Beds — Transfer Tables These roller tables move the steel product from the mill bay to further producing such as shearing, hot or cold saw, leveling or packaging. Linkages, roller or link chains, and pins and bushings on the tables are subjected to intense heat up to 400◦ F. Solid lubricants with graphite and molyboenum are used in these applications.
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18.11.5 Mill Stand Complex Mill Motors come in varying sizes, ranging from 8,000 to 10,000 hp direct drive dc motors in a slabber or rougher to 10 hp direct drive on runout table rolls. Large mills may use motors in tandem per stand to achieve 10,000 to 14,000 hp per stand. Many motor rooms have MG sets to provide the DC power required. These generators are lubricated by oil rings and flood lubrication with the speeds ranging from 360 to 3600 rpm. Main drive mill motors operate from 10 to 600 rpm and can be reversing in rotation. They trip out at 275% overload and routinely deliver 225% overload. Continuous drives can supply 125% overload for 2 h and trip at 175% overload. Oil filtration, sight flows, flow switches, thermocouples and the use of both rings or discs and flood lubrication, and circulating oil systems are common practices with oil film babbitted bearings. Hydrostatic lubrication at pressures up to 5,000 to 8,000 psi are also common to prevent slow-speed and start-up problems. All bearings are grounded to prevent stray current discharges through the bearings and oil. Fully inhibited R & O or antiwear type oils are used and sampled on a monthly basis for oxidation, contamination, and water. Mill drives or gear reducers not associated with all blooming, slabbing, and high-speed stands, utilize them ahead of the mill pinions. Two- and three-high mills use reduction gears and sometimes flywheels. Reversing mills spray oil on each side of the gear mesh. Continuous mills such as hot strip, cold strip, bar, billet, rod, and wire spray oil on one side only. Mill drives are on large circulating oil systems. The pinion drive gear smaller by comparison to the driven or bull gear and is hardened for extended tooth life. Mill drives for the most part are found in the motor room away from the water, dust, heat, and scale of the mill environment. EP gear oils are the most common type used in the reducers, although straight oils do show up in some applications. Both synthetics, especially PAO and PAG types are common instead of mineral type chemistries. Gear teeth are single helical, herringbone (double helical), or spur design. Through hardened teeth to AGMA quality 6 or 7 was the most common. Current technology is calling for case carburized and tempered gear teeth that are ground or hard finished to AGMA quality 11 or 12. This new technology increases the life of the tooth and increases the horsepower capacity of the teeth. Many gearing systems have vibration and temperature monitoring equipment mounted on the boxes. At the same time, monthly oil samples are taken and tested for oxidation, contamination, water, and TAN. Mill Pinions are mounted in a separate housing close to the mill spindles (Figure 18.7). They are usually a herringbone design with a 1:1 ratio and a single input and double output to the spindles. The spindles are driven at the same speed. Mill pinions are subjected to high impact forces from the strip entering the roll bite in the stand, thus requiring sufficient beam strength and surface durability. Many pinion gear arrangements will last for 20 to 25 yr if properly designed. The same AGMA specifications as in the reducers are used for the pinions. High viscosity EP type oils with up to 680 cSt viscosity are sprayed on the incoming mesh. A rule of thumb of 0.75 gpm/in of face with loads up to 10,000 lb/in. is used for the oil flow rates in a circulating oil system. Babbitt bearings have been used for the mill pinions as well as the reducers. Loads are generally below 1,000 psi. New installations use spherical roller and two-row tapered roller bearings. These bearings are fed oil from the same oil system as the gear mesh. Mill Spindles and Couplings are located between the mill stand and pinions. The main drive couplings are usually a gear type design, sometimes with a stub shaft or torque tube sometimes. Spindles can be a gear design, spade, and yoke (slipper type), or more recently universal joint type design. Lubrication of the couplings and the spindles is a challenge. The rolls are rotating and the top roll is moving up and down in the housing according to the roll bite requirements for gauge and shape. At the same time, water, scale, flying debris, and torque windup is occurring on the spindle. Slippers are generally made from a bronze such as manganese type. The geared spindle is enclosed and prepacked with grease. Water or coolant washout is a problem, so the spindle couplings are relubricated every 2 to 3 weeks. The slipper type is open or shrouded, but since lubricant is generally thrown out centrifugally from the point of lubrication, lubrication is a problem. The following techniques are used: oil spray off main oil system, spindle mounted lubricator, oil mist system through rifled holes, lubricant filled boot, drip oil feed, and
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manual. Varying speeds, angularities, torque transmitted, and space considerations will decide the proper application method. Mills with high angularity of the spindle and low speeds will use a slipper type or universal joints. Low angularity applications generally use geared spindles. In some applications, to minimize torque amplification, resilient elastomers are designed into the coupling. The Mill Stand or Roll Stand requires lubrication to the housing windows through which the roll bearing chocks and rolls must slide. Most mills now employ hardened wear liners on the chock side liner areas and the mill housing windows to control wear and maintain the position of the chocks during rolling operations. For the most part, grease that is water-wash-off resistant is fed via an automatic dispensing system to a series of grease grooves on the liners. At the same time, all slideways and guideways must be lubricated to ensure minimal corrosion and wear in order to reduce alignment problems. Mill screwdown arrangements as shown in the mill cross section in Figure 18.19, supply the roll separating forces against the top roll assembly are a combination of an acme power screw and worm gear arrangements. The hardened steel screws — 12 to 18 in. diameter — with bronze nuts extend up through the mill housing to the drive arrangement. The screws have high sliding velocities and high loads on the threads. EP gear oils with heavy base oil viscosities are used. On the bottom of the screw is a spherical seat that is matched to the seat on the breaker block or tapered roller thrust bearing assembly (Figure 18.17). These components are grease lubricated. The worm gear arrangement that drives the screw is fed oil from the pinion lubrication system. It is generally a double enveloping unit with a high ratio. Sometimes, a second smaller unit is also used to increase the reduction ratio and torque or load available. New rolling mills are being supplied with hydraulic pushup cylinders from the bottom of the mill. Some mills have a combination of the mechanical screw and a short stroke cylinder for gauge and shape control. The hydraulic systems used are servo type units using pressures up to 6000 psi, with polyol ester or water glycol oils for fire resistance. Load cells for the stand can be donut shaped in the top of the housing or a beam type in the bottom of the mill post. To keep the top roll assembly in position, most use have some form of balance beam arrangement or counterweight that is operated from a hydraulic cylinder. Bar mill stands have a problem with the small roller guides that guide the high-speed hot bars into the mill rolls. They are vulnerable to mill cobble damage. Bearing speeds can exceed 16,000 rpm. When damaged, the guides will skid on the bar or rod surface. Oil mist, air–oil, and grease systems have been used with low base oil viscosities of 68 to 100 cSt. The guide rollers generally use ball bearings while the mill stands use 2-row tapered roller bearings for the mill stand rolls. Today, most lubrication systems have been removed from the bearings and they are sealed with grease that will withstand water impingement, high speeds, and temperatures. The Mill Roll Neck Bearing holds the mill rolls in position during rolling. It sustains high shock during the initial bite of the rolls into the leading edge of the strip. The rolling forces are continuous for longer periods of time as the product lengthens with each pass. Separating forces can reach 5 million lb per screw in the hot strip mill. Work roll bearings in a hot strip mill see on the average 120 to 140 t load per bearing from the work roll bending systems in the chocks or in the mill windows (project blocks or Mae West blocks made up of hydraulic jacks under 2500 to 3000 psi pressure). This does not include the forces that appear on the bearing from thrust due to side shifting of the rolls in the mill window or pair cross rolling where the rolls are intentionally crossed for strip shape control. The bearings must also survive high temperatures, water, scale, and dust conditions. The types of roll bearings run the gamut from Babbitt, Babbitt with bronze grids, Babbitt with phenolic fabric inserts, phenolic resin fabric, oil film, and roller bearings. Modern mills use oil film and roller bearings for backup roll applications. Older mills may be using grease-lubricated metallic and water lubricated fabrics. The following details some of the aspects of the types of bearings: 1. Grease Lubricated Plain Bearings a. These bearings may be babbitted with tin or lead based babbitts (being phased out) or bronze, or brass bearing. Tin bronzes (80-10-10, SAE 660), aluminum bronze, manganese bronze, and some silicon bronze are used. There are new sintered bronze materials being used that are filled with graphite, moly and copper, or bismuth. These bearings do not need any lubrication when
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FIGURE 18.19
Typical mill stack cross-section.
used at low speeds. Graphite self-lubricating plain bearings are also being used for high heat applications. All of the materials are used to carry thrust and radial loads depending on the design. The thrust portion is usually in the form of a separate collar or flange. Inserts of phenolic resin fabric at the grease groove prevents Babbitt from wiping over the grease hole. The bearings come equipped with grease groove arrangements to move the oil or grease into the load zone or the bearing. Bronzes are also used extensively for wear liners on slideways in the rolling mills and processing lines. Many of these liners are grease lubricated. b. Scale guards on the bottom roll deflect scale away from the bearing. Water sprays are installed to keep the neck surface flushed or scale. These bearings are in small bar and billet mills.
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FIGURE 18.20 Typical pinion stand arrangement. Printed with permission Timken Co., Timken Rolling Mill Manual.
FIGURE 18.21 Typical screwdown bearing assembly. (Printed with permission NSK, Large Bearing Catalogue, p. c17.)
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FIGURE 18.22 Company.)
Handbook of Lubrication and Tribology
Typical oil film arrangement back-up rolls. (Printed with permission Morgan Construction
2. Fabric/Plastic Bearings a. Water lubricated bearings may be found on some plate mills, blooming, and old bar mills. They have low coefficient of friction and low maintenance costs. Roll neck surfaces must be lubricated prior to downtime to ensure there is no rusting, which can destroy the fabric. For reversing mills, grease is also used in addition to the water. Water only provides a thin hydrodynamic film. Grease prevents wear. b. Plastics, especially thermoplastics such as nylon, polyethylene therephal, and UHMW (ultrahigh density polyethylene), polyurethanes, PTFE (polytetrafluoroethylene), and composites of carbon are being used for bushing and wear liner materials in the steel mills. Many of these materials are self lubricating, making them capable of living in areas where lubrication would be a problem. Water, heat, loads, and contamination can affect the operation of the materials. Care must be made in the selection of the material for an application. 3. Oil Film Bearings (Figure 18.22) a. These bearings are found in rod mills, bar mills, hot and cold strip mills, plate mills, and blooming and slabbing or roughing mills. b. They are centrifugally cast Babbitt or nickel/Babbitt alloy. Oil is fed into the bearings and removed from the chock cavity with drain lines to a central reservoir. Turbine quality, high demulsibility, and high viscosity oils are used. Oil flows through these bearings can be up to 50 to 60 gpm depending on the size of the bearing, the loads, and the speeds. The circulating oil systems (Figure 18.23) usually have two reservoirs per system, one operating and one on standby settling or resting. The reservoirs may contain up to 10,000 gal each. The systems have pressure control valves, cooling devices, twin strainers with 150 to 100 µm baskets and orifices to control the amount of oil to each bearing. The systems also come with vacuum dehydrators or centrifuges to remove water. The assembly also has a thrust bearing to take the inertia forces from rolling. This is traditionally a two row steep angle tapered roller bearing or a ball bearing.
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FIGURE 18.23 Typical oil film bearing lubrication system. (Printed with permission Morgan Construction Company.)
FIGURE 18.24 Hydrostatic lift arrangement for oil film bearing. (Printed with permission Morgan Construction Company.)
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FIGURE 18.25 Company.)
Handbook of Lubrication and Tribology
Oil film sealing inboard seal with water quard seal. (Printed with permission Morgan Construction
FIGURE 18.26 Typical grease lubrication inlet design. Printed with permission, NSK Canada Inc., Large Bearing Catalogue, cat.# E125b, 1996, p. c-10.
FIGURE 18.27
Typical oil mist lubrication arrangement.
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FIGURE 18.28 Typical sealed TQO bearing.
Narrow Seal Concept
Unitized Seal Concept
FIGURE 18.29 Printed with permission, Timeken Limited, The Right Solution for the Rolling Mill Industry, 1996.
FIGURE 18.30 Typical arrangement for sealed work roll bearing/chock.
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FIGURE 18.31 Inboard lipseals and v-ring seal. Printed with permission, Timken Limited, The Right Solution for the Rolling Mill Industry, 1996.
FIGURE 18.32 Four-row tape and roller bearing with tapered core. Printed with permission, Timken Limited, The Right Solution for the Rolling Mill Industry, 1996.
The bearings and chocks are locked onto the roll journals with mechanical threaded locking arrangements or some form of hydraulic mount. c. Hydrostatic lift systems (Figure 18.24) with a separate pumping and control system for maintaining oil film during startup, mill sticking, and low speed (under 10 rpm). These systems have pads built into the babbitted bushing. They operate at 9,000 to 12,000 psi. d. The sealing of the chocks and bearing against water ingression and oil leakage (Figure 18.25) that could become a tramp oil in cold rolling operations, resulting in staining and rolling oil
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FIGURE 18.33 Cylindrical roller bearing assembly. Printed with permission, NSK Bearing Catalogue, p. c20.
FIGURE 18.34 Four-row tapered roller bearing assembly. Printed with permission, NSK Bearing Catalogue,p. c21.
contamination is especially critical. The oil film bearing has special sealing devices to do just that type of control. 4. Roller Bearings a. Roller bearings are used as the main support bearing for rolling mill work roll assemblies in hot rolling and cold rolling applications as well as bar and rod mills, and wire mills. They are also used as backup roll bearings for cold tandem mills and temper mills.
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Printed with permission, NSK Large Bearing Catalogue, p. c12, c13.
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FIGURE 18.35
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FIGURE 18.36 Schematic representation of the operating units making up the 2135-mm (84-inch) continuouspickling line at Gary Works of United States Steel Corporation. (Printed with permission Association of Iron and Steel Technology, The Making, Shaping, and Heat Treating of Steel, 10th Edition, p. 1088.)
b. Roller bearings are the main bearing type used to support the work roll in the mill window of a 4-Hi or 6-Hi mill. Steel strip is reduced by the work rolls. Water contamination is a major threat to this bearing, so much effort is put toward looking after the chock sealing devices, which are for the most part some form of lip seal design. The bearing type is a four row tapered roller bearing of the TQO designation. This bearing has a straight neck and is loose fitted onto the roll journal for the sake of ease of mounting. The bearings are prepacked with grease at each roll change (Figure 18.26). Roll changes can be as frequent as 15 min in a cold mill or after every 12 h in a hot strip mill finishing stand. It all depends on the roll body surface condition and the affects it has on the strip quality. Most roll assemblies are not connected to a centralized grease system. In order to reduce lubricant consumption and keep the contaminants out of the bearing, sealed bearings are now being used in most hot mills and many cold mills (Figure 18.28). The bearings are packed with a premium grease for up to 8 month mills pinion. Figure 18.29 shows two types of seal bearings used for the sealed work roll bearing. Figure 18.30 lays out the chock assembly requirements for a sealed bearing. c. The work roll bearing can also be lubricated with oil mist systems (Figure 18.27) and air–oil lubrication units. The backup roll assembly technology is now using four-row tapered roller bearings for cold rolling and tempering applications. The tapered journal (Figure 18.32) and straight neck designs (Figure 18.33) are both being employed. Large four-row cylindrical roller bearings combined with a two-row tapered roller bearing for thrust are being used in cold mill backup roll applications (Figure 18.34). Some of these bearings are up to 36 in. in bore diameter. The inner race of the bearing is interference fitted onto the journal and ground
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FIGURE 18.37 Schematic elevation of a modern continuous-gavanizing line. Printed with permission, Association of Iron and Steel Technology, The Making, Shaping, Heat Treating of Steel, 10th Edition, p. 1181.
Steel Industry FIGURE 18.38 Schematic of three types of electrolytic tin-plating lines. (Printed with permission, Association of Iron & Steel Technology, The Making, Shaping, and Heat Treating of Steel, 10th Edition, p. 1147.)
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with the roll body at the time of roll manufacture to enhance the accuracy of the rolls down to 0.0002 in. journal concentricity. Many mills operate at 5000 to 7000 fpm speeds and 2 to 3 million lb mill separating forces. The bearings are oil mist, air–oil, or circulating oil lubricated depending on the mill speed, loads, and the type of heat generated during the rolling operations. d. In bar mills that operate at up to 4,000 fpm, roller bearings are used for the work rolls. Grease lubrication is used with particular attention paid to the water resistance and shear stability in the presence of water properties of the grease. For rod mills running up to 20,000 fpm, oil film bearings are used with circulating oil systems where the oil is filtered to 5 to 10 µm. The oil viscosity on all bearings is 1,000 to 220 SSU depending on the load and speed. Mill shears installed in the primary or hot rolling side of the steel plant are generally drum type or crop shear design. They are driven by an electric motor and a double to triple reduction gearbox since the rotating speed is under one rpm. The shear drums are supported by two row cylindrical roller bearings that are grease lubricated. Electrically operated hot saws are used in bar and structural mills. In cold mills, pickle lines, and tin and galvanizing lines upcut hydraulic/mechanical shears are used. Coilers are employed in the hot mill to coil the hot strip. These rotating cantilevered expanding mandrels operate at up to 3000 fpm peripheral speed. Wedge type and link type designs for the expanding mandrel segments are the most common designs. They are grease lubricated. Water and heat are the main enemies of the mandrel lubricants. The mandrel is driven by a motor and herringbone gearbox arrangement that is oil lubricated. In the cold mills and process lines, expanding mandrels in mostly wedge designs are used to unwind and rewind the strip. Rod and wire mills use vertically mounted mandrels for rod and wire products.
18.12 Steel Finishing Facilities The steel mill would not be complete without further finishing operations that can be done to the steel. These include: wire coating equipment; plating lines for pipe, wire and strip that includes tinning, galvanizing painting; shear and trimming lines. Most of this machinery does not have major lubrication problems. Most lines have centralized grease-dispensing systems to lubricate bearings, pins, bushings, and slideways. Continuous annealing, hot dip galvanizing lines and electrolytic tinning lines can operate at speeds up to 2000 fpm. The lines may have grease systems or oil mist systems to lubricate the bearings. Gearboxes, including the ones for the payoff and recoilers have stand-alone circulating oil systems with them. Particular attention has to be paid to the spherical roller bearing pillow blocks used for the furnace applications and tank areas on these lines to ensure that high temperature, water resistant, and chemical resistant greases are used. Most mills also have mineral type hydraulic systems where required. Many galvanizing lines have single stand 4-HI temper mills to condition the strip. Most flat processing lines will have leveler roll and scale breaker sets (Figure 18.35), bridle rolls sets for steering and tensioning the strip, some form of shear, welder, and uncoilers and recoilers. Figure 18.36 shows the layout of the machinery and steps in a continuous pickling line for removing rust and seal from steel stripsss. Figure 18.37 has the schematic of a modern hot dip galvanizing line with the various sections of the process. Figure 18.38 shows three different types of tinning line technologies.
References [1] The Making, Shaping and Heat Treating of Steel, 10th ed., Lankford, W., Jr. et al. (Eds), AISE, Pittsburgh, 1985. [2] The Making, Shaping and Heat Treating of Steel, 11th ed., Casting Volume, Cramb, A.W., (Ed.), The AISE Steel Foundation, Pittsburgh, 2003. [3] The Right Solution for the Rolling Mill, The Timken Company, Canton, 1996.
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[4] [5] [6] [7] [8] [9]
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Large-Size Rolling Bearings Cat. #E125b, NSK Ltd., Japan, 1989. Morgil Bearings and Flat Mill Products, Morgan Construction Co., Worcester, 2003. Lubrication Engineers Manual 2nd ed., AIST, Pittsburgh, 1996. CRC Handbook of Lubrication, Vol.1, 1st ed., Booser, E.R. (Ed.), CRC Press/STLE, Boca Raton, 1983. Tribology Data Handbook, Booser, E.R. (Ed.), CRC Press/STLE, Boca Raton, 1997. The Making, Shaping and Heat Treating of Steel, 11th ed., Steelmaking and Refining Volume, Fruehan, R.J. (Ed.), The ASIE Steel Foundation, Pittsburgh, 1998.
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19 Aluminum Metalworking Lubricants 19.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19.2 General Overview . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19.3 Detailed Discussion of Lubricant Requirements . . . .
19-1 19-2 19-4
Formulation Chemistry • Regulatory and Customer Requirements • Process Equipment Lubricants • Filtration • Waste Treatment
19.4 Lubrication in Selected Aluminum Metalworking Processes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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Rolling • Lubricants in High-Temperature Processes • Lubricants for Forming of Finished Food and Beverage Containers
James R. Anglin Aluminum Company of America
Acknowledgments. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
19-19 19-19
19.1 Introduction This chapter addresses formulation concerns related to the metalworking of aluminum and its alloys, with an emphasis on processes commonly performed by or focused on by aluminum producers. Included is a general discussion as well as more detailed discussions for rolling, hot processing other than rolling, and forming of finished food and beverage containers. This information should provide a basis for understanding the requirements for process fluids used in other applications such as wire drawing, general sheet forming, and machining, which are not discussed here but are described elsewhere, such as in Schey’s monograph [1]. The following discussion will apply generally to aluminum and its alloys, except where specific alloys are indicated. While the primary function of an aluminum metalworking lubricant is to facilitate the fabrication of aluminum products, there is a lengthy list of requirements built into lubricant performance whose relative importance depends on the specific application. The lubricant must contribute to process robustness and metal surface quality as well as provide environmental, health, and safety (EHS) suitability, regulatory compliance, and acceptable cost. It must further be possible to remove lubricant residues as required for subsequent processing to be performed on the material. Finally, it must be possible to reclaim or dispose of the used lubricant economically and in compliance with applicable regulations.
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Handbook of Lubrication and Tribology -- Adsorbed water, lubricants, organic materials -- Oxide layer -- Near-surface metallurgical properties
-- Bulk metallurgical properties
FIGURE 19.1
Schematic representation of aluminum surface.
19.2 General Overview A properly formulated lubricant can be expected to act in concert with well designed and maintained tooling to enable the metalworking operation to proceed successfully. It will control friction in the toolworkpiece contact and resist the onset of adhesive wear, where aluminum adheres to the tooling and its surface is torn as it moves relative to the tool surface. This can occur when the oxide layer on the aluminum has been breached and the softer, more reactive underlying metal has been exposed. Certain alloys, such as 6262, incorporate moderate levels of lead and bismuth to facilitate metalworking. However, in general, only very mild metalworking of aluminum alloys is expected to succeed without lubricant. In certain hot forming processes, the lubrication can be provided by lamellar solids such as graphite, boron nitride, and molybdenum disulfide. These alone, or in combination with other components, provide reduced friction and reduced aluminum transfer to the tooling as the layers of the easily sheared solid slide over one another. On the other hand, certain forming operations with polymer-coated metal rely on thin layers of lubricant to reduce the friction between the polymer and the tooling, with no actual contact of the aluminum and the lubricant. In most instances, however, a liquid or dry film of lubricant directly lubricates the aluminum surface. The surface of the aluminum workpiece presents a complex system to the lubricant and tooling, as shown in Figure 19.1. The composition and properties of the metal will increasingly depart from the bulk material and reflect its processing history as its surface is approached. Similarly, the oxide layer can vary in composition and thickness based on processing history. As an example, alloys containing magnesium can have a higher concentration of that element at the surface. Friction coefficients have been linked to both the underlying strength of the alloy and to the levels of MgO found on the surface [2]. The surface may contain other elements, such as fluorine from furnace treatments to minimize hydrogen uptake by the metal in moist atmospheres [3]. In addition, adsorbed water, organic materials, or residual lubricants can be present on top of the oxide. Further, the topography of the surface can have a strong influence on the distribution of the lubricant, with depressions providing reservoirs and asperities controlling the flow of the lubricant in the contact. Asperities aligned across the direction of travel will tend to lead to thicker lubricant films [4]. The surface of the tooling has an important influence on the metalworking process. While generally outside the scope of this review, it should be noted that the surface composition and topography of the tooling are very important. The abrasiveness of the surface oxides on aluminum leads to the use of a variety of strategies for reducing tooling wear, including the use of hard coatings such as chromium or the use of wear-resistant materials such as chromium carbide. The tendency for adhesion of the aluminum to the tooling is dependent on the composition of the tooling surface as well as its topography, since rougher features can provide an opportunity for transferred aluminum to be held more securely. The directionality of tooling surface features will also influence the flow of the lubricant during the process and influence its film thickness [4]. In addition, the angle at which asperities rise from the tooling will determine the relative tendency to cut the workpiece and generate abrasive wear debris. Although machinery lubrication is designed where possible to occur under hydrodynamic lubrication conditions, this can be troublesome for aluminum metalworking. In hydrodynamic lubrication, the lubricant film thickness exceeds about three times the composite roughness of the lubricated surfaces.
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60
FIGURE 19.2 Transverse fissures.
FIGURE 19.3 Orange peel.
In rolling, for example, this will lead to slip in the mill and general loss of control of the metal being rolled. Low friction may also lead to bite refusals, especially with thicker slabs, where the rolls spin against the end of the workpiece without being able to draw it into the roll bite. The application of kerosene can be used to overcome bite refusals, but the possibility of ignition of the solvent must be anticipated. With the greater lubricant film thickness, transverse fissures (Figure 19.2) can be generated on the aluminum surface because of uneven expansion of the metal surface not in direct contact with the tooling. This imparts a grey appearance to the finished metal. Conversely, too thin a film can lead to primarily boundary lubrication, where higher friction leads to greater energy usage and heat generation, along with excessive transfer to the die or dislocation of surface material. This can give rise to scuffing or galling of the workpiece surface or a rough surface condition often referred to as “orange peel.” (Figure 19.3) For many applications, mixed lubrication, in which the lubricant film thickness is typically 1 to 3 times the aggregate surface roughness, appears to provide the best compromise.
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Liquid lubricants in mixed or boundary conditions rely for their effectiveness on fluid rheology and boundary additive performance, the combination of which is referred to here as film strength. Boundary additives have a strong affinity for existing and freshly formed metal surfaces under the conditions of the metalworking process. In work with steel, adsorbed fatty alcohols and fatty acids gave reduced friction coefficient with increasing chain length over the range C12 to C18 [5]. Work has also been performed with aluminum surfaces [6]. Guangteng and Spikes [7] have reported a tendency for esters in formulated products to concentrate preferentially at surfaces and, for thin films, to provide film thicknesses approaching that of the pure ester. In formulated process lubricants, just as in machinery lubricants, understanding the contributions and synergies among the various components remains a challenging area. Oleochemical products such as fatty acids, esters, and alcohols are the most commonly used boundary additives in aluminum metalworking along with phosphate esters and their derivatives in some instances. Molecules with long, linear chains afford the opportunity for a more densely packed layer of additive on the surface with the polar end groups of the acids and alcohols providing attachment and lateral forces among chains providing additional film strengthening. The esters appear to act synergistically in combination with the fatty acid or alcohol additives. With increasing temperature, the opportunity for chemical reactions involving the polar groups increases, such as soap formation with adsorbed fatty acids. Other mechanisms of action of these species have been proposed and are summarized by Schey [1]. One explanation presented by Rebinder [8] relates facile metalworking to a reduction of the surface energy of the workpiece with the adsorbed additives. It is likely that this effect is sensitive to the nature of surface oxides that are present. Many other types of compounds are also used in aluminum metalworking. Paraffin and other waxes have a crystalline structure that contributes to low friction, but hydrocarbon waxes are poor boundary additives and provide limited protection against adhesive wear. Other chemistries can be expected to provide breakdown products or reaction products at the aluminum surface that participate in the lubrication, including metal carboxylates, such as those of lead (now little used), tin, and bismuth. Boric acid, a lamellar solid, has also been discussed recently [9]. Evidence was presented that its performance varies with alloy, an observation that may reflect different interactions with the varying oxide compositions present on different alloys. However, the low temperature (169◦ C) for boric acid dehydration and loss of lamellar structure limits its effectiveness to relatively low temperature processes.
19.3 Detailed Discussion of Lubricant Requirements Key functions of lubricants in aluminum metalworking are the control of friction and the minimizing of wear, especially in sliding contact situations. The selection of lubricant components is determined in large part by temperature considerations. At higher temperatures where organic formulations may not be successful or ignition of the lubricant cannot be tolerated, water-based products or formulations based in part or wholly on inorganic solids can be used. In principle, inorganic liquids can also be used in such an application; however, products with melting behavior suitable for aluminum metalworking appear to be few. At lower temperatures, the formulation options are much broader but, as temperatures decrease, the tendency of some oleochemicals to solidify can be a limiting factor. The requirements for lubricant formulations are quite diverse and depend on the nature of the process in question. While lubricants used in many operations are consumed in the process, with residues helping to lubricate the tooling for subsequent parts, some processes, such as rolling and can bodymaking, use systems with large volumes of fluid that can continue to be used for many months without wholesale replacement. In these instances, additional concerns take on importance, such as oxidative and thermal stability, along with hydrolytic stability and biostability for water-based lubricants. The ability to filter well enough to control the level of debris becomes more important in these systems since continued increases in debris level will at some point adversely affect fluid handling or the metalworking process. It also becomes important to be able to monitor and control the composition of the lubricant so that performance over time remains as consistent as possible.
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19.3.1 Formulation Chemistry While many applications have typically used formulations based on petroleum-derived base oils, interest is increasing in the use of alternate materials such as oleochemicals that can reduce dependence on petroleum and provide improved biodegradability. Any changes need to be made in keeping with modern production strategy, which places a premium on consistent performance, enabling the process to be switched rapidly among products according to customer demand regardless of their sensitivity to lubricant system condition. The drive for lower costs may favor the use of oleochemicals containing significant levels of linoleic and linolenic acid and their derivatives, but it must be recognized that such products can detract from oxidative and thermal stability, a significant concern in systems where consistent performance over time is needed. Antioxidants can be a valuable tool in this situation as can the use of oleic acid based products with reduced levels of polyunsaturated components that are now becoming more routinely available. Care must be taken to choose antioxidants carefully, given that some can be water soluble [10] and amine-based products have the potential to react with fatty acids in the system to form amine salts that can affect metalworking performance and emulsification properties. The desire for biodegradable lubricants, whether process fluids or machinery lubricants that can leak into them, also presents a tradeoff, since this desirable property can lead to significant problems with microbial growth in water-based lubricants. The problems can include odor, chemical degradation, system corrosion, difficulty in filtering, and the costs as well as the hygiene considerations with the handling of biocides. The biocides themselves can also have a chemical influence on the lubricant [11]. These concerns must be balanced against the advantages afforded in waste fluid treatment and the management of spills. Fatty acids are well known to be excellent boundary additives, with commonly used products being those from oleochemical sources with chain lengths of 10 to 22 carbon atoms. Also well known are products with multiple acid groups, such as dimer acid and other oligomeric products obtained from unsaturated C18 fatty acids by thermal processes. Products of the latter type can also be partially esterified, generating a distribution of components ranging from the free fatty acids to fully esterified material. The linear saturated fatty acids have higher melting points and decreasing solubility as chain length increases, requiring that the formulator control the content of such materials to keep the lubricant fluid under the process conditions. Alternatives with improved solubility properties are oleic acid and erucic acid, along with isostearic acid, which is a complex mixture obtained by hydrogenating a highly unsaturated C18 fatty acid product known as monomer acid. Isostearic acid comprises primarily branched chain structures, some of which include a cyclic component. The use of fatty acids in lubricants is a two-edged sword from the perspective that their excellent metalworking performance is accompanied by a tendency to generate hard-to-filter fine debris and a tendency to undergo chemical reactions. They can form soaps or metal salts, including oil-soluble soaps based on aluminum or iron as well as poorly soluble soaps based on such water-hardness cations as calcium. Aluminum soaps are known to have a variety of possible structures that can be distinguished by differences in the infrared absorption energies of their carbonyl functions. Their molecular weights can range up to more than 1,000,000 [12]. As a result, they can significantly raise fluid viscosity. These materials, especially when derived from longer chain linear fatty acids, can have poor solubility near or below room temperature, a troublesome feature that does, however, afford an opportunity for controlling soap levels. Fatty acids can also be combined with amines to provide emulsifying behavior. Commonly used amines include triethanolamine, 2-amino-2-methyl-1-propanol, and isopropanolamines. Where triethanolamine is used, product that is low in diethanolamine and monoethanolamine is preferred for hygiene reasons, and where amines in general are used the use of nitrites needs to be avoided to eliminate nitrosamine formation. The levels of fatty acid amine salts can typically be monitored by infrared techniques. The amine salts of fatty acids, like the free fatty acids, are subject to depletion by reacting with water hardness ions and depositing on mill equipment or being filtered out of the system. This can lead to product quality concerns if deposits on rolling mill equipment subsequently fall or transfer onto the rolled product.
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Fatty acids and their salts are known to undergo additional degradation reactions at higher temperatures, including decarboxylation and formation of ketones [13]. Similarly, amine salts of carboxylic acids, other than those of tertiary amines, have the potential to be converted by heat to amides. They also can undergo reaction with other boundary additives as discussed below. The reactivity of these species increases the challenge of maintaining them at consistent levels for optimum formulation performance. The use of fatty alcohols is also well known, with the materials including both synthetic products and products formed by the reduction of species such as the methyl esters of the naturally occurring fatty acids. The natural products are highly linear but may contain unsaturation from the precursor compounds. On the other hand, the synthetic products can contain primary alcohols with differing amounts of branched chains as well as minor amounts of diols depending on the specific manufacturing method. While most synthetic procedures generate product with even numbers of carbon atoms similar to natural alcohols, some also produce the odd-numbered chain lengths. Like the corresponding fatty acids, saturated linear alcohols have the drawback of reduced solubility in typical organic media with increasing chain length, requiring that formulations with longer chain length alcohols be maintained warm. Some advantage in handling comes from blends of alcohols, such as mixed C12–C14 alcohols, which have a lower freezing point than the pure individual compounds and can be fluid at room temperature. Higher levels of branching would also decrease the freezing point, but compositional constraints associated with FDA compliance limit the extent of branching permissible for FDA-sensitive applications. The alcohols have a tendency to oxidize to aldehydes and other species, and are commonly used with an antioxidant present. Finally, they can react with fatty acids to form esters, thereby depleting the levels of both components in favor of the product. The third common type of additive, whether of oleochemical or synthetic origin, is esters. These are commonly used in combination with fatty acids or alcohols, where they may provide a synergistic boost in film strength performance, but are at times used alone. Esters are available in a wide range of structures and viscosities and range from the methyl esters of common fatty acids to higher viscosity polyol esters. The latter can be derived from synthetic polyols such as neopentyl glycol, trimethylolpropane (TMP), and pentaerythritol or from glycerine. Esters based on the above synthetic polyols have excellent thermal stability because the structures afford no opportunity for degradation by a β-elimination mechanism that releases an olefin. Esters can also be formed from multifunctional carboxylic acids, such as dimer and trimer acid, and very high viscosity products can be made from the combination of polyols with multifunctional carboxylic acids. It should be recognized that the contribution of esters to film strength can easily be overestimated when low to moderate levels of unreacted fatty acid are present as an impurity. Esters are capable of a variety of chemical reactions related to their use in lubricants. Among these are transesterification, in which either the acid or the alcohol moiety in the ester exchanges with other acids or alcohols present in the system to form different esters (Equations [19.1] and [19.2]). RC(O)OR + R C(O)OH → R C(O)OR + RC(O)OH
RC(O)OR + R OH → RC(O)OR + R OH
(19.1) (19.2)
In the presence of water and particularly in water-based formulations, esters can be prone to undergo hydrolysis, with the formation of the component fatty acids and alcohols. This can change the film strength and other properties of the formulation depending on the specific ester. The generation of a short-chain alcohol with a low boiling point can contribute to air emissions, whereas partial hydrolysis of a polyol ester can generate species that act as surfactants. In addition, as has been reported for turbine lubricants, the heating of TMP esters together with phosphate esters has been found to form the potent neurotoxin TMP phosphate [14]. The base oils used in aluminum metalworking are quite diverse. In cold rolling and foil rolling, linear paraffins are the leading products; however, hydrotreated kerosene streams, often with aromatics levels of less than 1% and very low heteroatom content, also have a place. The linear paraffins commonly
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used in aluminum metalworking are isolated from kerosene streams and fractionated to provide relatively narrow boiling range products. These cover a range of flash points and viscosities and have linear content typically above 97%. They are low in odor and staining tendency and have relatively high flash points together with relatively low viscosities compared with typical hydrocarbon streams of similar boiling point range. This combination can enable higher cold rolling speed where surface appearance issues related to thick lubricant films come into play. Some products have very narrow distillation ranges that, together with careful fractionation at the low end, further improve flash point [15]. Additive response can be expected to be very good in these highly nonpolar products. One disadvantage associated with the highly linear structures when compared to hydrotreated kerosene fractions is the relatively high freeze point. Fractions with substantial content of C15 or higher homologues need to be insulated or warmed under cold winter temperatures. In the future, suitably fractionated products may become available from natural gas or coal sources using the Fischer-Tropsch process to generate aliphatic hydrocarbons. These gas-to-liquids products are highly linear although not typically to the level of the linear paraffin products isolated from kerosene. While such base oils are generally resistant to oxidation, it has been shown that aluminum can catalyze the oxidation of base oils as well as additives [16]. As an approach to formulating lubricants with lower volatility, ionic fluids have been evaluated as base oils [17]. Staining of aluminum can occur when lubricant residues remaining on the metal during a hightemperature process such as annealing give rise to a discoloration of the aluminum surface. The color can range from a light stain, commonly yellow, to darker colored stains with a variety of colors. Stain is typically measured either visually or gravimetrically using samples with bright, clean surfaces. Among base oils, the lower boiling and fully saturated products such as cold rolling oils are most likely to have minimal stain. Among higher viscosity products, polybutenes yield minimal stain or residue if the hightemperature process is sufficiently high in time and temperature to convert the residues to low molecular weight, volatile products. Leakage of standard machinery lubricants into cold rolling oils can normally be tolerated only at low levels before stain becomes excessive, even where highly refined or other synthetic hydrocarbon oils are used. Naphthenic base oils are most commonly used in water-based emulsion products because of their relative ease of being emulsified. However, current refining practices are producing more severely hydrotreated products with improved hygiene characteristics and with improved stability. These changes can require adjustments to the emulsifier package. Naphthenic oils are available in a relatively wide range of viscosities, providing considerable latitude to the formulator to tailor the viscosity and film thickness of the lubricant film in the contact. A key additive type required for many water-based formulations is the emulsifier. The emulsifiers typically used in these applications enable insoluble or poorly soluble materials to be chemically stabilized in a water medium. Applicable fluid types, such as emulsions (often termed soluble oils), microemulsions, and micellar solutions have been reviewed by Laemmle [18]. These fluids differ in composition, droplet size, and thermodynamic stability, and ultimately in lubrication performance and waste treatability. Lubrication performance will generally be stronger with emulsions of relatively low stability, where the oil phase separates more readily from the emulsion. Lubrication performance may be weaker on moving to more stable emulsions, microemulsions, micellar solutions, and true solutions. Common types of emulsifiers include: • Nonionic, including such materials as ethoxylated alcohols and ethoxylated fatty acids • Anionic, commonly salts of longer-chain fatty acids or polyalkoxylated phosphate compounds with amines such as triethanolamine, 2-amino-2-methyl-1-propanol, and isopropylamines, but also including sulfonates • Cationic, such as quaternary ammonium compounds • Combinations of nonionic with ionic emulsifiers In addition, emulsifier properties can be built into other components of the system, such as boundary additives, to provide multifunctional performance. To assist in initial emulsification, lower molecular weight diols and triols can be included as coupling agents. The anionic emulsifiers generally have the
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drawback noted previously for amine carboxylates that they can combine with water hardness cations, such as calcium ions, to form insoluble calcium salts. This will deplete the emulsifier and, for amine carboxylates, alter the balance of fatty acid and emulsifier and further affect lubricant performance. For this reason, deionized water is preferred for blending with the oil and for water make-up since certain processes such as hot rolling require large water additions to replace evaporation losses. For water-based formulations, consistent performance is highly dependent upon consistency in the formulation chemistry. Changes in performance resulting from depletion of the emulsifiers or leakage of machinery lubricants into such systems can be difficult to correct through adjustments to the formulation chemistry. Either of these events can initiate a performance pattern where droplet sizes increase during usage, often giving improved metalworking performance until the product starts to become too unstable and performance degrades. Leakage of high viscosity products, such as gear oils, into a formulation can greatly affect lubrication performance. Skimming of unemulsified or separated material, often termed tramp oil, from the surface of the mixture is routinely performed. To the extent that the formulation components are removed with the tramp oil, they need to be replaced with fresh or possibly reclaimed product. Consistent emulsifier performance can be expected to require careful control of both the emulsifier content and composition. This is challenging because of the chemical complexity of many commercial emulsifiers and the difficulty in obtaining detailed compositional information on the emulsifiers present in emulsions using routinely available, inexpensive methods. Fortunately, levels of fatty acids and esters can be readily monitored by infrared techniques and adjustments made to the composition as necessary to keep these at target levels. This becomes more difficult if the formulation contains multiple products that cannot be distinguished by infrared or are too low in volatility for GC analysis. High levels of fatty acid can build up with rapid ester hydrolysis and high levels of ester can occur from contamination by ester-based mill equipment lubricants. The importance of the water phase chemistry in water-based formulations should not be overlooked. The pH affects equilibria involving, for example, fatty acid and amine species and ultimately emulsion stability and oil droplet size. It can also influence corrosion and the rates of hydrolysis of esters. Conductivity affects emulsion stability as well as corrosion, especially if certain anions such as halides are present at significant levels. Conductivity can be expected to rise gradually over time from pH adjustments and dissolution of alloy components, but will increase more rapidly if soft or potable water is used for additions. Water hardness cations, as previously noted, can react with and deplete certain components. In water-based formulations, certain problem areas, such as rust and corrosion, microbial growth, and foam, can require special attention. A variety of additives are available for addressing rust and corrosion issues and these can be tested for efficacy by using methods such as ASTM D 665 and D 130 or variants of them that simulate process conditions. Performance can be expected to depend on both oil and water phase chemistry. Special care should be taken to address any conditions that might give rise to pitting corrosion, which can rapidly lead to perforation of system hardware, especially if made of mild steel. While the preference is to formulate products that resist microbial attack, an array of biocides is available if needed. As noted above, amine-based products, including biocides, can react with fatty acids in the formulation [11]. Normal biocide strategy is to introduce a second biocide at intervals to reduce populations of microbes developing tolerance for the primary product. If foam issues occur, mechanical causes such as turbulence should be minimized first. If not successful, reformulation to minimize foam may be preferable to implementing formulations with significant foaming problems since excessive antifoam additive use can lead to air entrainment problems. Furthermore, the use of silicone-based products, even in seemingly inconsequential amounts, can lead to severe adhesion problems for products that will subsequently be coated or painted. Nevertheless, a range of defoamers, including ones free of silicones, is available. An additional concern with water-based formulations is the potential for formation of solution stain or water stain, which is commonly a whitish appearing surface blemish on the formed product. This is unacceptable for products requiring good surface appearance, and differences in friction performance over a partially stained surface can lead to problems in forming. The formation of this defect is related to the time of contact of aqueous materials with the aluminum surface at temperatures near 100◦ C. It can be addressed by rapid and efficient removal of aqueous metalworking products from the surface to limit
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the opportunity for the stain to form. Alternatively, reformulating or adjusting process conditions can be considered to reduce or eliminate the stain.
19.3.2 Regulatory and Customer Requirements It is important to be aware of a variety of considerations unrelated to lubricant performance that nevertheless are important when lubricant formulations are developed. For applications where aluminum products will be used in food or drug packaging, base oils and additives must comply with appropriate regulations. While this is a complex area, for United States and many international requirements, the information in the most recent issue of Title 21 — Food and Drugs of the Code of Federal Regulations (CFR) will apply. Key information can be found in 21 CFR 178.3910, which discusses surface lubricants used in the manufacture of metallic articles. Lists of substances and their usage limitations are presented, along with limits for the levels of residue on the formed product as it contacts the packaged product. The allowable levels depend on whether the lubricant function is to roll sheet or foil (0.015 mg/in.2 or 23 mg/m2 ) or to draw, stamp, or form articles from it (0.2 mg/in.2 or 310 mg/m2 ). Additional listings of potentially acceptable components can be found in Sections 172 (food additives permitted for direct addition to food for human consumption), 182 (substances generally recognized as safe), and 184 (direct food substances affirmed as generally recognized as safe). Other sections within other parts of Title 21 may provide additional materials for consideration, including 178.3570, which discusses lubricants with incidental food contact. These are components of lubricants used to operate machinery used in food applications, which differs from 178.3910 in that intentional contact with the packaged product is not expected. Careful interpretation of the sections and the precedents associated with their use is needed in making component selections. An actively managed program to ensure that good manufacturing practices are followed needs to be in place for the manufacture of food packaging products. It is very important to avoid the presence of components or contaminants that can cause the lubricant residues on the packaging material to be deemed unsafe and products contacting them to be deemed to be adulterated. Additional regulatory and customer requirements can also apply for food and drug packaging applications. For example, a need to meet kosher or halal requirements will require careful selection of oleochemical and other components as well as attention to all stages of processing and transportation of such materials to ensure suitability for use. Legislation enacted by the Coalition of Northeast Governors (CONEG) restricts the content of lead, cadmium, mercury, and chromium(VI) in packaging materials in many U.S. states. The presence of compounds listed under California Proposition 65 can require labeling. In addition, the use of products that are associated with allergies can be of concern, whether as lubricants or in other manufacturing applications where contact with the metal occurs. Among the most common allergens are latex, as well as food allergens, such as milk, eggs, soy, wheat, fish, shellfish, peanuts, and treenuts. Furthermore, it should be noted that certain families of esters and surfactants, such as phthalate esters and nonyl phenol derivatives contain members that are suspected to act as endocrine disruptors.
19.3.3 Process Equipment Lubricants Although equipment suppliers will normally have lubricant recommendations for use in their machinery, it may be desirable to lubricate certain process equipment with the metalworking formulation or a close approximation to it. This can, for example, apply to mill situations where connections in the spindles driving the rolls may be lubricated by the mill lubricant. Similarly, separate systems performing hydraulic or gear oil functions with a tendency to leak into the metalworking fluid or onto the aluminum part may be lubricated with the process fluid composition or simply the organic components of a waterbased product. This can provide advantages in the control of the process fluid composition and, through careful formulation, provide a means to address U.S. Food and Drug Administration (FDA) concerns or metal staining issues resulting from contamination. The formulations of these fluids may preclude them from fully meeting normal industry performance standards and their limited commercial use may not be
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sufficient to justify extensive performance testing. Nevertheless, they may provide adequate performance and be a good choice when the advantages provided on leaking into the process fluid are considered. An alternate approach to this challenge is the use of commercial machinery lubricants formulated to provide some of the following properties: low stain, FDA compliance, efficient misting, and good EP properties. As with the above products, their performance is likely to fall short of that of fully qualified machinery lubricants. When leakage into the process fluid or contact with the aluminum surface occurs, the impact of these products, based on such base oils as linear paraffins, hydrotreated light petroleum fractions, and polyalkylene glycols, should be less troublesome than conventional mineral-oil based products. Polybutenes can also be used as a base oil or, where compatible, as a thickener for other base oils. The potential effect of the polyalkylene glycol products on the emulsion stability properties of water-based formulations would need to be checked before use. One approach to the control of contamination, particularly for cold rolling formulations, is distillation and reuse of the recovered distillate in the process. For products whose components distill over a narrow range, this is an efficient means to separate the product from heavier machinery lubricants and higher molecular weight materials such as soaps that can form in the process. Distillation is typically performed under vacuum to minimize degradation reactions in the still that can reduce yield and also generate a burned odor that can be very persistent. Distillation may also be used to recover rolling components from control systems that use oil to capture organic materials in the mill exhaust. In either instance, it is important to analyze the distilled material to establish its composition since some components may not be fully recovered. Verification that the products still meet applicable FDA requirements, such as the UV absorbance limits for base oils, should be performed. A significant challenge occurs where the use of fire-resistant fluids in process equipment is deemed necessary and the potential exists for them to leak into the process fluid. The water-based products such as water-glycols, soluble oils, invert emulsions, and high water-based fluids, can be expected to alter the stability and performance of water-based process fluids and, therefore, not be preferred choices. Among anhydrous fluids, phosphate esters are subject to hydrolysis to generate phenols, a potential wastewater concern, along with acid phosphates, which are strongly acidic and can subsequently form salts similar to those of fatty acids. Other anhydrous fluid types are available that afford improved fire resistance over mineral oils in many situations. If leaked into a process lubricant, polyalkylene glycol derivatives (polyether polyols) and polyol esters have the potential to contribute to aluminum metalworking performance, although the former may also affect water-based fluid stability. The latter are subject to the side reactions noted above for esters and, if based on unsaturated fatty acids, are subject to oxidation and polymerization. Amine-based additives in these products, such as antioxidants in the polyol esters, have the potential to form salts with fatty acids in the formulation and alter its performance. An option to be considered is the implementation of improved equipment and personnel safeguards at the process to enable the use of mineral oil based products or other more compatible machinery lubricants with reduced fire resistance. The use of misting systems for process equipment lubrication can have the drawback of fugitive mist issues including inhalation concerns as well as potential deposition on the sheet. This may raise customer acceptance concerns in such products as foil for packaging applications. The use of air/oil systems to deliver metered volumes of oil in such applications may be a good alternative.
19.3.4 Filtration Metalworking processes typically generate debris that partitions among the metalworking fluid, the surface of the tooling, and the surface of the workpiece, where it may be termed smudge or smut. The debris consists primarily of particles of a mixture of aluminum or aluminum alloy and oxides as typically found at the metal surface. The composition of this surface material and the ease with which it is removed to generate debris will vary with alloy and processing history. The debris may also include particles from wear and corrosion of the tooling and system components as well as filter media. Most commonly, debris levels are controlled through filtration.
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The nature and size of the debris generated are closely linked to the contact conditions in the process and lubricant chemistry. In some processes, large amounts of relatively fine debris are formed, including submicron debris that settles very slowly and filters poorly. Fatty acids are known to contribute to the production of fine debris. The chemistry will also influence the degree to which those particles are dispersed individually or agglomerated into larger, more easily filtered particles. Increasing levels of debris in the fluid can lead to a number of concerns for the process. It can be expected at some concentration to lead to the generation of product with poor surface quality. In addition, high debris levels can lead to changes in the stability of emulsion products and the increased viscosity of debrisladen fluid will further reduce filtration speed. On the other hand, if much of the debris is relatively large, a settling step may be at least a partial answer to the need for cleaning the fluid. However, deposition of larger or agglomerated particles on mill equipment can open the door for accumulations to retransfer to the product and lead to surface quality problems. Larger and more numerous debris particles also may be more likely to lead to wear in any process equipment lubricated by debris-laden fluid [19]. Common types of filters include flatbed filters, with vacuum applied to the bottom side, and plate-andframe filters that operate under moderate applied pressure, often with the use of added filter media. The former can index the filter paper automatically based on the level of fluid on the bed, whereas the latter require “blowdowns” at intervals. When the pressure differential across the filter reaches the design limit, the filter is taken off line for renewal of the paper, followed by the deposition of a uniform layer of the filter media. Plate-and-frame filters are mostly used for neat oil streams and can be very efficient at removing even very fine debris. Commonly used filter media include diatomaceous earth, a product that must be handled carefully because of the inhalation hazard associated with its content of crystalline silica, and cellulose products. Diatomaceous earth is supplied in a variety of grades varying in their ability to filter out fine particles. Cellulose products have the advantage of being biodegradable and can be impregnated with citric acid for conversion of fatty acid salts in the lubricant back to the fatty acid. Two drawbacks with the use of filter media are increasing disposal costs and the value of the significant level of entrained lubricant in used product, which can be on the order of 30 to 40% by weight for used diatomaceous earth. Recovery of the lubricant or its energy value may be desirable. The conductivity of neat oil lubricants is of importance, since low conductivity can lead to static charge buildup during processes such as filtration. This leads to a risk of fire from static discharge, especially with products of relatively low flash point. Increasing the conductivity, either through fatty acid use or the use of conductivity-enhancing additives, along with grounding at key locations in the system, can help. The choice of filter paper needs to be made carefully so that its capacity and pore sizes are well matched to the needs of the system. The chemistry of the paper also needs to be tailored to the lubricants to be filtered to ensure that its wetting properties are well matched to the fluid and that it does not affect the stability of water-based products. In the event that solids loading and fluid viscosity are both high, the solids content may simply increase over time because conventional filtration is ineffective or impracticable. Continued development of alternate filtration techniques, such as electrostatic methods or membrane methods, may provide an answer to these shortcomings. In some instances with higher cost formulations, it may be economical to chemically remove the fines by treatment with acid or base, recognizing that this may lead to chemical changes in the lubricant, including changes in the relative levels of fatty acids and their soaps as well as ester hydrolysis.
19.3.5 Waste Treatment Minimizing the waste from metalworking operations is highly desirable from both an environmental and economic perspective. It may be feasible simply to burn waste oil or oil-solids mixtures in suitable burners to recover their heat value. For water-based formulations, it can be a significant challenge to achieve efficient separation of the oil and other organic components from the water to obtain suitably pure water and oil phases for subsequent treatment. A common approach is the use of heat together with acids or surfactants, or both, perhaps with the aid of flotation technology. Ultrafiltration is another
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option for concentrating water-oil mixtures, as is evaporation of the water phase, a process that may be economical using waste or renewable energy heat sources. The quality of the water obtained will depend on its level of contamination and the capability of the available treatment processes to purify it. In general, the water can be expected to contain a fraction of those components that are not exclusively oil soluble. This can include a portion of the surfactants, particularly those with high hydrophile-lipophile balance (HLB) values, as well as amine salts to the extent they are not associated with the organic phase. It can also include products such as low molecular weight alcohols generated by hydrolysis of esters based on them or materials present as coupling agents, or products of oxidation or microbial activity. Low levels of rust and corrosion inhibitors, antioxidants, and biocides from the process oil as well as additives from contaminant (tramp) oils might also be present, and can all contribute to the biological oxygen demand (BOD) of the water. After initial separation from any oil, further treatment of the water may be required to minimize disposal costs, meet local discharge regulations, or render it suitable for continued use in order to meet water conservation goals. This may be simply an adjustment of the pH to the desired range, a process that may precipitate excess levels of certain metal ions whose concentrations in the wastewater are regulated. Additional methods of wastewater treating to reduce BOD or oil and grease content can include microbial treatment or oxidation or adsorption processes using, for example, lime or activated carbon. In some instances, reverse osmosis may be a means to achieve the final purity needed. The choice of a treatment option could be based on both the water quality obtained and the options for disposal of the concentrated waste. It is important when formulating or evaluating a water-based lubricant to assess its ability to be treated by available wastewater systems. The compositions and amounts of the surfactants in the formulation have a strong influence on how successfully the system will separate the phases. It should also be recognized that surfactants vary considerably in their toxicity to aquatic organisms. The formulation with the most robust metalworking performance may not be the best for waste treating and a compromise is likely to be needed that balances those requirements.
19.4 Lubrication in Selected Aluminum Metalworking Processes 19.4.1 Rolling Modern rolling technology integrates a complex blend of lubricant technology with surface technology and metallurgy together with very precise process control. Cold mill sheet exit speeds of 1500 m/min are possible while sheet thickness is controlled to within a few microns. In modern continuous mills, feed coils can be joined without stopping the mill through the use of accumulators that maintain a continuing supply of metal to the mill during the coil joining process. Multi-stand or tandem mills provide very efficient production of many alloys at finished hot or cold mill gauge. In the rolling process, the lubricant is called upon to deliver consistent lubrication, to assist in controlling roll geometry by controlling roll temperature, and to aid in the management of debris generated in the contact. In hot rolling, water-based formulations are necessary in order to minimize fire risk. However, in cold rolling, the use of water makes possible increased reduction at high rolling speed through increased heat removal compared to neat oil formulations. The use of water-based formulations also provides an opportunity for reduced rolling emissions compared with neat oil formulations. As discussed earlier, water-based lubricants can lead to water stain on the finished product, especially if care is not taken to remove all water quickly and efficiently from warm aluminum surfaces. The surface quality of the rolled product at high levels of gauge reduction is very sensitive to the composition and the thickness of the oil film present in the lubricated contact. While marginal lubrication may be tolerated and can be beneficial in avoiding refusals in early hot rolling passes, as one progresses towards the final hot rolled gauge, the quality of the lubrication and its influence on the metal surface become increasingly important. Significant surface flaws from hot rolling may persist during cold rolling such that the metal will not meet finished product requirements. One approach to this need that is easier
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if more than one hot mill is available is the use of relatively stable emulsions for the early passes on a reversing mill, but less stable emulsions with better oil availability for later passes or coiling passes. After hot rolling, surface quality can be evaluated by anodizing, a process that highlights discontinuities in the surface that can be present even if the as-rolled surface appears to be uniform, and by examining for surface smoothness. A common surface feature that reflects weak lubricant performance is a topography resembling orange peel (Figure 19.3). This has been studied as a function of lubricant composition and properties [20]. An inadequate lubricant film is unable to prevent adhesion of the hot sheet to the work roll, leading to the dislodging of surface metal and oxide that subsequently can be repositioned on the surface as sliding occurs or transferred to the roll where it may be retransferred to the sheet. The casting of strip that is relatively thin through the use of belt or roll or block casters can dramatically reduce the need for extensive hot rolling compared with ingots that can range up to about 600 mm thick. In this streamlined process, there is much less opportunity for the elimination of surface flaws with the reduced number of hot rolling passes. In cold rolling, as in hot rolling, lubricant chemistry strongly influences how much gauge reduction is possible before lubricant failure occurs. Rolling in a mixed lubrication regime enables new surface to be generated under controlled contact conditions and generally provides acceptable surface quality. This minimizes problems related to the dull surface caused by transverse fissures generated with the thicker lubricant films present in hydrodynamic lubrication. It also limits the adhesion and abrasion that can lead to higher friction and greater heat generation occurring with minimal lubricant films. A pattern of marks in the sheet called herringbone [21] can signal lubricant failure and a likelihood of strip breakage. The greater power needed to drive the rolls under these boundary lubrication conditions may limit the output of mills operating at or near peak power. In the rolling process, the surface of the rolls also plays an important role. Rolls with freshly prepared surfaces will have a higher roughness and generally more angular surface asperities that will tend to cut into the aluminum surface more than rolls whose surface topography has become smoother in the course of an extended rolling campaign. Thus, as the nature of the roll surface changes, so will the characteristics of the debris formed. If a ground rolling surface is modified to provide improved wear life, through such means as chrome plating, changes in friction can result from the material change as well as changes in the topography resulting from the plating process. Both can affect the nature of the contact and debris formed. As roll wear proceeds, the lubrication regime will shift towards hydrodynamic and when the limit of the acceptable process window is reached, resurfacing of the roll will be required. Most rolling is performed using rolls whose surfaces are prepared using grinding wheels that generate a pattern resembling an array of canoes of different lengths aligned in the rolling direction. Alternate roll surfacing techniques are also available, such as shot peening, electron beam, electro discharge, and laser texturing. These methods provide an opportunity for alternate surface appearances in the finished part, including more isotropic surface topography, and have begun to find commercial application. The targeted tooling surface roughness will be based on the specific product and its surface finish requirements. Typical ground surface roughnesses range from an Ra of about 1 to 3 µm for hot rolling down to about 0.1 µm for foil. Very detailed decorative images can be imparted to sheet surfaces using embossing techniques. The amount and characteristics of debris formed in the rolling process can be expected to depend not only on lubricant chemistry, roll material and topography, alloy, and processing history but also on rolling conditions. In the rolling process, sliding of the workpiece against the roll will occur at all points in the arc of contact except for the neutral point, where the sheet and roll surface are moving at the same speed (Figure 19.4). At this point, the sheet is changing from moving more slowly than the roll to faster. For a given pass, higher levels of slip on entering and exiting the contact provide an opportunity for more abrasive wear, as does the use of larger diameter rolls providing a longer arc of contact. If rolling conditions are such that the roll surface develops and retains a significant level of aluminum alloy and oxides, this roll coating can leave an imprint in the metal contributing to the orange peel appearance. In hot rolling, where the roll coating is often substantial, scratch brushes are commonly used to perform a continuous scrubbing action on the roll surface. This provides control of the amount of roll coating and thus provides a roll surface to the rolling contact with more consistent topography and
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Handbook of Lubrication and Tribology Direction of rotation
Arrows adjacent to neutral point show relative amount and direction of metal speed compared with roll speed
Direction of travel Neutral point
FIGURE 19.4
Sliding of strip on the roll bite.
friction characteristics. The brushes can be made of steel wire, which can lead to product quality problems should any of the wires become detached and enter the roll bite, or a polymer such as nylon impregnated with abrasive particles. On the aluminum surface, some debris is relatively loosely bonded and can be removed and measured as smudge by various wiping or cleaning procedures [22]. The ease of removal of this material tends to diminish over a period of days. Some debris is pressed into the surface, but may be released by further rolling contacts or stretching of the surface in which it is embedded. Where subsequent processing of the rolled metal requires a cleaned surface, lubricant residues, including associated smudge, can be removed using mildly to strongly etching basic or acidic cleaning media. A variety of solvents from petroleum or renewable resources can be used to remove lubricant residues; however, such processes need to be designed carefully to minimize emissions and fire risk. The use of chlorinated solvents for this application is diminishing because of hygiene and environmental concerns. The cleanliness of the surface will reflect the level of residue present in the cleaning medium and the efficiency of its removal, and can be assessed by techniques measuring the water wettability of the surface. It is a significant challenge to provide an optimized lubricant film throughout the rolling process that enables finished product requirements to be met. Whether in hot or cold rolling, the lubricant or lubricants must accommodate the needs of multiple passes through the rolls, whether in single stand or tandem mills, as process conditions vary over a wide range. Wilson and Walowit [23] developed relations linking film thickness for the rolling contact to a series of parameters, with the relation for inlet film thickness being given in Equation (19.3). 3µ0 γ a(U1 + V ) h1 = (19.3) x1 (1 − e−γ (σ −s) ) where h1 is film thickness at the inlet edge of the work zone, µ0 is fluid viscosity at atmospheric pressure, γ is pressure coefficient of viscosity, a is roll radius, U1 is strip velocity at inlet, V is roll velocity, x1 is distance to inlet edge of work zone (from line connecting axes of top and bottom rolls), σ is material flow stress, and s is back tension stress. An inspection of the equation provides insight into the influence of not only viscosity and speed considerations, but other parameters, such as material flow stress, roll diameter, reduction, and the pressure–viscosity properties of the lubricant. Fluid viscosity at temperatures found in cold rolling can increase by roughly an order of magnitude under typical aluminum forming pressures. Until relatively recently, experimental verification of the thickness of oil films in the roll bite had not been achieved. However, x-ray work by Dow [24] and optical interferometry developments by Cameron and coworkers [25] have enabled new insights into this area. In the interferometry method, the analysis of interference patterns of light reflected from the contact between a roller and a partially reflecting transparent disk enables a calculation of the lubricant film thickness. The method has been refined by Spikes and coworkers [26] to enable the measurement of film thicknesses of under 1 nm under carefully
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Aluminum Metalworking Lubricants
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controlled conditions. This equipment, together with equipment that can monitor traction between metal specimens under similar contact conditions, enables the determination of the influence of formulation and surface topography over a moderate temperature range. In the past decade, interferometry work has led to significant new understandings in the performance of emulsions in aluminum metalworking. At lower speeds with oil-in-water emulsions, Zhu, et al. [27] observed the formation of an oil pool in the inlet zone. The measured film thickness initially increased with speed and was usually quite similar to the film thickness for the neat oil until the speed reached the “first critical speed.” At this point, the film thickness was observed to decrease with increasing speed, although in some instances it continued to increase at a reduced rate before beginning to decrease. As rolling speed increased further, it was observed that a “second critical speed” was reached, at which point the film thickness again increased with increasing speed. The observed film thicknesses remained less than those for neat oil but were greater than those for pure water. However, the approach to the values for oil was closer for destabilized emulsions or emulsions of higher concentration. The mechanism of formation of the oil pool observed at lower speeds at the inlet is not well understood in terms of the relative contributions of contacts of oil droplets with the surfaces outside the contact region and the alternate mechanism of entraining droplets in the inlet region to the contact. In emulsion lubrication, it has generally been believed that plate-out, which is the deposition of an oil film or the wetting of the metal surfaces by the oil, is a key to providing an adequate oil supply to the contact. The process of generating an oil-rich phase that enters the contact represents an inversion of the oil-in-water emulsion. The efficacy of the plate-out process, which will influence the first critical speed, has been related to the wetting of the metal surface by the oil droplets according to the nature and concentration of the emulsifier system, as well as pH and water hardness [28]. These properties also influence the droplet size and relative looseness or tightness of the emulsion. A preference for oil rather than water to be entrained into the contact has been shown in studies with single emulsion droplets [29,30]. However, above the first critical speed, which is variable and depends on the specific conditions, there is insufficient oil available to the contact for a full film to be present. At this point, the greater oil volume needed for a full film as both speed and film thickness increase could now be considered to exceed the volume that is plating out. Alternatively, for a point contact situation such as a bearing, a deficiency of lubricant or starvation might take place if repeated contacts by rolling elements occur before surface tension can draw back the lubricant displaced in the previous contact [31]. Speeds near or above the second critical speed values of 0.15 to 5.0 m/sec of Zhu, et al. are representative of commercial rolling speeds. For an adequate film to form at higher speed conditions, a mechanism of wetting and film formation is needed that applies under relatively high shear conditions. A dynamic concentration theory has been proposed that particularly applies to line, rather than point, contacts [29]. Since the film thickness in rolling contacts is commonly less than the emulsion droplet size, droplets approaching the inlet will bridge the gap between the roll and the sheet. This is shown schematically in Figure 19.5. With the narrowing of the gap as the contact zone is approached, the oil droplets will be squeezed to a flatter shape and continue to be drawn in with their viscosity increasing exponentially with pressure. As droplet concentration and size increase, the film thickness also will increase. The water is preferentially left behind, inverting the emulsion. Continued work is needed to provide a clearer picture of the mechanism(s) of lubrication at higher speeds.
Oil Water
Phase inversion
FIGURE 19.5 Schematic representation of two-phase lubricant in the rolling contact.
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Handbook of Lubrication and Tribology
19.4.2 Lubricants in High-Temperature Processes Lubricants are used as forming aids and release aids in a number of higher temperature aluminum metalworking processes that involve metal temperatures ranging from about 300 to over 700◦ C. Such processes include casting, extrusion, forging, and superplastic forming. Aluminum is cast commercially in a variety of ways that place a range of demands on the lubricant. Often, it is important that lubricants for casting act as release agents. However, where sliding occurs, they must also provide lubrication and, in combination with the choice of tooling material, serve to minimize adhesion of the very soft hot metal to the tooling and erosion of the tooling by the metal. In roll casting, lubricant is continuously applied to the rolls that contact and squeeze the solidifying metal to the desired thickness. Materials used in such applications often utilize carbon, whether as graphite, which may be applied as an aqueous suspension, or soot deposited from a smoky flame translated across the roll surface. Similarly, lubricants containing solids are commonly used with block or belt casters, with a significant concern being the avoidance of excessive levels of residue that may impart a dirty appearance to the aluminum surface after rolling. In die casting, a variety of water-based products are used to lubricate moving parts used for metal transfer and to facilitate release of the solidified part with good surface appearance. Lubricants used in direct chill ingot casting molds facilitate the formation of a uniform surface layer on the solidifying ingot and the smooth travel of that surface against the mold while controlling the formation of flaws that can lead to cracks in subsequent processing [32]. Although greases have been used, commonly used lubricants include vegetable oils and derivatives based on oleates and ricinoleates that have high flash points and are supplied to the metal-mold interface. Preferred materials have high decomposition temperatures and function without leaving undesirable residues on the mold surface or tenacious stains on the ingot surface. In the hot extrusion process, lubricants containing solids such as graphite or boron nitride can be used to lubricate the die surfaces, but frequently no lubricant is used [1]. With no facile way to apply lubricant continuously to the die surface during the extrusion process, a functionally useful amount of lubricant can be expected to persist only for a limited length of extrusion. The use of excess lubricant may give surface quality defects on the extruded part. Cleaning the die with caustic material can provide the benefit of removing aluminum pickup; such compositions can be combined with graphite. Tooling considerations, such as the die surface composition and roughness, are important factors for minimizing adhesion of the hot aluminum to the die. The hot forging process also provides an array of significant lubrication challenges, which has led to the development of products tailored for specific forging jobs. In certain applications, the lubricant must provide excellent forming performance in parts where large amounts of fresh aluminum surface are generated at die temperatures that can exceed 425◦ C. In precision hydraulic forging applications, the lubricant needs to provide good performance for net or near-net finish on key surfaces and may be called upon to function during a contact time that can extend to tens of seconds. For such products, the ability to capture fine detail is more important than excellent forming performance, but in either instance it is important to avoid adhesion of the workpiece to the die and subsequent scuffing of the surface as it slides over the die surface. The die surface material and topography also play a role here. The surface requirements of many forgings require facile lubricant residue removal that does not rely on aggressive etching. Oil-based forging lubricants containing lubricious suspended solids such as graphite have typically been applied to the die and often to the billet to facilitate forming. Hygiene and environmental improvements have been made by switching from lead-based additives to products containing other carboxylates such as those of tin and bismuth and by adjusting formulations so that they generate lower levels of particulate and organic emissions. Increasingly tight environmental limits have driven an aggressive pursuit of waterbased products in recent years. A partial answer comes from the use of water-based precoats containing solid lubricants that can be applied to billets that have been warmed to assist in forming an even coating. Precoats can assist in making preforms or, where the forging requirements are demanding, can act as a supplement to the usual lubricant.
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With water-based lubricants, the rapid boiling and draining of the water upon contact with hot surfaces inhibits efficient and uniform deposition of the lubricating components. Therefore, a key to success in this process is the efficient deposition of lubricating components on hot die and billet surfaces without resorting to high levels of organic components that contribute to emissions. Progress continues to be made in the application and composition of water-based products such that they are approaching equivalence to the older generation oil-based products. Hot forming processes, such as superplastic forming of aluminum alloys for aerospace or automotive applications, are typically lubricated with formulations based on boron nitride. Graphite has also been used in recent works by Davies, et al. [33] and by Osada and Shirakawa [34]. In this process, the periphery of the sheet is sealed in the die and the forming can be done using a gas pressure of no more than a few megaPascals. The lubricant must facilitate the smooth movement of the metal over the die features during a time period that may last up to tens of minutes. It must complement the die design to enable forming without excessive part thinning, and should enable facile release of the hot and very soft formed part from the die without distorting its shape. In addition, for faithful replication of the detailed shape and good surface appearance, the lubricant must be applied in a very thin evenly distributed layer that needs to adhere well to the aluminum surface. It must function without allowing a transfer or buildup of aluminum on the die that can scuff the soft part as it forms and without the generation of accumulated lubricant on the dies that can lead to imprinting of those residues into the formed part.
19.4.3 Lubricants for Forming of Finished Food and Beverage Containers As noted previously, the lubricants used for the manufacture of sheet and foil products for food and beverage packaging must comply with regulatory requirements. In the United States, the components and limitations in 21 CFR 178.3910(b) provide a starting point for formulating lubricants for drawing, stamping, or forming these articles. For components listed therein, the lubricant residue limit on surfaces contacting the food or beverage is 0.2 mg/in.2 (310 mg/m2 ). In addition to good metalworking performance, customers will frequently have further requirements based on dietary or product compatibility considerations, or on organoleptic requirements involving taste and odor evaluations. A major share of the sheet rolled for food and beverage applications is used in the manufacture of drawn and ironed cans, with U.S. annual production being in excess of 100 billion cans in recent years. To make the can bodies, a series of lubricants are used that in most instances are not FDA compliant; however, the finished can bodies are cleaned using etching solutions and then coated with an FDAcompliant polymeric film that serves as a protective barrier to minimize any potential interaction with the packaged product. Commonly, a reoil is applied to can body stock after rolling that serves to protect the metal from scratching and moisture during handling, transporting, and storing. Together with the cupper lubricant, it aids in the forming of a cup or preform that is shorter and wider than the final shape. The cups are then converted into cans having the final diameter and full wall height by redrawing and ironing, a high-speed operation that uses water-based lubricants [35]. These bodymaking lubricants can range from soluble oils to synthetics (oil-free water-based formulations). Like some hot rolling formulations, these compositions are often based on an emulsifying combination of fatty acids and amines, and can also contain additional surfactants, such as nonionics, along with esters and oxidation and corrosion inhibitors. Satisfactory performance in the bodymaker process requires compatibility with the reoil and cupper lubricants, which are brought into the lubricant with the incoming cups, as well as an ability to function well in the presence of gear oil that can leak in from the tooling. The avoidance of any additional contaminants and the skimming of insoluble contaminants and byproducts from the lubricant surface can aid in prolonging lubricant life. Microbial control is also a key for these products as is good control of debris through filtration. The necking operation, in which the trimmed top of the cup wall is reduced in diameter to enable the use of a smaller, lighter lid, can be aided with a light hydrocarbon solvent as lubricant.
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To the extent possible, good control of the composition is a key to consistent performance in the bodymaker. As in the rolling process, the details of the lubricant composition and the droplet size distribution of emulsion products can be expected to influence the lubricant film thickness and the tribological conditions in this high-speed process. One flaw, referred to as bleed through, is a dark and irregular undercast on the decorated sidewall. This has been attributed by Knepp [36] to a high degree of hydrodynamic lubrication during the deforming of the metal; however, Japanese workers have ascribed the problem to lack of lubrication in the drawing and redrawing process [37,38]. High productivity in can plants requires a very low level of down time through the minimization of problems such as unsuccessfully formed cans, known as tearoffs, or excessive wear of the tooling, which is made of highly wear resistant material such as tungsten carbide. Consistent metal surface topography, carefully tuned metallurgy, and careful filtering of the molten metal to limit inclusions are important for both high productivity and the formation of cans with good surface characteristics. A second surface flaw is looper lines, which are visible patterns in the can wall showing the roughness in the sheet imparted by the ground rolls. If pronounced, they may not be satisfactorily covered by the can decoration. An alternate strategy for the formation of drawn and ironed cans is the application of a polymer to the metal prior to forming [39]. In many instances, the lubricant used for the forming operation is combined with or applied over a polymeric coating on the aluminum. For example, aluminum can ends typically have a coating based on epoxy, acrylic, or vinyl polymers that can contain an internal lubricant that blooms to the surface as the polymer is cured. Such lubricants include waxes, lanolin, and polyethylene. External lubricants or postlubricants can be applied onto the cured polymer, using electrostatic or roll-coating techniques, with warming as necessary to facilitate application. Application from solvent solution is well suited to high-speed lines but requires the management of significant environmental, health, and safety concerns associated with the solvent. To a certain extent, the polymeric coating reduces the need for good film strength in forming lubricants applied over it. Commonly used compositions include petrolatum and paraffin, or blends of paraffin with materials such as lanolin. For waxy compositions, such as those containing sizeable levels of paraffin, the potential for the build up of tenacious residues on the tooling must be recognized. The lubricant, in combination with the coating and the metal, can be looked at as a system that enables facile high-speed formation of high-strength ends with very low levels of flaws such as splits in the more severely formed areas. The lubricant requirements include: • • • • • • • •
Regulatory compliance Sufficient lubrication performance Ease of uniform application Absence of deleterious effects on coating performance Absence of undesirable interaction with the packaged product Absence of negative organoleptic effects Reliable supply of consistent product Ability to decorate ends over the lubricant if necessary
In the instance of packaged beverages, undesirable interactions can include such visual concerns as an appearance of higher turbidity, undesirable changes to the consistency and longevity of foam on beer, or the occurrence of “film float,” which is a visible film of lubricant on the surface of the product. The presence of polyunsaturated fatty acids and their derivatives, typically linoleates, has been linked to the formation of unsaturated aldehydes, which in trace amounts can have a very deleterious effect on beer taste. Such products must be scrupulously avoided as lubricants or potential contaminants in this application [40]. Minimizing or eliminating lubricants on the surface of finished can ends that contacts the product can aid in minimizing undesirable product interactions. Where no polymeric coating is present, the lubricant may be relied on to provide a higher level of performance or film strength than is normally provided by such products as petrolatum and waxes. This applies to the multi-step formation of the tabs used for easy-open aluminum can ends, since much of the metal supplied currently is not coated with polymer. Because the potential exists for transfer of lubricant
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from the tab on one lid to the product side of the end nested next to it in stacks of finished ends, the above concerns about product compatibility can extend to tab forming lubricants. In certain other applications, such as aluminum pie pans, the lubricant is called upon to minimize the release of any smudge remaining on the sheet and, in effect, also serve as a coating. This provides a difficult challenge since boundary additives tend to generate smudge and in many instances serve to solubilize existing smudge. On the other hand, waxy materials, such as paraffin, can serve as a coating but provide little film strength by themselves and are prone to cause die buildup.
Acknowledgments The author thanks Alcoa Inc. for permission to publish and Simon Sheu for assistance with the figures.
References [1] Schey, J., Tribology in Metalworking, American Society for Metals, Metals Park, 1983. [2] Overfelt, R.A., Wert, J.J., and Hunt, W.H., Jr., The influence of thermal oxide characteristics on the friction behavior of aluminum auto body sheet alloys, ASLE Trans., 24, 175, 1981. [3] Hatch, J.E., Aluminum: Properties and Physical Metallurgy, American Society for Metals, Metals Park, 1984, 156. [4] Patir, N. and Cheng, H.S., An average flow model for determining the effects of three-dimensional roughness of partial hydrodynamic lubrication, ASME J. Tribol., 100, 12, 1978. [5] Jahanmir, S., Chain length effects in boundary lubrication, Wear, 102, 331, 1985. [6] Underhill, R. and Timsit, R.S., Interaction of aliphatic acids and alcohols with aluminum surfaces, J. Vacuum Sci. Technol., A10, 2767, 1992. [7] Guangteng, G. and Spikes, H.A., The control of friction by molecular fractionation of base fluid mixtures at metal surfaces, Tribol. Trans., 40, 461, 1997. [8] Rehbinder, P.A. and Shchukin, E., Surface phenomena in solids during deformation and fracture processes, Prog. Surf. Sci., 3, 97, 1972. [9] Wei, J., Erdemir, A., and Fenske, G.R., Dry lubricant films for aluminum forming, Tribol. Trans., 43, 535, 2000. [10] Rohrbach, P., Hamblin, P., and Reyes-Gavilan, J., The benefits of a high molecular weight phenolic antioxidant compared with BHT and 2,6-DTBP, Tribol. Lubr. Technol., 60, 56, 2004. [11] Burrell, R.E., Firko, H.T., and Heenan, D.F., Biocide replacement in an aluminum cold rolling mill. A case history, Lubr. Eng., 56, 675, 1990. [12] Leger, A.E., Haines, R.L., Hubley, C.E., Hyde, J.C., and Sheffer, H., The structure of aluminum diand tri-soaps, Can. J. Chem., 35, 799, 1957. [13] Buehler, C.A. and Pearson, D.E., Survey of Organic Synthesis, John Wiley & Sons, New York, 1970, 686. [14] Wright, R.L., Jr., Formation of the neurotoxin TMPP from TMPE-phosphate formulations, Tribol. Trans., 39, 827, 1996. [15] Serna, P.F. and Louis, E., Linear paraffin-based cold rolling lubricants, Aluminium, 52, 120, 1976. [16] Hombek, R., Heenan, D.F., Januszkiewicz, K.R., and Sulek, H.H., Oxidation of aluminum cold rolling base oils, Lubr. Eng., 45, 56, 1989. [17] Reich, R.A., Stewart, P.A., Bohaychick, J., and Urbanski, J.A., Base oil properties of ionic liquids, Lubr. Eng., 59, 16, 2003. [18] Laemmle, J.T., Metalworking lubricants, in ASM Handbook, Volume 18, Friction, Lubrication, and Wear Technology, Blau, P.J. (ed.), ASM International, 1992. [19] Rowe, C.N., Lubricated wear, in Handbook of Lubrication. Theory and Practice of Tribology. Volume II. Theory & Design, Booser, E.R. (ed.), CRC Press, Inc., Boca Raton, 1994, 218. [20] Kurachi, T., Yoshida, T., and Suzuki, K., Influence of oil composition and emulsion properties on the rollcoating buildup during hot rolling of aluminum, Lubr. Eng., 43, 660, 1987.
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[21] Visual Quality Characteristics of Aluminum Sheet and Plate, 4th ed., The Aluminum Association, Washington, 2002, 28. [22] Bekmesian, G. and Januszkiewicz, K.R., Measurement of aluminum wear debris on rolled sheet, Lubr. Eng., 51, 901, 1995. [23] Wilson, W.R.D. and Walowit, J.A., An isothermal lubrication theory for strip rolling with front and back tension, Tribol. Conv. 1971, Institution of Mechanical Engineers, London, 164, 1972. [24] Dow, T.A., A rheology model for oil-in-water, SME Technical Paper No. MS77-339, Society for Manufacturing Engineers, Dearborn, 1977. [25] Hamaguchi, H., Spikes, H.A., and Cameron, A., Elastohydrodynamic properties of water in oil emulsions, Wear, 43, 17, 1977. [26] Glovnea, R.P., Forrest, A.K., Olver, A.V., and Spikes, H.A., Measurement of sub-nanometer lubricant films using ultra-thin film interferometry, Tribol. Lett., 15, 217, 2003. [27] Zhu, D., Biresaw, G., Clark, S.J., and Kasun, T.J., Elastohydrodynamic lubrication with O/W emulsions, Trans. ASME, 116, 310, 1994. [28] Ratoi-Salagean, M., Spikes, H.A., and Reiffe, H.L., Optimizing film formation by oil-in-water emulsions, Tribol. Trans., 40, 569, 1997. [29] Schmid, S.R., and Wilson, W.R.D., Lubrication of aluminum rolling by oil-in-water emulsions, Tribol. Trans., 38, 452, 1995. [30] Reich, R.A., Epp, J.M., and Festa, R.P., A method to study the mechanism of lubrication of an O/W emulsion in cold rolling aluminum using H18 2 O, SIMS and TOF-SIMS, Lubr. Eng., 50, 31, 1994. [31] Chiu, Y.P., An analysis and prediction of lubricant film starvation in rolling contact systems, ASLE Trans., 17, 22, 1974. [32] Laemmle, J.T. and Bohaychick, J., Mold lubricants for casting aluminum and its alloys, Lubr. Eng., 48, 858, 1992. [33] Davies, R.W., Khaleel, M.A., Pitman, S.G., and Smith, M.T., Experimental determination of the coefficient of friction during superplastic forming of AA5083 aluminum alloy, Proceedings from International Symposium on Superplasticity and Superplastic Forming Technology, 7–9 October, 2002, Columbus, OH, ASM International, Materials Park, 2003, 39. [34] Osada, K. and Shirakawa, K., Mass production of a spare tire housing for an automobile, Advances in Superplasticity and Superplastic Forming, Taleff, E.M., Friedman, P.A., Krajewski, P.E., Mishra, R.S., and Schroth, J.G., (eds), The Minerals, Metals and Materials Society, Charlotte, 2004, 41. [35] Knepp, J.E., Lubrication in the manufacturing of drawn and ironed aluminum cans, Lubr. Eng., 40, 554, 1984. [36] Knepp, J.E., Causal mechanisms of the bleed-through defect on drawn and ironed cans, Presented at the American Society of Lubrication Engineers National Conference, Las Vegas, May 6, 1985. [37] Koyama, K., Kanbayashi, K., Inabayashi, Y., and Fujikura, C., A formation mechanism of the bleed-through defect of aluminum D&I cans, J. Japan Inst. Light Metals, 41, 18, 1991. [38] Ito, K., Tsuchida, S., and Takeshima, Y., Influence of redrawing on surface properties of aluminum D&I cans, Sumitomo Light Metal Tech. Rep., 32, 27, 1991. [39] Jaworski, J.A. and Schmid, S.R., Survivability of laminated polymer lubricant films in ironing, Tribol. Trans., 42, 32, 1999. [40] Kipers, K., Autoxidation of fatty material, Lubr. Eng., 46, 418, 1990.
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20 Mining Industry 20.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 20.2 Contamination Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
20-1 20-2
Sealing • Filtration
20.3 Bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
20-5
Rolling Contact Bearings • Plain Bearings • Bushes and Pins
20.4 Gears . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
20-9
Gear Lubrication • Determination of Viscosity for Spur and Helical Gears • Specific Film Thickness • Calculation of Elastohydrodynamic Lubrication Film Thickness • Load Carrying Tests • Full Scale Tests • Gear Test Rigs • Bench Tests • Reason for Pitting Fatigue • Surface Roughness • Influence of Lubricants on Pitting Fatigue
20.5 Wire Ropes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
20-14
Lubrication of Wire Ropes • Lubrication During Manufacture • Lubrication During Service • Lubricant Specifications
20.6 Power Hydraulic Systems. . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
20-17
Hydrokinetic Transmission • Hydrostatic Transmission • Hydraulic Fluids • Types of Hydraulic Fluid • Fire Resistant Hydraulic Fluids
20.7 Manufacturers’ Lubricant Specifications . . . . . . . . . . . . .
Will Scott Queensland University of Technology
20-20
Multipurpose Extreme Pressure (EP) Grease • Multiservice Mining Lubricant • Enclosed Gear-Case Oil • Wire Rope Lubricant • Antiwear Hydraulic Oil
References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
20-24
20.1 Introduction Mining is an expansive industry that may be categorized according to surface (open cut), underground, process plant, and exploration. Common to all categories is the dust-laden environment, rough loading of machines, supervisory difficulties, and health and safety of the workers. In many cases these are exacerbated by space constraints and humidity. Collectively, these conditions conspire against the tribologist and plant maintainer. Machines and equipment connected with mining and requiring lubrication are: 1. Fans, pumps, and compressors 2. Winders and haulages 20-1
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3. Longwall face equipment for underground coal mining 4. Continuous miners and shuttle cars that are now mostly used for underground roadway development 5. Draglines for overburden removal in surface mining 6. Drilling rigs 7. Crushers, comminution mills, and conveyors 8. Mobile equipment such as dozers, excavators, and load-haul-dump trucks 9. Bucket wheel excavators 10. Vibratory screen and centrifuges 11. Stacker-reclaimers Although the machines themselves may be most commonly associated with mining, their elements are ubiquitous. It is their size, configuration, and conditions under which they operate that pose a challenge to tribologists. Here we consider the lubrication of machine elements as they are found in the mining industry. In so doing, the roles of the designer, the manufacturer, and the operator/maintainer will be acknowledged. Most mining personnel and machinery manufacturers recognize the value of proper lubrication in achieving high reliability but few know how to apply lubrication technology effectively. Too often chronic lubrication problems, such as recurrent premature bearing failures, are treated in a superficial manner by repair or replacing and perhaps a change of lubricant when the real cause is that of an inadequate tribological design. Since a major source of machine failure in mining is solid contaminants it would seem logical to design so as to prevent their ingress. This may be achieved by special attention to filtration of the lubricant and sealing of the lubricated elements. It is seldom possible to eliminate the dirt completely and so it becomes necessary to establish the effect of the contaminated lubricant on the system. Recognizing this fact, SKF have developed a nonseizure bearing for conveyor idler rollers [1]. Particulate sensitivity of a lubricated component is an important parameter influencing system reliability. The sensitivity can be related to the minimum thickness of fluid films between adjacent surfaces. It can be appreciated that a large film thickness to particle size ratio will be less harmful to the system. Thick films are also beneficial under shock loading conditions. It should also be noted that the concentration level of contaminating particles, even where these are significantly smaller than film thickness, can be damaging due to silting, filter blocking, and erosion. Where the film results from a fixed clearance, and is therefore of a constant thickness, the size of the particle that may be tolerated is relatively easy to ascertain. However, where the film is generated by the motion and geometry of the surfaces, several variables are involved in establishing the film thickness and the particle size that may be tolerated is dependent upon pressure, velocity, and lubricant properties. Systematic tests to determine component particulate sensitivity have been few but papers relating to hydraulic components are relevant to mining applications [2,3]. Water is another source of contamination commonly found in underground mining machinery. It has been shown that even small amounts of water contamination can significantly reduce the life of machine components, especially rolling contact bearings [4].
20.2 Contamination Control The definitive work on this subject is due to the team at the former Fluid Power Research Center in Oklahoma [5]. In general, the life of a lubricated component is the operating time during which the magnitude of performance degradation is still acceptable. Since this degradation is the end result of wear, it is common to express component wear in performance related terms. Due to the design complexity of most lubricated components, it is virtually impossible to evaluate service life in terms of the material loss or dimensional changes that accompany the wear process.
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20.2.1 Sealing Seals may be used to keep a fluid from escaping from an enclosure or to prevent contaminants from entering and are classified under two main headings, that is, static and dynamic. Of the latter group, sealing that takes place between surfaces in sliding contact is the type that will be considered here. These seals operate on lubricant films usually in the fluid film mode. Where boundary lubrication is the predominant mode seal life is short, especially in mining applications where abrasives also contribute to reduce life. Limits of temperature, relative velocity between seal and counterface, pressure, size, fluid compatibility, and acceptable leakage rate will influence the choice of seal. Due to the small size of the fluid film in mechanical face seals, they require a flushing system where abrasives may be present. They are seldom used in a mining environment but may be found in centrifugal pumps as an alternative to packing glands. With rotary lip seals the sliding interface between the lip and shaft is normally separated by a thin film of lubricant. A general check list for their operating conditions is given in Table 20.1 [6]. Lubricant starvation often occurs where two or more seals are fitted in series even though the sealed fluid may possess good lubricating properties. This is a common “design modification” carried out on mining equipment on-site or at overhaul where it is considered that extra sealing is necessary. The inboard seal may function very efficiently and, in so doing, starve the remainder of the seal assembly of lubricant. The result is high friction, an overall rise in temperature of the shaft and housing, and in many cases excessive wear of the shaft. Soft piston seals and O-rings may or may not operate under fluid film conditions depending on the application (e.g., pneumatic cylinder piston seals operate in the predominantly boundary lubrication mode). Material selection and seal life will vary accordingly. Controlled tests on cylinder wiper seals showed a greater than tenfold difference in effectiveness [5]. Typical failures of soft piston seals are given in Table 20.2 [7].
20.2.2 Filtration Filtration is an effective way of removing solid particles from a circulating lubrication system. The provision of such a system can be expensive and, due to the extent of contamination in the mining environment, regular filter cleaning or element replacement is necessary to ensure adequate performance.
TABLE 20.1
Operating Conditions for Rotary Lip Seals
Up to 35 mm (1 12 in.) dia. 75 mm (3 in.) dia. Over 75 mm (3 in.) dia.
Approx. 8000 rev/min Approx. 4000 rev/min Approx. 15 m/sec (50 ft/sec) peripheral speed
Up to 75 mm (3 in.) dia. Over 75 mm (3 in.) dia.
Approx. 0.6 bar (10 psi) Approx. 0.3 bar (5 psi) By using a profiled backing washer to support the lip, pressures up to 6 bar (100 psi) can be accommodated
See table of rubber materials
Permissible oil temperatures are set by the sealing lip material. Do not ignore low-temperature conditions.
Housing
Better than 0.25 mm (0.01 in.) total indicator reading when clocked from shaft.
Shaft
Depends on speed. Aim for better than 0.025 mm (0.001 in.) total indicator reading when rotated in its own bearings
Housing
Fine turned. Provide lead-in chamfer
Shaft
Grind and polish to better than 20 µ in cla (0.5 µm Ra ) Surface must be free from all defects greater than 0.0025 mm (0.0001 in.) deep. Use cardboard protection sleeve during manufacture
Source: Reprinted from Jagger, E.T., Lip Seals, Tribology Handbook, 2nd ed., M.J. Neale (ed.), 1975, A37.
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20-4 TABLE 20.2
Handbook of Lubrication and Tribology Types and Causes of Failures of Soft Piston Seals
Type
Usual symptom
Cause
Channeling (fluid cutting)
Small, straight grooves across the Sealing surface
Fluid leaking across seal at high velocity
Abrasive wear
Flat on “0” ring. Circumferential groove on lip seal. Sharp sealing edge on lip seal
Pressure too high or abrasive mating surface Pressure too high or too much clearance
Extrusion
Surface broken. Slivers of rubber
Chemical attack
Softening or hardening — may break up
Incompatible fluid
Temperature effects
Hardening and breaking up Breaking up
Too hot and/or excess friction Too cold
Source: Reprinted from Lawrence, R.T., Tribology Handbook, 2nd ed., M.J. Neale (ed.), 1975.
In specifying the requirement of a fluid filter for a particular application, the following points must be taken into account: 1. Maximum acceptable particle size downstream of the filter. This will be influenced by the thickness of fluid films within the components. 2. Allowable pressure drop across the filter. Excessive losses can cause fluid starvation at certain bearing locations within the system. 3. Range of flow rates. 4. Range of operating temperatures. 5. Viscosity range of fluid to be filtered, that is, with respect to temperature. 6. Maximum working pressure. 7. Compatibility of the fluid and filter element material. This is especially important where there is a change of fluid, say from a mineral oil to one of the synthetic or fire resistant types. The nominal filter rating was one of the earliest rating methods formalized in late 1958. It is somewhat arbitrary and is not in favor due to its variability among manufacturers. It is a measure of the retention of a specified contaminant (usually glass beads) by weight, expressed as a percentage. Mean filter rating is a measure of the mean pore size of the filter measured using a bubble point test. It is a measure of the particle size above which the filter starts to be effective. The beta ratio, βx , is determined from a multipass test, which establishes the ratio of number of particles larger than a given size upstream of a filter to that downstream. Obviously its value must be greater than 1. This is the rating system used today. βx = (Nx )u /(Nx )d where, Nx is concentration of particles measured as number/ml greater than a size, x µm, u, d are subscripts to indicate position upstream or downstream respectively. Filter efficiency, Ex = βx − 1/βx . For example, a beta 5 with a value of 100 would mean that the filter is capable of removing 99% of particles greater than 5 µm in size. The lubricating fluid will flow under pressure around a closed circuit from, and returning to, a reservoir. For systems where the reservoir is at atmospheric pressure the air filter effectiveness should be comparable with that of the fluid filter. Positioning of the fluid filter is important and whether it be placed on pump suction line, pressure line, or return line depends on application. It is common to use a coarse filter (strainer) at pump suction and a finer filter on pressure or return line. Due to the large amounts of solid contaminants, common in mining applications, most filters are fitted with a by-pass valve, which ensures that the filter cannot cutoff the lubricant supply completely when
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choked. Under these conditions contaminants can accumulate rapidly in the system. An external by-pass filtering system in the form of a mobile cart has become popular in recent years. Throwaway elements have been favored for mining machinery where fine filtration is required since the more expensive elements, such as sintered metal, are seldom recycled when clogged. Simpler methods for cleaning the latter type of element may see an increase in their usage.
20.3 Bearings A bearing is a support or guide by means of which one machine component is located with respect to others in such a way that prescribed relative motion can occur while the forces associated with the operation are transmitted smoothly and efficiently. For the majority of bearings, transmitting the forces smoothly and efficiently requires the use of a lubricant. Bearings can be classified in several ways, for example, according to: • The mode of operation (rubbing, hydrodynamic, hydrostatic, or rolling element) • The direction and nature of the applied load (thrust, journal, etc.) • Geometric form (plain, taper land, tilting pad, roller, ball) All types of bearings can be found in mining machinery.
20.3.1 Rolling Contact Bearings The fatigue life of rolling bearings is a statistical function of load and speed. Internationally agreed data from which the life for any particular operating conditions can be calculated are available from the manufacturers’ catalogues [8,9]. This gives standard L10 life, which is defined as the number of hours at which 90% of the bearings will have survived. An analysis of long-term field experiences in mining has demonstrated that the majority of bearing failures result from factors other than the bearing reaching its designed fatigue endurance limit. The lack of proper lubrication is commonly cited as the reason for failure, which is perhaps unfair as in most cases a detailed root cause analysis is not carried out. The ingress of solid contaminants will curtail the life of rolling contact bearings [10]. There are some excellent publications, which address the lubrication of rolling bearings in general [11,12]. For the mathematically inclined, actual lubricant film thicknesses between the rolling elements and the raceways can be calculated according to the elastohydrodynamic lubrication (EHD) theory [13]. Rolling contact bearings may be lubricated by oil, usually when they are part of an assembly such as an enclosed gearbox, or by grease, mostly when they are self-contained. When selecting the lubricant, the operating conditions, such as load, speed, and temperature are of prime importance. The environmental conditions must also be taken into account and, with respect to mining, this may be the predominant factor. Mineral oil is the most commonly used liquid lubricant for rolling contact bearings. Where temperatures are above about 150◦ C then a synthetic, for example, polyalphaolefin, lubricant is normally recommended. Although a lithium based NLGI No. 2 consistency grease is satisfactory for the majority of applications in mining machinery, softer grades may be required for automatic lubrication systems where pumpability is as important as lubricating ability. For large, low speed, heavily loaded bearings that are common in mining machinery, greases with the more viscous base oils, given the same NLGI consistency, have proven to give better performance [14]. Adequate lubricant supply in both quantity and frequency is necessary for satisfactory operation. This can be accomplished by taking cognizance of the lubrication system, the bearing design and location, and the service life of the lubricant. Methods of supplying oil to rolling bearings may be divided into two categories, total loss systems and circulating or bath systems. The latter of the two is predominant in modern mining machinery. Circulating
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systems are particularly suitable for situations where the oil has a major function as a coolant in addition to lubrication. If necessary the oil can be passed through a filter and heat exchanger before returning to the bearing. This also facilitates sampling for condition monitoring purposes. An average maximum oil level for satisfactory operation with a horizontal shaft is the center line of the ball or roller at the lowest position (this also applies to bearings on vertical shafts but in a multirow bearing the bottom row of balls or rollers). Too high a level will lead to unnecessary heat generation. In circumstances where the bearings are incorporated in a gear-case, in which the gears are splash fed, special galleries and pockets are fabricated or machined into the casing to distribute the oil to the bearings [12]. In such cases, care has to be taken with the orientation of the gearbox on site as mounting it on an incline can cause lubricant starvation due to lack of gravity feed. Initial grease charge should be carried out as follows: 1. Pack bearings to capacity with grease to ensure that all functional surfaces are supplied with grease. 2. Fill the housing space on both sides of the bearing with grease to such an extent that it can still accommodate the grease expelled from the bearing. Excessive circulation of the grease is prevented. 3. Fill high-speed bearings, electric motor only partially (30 to 40% of the free space) in order to facilitate and accelerate the grease distribution during bearing start-up. 4. Pack low-speed bearings (n · dm < 50,000 min−1 mm) and the housing cavities to capacity with grease. The lubricant friction due to working is negligible. Grease replenishment or exchange is required if the service life of the grease is much shorter than the bearing life (which is usually the case). Grease may be replenished either by grease gun or by a centralized automatic lubrication system. Figure 20.1 shows the lubrication interval tf as a function of the speed index of the bearing (kf · n · dm ) where kf is a factor determined by bearing type, n is rpm, and dm is the mean diameter of the bearing [11]. They apply to a failure probability of 10 to 20%. A kf value range for some bearing types is given in Figure 20.1. The higher values apply to the heavier series (higher load-carrying capacity) and the smaller values to the lighter series of a bearing type. The lubrication intervals of Figure 20.1 apply to lubrication with a lithium soap base grease and temperatures up to 70◦ C, measured at the bearing outer ring, normal environmental conditions, and a mean bearing load corresponding to P/C < 0.1 where P is actual load and C is dynamic load capacity of the bearing. For higher bearing loads and temperatures lubrication intervals are shorter. Every rise in temperature by 15◦ C over 70◦ C halves the lubrication interval of lithium soap greases with mineral base oil. Also vibrations acting on the bearing reduce the lubrication intervals because they result in a separation of the grease into thickener and base oil. Contaminants (including water) penetrating through the seals also affect the lubrication intervals. With gap-type seals, an air current passing through the bearing considerably reduces the lubrication interval. The air current deteriorates the lubricant, carries oil or grease from the bearing, and conveys contaminants into it. For poor operating and environmental conditions, a reduced lubrication interval tfq is obtained from the equation tfq = f1 · f2 · f3 · f4 · f5 · f6 · tf The reduction factors f1 to f6 take into account the effect of contamination, shock loads, vibrations, increased temperatures, higher bearing loads, and air current. Table 20.3 shows the corresponding reduction factors. It can be seen that tfq would be very much reduced for bearings operating in the mining environment as f1 , f2 , and f4 would be low. An overall reduction factor, q, which takes into account all poor operating and environmental conditions can be applied to certain bearing applications. The reduced lubrication interval is obtained from, tfq = q · tf
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10,000 Lubrication 5,000 interval 3,000 tf (h) 2,000 1,000 500 300 200 20
30
50 70 100 150 200 300 500 70010001500 2000 kf · n · dm [103 min–1·mm]
Bearing type
kf
Bearing type
Deep groove ball bearing
0.9...1.1 1.5 1.6 2 0.75 0.9 1.6 1.3...1.6 5...6 1.4
Cylindrical roller bearing Barrel roller bearing
Angular contact ball bearing Spindle bearing Four-point bearing self-aligning ball bearing Thrust ball bearing Angular contact thrust ball bearing
Single row Double row Single row Double row a = 15° a = 25°
Double row
kf
3...3.5a Single row 2 Double row Full complement 25 Cylindrical roller thrust bearing 90 Needle roller bearing 3.5 Tapered roller bearing 4 10 Barrel roller bearing Spherical roller bearing without lips (E design) 7...9 9...12 Spherical roller bearing with centre lip a For bearing which are loaded radially and constantly axially; for varying axial loads kf = 2.
FIGURE 20.1 Lubrication intervals under favorable environmental conditions (Anon. The Lubrication of Rolling Bearings, Publ. No. WL 81 115/2 EF/98/10/87, FAG Australia Pvt. Ltd.)
A typical value of q for vibratory rolls and screens, slew bearings of excavators, belt conveyor pulleys, bucket wheel excavators, and so on, is 0.1. The appropriate amounts of grease for replenishment are given in Table 20.4.
20.3.2 Plain Bearings Plain bearings are those in which the load is transmitted between the moving parts by sliding contact, without the use of balls or rollers. They may operate in hydrodynamic, hydrostatic, or boundary lubrication modes. For most applications the mode will change with different stages of the duty cycle. Where hydrodynamic or hydrostatic is the predominant mode, the viscosity of the lubricant is a major design factor and the lubricant used in service must comply. Large plain journal and thrust bearings are used for the propel mechanism of draglines, the main bearings of cone crushers, and for some pumps and compressors. They are also favored over rolling contact bearings where the presence of water can reduce the fatigue life of the latter by an unacceptable amount, for example, in hydraulic pumps and motors operating on a water-based fire resistant fluid. As plain bearings operate on a “full fluid film” an adequate supply of lubricant is essential. Under starved lubrication conditions, seizure takes place in a time determined by the type of lubricant and bearing material. It has been shown that, for a copious supply of lubricant, grease and oil perform in much the same way. A common cause of dragline bearing failures is that of maintainers altering injector settings from manufacturer recommendations in the belief that too much grease is being supplied, as judged by the
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Handbook of Lubrication and Tribology TABLE 20.3 Conditions
Reduction Factors f1 to f6 for Poor Operating and Environmental
Effect of dust and moisture on the bearing contact surfaces Moderate Strong Very strong
f1 = 0.7 · · · 0.9 f1 = 0.4 · · · 0.7 f1 = 0.1 · · · 0.4
Effect of shock loads and vibrations Moderate Strong Very strong
f2 = 0.7 · · · 0.9 f2 = 0.4 · · · 0.7 f2 = 0.1 · · · 0.4
Effect of high bearing temperature Moderate (up to 75◦ C) Strong (75 to 85◦ C) Very strong (85 to 120◦ C)
f3 = 0.7 · · · 0.9 f3 = 0.4 · · · 0.7 f3 = 0.1 · · · 0.4
Effect of high loads P/C = 0.1 to 0.15 f4 = 0.7 · · · 1.0 P/C = 0.15 to 0.25 f4 = 0.4 · · · 0.7 P/C = 0.25 to 0.35 f4 = 0.1 · · · 0.4 Effect of air current passing through the bearing Slight current f5 = 0.5 · · · 0.7 Strong current f5 = 0.1 · · · 0.5 Centrifugal effects on vertical shafts depending on the sealing
f6 = 0.5 · · · 0.7
Source: Reprinted from Anon, The Lubrication of Rolling Bearings, Publ. No. WL 81 115/2 EF/98/10/87, FAG Australia Pvt. Ltd.
TABLE 20.4 Amounts of Relubrication Grease Relubrication quantity m1 for weekly to year relubrication m1 = D · B · x (g) Relubrication Weekly Monthly Yearly
x 0.002 0.003 0.004
Quantity m2 for extremely short relubrication interval m2 = (0.5 to 20) · V (kg/h) Relubrication quantity m3 prior to restarting after several years of standstill m3 = D · B · 0.01 (g) V = free space in the bearing G = π/4 · B · (D 2 − d 2 ) · 10−9 − (m3 ) 7800 d = bearing bore diameter (mm) D = bearing outside diameter (mm) B = bearing width (mm) G = bearing weight (kg) Source: Reprinted from Anon, The Lubrication of Rolling Bearings, Publ. No. WL 81 115/2 EF/98/10/87, FAG Australia Pvt. Ltd.
amount ejected at the bearing. Investigation into the failure of main propel shaft bearings of a dragline highlighted the fact that there is no standard method to determine the optimum grease supply rate for this type of bearing. The lubricant supply grooving is an influential design factor for hydrodynamic bearings. Cone crusher plain bearing failures are often diagnosed as lack of lubricant. In a misguided attempt to correct the situation, maintainers have been known to increase the depth and width of the oil grooves. Instead of this increasing the flow to the contact area it has the opposite effect of providing an easier path for the oil to escape. Guidance on flow rates and grooving can be found in the literature [15,16].
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Hydrostatic bearings are those for which the lubricant is supplied to a recess by a positive displacement pump and allowed to build up to a high pressure, which will completely separate the surfaces. They are used, for example, in crusher drum supports for large grinding mills. This type of bearing provides for very low friction forces and is therefore energy efficient. Control of viscosity, pressure, and flow rate is essential for stable, satisfactory operation.
20.3.3 Bushes and Pins Plain bearings that operate predominantly in the boundary lubrication mode are numerous in mining machinery as pivots, pins, and bushes, many of which have oscillating motion. The lubricant is usually applied on a total-loss basis whether fed individually by grease gun or as part of a centralized system. Despite the fact that these are simple components their failure can have a devastating effect on production and safety. This was evidenced at Markham colliery in 1973 where 18 men were killed and 11 seriously injured when the cage in which they were traveling plummeted to the bottom of the shaft. This was the result of a pin/bush seizure on the winder brake [17]. A common weakness with pins and bushes is the lack of provision for the lubricant to access the contact area. Many premature failures have occurred as a result of the lubricant feeder holes not being aligned with the distribution grooves or by the drilled hole in the housing being omitted during manufacture.
20.4 Gears The gears used in mining machinery include both open and enclosed types and are usually of the external spur, helical, or worm configuration. Open gears are those whereby the casings are not sealed and lubrication is by an adhesive lubricant on a total loss basis. Mining machinery gears do not differ substantially from other industrial heavy drives but in mining the percentage of very large gears is high with the consequential problems of manufacturing, measuring, testing, and aligning. When these problems are coupled with duty cycles that vary in speed, load, and direction (as in draglines) serious problems can be encountered.
20.4.1 Gear Lubrication Gears, like any other machine elements that require lubrication, depend upon the properties of the lubricant, the means of application, and the quantity and frequency of supply. It must be remembered that, for successful gear lubrication, the correct lubricant, in the proper condition, has to be applied in the right quantity to the right location at the right time. This emphasizes the importance of the means of application as well as the choice and maintenance of the lubricant. At times, a compromise has to be made. A combination of sliding and rolling action takes place in varying proportions with all meshing gear teeth regardless of type. With conventional spur gears the theoretical lines of contact run straight across the tooth faces. The direction of sliding is then perpendicular to the lines of contact. Because of the twisted shape of the teeth in helical and herringbone gears, the theoretical lines of contact slant across the tooth faces thus introducing some side sliding along the lines of contact. In the selection of lubricants for gears, tooth sliding is important for two reasons: • It increases the operating temperature because of the frictional effects. • Sliding along the line of contact tends to wipe the lubricant away from the convergent zone and thus makes it more difficult to form lubricating films. Longevity of gear teeth as influenced by a lubricant may rely on physical (viscosity) protection or chemical (additive) protection depending on the mode of lubrication prevailing. The lubricant may also be utilized as a heat transfer medium and this can be important in establishing adequate viscosity at the gear tooth contact.
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Empirical value of viscosity grade can be found in several references, for example, AGMA Standard 250.02 [19]. Similar recommendations are given in other standards and by oil companies but few go into the detail of the actual lubricant specifications. The effect of pressure, temperature, and shear rate on viscosity as well as the limitations of boundary lubricants has to be considered. Since, for most gear applications in mining, both fluid film and boundary lubrication modes are prevalent a lubricant offering viscosity and boundary protection is used. It is clearly desirable to achieve full fluid film lubrication as rapidly as possible, although in certain circumstances it may not be achieved at all. For gears, the criteria establishing full fluid film lubrication are: • • • • • •
elastic contact of the tooth surfaces modest sliding speed adequate rolling speed adequate viscosity of the lubricant at the conditions of contact an adequate supply of lubricant a load that is not too great
Viscosity and its changes with pressure, temperature, and shear rate is treated in detail in another section of this handbook [19].
20.4.2 Determination of Viscosity for Spur and Helical Gears It is often stated that viscosity is the most important property of a lubricant. This may be true; but only for the conditions appertaining to the contacting surfaces. Since viscosity changes with temperature, pressure, and in some cases with shear rate, it can be appreciated that determination of the effective viscosity is difficult. To achieve viscosity protection we must ensure an adequate fluid film at the contact between the gear teeth. The problem resolves to that of estimating the viscosity at the contact by elastohydrodynamic lubrication (EHL) theory which, in turn, allows a suitable viscosity grade to be selected. The problem is compounded in gear trains having several reductions and where the bearings are lubricated with the same oil. In such cases a compromise has to be made. The question of what viscosity is considered adequate is normally related to the surface roughness of the contacting teeth. This prompts a further question as to whether the roughness should be taken to be that as manufactured or after run-in. Also, surface roughness may be measured in the direction of sliding or at right angles to the direction of sliding, thus applying a two-dimensional measurement to a three-dimensional geometry. Normal practice is to use the “as manufactured” roughness in the direction of sliding, unless a suitable run-in procedure can be assured.
20.4.3 Specific Film Thickness The specific film thickness describes the thickness of the lubricant film in relation to the roughness of the lubricated surfaces. That is, specific film thickness, λ = h/σ , where h is the minimum film thickness and σ is the composite roughness of the two surfaces given by σ = (σ12 + σ22 )0.5 where σ1 and σ2 are arithmetical average roughness values, Ra . (Note root mean square roughness, Rσ ∼ = 1.3Ra .)
20.4.4 Calculation of Elastohydrodynamic Lubrication Film Thickness Several solution schemes for EHL film thickness are available in the literature. The isothermal Grubin equation is used in a form developed by Cheng for the general case of an elliptical Hertzian contact [13]. The Cheng equation is simplified into two forms for elliptical contact and line contact. Although elliptical contact occurs in ball bearings, helical and spiral bevel gears, and in some cam-follower systems, in all cases
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the long axis of the ellipse is perpendicular to the direction of motion and line contact can be assumed with very little loss in accuracy. Comparison of the specific film thickness with Wellauer and Holloway’s probability of distress graph will give an indication of the protection afforded to the gears [20].
20.4.5 Load Carrying Tests The wear and scoring protection given by a lubricant is often referred to as its load-carrying capacity. However, this is an elusive term because it is not a fixed quantity even for a particular lubricant. Lubricants offer protection in different ways depending on the operating conditions and environment. Mineral oils without antiwear or EP additives derive their load-carrying capacity from their viscosity and the maximum load at which they may be safely used can be determined. However, the load-carrying capacity of EP oils and those compounded with a friction modifier is enhanced by the reactions of their additives that are determined, at least in part, by the conditions under which they are used. Therefore, the indicated level of load-carrying capacity of EP and compounded oils is usually determined experimentally in one or more of several tests, some of which are described below. Because the reactions of additives vary with the test conditions it is not generally possible to use the results of one test to predict the results of another.
20.4.6 Full Scale Tests The only reliable evaluation of a gear lubricant for a given application is its performance in service. Manufacturers sometimes set up full scale rigs and operate them under the most severe conditions that are likely to be encountered in order to determine the lubricant suitability for the application. In the case of large mining equipment the expense of setting up such tests is usually prohibitive and other means of evaluating lubricants are necessary. Nevertheless, full scale tests were carried out on replacement propel gearboxes for an existing dragline using a “back-to-back” or four square testing method [21]. This method of testing requires the units to be operated with a locked-in or recirculating torque, which is applied statically. Input power to rotate the system then only has to overcome frictional losses. The configuration of the two propel boxes lent themselves readily to this type of test since they contained identical gearing but were opposite handed. The low-speed shafts were coupled directly with a splined connector and high-speed shafts were coupled with a torque wind-up coupling. It was confirmed that the operating temperature was sensitive to the viscosity of the oil. A short no load test of only 2 h duration showed the bulk oil temperature to be 11◦ C higher for a heavy synthetic oil (3000 cSt at 38◦ C) compared to a test conducted with a lubricant having a viscosity of 100 cSt at 38◦ C. Bearing temperatures likewise ran about 15◦ C higher. The operating temperature was sensitive to the amount of submersion of the gears in the viscous oil. A 25 mm higher oil level accounted for an increase of 5.5◦ C in the bulk oil temperature. It was concluded that the four square test was an economical and effective method to run these boxes under full load. Attempting to apply a torque load at the output shaft of 895 kN m by any other method would have been impractical.
20.4.7 Gear Test Rigs Gear test rigs in which the gear design and operating variables can be adapted to simulate conditions encountered in service offer an economical method of evaluating lubricants. They offer better control of the conditions than a full scale or actual service test. The FZG gear rig, which has been widely used in Germany for many years for evaluating gear oils, is now the basis of many standard test methods including ASTM D 4998 and D 5182 [22].
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Handbook of Lubrication and Tribology
20.4.8 Bench Tests Although gear rig tests are less expensive and quicker than full scale tests, they are still relatively costly when one considers that in a lubricant development program, many candidate formulations may have to be tested in order to get the best balance of properties that are required for the intended purpose. Consequently, several small scale test methods have been devised that are useful for evaluating oils. The more familiar of these are the Timken and Four-Ball EP Tests, which are included in many mining machinery manufacturers’ specifications. These tests were originally developed to distinguish only between EP and non-EP oils and so their precision is not always sufficient to establish levels of EP activity. Furthermore, each test appears to respond best to a particular type of additive, and consequently, the results seldom indicate the best lubricant with any degree of certainty. In addition, erratic results are likely if details in the testing technique are not carefully observed. Either of the EP tests may duplicate one or two service conditions and thus may be useful as a screening test. Since either of these tests may have some degree of significance in certain applications, users must exercise their best judgment as to which are the most meaningful and reliable for their purposes. It is obviously unwise to attempt to predict performance solely on the basis of EP bench tests.
20.4.9 Reason for Pitting Fatigue Due to their slow speed, heavy and shock loading conditions, mining machinery gears generally experience pitting as the mode of tooth surface failure. Exceeding the critical contact (Hertzian) pressure allowable for a specific material is the main reason for pitting fatigue. Other theories have been proposed but it is generally accepted that pulsating and alternating stresses will lead to fatigue of material followed by pitting. The following factors will influence the development of pits on gear tooth surfaces (i) material, (ii) surface treatment, (iii) surface finishing, (iv) tooth proportions and tooth profile, (v) operating conditions, and (vi) lubricants.
20.4.10 Surface Roughness Intuitively one would expect the microtopography of the tooth flank surfaces to have some influence on gear durability. This has been proven to be the case [23]. However, the exact nature of the surface interactions is yet to be determined. Here we will consider the specification, production, and measurement of surface topography and show why it is important to the reliability of lubricated gears. The most common method of measuring surface roughness is by profilometer, whereby a lightly loaded sharp diamond stylus is traced across the surface and its movement converted to an electrical analog of the surface profile. The signal so obtained is analyzed to describe the surface profile by statistical values. In general, surfaces will contain irregularities with a large range of spacing but profilometers are designed to respond only to irregularity spacing less than a given value. This is called the “cut-off.” The effect of reducing the cut-off is to suppress more of the longer wavelengths so the signal being processed is substantially different in each case highlighting the importance of specifying the cut-off when taking a measurement [24]. The microgeometry of most surfaces of engineering interest is very complex and it is not possible to uniquely describe a surface by a single number. Nevertheless, a single value such as arithmetical average height (Ra ) is often all that is given to specify surface topography. In some cases (including gears) no values are given and the surface microgeometry is inherent in the specified finishing process, for example, hobbed, ground, planed. It is now generally agreed, however, that at least two parameters are necessary to give a good indication of the properties of a surface. One of these parameters relates to asperity heights and the other to asperity spacing.
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Mining Industry
20-13 TABLE 20.5 Measured Roughness in Radial Direction (µm) Ra
Rt
Rp
Rtm
Rpm
Average cut-off 2.5 mm 5.5 30.0
13.47
23.57
Average cut-off 0.8 mm 4.36 16.46
7.9
13.467
6.13
Average cut-off 0.25 mm 2.04 11.6
5.1
8.35
4.15
10.7
Although the terms texture, roughness, and finish, in relation to surface topography, are sometimes used synonymously, there is a distinct difference in their meaning. These terms are used for the quantitative assessment of surface “finish,” which may be described by the finishing process. Measurements of the tooth surfaces of “as machined” gears were taken using different parameters and cut-offs. As an example, the roughness profile of an AGMA Class 10 pinion tooth for a dragline hoist gearbox, taken in the radial direction, is illustrated by the numerical results for three different cut-offs given in Table 20.5. The results of these measurements, which were carried out by profilometry, indicate that the chosen instrument cut-off value is important and should be selected with reference to the Hertzian contact bandwidth. Values for as-machined and run-in conditions are generally quoted as being much lower than those given in Table 20.5. Elastohydrodynamic (EHD) lubrication theory is being used by some designers to predict the performance of gears with regard to tooth surface durability. Since this theory involves the tooth surface texture then this property should be specified before manufacturing and measured thereafter. In many cases this does not occur and the resultant surface texture is inherent in the finishing process [24]. Application of this theory to dragline gears gave the values shown in Table 20.6. It can be seen that surface stress is extremely high and specific film thickness is very low, which implies a dependence on the boundary lubrication properties of the lubricant.
20.4.11 Influence of Lubricants on Pitting Fatigue Since the majority of failures of mining machinery gears appear to be due to pitting fatigue, it is appropriate to consider the influence of lubricants on this failure mode. Controlling pitting fatigue due to cracks initiated beneath the surface seems to be possible only by reducing the contact (Hertzian) pressure thus reducing the pulsating and alternating stresses. At given stress conditions this can only be achieved by increasing the elastohydrodynamic (EHL) component of force transmitted. One possibility for retarding the formation of cracks initiated at the surface is to reduce the sliding friction coefficient at the point of contact, leading to lower tangential stresses at the surface. Ways of achieving this are given in Table 20.7. After cracks have been initiated, breaking out of particles may be accelerated by splitting action of the oil. During the rolling and sliding action of the two meshing gear tooth surfaces, lubricant will be forced into cracks and pressurized. When the pressure is released suddenly behind the point of contact, some particles may be broken out by the oil trapped within the cracks. During the operation of a driving pinion and driven gear wheel the cracks will be “opened” by tangential forces near the roots of the teeth, in the course of which the lubricant can be forced into the cracks. Conversely, near the tip of the teeth these cracks will be compressed and “closed” leading to much less pitting in this area of gear tooth surface. Tests were carried out on an FZG rig by Bartz and Kruger [26] for thin lubricant film conditions comparable with that given in Table 20.6. They highlighted the importance of additive type and quantity, method of application, and operating conditions on the fatigue life of the gears.
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20-14
Handbook of Lubrication and Tribology TABLE 20.6
Dragline Hoist Gears at Half Full Speed
Tooth loading W/l N/m (lbf/in.) Hertz bandwith at pitch line (mm) Max Hertzian stress (N/m2 ) Max sliding speed (m/sec) Slide/roll ratio Composite roughness estimate (σ ) Lubricant viscosity (cSt) Pressure-viscosity coefficient (m2 /N) Estimated tooth temperature (◦ C) hmin , pitch line (µm) Specific film thickness
Input motor
Final
5.57E+05 (2414) 1.16 7.20E+08 (SAP) 1.12 (EAP) 0.455 (EAP) 4.84 60 1.8E−08 100 1.056 0.218
2.60E+06 (14836) 3.66 1.08E+09 (SAP) 0.805 (EAP) 0.904 (EAP) 4.84 70 1.2E−08 90 0.528 0.109
SAP – Start of active profile, EAP – End of active profile. Source: Reprinted from Scott, W., Report on the Lubrication of Dragline Gears for BMA, August 2004.
TABLE 20.7
Influence of Lubricants on Pitting Fatigue
1.
Reducing the development of internal cracks Reducing the Hertzian pressure by increasing the elastohydrodynamic transmitted load by, Higher rated viscosity Higher pressure–viscosity coefficient Lower temperature dependence of viscosity
2.
Reducing the development of cracks at the surface Reducing solid body friction coefficient by, Higher operational viscosity Suitable type of lubricant Suitable additives Reducing splitting effects of cracks by, Higher operational viscosity Suitable additives Source: Reprinted from Bartz, W.J. and Kruger, V., Test Method for Evaluating the Influence of Lubricants and Lubricant Additives on Fatigue Failure of Quenched and Tempered Case-Hardened Spur Gears, Rolling Contact Fatigue-Performance Testing of Lubricants, R. Tourret and E.P. Wright (eds), Institute of Petroleum, London, 1977.
Compounding the problem of high stresses and low lubricant film thickness predicted for the final drive gears of Table 20.6 is the double helical (herringbone) configuration of the gears. This type of gear was developed to overcome the effects of end thrust. The thrust from one helix supposedly balances that from the other helix. However, even a single pair of helical gears is difficult to match, so double helicals are virtually impossible to match. Usually one gear (the pinion) is allowed to float and find its own center but, since this gear is moving axially, the loads on the helices cannot be equal. Also, with heavily loaded double helical gears friction between the tooth faces heavily damps the free axial floating action that is essential if the load is to be continuously shared between the helices.
20.5 Wire Ropes Wire rope is one of the most useful and ingenious machine elements used today. They are the sinews of the mining industry. Although a wire rope may appear to be simple device, it is, in fact, a complicated tool with many moving parts. A typical 6 × 49 rope, for example, encompasses 343 wires, all of which move together and independently as the rope bends.
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Mining Industry
20-15
20.5.1 Lubrication of Wire Ropes Many claims have been made (especially by lubricant suppliers) that efficient lubrication in service can increase the life of a rope by as much as 300%. Rope manufacturers will state that too much lubricant is better than too little. On the other hand, some operators insist that, under certain conditions, the application of a lubricant to the external wires of rope merely acts to pick up solid particles that hasten failure by abrasion. It is also possible that since abrasion is a form of cutting, the lubricant may act as an efficient cutting fluid in which case more abrasive wear would be produced in the presence of the lubricant. Many operators of friction winders do not apply any external lubricant to the ropes as they may reduce the coefficient of friction. There appears to be no clear understanding between rope manufacturer, machine supplier, lubricant supplier, and machine operator as to the responsibility for the quantitative specification of rope lubricants. Few manufacturers give firm recommendations on lubricants and even fewer give any information on the lubricant used in rope manufacture. If, as is often recommended, the lubricant applied externally should penetrate the rope, then the operator could reasonably expect some information on the compatibility between the applied lubricant and the lubricant used during manufacture. Where solvents are used in the external lubricant, or if the external lubricant is of extremely low viscosity, excessive bleeding and early exudation of the internal lubricant may occur. The external lubricant must also be compatible with the rope core materials, for example, synthetic cores, and with elastomeric sheave liners where these are used. Lubricant recommendations tend, in most cases, to be on a qualitative “wish list” of desirable properties. Frequency and method of application are determined by experience. The results of an extensive investigation into the lubrication of wire ropes for mine hoists confirm what many believe to be true, that is, there is no coherent quantitative approach to the subject [27].
20.5.2 Lubrication During Manufacture Formulations for internal lubricants are based on petroleum jelly or asphaltic residue and they usually contain corrosion or rust inhibitors and antiwear or extreme pressure additives. High drop point, high melting point, and adhesion are desirable properties. Tests to discover the effect on rope fatigue of (i) full lubrication, (ii) core lubrication, and (iii) no lubrication (except for cordage or batching oil) showed, 1. “Full” lubrication of the strands during stranding and closing and the core during closing resulted in an axial fatigue life of more than 2.5 times that of lubricating the core only. 2. Core lubrication only (of the core during closing) resulted in an axial fatigue life three times that provided by the cordage oil alone. 3. Ropes containing cores saturated with lubricant (about 25%) lose about 40% of this lubricant due to compression when the rope is in pure tension. These tests were run with no addition of external lubricant, which could have influenced the amount of core lubricant [28]. It may be expected therefore, that ropes with a wire rope core would also benefit from “full” lubrication. It has been suggested that the main problem with internal rope lubrication is not attributed to the lubricants available but on the means of application. During stranding, the lubricant is thrown off instead of reaching the interior. A patented stranding head that allows the rope voids to be completely filled with lubricant was developed to overcome this problem [29]. There appears to be no distinction made by rope manufacturers between the type of lubricant used for different rope specifications with the possible exception of locked-coil ropes.
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Handbook of Lubrication and Tribology
20.5.3 Lubrication During Service The antiquated view that external wire rope lubricants (or dressings) are applied to act as a seal to prevent ingress of contaminants, such as moisture and dirt, and egress of the internal lubricant, has now been largely superceded by a more enlightened approach. The general consensus now is that the external lubricant should penetrate the rope wires to lubricate them and to replenish the internal lubricant to prevent drying of the core and internal corrosion.
20.5.4 Lubricant Specifications The survey and tests carried out by the Battelle Columbus Laboratories [27] in 1978 have shown that low viscosity lubricants can be just as effective in preventing corrosion as extremely viscous materials. This definitive work on wire hoist rope lubrication gives the ideal lubricant, for the majority of such applications, as that described quantitatively in Table 20.8. The nonstandard method of evaluating corrosion inhibition was chosen to simulate the washing effects of rain (or mine water) on thin lubricant films on steel substrates. Steel-wool was compressed and formed by gloved hands into balls of about 25 mm diameter. The balls were washed successively in xylene, toluene, and twice in methanol. Then they were flung out and dried in air for 1 h. The balls were carefully submerged in the lubricant to be evaluated and the excess lubricant was expressed by hand-squeezing each ball using rubber gloves. The lubricated balls were placed on galvanized hardware — cloth trays on the platforms in dessicators to which 25 mm of distilled water had been added to provide a humid environment. In order to keep condensed water from dripping nonuniformly on the specimens from the tops of the dessicators, a piece of stainless steel hardware cloth was placed between the dessicators and their lids so as to vent moisture that would otherwise condense on the lids. Each day for five days per week each specimen was sprayed directly with distilled water; three bursts from a small atomizer provided direct contact with discrete water somewhat simulative of the action of rain or that of occasional contact
TABLE 20.8
Prototype Specification for External Lubricants for Mine-Hoist Ropes in Servicea
Origin of specification tests and requirements
Laboratory test Viscosity, ASTM D 445 and D 446
Corrosion prevention, ASTM D 665 Standard methods Requirements determined and suggested by Battelle
Nonstandard methods Methods and requirements developed, determined and suggested by Battelle
Requirements Not greater than 150 cSt at 37.8◦ C; preferably not greater than 37 cSt at 57.8◦ C; not less than 3 cSt at 99◦ C Pass
Wear prevention, shell 4-ball Testb EP activity, shell 4-ball EP weld point Consistency at low temperature ASTM D 217 Flash point, ASTM D 92
Not greater than 0.80 mm; preferably, less then 0.50 mm Not less than 200 kg
Corrosion prevention Adhesion Water washout Rope-core penetration
No corrosion evident in less than 90 days Not less than 55% Not greater than 11% Not greater than 167 min
No significant change from room temperature to −22◦ C Not less than 163◦ C
a This specification does not apply to head rope dressings for friction (Koepe) hoists. b 20 kg, 1800 rpm, 1 h, 54◦ C.
Source: Anon, Hoist rope lubrication criteria, Battelle Columbus Laboratories, Report No. PB80-182959, Bureau of Mines, Washington, D.C., 1978.
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Mining Industry
20-17
of a lubricated rope with mine water. In each dessicator, an unlubricated control specimen was included for comparison. The specimens were observed each day and the incidence and character of corrosion was recorded. The nonstandard test method for assessing adhesion involved the use of about 25 mm diameter steelwool balls. After thorough cleaning and drying, each ball was weighed and immersed in the lubricant to be tested. Following thorough wetting, the balls were allowed to drain for 1 h and then reweighed. The ball was squeezed by hand to remove excess lubricant then weighed again. The net weight of lubricant retained was used as a measure of the adhesive property of the lubricant. The nonstandard water washout test used similarly prepared steel-wool balls. They were submerged in the lubricant being studied, withdrawn and squeezed out by hand, then weighed and stored in racks. Once each week for five weeks, these balls were “washed” with 100 bursts of distilled water delivered through an atomizer. After each washing, the specimens were oven dried for 1 h at 80◦ C to remove excess water in a manner similar to air drying of ropes after rain. After five weeks, the balls were reweighed and the percentage loss of lubricant caused by the washing used as a measure of water wash-out. The rope-core penetration test involved the use of 300 mm lengths of 19 mm diameter 6 × 19 fiber core wire rope. An extractor was used to remove the internal lubricant. These rope specimens, which contained a vegetable fiber in the center were oven dried and placed in a horizontal position. The lubricant to be tested was added at the rate of one drop every 3 to 5 sec to the mid-position of the rope. The time taken for the lubricant to penetrate to the core and to migrate to each end of the specimen was recorded. The lubricants exhibiting the shortest migration time represented the highest penetration. Another test not reported to date, which could be used to assess the adhesion properties of a lubricant would be to use the Timken machine. A prescribed amount of the lubricant could be evenly smeared onto the cup. The cup would be rotated at a surface speed similar to that of wire ropes winding onto or from rope drums for a nominated time. The net loss of lubricant from the cup would be indicative of the adhesive quality of the lubricant.
20.6 Power Hydraulic Systems The nomenclature for power hydraulics varies between different users. Here we will include hydrokinetic applications such as fluid couplings and torque converters as well as the more common hydrostatic systems. Most of the mobile plant used in mining such as continuous miners, shuttle cars, excavators and loadhaul-dump trucks incorporate hydraulic systems. Hydraulics are also common in fixed plant such as roof support systems, shearers, hopper doors, belt conveyor loop take-up, bunkers, and drilling rigs. Solids contamination is a particular enemy of hydrostatic systems and so its control is of paramount importance. Another troublesome area is the numerous incidences of burst flexible hoses. Attention to simple design guides on their selection, fittings, installation, and limitations and adherence to proper operating procedures can reduce loss of production due to hose related failures. The discharge of hydraulic fluid to ground has become an ecological issue with subsequent renewed interest in environmentally friendly fluids.
20.6.1 Hydrokinetic Transmission Fluid couplings and torque converters have a wide application in power transmissions that are subject to pronounced and sudden fluctuations in load. Since there is no mechanical connection between the input and output, power take-up and acceleration are very smooth. Applications include rail and road vehicle traction (coupled with gear-boxes), conveyor, and pump drives, providing a smooth power output from a pulsating power source and torsional damping. Because a fluid coupling will provide a certain amount of slip at all speeds, one form of this coupling includes a centrifugal friction clutch that locks the two moving parts of the coupling together when the
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Handbook of Lubrication and Tribology
output shaft has picked up sufficient speed. Another form of coupling employs an adjustable scoop, which controls the slip by varying the amount of fluid in the coupling; fluid is allowed to leak into an outer rotating casing where it forms a fluid belt. The depth of the belt and, therefore, the quantity of fluid in the coupling, is controlled by the position of the scoop, which can be adjusted externally. The quantity of oil used in these couplings governs the coupling effect and the quality of the oil governs the transmission characteristics. The primary requirement is low viscosity and a high viscosity index, for the oil may well be subject to low-temperature conditions. High temperature must also be taken into account, for, although a hydraulic coupling can run stalled without harm, all the input is dissipated in the form of heat and the oil temperature rises; under certain operating conditions coolers are necessary. The amount of slip is largely determined by the oil viscosity and increases as the oil gets thinner. Higher oil viscosity decreases slip but this may be offset because the more viscous liquid generates heat and the working temperature will rise. Fluid couplings that do not have a scoop control usually incorporate a fusible plug, which will eject the fluid when it reaches a specified temperature. Because of the safety implications of ejecting hot fluid, a fire resistant type (usually phosphate ester) is preferred to mineral oil for mining applications. Where this is the case less fluid is required to transmit the same power due to the higher density of the phosphate ester fluid.
20.6.2 Hydrostatic Transmission Although motion is involved in this type of hydraulic system, the speed is relatively low and it is pressurized fluid that moves the load. As long as the rate of flow is small, it can be called a hydrostatic system. At the heart of the system is a positive displacement pump that creates fluid flow from a reservoir supply and directs it into a sealed system such as a hydraulic plunger. Once the plunger has moved to its extremity the pressure will increase until the fluid is released by failure of a component or by some other means. A simple method of protecting the sealed system from overload is to include a pressure relief valve. When the applied force rises to a value corresponding to the setting of this valve, fluid flows past the relief valve and out of the system. However, the hydraulic balancing force is still maintained. This is exemplified by the hydraulic roof support, which is used extensively in underground coal mines. If the mine roof tends to give way, the support, which is still resisting an enormous force, will yield until the load has readjusted itself equally on all the supports in the vicinity. The accuracy and sensitivity with which the pressure and the applied forces can be limited is a great advantage that can be used for clamping of delicate components or in limiting the applied forces on machines. It must always be borne in mind that a pump produces flow and not pressure. Pressure results from resistance to the flow. If there is no resistance to the flow there will be virtually no pressure in the system.
20.6.3 Hydraulic Fluids The following fluid physical properties are important when used in hydraulic systems: 1. Viscosity — If this is too high it will cause unnecessary viscous friction and if too low it will allow excessive leakage at clearances and thereby reduce volumetric efficiency. It is usually chosen in accordance with the requirements of the hydraulic pump. 2. Viscosity index — The rate at which viscosity reduces with increased temperature, for liquid lubricants, is not the same for all. The change is specified by the lubricant’s viscosity index, which is an arbitrary scale and ranges from about 50 to 140 for petroleum oils with the higher values indicating less change. A VI of about 100 is normally suitable for hydraulic systems but for a large ambient temperature range it may be as high as 140. 3. Vapor pressure — When the molecules escaping from the surface of an enclosed liquid equals the number of molecules bounding back in to the liquid, the vapor is then said to be saturated and the pressure exerted on the container is the saturation vapor pressure. Vapor pressure increases as the temperature increases. The boiling point of the liquid is the temperature at which the vapor
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pressure is equal to the atmospheric pressure. High vapor pressure can cause cavitation problems at pump suction. 4. Bulk modulus — When a fluid is placed under pressure its volume may reduce. The extent to which the fluid changes in volume with pressure is termed its bulk modulus. Entrained gas in a liquid will greatly reduce its bulk modulus, that is, make it more compressible. The higher the bulk modulus of a hydraulic fluid, the more efficient the power transmission. 5. Air solubility and entrainment — A typical hydraulic oil dissolves about 9% by volume of air at 20◦ C and atmospheric pressure. This solubility is unaffected by any of the commonly employed additives. It is, however, affected by viscosity and doubling this reduces the solubility by 5% of its value. Dissolved air has no measurable effect on compressibility, but it is released at low pressures and this markedly affects the performance of the hydraulic system. It is of considerable practical importance as the air becomes entrained as bubbles which can cause instability of servo systems, noisy and erratic action, general loss of power, and damage to pumps. The fluid must also have good chemical properties such as anticorrosion, antioxidation, and antiwear.
20.6.4 Types of Hydraulic Fluid Classifications for the different types of hydraulic fluids are given in ISO 6743/4 and are listed in Table 20.9.
20.6.5 Fire Resistant Hydraulic Fluids The only fireproof hydraulic fluid is water. To determine the degree of fire resistance offered by a fluid, in comparison to mineral oil, many different test methods have been established. Typical of these are: • • • • • • • • •
open-cup flash point and fire point autoignition temperature linear flame propagation test heating rate “Monsanto” molten metal test hot manifold test Factory Mutual hot channel ignition test spray ignition flammability tests evaporation (wick) tests
TABLE 20.9
Hydraulic Fluids Classifications Per ISO 6743/4
Symbol HH HL HM HR HV HG HS HFAE HFAS HFB HFC HFDR HFDS HFDT HFDU
General characteristics Non inhibited solvent refined mineral oils (−10 to 90◦ C) Refined mineral oil with improved antirust and antioxidant properties (−10 to 90◦ C) Oils of HL type with improved anti-wear properties (−20 to 90◦ C) Oils of HL with improved viscosity/temperature properties (−35 to 120◦ C) Oils of HM with improved viscosity/temperature properties (−35 to 120◦ C) Oils of HM with antislip antisticking properties Synthetic fluids with nonfire-resistant properties Fire-resistant fluids oil in water emulsions with a max of 20% by weight of combustible materials Fire-resistant solutions of a chemical in water with a minimum content of 80% by weight of water Fire-resistant emulsions of water in oil Fire-resistant fluids of water-polymer solutions with a minimum content of 35% in weight of water Synthetic fire-resistant fluids phosphate-ester based Synthetic fire-resistant fluids chlorinated hydrocarbon based Synthetic fire-resistant fluids consisting of blends of HFDR and HFDS Synthetic fire-resistant fluids from other types
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20-20 TABLE 20.10
Handbook of Lubrication and Tribology Comparison of Some Fire Resistant Hydraulic Fluids
Property or quality Specific gravity Typical viscosity at 40◦ C (CS) Viscosity index Thermal conductivity (kJ/h/m◦ C) Specific heat (Cp ) (kJ/kg/◦ C) at 25◦ C Lubricity Fire resistance Corrosion protection Stability Price
Petroleum oil
Oil-in-water emulsion
Water-in-oil emulsion
Water glycol
Phospate ester
0.85–0.9 40
1.0 4 at 20◦ C
0.9–0.95 70
1.1 35
1.15 50
95+ 0.4
— 1.6
160+ 1.2
160+ 1.5
20 0.47
1.9
4.0
2.65
3.0
1.6
Good to excellent Poor Very good Excellent 1
Limited Excellent Limited Poor 0.1 (Depends on dilution)
Good Good Good Poor 1.5
Good Good Good Good 3–4
Good to excellent Good Very good Excellent 8
A comparison of the properties of the most common fire resistant fluids with those of mineral oil is given in Table 20.10.
20.7 Manufacturers’ Lubricant Specifications Most manufacturers’ lubricant specifications/guidelines currently found at mine sites are obsolete and many still quote viscosity units in Saybolt Universal Seconds (SUS). Nevertheless typical requirements issued by manufacturers of mining equipment for various categories of lubricant are given in this section.
20.7.1 Multipurpose Extreme Pressure (EP) Grease These are intended to lubricate rolling element bearings, bushings, and plain bearings that require grease. They may be applied by hand pressure guns or centralized systems. Lithium 12-hydroxystearate or lithium complex bases are usually recommended although other formulations may also be acceptable provided care is taken with compatibility. NLGI consistency should be appropriate for the prevailing ambient temperature range. Typical manufacturers’ performance requirements are given in Table 20.11.
20.7.2 Multiservice Mining Lubricant These lubricants are intended for open gears, racks, bushings, rails, rollers, dipper handles, slew bearings, and low to moderate speed plain and rolling element bearings. They are usually partially or fully synthetic semifluid greases or polymer/gel thickened with excellent adhesive and cohesive properties. They are intended to be dispensed intermittently from centralized lubrication systems. Typical manufacturers’ performance requirements are given in Table 20.12.
20.7.3 Enclosed Gear-Case Oil These are intended to lubricate spur, helical, worm, bevel, spiral bevel, and gear assemblies that may include plain and rolling element bearings, bushes, sprockets, chain drives, and other components enclosed in oil-tight houses. They have to be capable of operating under low speed — high torque or high speed — low torque as well as high shock load conditions. Care has to be taken with the compatibility of oils with EP additives or other friction modifiers in applications containing brakes, backstops, clutches, and bronze
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Mining Industry TABLE 20.11
20-21
Performance Requirements for Multipurpose EP Grease
NLGI gradea Worked penetration ASTM D217/ISO 2137 at 60 strokes, 25◦ C/77◦ F Dropping point, ASTM D566/ISO 2176 or ASTM D2265, min. ◦ C/◦ F Base fluid viscosity, ASTM D445/ISO 3104; D2161, min. mm2 /sec at 40◦ C Oxidation stability, ASTM D942, max. pressure drop at 100 h, kPa/psi Roll stability, ASTM D1831 max. points change Water washout, ASTM D1264, max. % loss at 79◦ /C/175◦ F Rust protection, ASTM D1743 rating EMCOR rust protection DIN 51 802/1P 220, min. rating Copper strip corrosion, ASTM D4048 Max. rating, 24 h at 100◦ C/212◦ F Timken EP Test, ASTM D2509 OK value, min. kgf/lb Four ball EP, ASTM D2596 — Weld load, min. kgf Load wear index, min. kgf Deleterious particles, ASTM D1404, max. number of scratches Pumpability,b Lincoln ventability test, nominal max. psi at lowest anticipated ambient temperature
0
1
2
2
355–385
310–340
265–295
265–295
Not applicable
177/350
177/350
260/500
68
100
220
220
35/5
35/5
35/5
35/5
30 Not applicable
30 10
30 10
30 10
Pass 0.0
Pass 0.0
Pass 0.0
Pass 0.0
2
2
2
2
20/45
20/45
20/45
20/45
315 45 20
315 45 20
315 45 20
315 45 20
400
400
400
400
a The preferred grade is NLGI #2. However, an NLGI #1 or #0 may be required for proper dispensing at low temperatures, or
by centralized lubrication systems. b This is applicable for use only in centralized lubrication systems.
TABLE 20.12
Performance Requirements for Multi-Service Mining Lubricant
Flash point Copper strip corrosion Rust protection Four ball wear test Four ball EP test Load Wear Index Timken EP test US steel retention test (Timken) Base oil viscosity Solid lubricant particle size Asphaltene content Pumpability
ASTM D92/ISO 2592 ASTM D4048, 24 h at 100◦ C ASTM D1743 ASTM D2266, scar diameter, 60 min at 40 kgf ASTM D2596, weld point ASTM D2509, no score US steel DM51, 4 g, no refill, 33 lb for 30 min ASTM D445/ISO3104/D2161 ASTMD3279 US steel mobility test, DM 43, test run at 150 psi Lincoln Ventmeter test, VE-1, test run for 30 sec
130◦ C (266◦ F) mina 1b max. Pass 0.7 mm max. 500 kg min. 70 min. 50 lb min. Pass 300 cSt min. at 40◦ Cb 20 µm max. Less than 0.10% 0.30 g/m min. at lowest operating temperature 400 psi max. at lowest operating temperature
a Note for diluent containing lubricants: Although the flashpoint requirement applies to the mixture of
diluent and lubricant, it should be noted that the flashpoint of a diluent alone (i.e., the flashpoint of the diluent when it is not mixed with the lubricant) may be considerably lower than the flashpoint of the mixture of lubricant and diluent. b While 300 cSt is the minimum viscosity required for most climates, it may be necessary to formulate products with slightly lower base oil viscosity for use in arctic conditions. In such cases these products should only be applied within their intended temperature range.
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20-22 TABLE 20.13
Handbook of Lubrication and Tribology Enclosed Gear-Case Oil Requirements
Viscosity grade Viscosity Viscosity index, ASTM D2270/ISO 2909, min. Cleanliness Additive solubility Pour point, ASTM D97/ISO 3016 max. Flash point, ASTM D92/ISO 2592, min. Rust test, ASTM D665/ISO 7120, Procedure A Procedure B Copper strip corrosion, ASTM D130/ISO 2160 3 h at 100◦ C/212◦ F, max. Oxidation stability, ASTM D2893, at 121◦ C/250◦ F max. % viscosity increase Foam suppression ASTM D892
ISO VG 68 — 1500, AGMA 2EP-9EP. In accordance with ASTM D2422/ISO 3104/D2161 In accordance with ASTM D445/ISO 3104/D2161 90 Must be free of visible contaminants Must be filterable to 25 µm (Beta 25 = 200 filter rating) without loss of additive(s) 5◦ C/9◦ F below the lowest anticipated ambient temperature 204◦ C/400◦ F Pass Pass 1b 6 (Viscosity grades 68 to 6880) 10 (Viscosity grades 1000 to 1500) Max volume of foam (ml) after 5 min blow
Seq. I 24◦ C/75◦ F Seq. II 93.5◦ C/200◦ F Seq. III 24◦ C/75◦ F
75 75 75
Demulsibility ASTM D2711 Must be within these limits Max % water in oil after 5 h test Max cuff after centrifuging Min total free water collected during entire test (start with 90 ml water) Four ball EP test, ASTM D2783 Weld load, kgf, min Load wear index, kgf, min Timken EP test, ASTM D2782 OK value, kgf, min FZG test, ASTM D5182 with A/8.3/90◦ C parameters, Fail stage, min
10 min blow 10 10 10 Viscosity grades
68–460 2.0 1.0 ml 80 ml
680–1500 2.0 1.0 ml 50 ml
250 55 27 >12
gears. They may be used for splash, recirculating, or total loss applications and the viscosity grade has to match the conditions of the gearbox. Typical manufacturers’ performance requirements are given in Table 20.13.
20.7.4 Wire Rope Lubricant A typical manufacturer’s performance specification is given in Table 20.14. This specification covers wire rope lubricants manufactured from mineral oil or synthetic base stocks, or a blend of mineral oil and synthetic base stocks. These materials are intended to lubricate large-diameter wire ropes such as those used for hoist or drag functions on draglines or shovels. It covers all grades of wire rope lubricants that are usable from −46 to 49◦ C. Materials furnished under this specification may contain a diluent to enhance penetration to the rope core, or improve sprayability. The specification states that these wire rope lubricants must have excellent adhesive and cohesive qualities, must not chip or throw off, have excellent water resistant and rust-preventive qualities, and be
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TABLE 20.14 Wire Rope Lubricant Minimum Performance Requirements Base fluid viscosity, ISO 3194/ASTM D 445;D2161 min. mm2 /sec at 40◦ C Flash point, ASTM D 92/ISO 2592, min. ◦ C/◦ F (as applied) 4-Ball EP Test, ASTM D 2596, weld point, min. kgf, 4-Ball wear test, ASTM D 2266, 60 min at 40 kgf, max. scar, mm Rust protection, ASTM D 1743 Deleterious particles, ASTM D 1404, max. number of scratches Adhesive properties Pumpability, Lincoln ventability test, nominal max.
320 79/175 250 1.0 Pass 20 Must adhere to surfaces at lowest anticipated temperature 400
Psia at lowest anticipated ambient temperature
a Consult lubricant system component manufacturer regarding any deviations form this requirement.
TABLE 20.15 Antiwear Hydraulic Oil Minimum Performance Requirements
Viscosity grade, ISO 3448/ASTM D 2422 Viscosity index, min. ISO 2909ASTM D 2270 Pour point, ISO 3016/ASTM D 97, ◦ C/◦ F max. Flash point, ISO 2592/ASTM D 92, ◦ C/◦ F min. Copper strip corrosion, ISO 2160/ASTM D 130, 3 h at 100◦ C/212◦ F, max. Rust test, ASTM D 665 Procedure A Procedure B Oxidation stability, ISO4263/ASTM D 943, min hours to reach a neutralization number of 2.0 Foam suppression, ASTM D 892, max ml after 5 min/10 min blow Seq. 1 24◦ C/75◦ F Seq. II 93.5◦ C/200◦ F Seq. III 24◦ C/75◦ F Demulsibility, ISO 6614/ASTM D 1401, max. minutes to obtain an emulsion of 40-37-3 at 54◦ C/129◦ F Vane pump wear test, ASTM D 2882 FZG test, DIN 51 354, with A/8.3/90◦ C parameters, min fail stage Must be free of visible contaminants
ISOVG32
ISOVG46
ISOVG68
ISOVG100
32 95 −29/−20 177/350
46 95 −26/−15 177/350
68 95 −23/−10 204/400
100 95 −18/0 204/400
1b
1b
1b
1b
Pass Pass
Pass Pass
Pass Pass
Pass Pass
3,000
3,000
3,000
3,000
75/10 75/10 75/10
75/10 75/10 75/10
75/10 75/10 75/10
75/10 75/10 75/10
30 Pass >10
30 Pass >10
30 Pass >10
30 Pass >10
Note: Additive solubility — must be filterable to 3 µm (beta = 200 filter rating) without loss of additive(s).
capable of penetrating into the body of wire rope, yet maintain retarded dripping qualities for operation over wide temperature ranges. These products shall contain chemical EP and solid film additives to improve film strength and control fretting and rubbing friction during operation.
20.7.5 Antiwear Hydraulic Oil This specification covers premium circulating oils produced from refined mineral oil-base stocks, and compounded with antiwear additives for high load-carrying ability. These materials are primarily intended for use in hydraulic systems operating within an ambient temperature range of −18◦ C/0◦ F to 54◦ C/130◦ F. They may also be used to lubricate high-speed plain or rolling element bearings, lightly loaded enclosed gear drives, and miscellaneous items such as links, pins, and bushings operating in circulating, sump (splash), or total loss applications. Typical manufacturers’ performance requirements are given in Table 20.15.
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References [1] Anon, The World of SKF, SKF Publication 3277E Reg 0770.75.000.19803, Sweden, p. 202. [2] Lehner, S. and Jacobs, G., Contamination sensitivity of hydraulic pumps and valves, Tribology of Hydraulic Pump Testing, ASTM STP 1310, George E. Totten, Gary H. Kling, and Donald J. Smolenski (eds), American Society for Testing and Materials, 1996. [3] Tessman, R.K. and Hong, I.T., Hydraulic pump contaminant wear, Hydraulic Failure Analysis: Fluids, Components and System Effects, ASTM STP 1339, G.E. Totten, D.K. Wills, and D. Feldman (eds), American Society for Testing and Materials, West Conshohocken, PA, 2001. [4] Roberts, W.H., The lubrication of rolling bearings, 3rd Annual Conference on Industrial Tribology, Caulfield Institute of Technology, Melbourne, Australia, 48, 1982. [5] Bensch, L.E., Fitch, E.C., and Tessman, R.K., Contamination Control for the Fluid Power Industry, Pacific Scientific Company, Montclair, CA, 1978. [6] Jagger, E.T., Lip seals, Tribology Handbook, 2nd ed., M.J. Neale, (ed.), Newnes-Butterworth, 1975, A35. [7] Lawrence, R.T., Soft piston seals, Tribology Handbook, 2nd ed., M.J. Neale, (ed.), London, 1975, A37. [8] Anon, SKF Electronic Handbook, SKF Publication 4485E 10000, Sweden, 1995. [9] Anon, FAG Technical Publication and General Catalogue, Compact Disc, FAG Australia Pvt. Ltd., 2000. [10] Ioannides, E. and Jacobson, B., “Dirty lubricants — reduced bearing life,” Ball Bearing Journal, Special Issue 89, 22, 1989. [11] Anon, The Lubrication of Rolling Bearings, Publ. No. WL 81 115/2 EF/98/10/87, FAG Australia Pvt. Ltd. [12] Harris, J.H., The Lubrication of Rolling Bearings, Shell-Mex and B.P. Ltd., London, 1967. [13] Cheng, H.S., Elastohydrodynamic Lubrication, Handbook of Lubrication, Volume II, Theory and Design, E.R. Booser (ed.), CRC Press, Boca Raton, FL, 1983, 139. [14] Anon. Roller Bearing Lubrication Handbook, Fuchs Petrolub AG Oel + Chemie, GfT Work Sheet 3, 1994. [15] Ribble, H.C., Cast Bronze Bearing Design Manual, 2nd ed., Cast Bronze Bearing Institute Inc., Cleveland, Ohio, 1965. [16] Wills, J.G., Lubrication Fundamentals, Mobile Oil Corporation, Marcel Dekker Inc., New York, 1980, 112. [17] Scott, W., Design detail affecting reliability, Presented at International Mechanical Congress, MECH ’91 I.E. Aust., Sydney, Australia, 1991. [18] AGMA Specifications, Lubrication of Industrial Open Gearing. (AGMA 250.02), The American Gear Manufacturers Association, Washington, D.C. [19] Klaus, E.E. and Tewksbury, E.J., Liquid lubricants, Handbook of Lubrication, Volume II, Theory and Design, E.R. Booser (ed.), CRC Press, Boca Raton, FL, 229, 1983. [20] Wellauer, E.J. and Holloway, G.A., Application of EHD oil film theory to industrial gear drives, Trans. ASME, J. Eng. Ind., 98B, 626, 1976. [21] Roth, H.E., Design and manufacture for load distribution, Presented at International Conference on Mining Machinery, I.E. Aust, National Conference Publ. No. 79/5, Brisbane, Australia, 368, 1979. [22] Annual Book of ASTM Standards, Section 5, Petroleum products, lubricants and fossil fuels, Vol. 05.03 Petroleum Products and Lubricants (III), D4636-latest; Catalysts, 1997. [23] Williamson, J.B.P., The shape of surfaces, Handbook of Lubrication, Volume II, Theory and Design, E.R. Booser (ed.), CRC Press, Boca Raton, FL, 3, 1983. [24] Scott, W. and Hargreaves, D.J., Specifying surface roughness for spur and helical gears, Tribology for Energy Conservation, D. Dowson et al. (eds), Elsevier Science B.V. 267, 1998. [25] Scott, W., Report on the Lubrication of Dragline Gears for BMA, Scott Tribology Services Pty. Ltd., Brisbane, Australia, August 2004.
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[26] Bartz, W.J. and Kruger, V., Test Method for Evaluating the Influence of Lubricants and Lubricant Additives on Fatigue Failure of Quenched and Tempered Case-Hardened Spur Gears, Rolling Contact Fatigue-Performance Testing of Lubricants, R. Tourret and E.P. Wright (eds), Institute of Petroleum, London, 161, 1977. [27] Anon, Hoist rope lubrication criteria, Battelle Columbus Laboratories, Report No. PB80-182959, Prepared for Bureau of Mines, Washington, D.C., 1978. [28] Critchlow, J.P. and Flynn, R.W., Wire rope lubricants and lubrication, Lubrication Engineering, August 1951, 178–181 and 195. [29] Kaderjak, G., Stranding head for the internal lubrication of steel wire ropes, Wire World International, 18, 35, Jan–Feb 1976.
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21 Farm and Construction Equipment 21.1 21.2 21.3 21.4 21.5
Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Types of Equipment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Development of Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . Know the Machine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
21-2 21-2 21-2 21-3 21-3
Diesel • Gasoline • Liquid Propane (LP) Gas • All Engines
21.6 Clutches, Fluid Couplings, and Converters . . . . . . . . . .
21-4
Clutches • Fluid Couplings • Converters
21.7 Electric Motors, Generators, and Rectifiers . . . . . . . . . .
21-5
Electric Motors, Generators • Motors, Generators • Rectifiers
21.8 Gear Drives . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
21-6
Powershift and Automatic Transmissions • Manual Transmissions, Enclosed Gear Drives • Open Gear Drives
21.9 21.10 21.11 21.12 21.13 21.14
Couplings. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chain Drives . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Wire Ropes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Crawler Mechanisms . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Hydrostatic Hydraulic Systems . . . . . . . . . . . . . . . . . . . . . . . Bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
21-9 21-9 21-10 21-10 21-10 21-11
Antifriction Bearings • Plain Bearings • Bearing Lubricants
R. Lal Kushwaha Professor, Machinery Systems Ag. and Biosource Eng. Dept. University of Saskatchewan
Jude Liu Post-Doctoral Fellow Ag. and Biosource Eng. Dept. University of Saskatchewan
21.15 Pneumatic Equipment. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
21-13
Air Compressors • Air Cylinders • Air Motors • Pneumatic Tools
21.16 Good Lubrication Practice . . . . . . . . . . . . . . . . . . . . . . . . . . . .
21-14
Management • Operators • Maintenance • Supplier • Oil Analysis
Acknowledgment. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
21-18 21-18
21-1
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Handbook of Lubrication and Tribology
21.1 Introduction Farmers and contractors in the United States currently claim assets of close to $2 trillion and employ over 8 million workers. The value of farm cash receipts exceeds $240 billion while the value of construction is well over a trillion dollars per year. The equipment currently consumes over 30 billion gal of fuel per year and uses more than 300 million gal per year of oils and greases (www.nass.usda.gov/census/ and www.census.gov/mcd/). Easily, farming and construction together are this country’s biggest business and the world’s largest and most important endeavor, from providing the four Rs of road work (reconstruction, resurfacing, rehabilitation, and recycling) to producing food and fiber. Their machines are built for utility and the ultimate use of raw power, but are still beautiful in their complexity. Any farm tractor can be fitted with a dozer blade, bucket, or backhoe to move earth for any construction project on the spread and the farm equipment manufacturers have naturally progressed to the manufacture of heavier units for construction of buildings, pipelines, and roads. Whereas each industry has machines particular to certain work, all employ the same basic power and drive train systems, all work off the road, and all are subject to the same heavy-duty service and inhospitable outdoor working conditions.
21.2 Types of Equipment There are more than 100 distinctive equipment types, but they can be put into a few basic categories according to the work they do: • Digging, filling, earth-moving Backhoes, shovels, scrapers, excavators, dozers, draglines, tractors, trenchers, and graders • Loading Loaders, shovels, excavators, and lift trucks • Hauling Trucks and trailers • Drilling, breaking Rippers, drills, hammers, and planers • Processing Harvesters, crushers, screens, and mixers (asphalt, concrete) • Placing, laying Cranes, pipelayers, layers (asphalt, concrete), and rollers • Auxiliary Air compressors, pumps, pneumatic tools, welders, finishing tools, and generating plants Since attachments can be made to many of the basic machines, most do more than one job. If the changes could be made overnight, a crane could be digging one day, driving piles the next, and placing steel the day after that. An excavator could be digging a basement one day and punching holes in old concrete the next. A typical farm tractor with a variety of add-ons can plow, load and haul, or dig a trench.
21.3 Development of Lubricants The agencies involved in engine oil classifications: 1. 2. 3. 4. 5. 6. 7.
American Automobile Manufacturers Association (AAMA) Japan Automobile Manufacturers Association (JAMA) International Lubricant and Standardization and Approval Committee Society of Automotive Engineers (SAE International) American Society for Testing and Materials (ASTM) American Petroleum Institute (API) Federal Government — Environmental Protection Agency (EPA)
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TABLE 21.1 API Oil Classifications Year
Light-duty engine oil
Heavy-duty engine oil
1955 1968 1972 1980 1985 1988 1991 1993 1994 1995 1996 1998 2001 2002 2004
— SD SE SF — SG — SH — — GF-2 — GF-3 — GF-4
CD — — — CD-II CE CF-4 — CF, CF-II CG-4 — CH-4 — PC-9 —
Table 21.1 lists the development of various lubricants classifications according to API for the light-duty engines (spark ignition) and the heavy-duty engines (compression ignition) (Imperial oil, 1999).
21.4 Know the Machine Good lubrication practice requires a thorough knowledge of the machine, from the prime mover through to the final application of the force to get a job done. Stationary equipment has essentially one power flow, to crush rock, size aggregate, convey materials, etc. Mobile equipment incorporates two power trains — one to propel the machine, another to do the work intended. A careful trace of the flow of power in the machine is vital, not only to recommend the proper lubricants, but also to assure that no elements in the system are missed. Modern machinery employs every conceivable method of transmitting power to the load or the road. Electric motors or internal combustion engines exert power through transmissions or gear boxes to final drives, or through converters to chain and gear drives, or send power to do work through hydrostatic pump and motor systems. Loads are handled by wire ropes, hydraulic cylinders, and gear-cases. Power is transmitted through couplings and shafts. All machine makers provide operator’s manuals with a lubrication chart; if not for the composite machine, at least for the various components. Many of these unfortunately become outdated or get lost. It falls on the lubrication engineer, even if a metal plate or a decal is attached to the machine, to use his experience, knowledge of power transmission, and familiarity with lubricants and lubrication to see that best lubrication practice is performed.
21.5 Engines 21.5.1 Diesel Most farm and construction machines are powered by diesel engines since this type of work calls for low-speed, high-torque operation, and lower rates of acceleration than found in many gasoline or gas (Otto cycle) engine applications. The diesel combustion cycle is well fitted for this type of service with high-torque rise at lower engine speeds being part of the basic design. The engines may be of two- or four-stroke cycle, direct injected or prechamber, turbocharged or naturally aspirated, and air or water cooled. They are a rugged and dependable power source, operate many hours before major overhaul is required, and cost per brake horsepower–hour is usually lower than for other engines.
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Because of the high compression ratio of diesel engines, 14 to 20 : 1, protection of the piston ring belt area from deposits is of prime importance. As a general rule, engine oils of American Petroleum Institute (API) Service Classification suitable for naturally aspirated engines and for supercharged or turbocharged engines should be used. From the environmental viewpoint, it is important to match the acid neutralizing capability of the oil, indicated by total base number (TBN), to the fuel used.
21.5.2 Gasoline These engines are used in smaller horsepower ranges for mobile and stationary equipment and may also be of the two- or four-stroke cycle design, air or water cooled. In general, more appropriate oils of API Service Classification suitable for four stroke engines as well as for two-stroke engines (oil premixed with the fuel) should be used. Some engine manufacturers may require the use of very low ash content oils (≈40 ppm), specifically formulated for this type engine.
21.5.3 Liquid Propane (LP) Gas Engines in lift trucks and smaller unit applications may be of this design, using fuel commonly called “propane.” This fuel burns cleanly and causes little or no crankcase oil dilution, so oils with mild detergentdispersancy and with oxidation and bearing corrosion inhibitors may be used in most cases. To offset increasing viscosity due to oxidation, make-up oil of the next lower SAE number may be required.
21.5.4 All Engines While, in general, the service (“S”) classifications of engine oils are used to lubricate gasoline and LP gas engines, and commercial (“C”) classifications are recommended for diesel engines, all engine manufacturers recognize the specific ASTM test requirements of all classifications and have experience with particular formulations. They seek to incorporate in their specific recommendations the properties required by their engines for their design purposes. Therefore, most manufacturers will cross classify their recommendations for engine oil. Many engines are also critical as to oil ash content and additive chemistry. To prolong engine life, most manufacturers will specify the ash percentage (as a maximum, range, or minimum) and require, or effectively dictate, the percentage of certain detergent-dispersant and antiwear additives. Oil change intervals vary with many parameters: engine application, climatic conditions, degree of turbocharging, oil quality, fuel quality, type of oil filtration, fuel consumption, oil consumption, and crankcase capacity.
21.6 Clutches, Fluid Couplings, and Converters 21.6.1 Clutches Conventional dry face type or friction clutches may require lubrication of the pilot and release bearings. Most require the use of high-temperature grease, sparingly applied, but applied often enough. Wet type clutches run in oil which is usually the transmission fluid, may be the engine oil, or may be a separate fluid. Lubricant should be changed periodically.
21.6.2 Fluid Couplings These units absorb the shock loading forces between engine and drive, using oil as the “cushion” and connection between the rotors. The engine oil or other recommended fluid requires regular changing.
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Farm and Construction Equipment
21-5 Boom hoist brakes
Worm gear
Boom hoist Boom hoist drum
Boom hoist drive shaft Drum drive-floating on shaft Chain drives Horizontal independent swing shaft Drum shaft Main drive shaft Swing converter drive
Slide pinion shaft Vertical travel shaft
Hoist converter drive Vertical independent swing shaft Swing shaft Ring gear & roller path Horizontal travel shaft Crawler drive sprocket
FIGURE 21.1 Work and propel power trains showing variety of machine elements to rotate wire rope drums, swing the upper works, and propel the lower works of a crane. (Courtesy of Manitowoc Engineering Company, WI.)
21.6.3 Converters These units have the feature of multiplying engine torque and provide greater flexibility in power train design and operation. In the propel power train of mobile machines they may or may not be integral with the transmission and, in most cases, will automatically lock up into direct drive at specified higher revolutions per minute to increase speed and fuel efficiency. In many machines, for example, cranes, the load power train will find torque converters as separate, controlled units ahead of chain or gear drives (Figure 21.1). Depending on the maker and whether of single or multiple stage design, these units list a variety of oils that should be used. In one type of installation, diesel fuel is the hydraulic medium, the supply and return being piped to the fuel storage tank. Most units, however, either use inhibited oils of various viscosities specially formulated for the purpose, or engine oils and automatic transmission fluids. Since efficiency is affected by viscosity, the manufacturer’s and machine maker’s recommendations should be strictly followed. Converters, along with fluid couplings, are examples of the hydrodynamic principle of fluid in motion.
21.7 Electric Motors, Generators, and Rectifiers 21.7.1 Electric Motors, Generators Some of the largest equipment employs electric motors, both as the prime mover to propel the machine and as the source of power to do the work. Generators may be driven by electric motors or internal
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combustion engines. Prime examples are shovels, cranes, pumps, crushing and screening plants, and the largest loaders and haulage trucks. In the latter examples, a diesel engine is the prime mover for a generator, with electric motors supplying the final application of power to the wheels.
21.7.2 Motors, Generators These rotate on ball or roller bearings which are generally grease lubricated. In some applications, particularly where thrust bearings are necessary, various viscosities of inhibited oils are recommended. Greases should be sparingly applied and at relatively long intervals. The bearing housing should be maintained at one-quarter to one-third full. Since overfilling causes heat due to internal friction in the grease, the usual procedure is to remove the relief plug, apply grease until new grease is evident, then run the unit for a few minutes to expel excess lubricant. Following the manufacturer’s recommendations is mandatory. Greases should be of high quality, with dropping points suitably high for the operating temperature, exhibiting high degrees of oxidation resistance and water tolerance, and having excellent corrosion protection properties. Another determining factor in the grease formulation is the insulation temperature class of the unit, and this should be known before recommending the lubricant to be used. Periodically, the bearings should be removed, cleaned, and inspected, depending on the severity of service.
21.7.3 Rectifiers These units change AC to DC for the final power source. Transformer oils are used for cooling and require replacement at dictated intervals.
21.8 Gear Drives In both the propel and load power trains, gear drives provide intermediate and final applications of desired torque and speed. Drive trains consist of transmissions, transfer cases, differentials, and final drives, or planetaries (Figures 21.2–21.4). Load trains incorporate transmissions, power take-offs, pump drives, reduction gear cases, and open gears (see Figure 21.1 and Figure 21.5). Gear designs used are spur, spiral bevel, bevel, helical, hypoid, herringbone, and worm. Planetary types of gear drives are used in both the propel and load trains.
21.8.1 Powershift and Automatic Transmissions These units have contributed much to lessening operator fatigue by reducing the need to use the clutch foot pedal and to the life of machine components by the easier selection of the proper gear for the job at hand. In the automatic, gear ratios are selected automatically by the transmission, sensing the proper ratio needed, whereas in the powershift the ratios are selected by the operator moving a lever. Many newer units incorporate the features of both designs. In all these types of transmissions, internal clutches, applied hydraulically, route the power through the proper gear sets to the output shaft. Manufacturers of these transmissions have definite recommendations on lubricants since additive treatment plays a large role in the life and efficiency of the components. Additive systems must be compatible with the frictional material used and maintain the friction-retention properties designed into the clutch packs for the amplitude of torque applied. Fluids of the Type A, Suffix A; DEXRON®; DEXRON®-II and Type C-1 and C-2 specifications are the generally recommended lubricants for automatic and hydraulic transmissions, with provisions by each manufacturer that compatibility and friction-retention properties must be met (refer to SAE J311 Standard). Viscosity characteristics at various temperatures are a consideration in applying the proper lubricant. The last decade has seen the development of many farm and construction wheeled tractors which simplify lubrication by the use of the same oil (in a common sump) for the lubrication of the gears
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Farm and Construction Equipment
21-7
Planetary final drive
Differential Drive shaft
Transmission
FIGURE 21.2 The power train of John Deere 9020 Series Wheel Tractors. (Courtesy of John Deere, Deere and Company, Moline, IL.)
Hydraulic motor Drive sprocket Reduction gears
FIGURE 21.3 Drive train showing hydraulic motor, reduction gears, and drive sprocket of a crawler (hydrostatic drive). (Courtesy of FMC Corporation, Cedar Rapids, IA.)
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Handbook of Lubrication and Tribology
Reduction gears Clutch
Drive shaft Transmission oil
FIGURE 21.4 The Powershift transmission of John Deere 8020 Series Tractors. (Courtesy of John Deere, Deere and Company, Moline, IL.)
Hydraulic pump drive Hydraulic motors
Hydraulic pumps Swing gear
FIGURE 21.5 Pump drive, pumps for propel and tool hydraulic systems, swing pump and swing motors on a hydrostatic crawler-excavator. (Courtesy of FMC Corporation, Cedar Rapids, IA.)
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21-9
in the transmission, differential, and planetaries, and also as the hydraulic medium to operate the tools (Figure 21.4). These lubricants, commonly called “tractor hydraulic fluids” or “tractor universal fluids,” incorporate additives to satisfy the antiwear and lubricity requirements of gears, wet brakes, and, hydraulic pumps.
21.8.2 Manual Transmissions, Enclosed Gear Drives Because of the heavy-duty nature of farming and construction work, gear teeth must be protected by antiwear or extreme-pressure type lubricants. In some designs, however, where increased loadcarrying capacity is not needed,“regular type”lubricants are satisfactory. Recommendations range from API Service Classification GL-l through GL-6 for the gears of drive trains and load trains on mobile equipment to the American Gear Manufacturers’ Association (AGMA) lubricants of rust and oxidation inhibited, fatty oil compounded, or mild extreme-pressure types for some reduction units on stationary machinery (refer to SAE J306 and SAE J308 Standards). Many machine manufacturers allow the use of engine oils in some gear boxes, the antiwear or extreme-pressure quality being sufficient for these designs. In many steering gear boxes and some planetary type gear cases, greases of low to medium consistencies and containing extreme-pressure additives or friction-reducing solids are recommended.
21.8.3 Open Gear Drives This type of power transmission is found on most mobile and stationary equipment. Design is usually of the spur gear type and in most cases calls for manual application of heavy, “residual type” lubricants of seasonal consistency, but usually needing heat to apply. Solvent cut-back types, wherein a solvent eases the application and then evaporates to leave a durable film, are the more convenient lubricants. Some designers call for the additional protection of extreme-pressure additives for their applications. In many cases, the lubricant is applied from a remote-point automatic system, in which case the lubricant must be able to be pumped through long lines. Many machine manufacturers find extreme-pressure greases or those made with heavier oils to be satisfactory lubricants.
21.9 Couplings Many drivelines employ a coupling between the power source and the gear drive to protect against shock loading and allow for possible misalignment of the major units. Those with elastomeric inserts or of flexible plate type require no lubrication, but many couplings are of the grid spring, chain, or gear design and require periodic lubrication with grease or heavy, residual oils depending on the type and conditions of service. Much work has been recently completed on the properties needed by greases. 1. High-speed service — low oil separation, high base oil viscosity at 40◦ C (104◦ F), high dropping point. 2. High-torque, high misalignment (low-speed) service — low oil separation, good lubricity, low dropping point. The latter type of service will be the most likely case in farm and construction machinery, but both types of service will be encountered.
21.10 Chain Drives This type of power transmission is used extensively in off-the-road equipment. Rapid wear will occur if the internal and external parts are not properly lubricated and protected from contamination and corrosion. Larger open chains, such as crawler drive chains, are lubricated manually and usually with engine or gear oils of proper viscosity, but many other applications are satisfied with open gear type lubricants or with
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greases. Such open chains, particularly where exposed to excessive contamination, may be better run with frequent removal for cleaning in fuel oil and soaking in lubricant. Even under ideal conditions, lubricated chains should receive the cleaning and soaking treatment periodically. Enclosed chain drives, being built to greater precision, may be designed to run in an oil bath or be lubricated by drip oilers. Lubricants recommended are seasonal viscosities of engine, extreme-pressure gear, or antiwear oils. To be found on some machines are grease lubricated chain cases, the lubricant being extreme-pressure greases of semifluid or low consistency numbers.
21.11 Wire Ropes These are the prime tools in the construction industry for operating a tool at a distance from the machine, such as in drag line operation or in placing steel for tall buildings. Ropes require not just protection from rust and corrosion, but also wear protection of the individual strands as they rub against each other during flexing over sheaves and around drums. Regular lubrication is necessary because of the continual squeezing out of the lubricant from the inside strands to the outer. Application of lubricants by brush or swab while on the machine, while protecting outer strands and sealing in the inner lubricant, does not effect proper penetration to the inner wearing surfaces. Periodically wire rope should be removed from the machine, cleaned in solvent or kerosene, and then relubricated by passing it through a device which will bend the rope over sheaves in a heated bath or by soaking it therein. Moving cables should receive the most attention, but all ropes need frequent applications of lubricant. One exception is dry rope used in such jobs as drag line operation, where abrasives will tend to be collected by the lubricant. Depending on the operation, a variety of lubricants are called for. The heavy, residual type compounds which may be applied at ambient temperatures or may be heated are the most popular recommendations. These may be solvent cut-back and may also contain extreme-pressure agents. In some dusty atmospheres, lighter-bodied oils better serve the purpose. Gaining wider use are specially formulated grease types, compounded with inhibitors and an effective penetrant.
21.12 Crawler Mechanisms Track laying assemblies or crawlers are used extensively to provide propulsion and tractive efforts for tractors, shovels, cranes, etc. used in construction, with still a few being found on the farm. The mechanism consists of an endless steel belt, driving sprockets and idler tumblers, carrier rollers, and the weightsupporting track rollers. Older types require frequent lubrication with special equipment and special low consistency greases of the tacky, water-resistant type, incorporating a heavy oil. Under some conditions, appropriate viscosities of engine oil or extreme-pressure gear oil are used. Newer designs may also require special equipment for the application of greases or engine oils. Lubrication intervals vary, depending on contaminating conditions and seal design. Some units are designed to hold lubricant until overhaul; others call for daily relubrication to aid in flushing out contaminants. When traveling distances during operation, some of the latter should be greased every 0.8 km ( 12 mi) or 12 h.
21.13 Hydrostatic Hydraulic Systems Fluid power is transferred to mechanical action in farm and construction equipment by two methods: inline by means of a cylinder and piston, or rotary by means of a rotating motor (see Figures 21.2–21.5). The flexibility of the system lies in the prime power being routed to the load through lines rather than through gear trains or wire ropes; power is smoothly applied with infinite control of speed and torque. In the open loop hydraulic system the fluid is pumped from a reservoir by a vane, piston, or gear type pump through lines containing pressure and flow control valves to cylinders or motors that perform the mechanical work, then back to the reservoir. In many systems, pressure and flow are controlled by variable volume/pressure
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compensated pumps and speed is controlled by a variable motor. Many machines utilize closed loop systems, wherein oil from the discharge port of the pump flows to the motor and then directly back to the pump inlet port. This arrangement is called a hydrostatic transmission. Both open and closed loop designs may be found on the same machine. The most readily identified hydraulic system on any machine uses cylinders and pistons for straight-line back and forth motion. These systems are found mainly in the work train to manipulate tools such as the backhoe or bucket, and to operate controls such as those for clutches and brakes. In the propel train, the steering of many wheeled vehicles is done by cylinders, either as an assist or fulltime. Those hydraulic systems employing rotating motors to do the mechanical work are being used increasingly in both the work and propel trains, and in both wheel and track vehicles. On a single machine, for example, some motors may power wire rope drums while others travel the lower and others swing the upper works. Hydraulic motors are of the piston, vane, or gear design and usually (but not always) match the pump as to type. Oil recommendations run the gamut of types produced, depending on the application: mineral oils (nonantiwear and antiwear), water-in-oil invert emulsions, oil-in-water emulsions, water–glycols, phosphate esters and blends, synthesized hydrocarbons, engine oils and the transmission fluids such as General Motors’ Type A, Suffix A, DEXRON®, and DEXRON®-II; Ford’s M2C 33F; Allison’s C-3; and the so-called tractor fluids. With most modern hydraulic systems operating well above 14,000 kPa (2,000 psi), most recommendations will be for those fluids fortified with appropriate degrees of antiwear. Because of the variety of recommendations and the alternates allowed, a farmer or contractor with a large variety of equipment must pay particular attention to hydraulic oil stock and application. Since viscosity is the most important single property, makers of equipment and hydraulic system components recommend use of specific viscosities and viscosity indexes to obtain the best efficiency at the operating temperature of the fluid. Limits take many forms and, since the operating temperature may vary from one point in the system to another, are compromises. Machine makers usually recommend a certain viscosity number for a range of temperatures, some indicating that the lower temperature is the minimum ambient for start-up and the higher one the maximum oil operating temperature. Component manufacturers (and some machine makers) recommend a minimum viscosity index (usually 90), a maximum viscosity at cold start-up, a viscosity range at operating temperature of the fluid, and an optimum operating viscosity. Recommendations vary according to the experience and experimentation of each equipment and component manufacturer for each pump, motor, and cylinder design, so it is paramount that the users of hydraulic fluids be familiar with the requirements of the particular system. In general, hydraulic fluid must have the viscosity (and film strength) to adequately lubricate the closely machined parts of the system, yet not be too high in viscosity to lower efficiency or cause cavitation in the pump (and resulting noise), nor yet too light to lower efficiency or promote leakages. Very few mobile equipment hydraulic systems using mineral oils operate in the ideal range of 54 to 60◦ C (130 to 140◦ F), most of them reaching temperatures of 82◦ C (180◦ F) or even 100◦ C (212◦ F). Water-containing fluids generally should be maintained at operating temperatures below 49◦ C (120◦ F), but in many sealed systems such fluids are used at temperatures up to 82◦ C (180◦ F).
21.14 Bearings Because of the variety of stationary and rolling equipment in the farm and construction industries, every type of bearing design is used to support rotating parts in place and provide their free and efficient motion. While most bearings in enclosed units such as engines and gear boxes receive lubrication by splash, mist, or an internal circulating system, even these units may have bearings that need periodic special attention. Some gear reducers employ an upper bearing that does not receive lubrication by splash or carry-up, and so needs periodic application of grease. Many older engines call for regular attention to accessories such as starter motors, generators, distributors, fan drives, and water pumps. Engine accessories on many newer engines are equipped with nonrelubricatable bearings.
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Wheel bearings, drive lines, and steering and suspension systems of mobile farm and construction equipment require much more frequent lubrication than on-road vehicles. While wheel bearings and many bearings in the work train are repacked or run in an oil bath, most bearings are equipped with fittings or oil or grease cups for daily, weekly, and monthly applications. However, many bearings that demand fairly frequent lubrication have threaded plugs in their housings — these are to be removed and a fitting installed. It is necessary to become familiar with this practice by knowing the drive or load trains of each machine. Not to be overlooked are those moving points such as linkages and pivots that require hand oiling.
21.14.1 Antifriction Bearings Best lubricant recommendation and application for rolling bearings requires knowing the bearing operating temperature and speed. Generally, for operating temperatures below 93◦ C (200◦ F) and speeds below 3000 rpm, the usual greases are used. Above these figures, oils are generally used and oil circulation or mist is considered. However, many of the complex soap, polymer, or inorganic thickened greases perform well up to about 204◦ C (400◦ F) with synthetic formulations being recommended for up to 246◦ C (475◦ F). The caveat in all grease applications is to use greases at maximum operating temperatures of 56◦ C (100◦ F) below the grease dropping point. The operating temperature of the lubricant itself will normally be between 3 and 11◦ C (5 and ◦ 20 F) higher than the housing temperature, which can be fairly easily determined with a contact thermometer. The required viscosity of the oil, or oil component in a grease, can be determined from viscosity–temperature charts for that particular oil. When the bearing load is unknown, the following may serve as a guide for minimum viscosity at the lubricant operating temperature: Bearing type Ball, cylindrical roller, needle Spherical roller Thrust
Minimum oil viscosity (cSt) 13.00 20.45 31.85
Bearing housings should not be overfilled, because overheating may result due to churning of the lubricant. The rule for oil bath housings calls for the lubricant level to be no higher than the center of the lowermost ball or roller when the bearing is at rest; grease lubricated bearings are satisfied with the bearing or bearing housing one-third to half full, keeping in mind that the higher the speed the lesser grease quantity, so long as a sufficient amount is present. On many bearings lubricated through fittings, a vent plug may be provided. This control serves to prevent overfilling since grease is applied during running, with the plug removed until equilibrium occurs. Vent plugs (and pressure relief type fittings) also prevent the blowing out of seals during pressure gun greasing. Many drive line universal joints and other applications not having vents require the use of hand guns to prevent seal or bearing damage.
21.14.2 Plain Bearings Bushings used in farm and construction equipment are not sealed as well as antifriction bearings, and rely on grooves in the bearing surface to supply the correct amount of lubricant over the surface of the journal. They are lubricated five to ten times as often as rolling bearings, using flushing action to keep out contaminants. Maintaining a bead of grease at the seals is the usual method of assuring adequate greasing intervals. Oil cups serve the same purpose in some machine designs.
21.14.3 Bearing Lubricants All classifications and formulations of oils and greases are recommended, taking advantages of specific properties or reflecting factory and field experience with available products. The high-torque and shock
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loading nature of this machinery, however, in general demands lubricants of high loadcarrying capacity and capable of operation at temperature extremes and in contaminating atmospheres. To lubricate all bearings on a farm, and particularly on a large construction site, requires knowledge of the general formulations of the lubricants and their physical characteristics. Since equipment is increasingly being lubricated through centralized systems, the pumpability of the lubricant should be questioned; in the case of greases, the ability to show a minimum of ingredient separation is paramount.
21.15 Pneumatic Equipment Compressed air is used extensively in heavy equipment to apply brakes or shift transmissions, operate gates and valves, apply clutches, operate air tools, and rotate rope drums, to mention just a few uses. Air power is as useful in many cases as hydraulic power to control and power implements.
21.15.1 Air Compressors On large farms and on construction sites, the most noticeable machines are mobile, self-contained air plants with the power source and compressor under one protective housing. They employ gasoline or diesel engines as the prime mover, generally connected to the compressor through a flexible coupling. In many designs, the compressor and engine share the same crankcase and oil pan, half of the cylinders being engine and half being compressor. Compressors in the work train of, say, cranes or concrete batching plants are smaller units, and may be belt driven off the main engine or by an electric motor. All compressors are of the positive displacement type, either reciprocating piston, rotary screw, or rotary sliding vane, or are usually limited to a maximum pressure output of 860 kPa (125 psig), with outputs ranging from 0.017 m3 /sec (35 cfm) to 0.85 m3 /sec (1800 cfm). However, there are uses for compressors with pressure outputs up to 70,000 kPa (10,000 psig) and higher. Problems with air compressors are usually due to two maintenance faults: too long a period between oil changes and ineffective air filtration. Because of the heat of compression (and heat of combustion in the case of integral engine compressors) and contaminants from intake air, manufacturer oil change recommendations should be strictly adhered to, with special consideration to high ambient temperatures and dusty atmospheres possibly calling for shorter periods for oil changes and filter maintenance. Because of the variety of designs and operating conditions, a variety of oils are recommended. Generally, engine oils of SAE 20 or 30 will lubricate satisfactorily in most cases, but particular conditions may call for ashless additive turbine type oils or even naphthenic oils of the soft, fluffy carbon deposit nature. Automatic transmission fluids are alternate recommendations in many cases, as are synthetic fluids.
21.15.2 Air Cylinders These units are lubricated and protected from corrosion by oil fed into the air stream by oilers. Lubricants are the lighter viscosity grades of inhibited hydraulic oils, engine oils, or the specially formulated pneumatic tool lubricants.
21.15.3 Air Motors As with hydraulic motors, these are usually of the vane or piston type and can do practically the same jobs and with less maintenance. They receive lubrication by airborne oil mist (vane type) or by built-in splash oilers (piston type). Most air motors are geared or work through gear reduction boxes to reduce their high revolutions per minute to the more usable speeds. Lubricants used are generally the same types as those for compressors. When the gear head is an integral part of the motor, the same oil is used, but many motors operate through separate gear boxes which may call for gear oils.
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21.15.4 Pneumatic Tools Whether hand-held or mounted on a rig, these are complex machines in the variety of machine components used to provide feed, reciprocating motion, and rotation to the working piece. Pneumatic tools employ air power in several ways. Air cylinders are used to feed the tool to the work, as in feed leg drills and stoppers. Air motors, operating through gear trains, provide rotation; through a screw arrangement or gearbox and chain, feed pressure. These applications normally call for lubricants as discussed before. It is the use of percussion force in many drills; however, that demands special lubricants. In these tools, rapid reciprocating motion of the air-driven piston hammer, the splines on the hammer and rifle bar, and the sliding drill rod shanks demand lubricants containing extreme pressure additives. The wide range of operating temperatures encountered and the presence of water, either from moist air or the water flush, call for high oxidation resistance and emulsifying (and demulsifying) capabilities of the oil. A full line of “rock drill oils” for ambient temperatures of −37 to over 43◦ C (−35 to over 110◦ F) might require oils of ISO VG 32 to ISO VG 100. Lubrication is provided by hand oiling, oil reservoirs, airline oilers, or central systems.
21.16 Good Lubrication Practice It falls on farm or construction management, the machine operator, and the fuel and lubricant supplier to see that the best lubrication practice is performed to keep complex equipment operating with a minimum of downtime, and working efficiently and safely.
21.16.1 Management Each work site must be provided with the proper number and sizes of machines and tools to do the job in the time allotted and trained operators selected. Maintenance must be accounted for in the plan, and this requires effective scheduling of fuel, lubricant, filter, and parts inventories to fit the mix of equipment on the job. Proper recordkeeping forms should be provided to control scheduled maintenance and monitor the performance and life of machine components. Lubrication schedules must incorporate the myriad oil and grease specifications and the differing relubrication intervals brought on by the large variety of machine components and configurations. Such a program may of necessity contain compromises.
21.16.2 Operators Trained not only in working his machine to its fullest utility, the operator should also thoroughly know his equipment so that he can spot slight malfunctions and recognize the need for maintenance, adjustment, or repair. Even though he may not do the actual servicing, a good operator will check all fluid levels; drain the water and dirt from oil and fuel filters and strainers; check the condition of air, oil, and water filters; and inspect compartments and lines for leaks before starting the day’s work. During operation, machine gauges and meters and the air restriction indicator should be checked frequently and engine exhaust should be monitored for color. Any highly unusual situations should be reported to the maintenance crew immediately.
21.16.3 Maintenance Preventive maintenance is receiving increasing attention for tighter cost and downtime control and to keep older machines in good shape longer. Lubrication personnel should be thoroughly familiar with the machinery drive trains and armed with the proper charts, should miss no lubrication points or filter maintenance. The operators’ reports should be thoroughly investigated. Many malfunctions may be traceable to improper lubricant or lubrication, lack of maintenance of filters, strainers and breathers, or contamination of a lubricant or fuel.
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Diagnosis of present or potential problems can often be made by examination of filters, magnetic drain plugs, engine exhaust, used lubricants, and the sound of the various components in operation. Engine oil filters may show faulty combustion and wear particles, and to the experienced person, a glycol leak. Other component filters can show excessive oxidation and contamination and wear particles that may be circulating. Diesel fuel filters may point out the presence of water in the fuel storage and the ever-present microorganisms whose waste leaves slime and sludge. Ferrous wear is readily detected by the magnetic drain plugs used in almost all components. A blue casted engine exhaust reveals excessive lubricating oil consumption; a black exhaust shows faulty combustion caused by too rich fuel–air mixtures. In the case of diesel engines, black exhaust can be caused by restricted air intake or faulty injectors. A white exhaust signals water vapor, but in diesel engines may signify raw unburned fuel or a cold smoke condition with low cetane fuel. Used lubricants, of course, readily show by their condition, color, or smell the effects of contamination and high temperature. Many test kits are available for on-the-site analysis of used oils and their judicious use may prevent some future failures. As discussed later, lubricant laboratories are well equipped with the instrumentation and technicians necessary for thorough investigation and analysis of test results. Gear or bearing whine may be a sign of wear; clunking of gears shows the presence of broken teeth. Engine sound changes may indicate fuel or air system problems. Unusual sounds in any compartment can often be traced to improper lubrication practice, either the wrong lubricant or too long a drain or greasing interval for the operating conditions, or faulty filter and strainer service. The maintenance crew, then, should not only be capable of making repairs, but should also be able to diagnose symptoms and correct the problems before catastrophic failures occur.
21.16.4 Supplier Whenever new equipment is put into operation or a new project is started on the farm or building site, the petroleum supplier should be a part of the start-up. The supplier should know the equipment, be able to follow its drivelines through, be able to interpret the lubricant designations of the machine manufacturer, and work with management and maintenance in setting up a consolidated lubricant and lubrication program for the machinery on that job. If the machines of many different manufacturers are represented on a construction site or large farm, there will possibly be a large variety of lubricants recommended (Table 21.2) because different machine makers may call for different types of lubricants for the same application (Table 21.3). Also, the intervals for the services listed in Table 21.4 will vary from one machine maker to another. Any compromises in lubricants or application intervals in the consolidated program deserve the expertise of a fuel and lubricants supplier who knows his products’ physical and chemical properties. What quality of fuel is available; what fuel system or fuel storage additives may be needed? Which multipurpose oils may be used in more than one engine make, more than one transmission make; which multipurpose greases may be used for more than one bearing application? These are just a few questions for which the supplier must have the answers.
21.16.5 Oil Analysis While the appearance and odor of used oils and filtering media may give some qualitative indication to experienced personnel of the operation of engines and other compartments, the use of field kits for more sophisticated analysis is widespread. For over four decades, methods and portable equipment have been available for field testing for water, glycol, and fuel contamination, nature of solid contaminants, and viscosity of the sample. Everyone is familiar with the hot plate test for water, the “blotter” test for solids, dispersancy and oxidation; chemical tests for glycol, microscopes for closer identification; centrifuges for solids; and various equipment for viscometry. While field tests give qualitative answers to general questions, more accurate determinations, quantitative answers and chemical analyses require well-equipped and staffed laboratories. Indeed, the past decade has seen the laboratory become a valuable consulting service
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TABLE 21.2
Farm and Construction Equipment Lubricantsa
Symbol EO
Type Engine Crankcase Oil (diesel and gasoline), described in SAE J183, J300, and J304 SA CA SB CB SC a CC a SD a CD
ATF
HTF
BF
HYDO
Automatic Transmission Fluid, described in SAE J311 Type A Type A, Suffix A Dexron® Type F Hydraulic Transmission Fluid Type C-1 a Type C-2 Brake Fluid, described in SAE J1702 and SAE J1703 SAE J1702 a SAE J1703 (formerly SAE 70R3) Hydraulic Oil MIL-H-5606 Industrial hydraulic oil resistant to rust, oxidation, and foaming industrial hydraulic oil with antiwear additives, resistant, resistant to rust, oxidation, and foaming
Symbol
Type
FRF
Fire Resistant Fluid (hydraulic) Oil/Water Emulsion Water Glycol Fluid Phosphate Ester Type Fluid Regular Type Gear Lubricant, described in SAE J306, SAE J308, and ASTM RR25-D2 (addendum 10/68) Straight Mineral Oil or API Service GL-1 Multipurpose Type Gear Lubricant, described in SAE J306, SAE J308, and ASTM RR25-D2 (addendum 10/68) API Service GL-4 or MIL-L-2105 a API Service GL-5 or MIL-L-2105B API Service GL-6 Open Gear Lubricant Track Roller Lubricanta
RGL
MPL
OGL TRL MPG MPGM WBG HTG SPC
Multipurpose Type Grease, described in SAE J310 Multipurpose Type Grease with Molybdenum Disulfide Wheel Bearing Grease, described in SAE J310 a High Temperature Grease Special Lubricant
a The specifications, classifications, or lubricants marked with an asterisk are found in common use today. It is strongly
recommended that on any single machine a minimum number of lubricants be used. It is further recommended that engine oil, multipurpose type grease, and multipurpose type gear lubricant be used wherever possible. These lubricants may be known by specific trade names or performance specifications. Military standards, MIL-H-2105, MIL-H-5606, MIL-H-21058, are available from U.S. Government, DODSSP, Standardization Documents Order Desk, Building 4D, 700 Robbins Avenue, Philadelphia, PA 19111-5094. Source: Extracted from SAE Recommended Practice J754a, in SAE Handbook, Part 2, Society of Automotive Engineers, Warrendale, PA, 2004, 40.88. With permission.
for the contractor (and possibly the large farmer), whether the consultant is the equipment dealer, the petroleum supplier, or an outside firm specializing in the service. Routine programs of analysis serve three purposes: 1. Detection of the onset and the type of abnormal wear 2. Detection of the rate and type of contamination 3. Determination of practical or reduced drain intervals Such programs call for routine sampling and testing because in all three purposes “rate” is inherent. Since no two pieces of equipment or two units are exactly the same, all machines should be sampled. This is ideal, of course, and some compromises may have to be effected by time and economics. Possibly one hydraulic system, one engine, or one gearcase may have to serve as the basis for all others. This spot approach will indicate if routine maintenance procedures are satisfactory. It may not, however, detect the machine or component in the fleet with an abnormal condition. Samples should be taken from the various systems using, at least initially, the manufacturers’ recommended drain intervals: 100 to 250 h for engines and 250 to 2000 h for hydraulic systems and other components, since these ranges cover the average change interval recommendations. Professional laboratories are better equipped for accurate qualitative and quantitative determinations; wear metals must be determined
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Farm and Construction Equipment TABLE 21.3
21-17
Lubrication of Construction Equipment Components
Component
Lubricants used
Engine crankcase (diesel and gasoline) Diesel fuel injection pump housing Air cleaner, oil bath Clutches and brakes (wet) Hydraulic wheel brake systems Hydraulic control systems Hydraulic transmissions Transmissions Bevel gear and final drive gears Limited slip differentials Gear components (other than above) Open gears Wheel bearings Bearings, shafts, levers, drivelines Track rollers Alternators, generator, electric motor
EO EO EO EO, ATF BF, EO EO, ATF, HTF, HYDO, FRF EO, ATF, HTF EO, RGL, MPL EO, RGL, MPL MPL EO, RGL, MPL, MPG MPL, OGL MPG, MPL, WGB MPG, MPGM, EO EO, TRL, MPL, MPG EO, MPG, HTG
Source: Extracted from SAE Recommended Practice J754a, in SAE Handbook, Volume 3, Society of Automotive Engineers, Warrendale, PA, 2004, 40.88. With permission. 1. Several lubricants may be shown. They should not be mixed. 2. To minimize the number of lubricants used, specify engine oil, multipurpose type grease, and multipurpose type gear lubricant wherever possible. 3. Special lubricants may be required in any of the mentioned components. 4. For specific recommendations consult equipment manufacturer. 5. Maintenance intervals are listed in Table 21.4.
TABLE 21.4
Maintenance Intervals
Interval time in hours 10 50 100 250 500 1000 2000
Equivalent time Each shift Weekly 2 weeks Monthly 3 months 6 months Yearly or whichever occurs first
Note: It is recommended to consult manufacturer and/or the operations & maintenance manual. Source: Extracted from SAE Recommended Practice J753, in SAE Handbook, Volume 3, Society of Automotive Engineers, Warrendale, PA, 2004, 40.78. With permission.
in them because of the sophisticated spectrographic equipment needed. With data on the metallurgical design of the various engines and other systems, laboratories can more readily detect sources of wear and contamination, analyze the rates of generation, and provide consultation on maintenance and repairs needed to prevent catastrophic failures. “Catastrophic” is a widely used term or adjective to denote sudden, violent, widespread damage as opposed to normal wear over a period of time. Oil analysis, consultation on maintenance, and repairs are to prevent catastrophies. Best lubrication practice, then, keeps machines running with a minimum of downtime for maintenance and repairs, for the longest possible life. Farm and construction management, their maintenance
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personnel, and the petroleum supplier all have vital roles in this effort, and they must be cooperative if the production intended for man’s machines is to be reached and improved.
Acknowledgment We acknowledge the initial work by William J. Hanley in the previous volume.
References [1] SAE Recommended Practice J183 Engine Oil Performance and Engine Service Classification (other than "energy conserving"), in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.2. [2] SAE Recommended Practice J300 Engine Oil Viscosity Classification, in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.33. [3] SAE Recommended Practice J304 Engine Oil Tests, in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.36. [4] SAE Recommended Practice J311 Fluid for Passenger Car Type Automatic Transmissions, in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.62. [5] SAE Recommended Practice J1702 Self-propelled Sweepers Sweep-Ability Performance, in SAE Handbook, Volume 3, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 40.23. [6] SAE Recommended Practice J1703 Motor Vehicle Brake Fluid, in SAE Handbook, Volume 2, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.54. [7] SAE Recommended Practice J306 Automotive Gear Lubricant Viscosity, in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.55. [8] SAE Recommended Practice J308 Axle and Manual Transmission Lubricants, in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.52. [9] SAE Recommended Practice J310 Automotive Lubricating Greases, in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.56. [10] SAE Recommended Practice J754a Lubricant Types — Construction and Industrial Machinery, in SAE Handbook, Volume 3, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 40.88. [11] SAE Recommended Practice J753 Maintenance Interval Chart, in SAE Handbook, Volume 3, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 40.78. [12] Imperial Oil, Product Information: Lubricants and Specialties, 10th ed., 1999. [13] http://www.nass.usda.gov/census/. Accessed on 14 March 2005. [14] http://www.census.gov/mcd/. Accessed on 14 March 2005.
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22 Industrial Lubrication Practice — Wheel/Rail Tribology 22.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 22.2 Wheel/Rail Contact Mechanics . . . . . . . . . . . . . . . . . . . . . . .
22-1 22-2
Contact Position • Friction and Creep • Contact Stress • Wheel and Rail Profiles
22.3 Wheel and Rail Materials . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 22.4 Wheel and Rail Damage Mechanisms . . . . . . . . . . . . . . . .
22-7 22-8
Wear and Plastic Deformation • Wear Transitions • Wear Mapping • Rolling Contact Fatigue • Interaction of Wear and Fatigue • Modeling Damage Mechanisms
Roger Lewis and Rob Dwyer-Joyce Department of Mechanical Engineering, The University of Sheffield
22.5 Friction Modification . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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Increasing Friction • Reducing Friction
22.6 Maintenance of Wheels and Rails . . . . . . . . . . . . . . . . . . . . 22.7 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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22.1 Introduction The wheel/rail contact is a crucial component in the successful operation of railways. A large variety of loading conditions and contact geometries exist due to the many different rail and wheel profiles, rail cant and curve radii, and railway vehicles running on a network. Contact conditions vary considerably between the two main contact areas: wheel tread/railhead and wheel flange/rail gauge corner, but are usually more severe in the latter, where greater wear and fatigue cracking is seen to occur. Friction and creepage in the contact are also highly variable. Natural lubricants such as humidity, precipitation, and leaves can negatively influence the friction in the wheel/rail contact, causing braking problems and wheel slip in traction. These problems can be overcome by using applied lubricants to reduce wear in curves and friction modifiers to increase adhesion. This, however, further adds to the complexity of the wheel/rail contact system. Effective management of the wheel/rail contact is an important aspect of rail infrastructure operations. Rail maintenance was estimated to have cost 300 million Euro within the European Union in 1995 (Cannon, 1996). All the influencing factors have to be taken into account as they interact closely (as indicated in Figure 22.1) (adapted from Kalousek and Magel, 1997). For example, measures used
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Handbook of Lubrication and Tribology
Rail and wheel life up
Wear
Plastic flow
Damage modes Rail rollover
RCF
Hollow wheels
Profit up
Friction management
Thermal cracks
Wheelset dynamics
Maintenance costs down
Wheel/rail material
Contact mechanics Spending down
FIGURE 22.1 Systems approach to wheel/rail interface management and research. (Adapted from Kalousek, J. and Magel, E., 1997, Railway and Track Structure, January.)
to reduce wear, such as lubrication, can influence fatigue and adhesion, and the measures used to increase adhesion, such as sanding, can have a detrimental effect on wear. A fine balance has to be found in determining maintenance schedules and lubrication regimes to keep railway networks running smoothly. This is becoming increasingly difficult as new specifications on wear and reliability are being imposed to increase the time between reprofiling, increase safety, and reduce total life-cycle costs. In parallel with these requirements, vehicle missions are changing due to the need to operate rolling stock on track with low radius curves, as well as the high radius curves found on high speed lines and increasing speeds. These are leading to an increase in the severity of the wheel/rail contact conditions (Stanca et al., 2001).
22.2 Wheel/Rail Contact Mechanics 22.2.1 Contact Position The position of the wheel/rail contact, which is typically 1 cm2 in size, varies continuously as a train progresses down a section of track. The exact position will depend on the wheel and rail profiles and the degree of curvature of the track and whether the wheel is the leading or trailing wheelset on a bogie, as well as other factors determined by the bogie design. In straight track, however, it is likely the wheel tread and railhead will be in contact with wheel flange and rail gauge corner contact occurring in curved track. Figure 22.2 shows how the contact position and stress varies for the two wheels on a leading wheelset entering a right-hand curve. Three possible regions of wheel/rail contact have been defined (as shown in Figure 22.3) (Tournay, 2001): 1. Region A — Wheel tread/railhead. The wheel/rail contact is made most often in this region and usually occurs as the railway vehicle is running on straight track or very high radius curves. This region yields the lowest contact stresses and lateral forces. 2. Region B — Wheel flange/rail gauge corner. The contact in this region is much smaller than that in region A and is often much more severe. Typically contact stresses and wear rates are much higher. If high wear and material flow occurs, two point contacts may evolve, where tread and flange contact is apparent.
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Contact stress
FIGURE 22.2
Leading wheelset entering a right-hand curve.
Region C
Region A
Region B
FIGURE 22.3 Wheel/rail contact zones. (Adapted from Tournay, H., 2001, Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA.)
3. Region C — Contact between field sides of wheel and rail. Contact is least likely to occur here and if it does, high contact stresses are induced, which will lead to undesirable wear features causing incorrect steering of the wheelset.
22.2.2 Friction and Creep The friction between the wheels and rail is extremely important as it plays a major role in the wheel/rail interface process such as adhesion, wear, rolling contact fatigue, and noise generation. Effective control of friction is managed through the application of friction modifiers to the wheel/rail contact. The aim of friction management is to maintain friction levels in the wheel/rail contact to give: low friction in the wheel flange/rail gauge corner contact; intermediate friction wheel tread/railhead contact (especially for freight trucks); and high friction at the wheel tread/rail top contact for locomotives (especially where adhesion loss problems occur) (Zakharov, 2001). Ideal friction conditions in these contact regions for high and low rails are shown in Figure 22.4 (Roney, 2001; Sinclair, 2004). The wheel/rail contact occurs in the incipient sliding regime, hence there is not complete adhesion within the contact, and stick and slip zones are apparent, as shown in Figure 22.5. Three types of slip (or creepage) occur in the wheel/rail contact; lateral (perpendicular to the direction of wheel motion), longitudinal (in the direction of wheel motion), and spin creepages and these are generated as the wheel deviates from a pure rolling motion. These vary in the same way as contact stresses change as the wheel/rail contact position moves and they are more severe (particularly lateral creepage) in curves. The degree of creepage in the contact depends on the normal load and friction in the contact, as shown in Figure 22.5, which illustrates a creep curve for a wheel/rail contact. The greater the traction force, the larger the slip region in the contact. This curve varies dramatically with the introduction of a third body layer to the contact, such as a lubricant (natural or applied) or a friction modifier.
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Handbook of Lubrication and Tribology 0.25 < m < 0.4 0.25 < m < 0.4 m < 0.1
Low rail
High rail
FIGURE 22.4 Ideal friction coefficients in the wheel/rail contact. (Adapted from Sinclair, J., 2004, Vehicle Track Interaction: Identifying and Implementing Solutions, IMechE Seminar, February 17th.)
Creep = 0.01 to 0.02
Rolling direction
Tractive force (= mN)
Slip
Tractive forces Slip
Stick
Stick Stick
Slip
Slip Creep
FIGURE 22.5
Relationship between traction and creep in the wheel/rail contact.
22.2.3 Contact Stress As a result of the fact that the contact position is not spread evenly over the entire wheel or rail profile, the shape of the profiles will change as time progresses, due to wear and material flow (processes largely controlled by the loads and creepage within the contact). In order to be able to predict how profiles may evolve, a good understanding of the contact stress is therefore required. The simplest solution for determining wheel/rail contact geometry and stress is Hertz analysis (see Johnson [1985]), where the wheel and rail can be equated to two cylinders in contact perpendicular to each other. The maximum contact pressure, p, is given by: p=
3
3PE 2 2π 3 R 2 (1 − ν 2 )2
(22.1)
where P is the normal load, E and ν are the Young’s modulus and Poisson’s ratios respectively (assumed to be same for wheel and rail materials in this case), and R is the equivalent radius given by: 1 1 1 = + R R1 R2
(22.2)
where R1 and R2 are the contact radii of the wheel and rail. This approach is, however, limited in accuracy due to the assumptions made in the analysis, such as smooth contacting surfaces, a linear elastic material response and that the contact dimensions must be small compared to the radii of curvature of
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FEM y
CONTACT
Hertz
(Case 1, N = 80377 N) x
PMAX = 1500 MPa
PMAX = 3457 MPa
Dx(max) = 20 mm Dy(max) =17 mm
Area = 134.8 mm2
PMAX = 3601 MPa
Dx(max) =18.1 mm Dy(max) =11.3146 mm
Area(sum) = 46.7181 mm2
Dx = 19.6 mm Dy = 12 mm
Area = 45 mm2
(Case 2, N = 80377 N) PMAX = 665 MPa Area = 172.8 mm2
PMAX = 865 MPa Area = 134.21 mm2
PMAX = 1080.53 MPa Area = 111.58 mm2 x y
Dx = 12.53 mm Dy = 18.58 mm
Dx = 10.12 mm Dy = 15.45 mm
Dx = 10.34 mm Dy = 13.74 mm
FIGURE 22.6 Comparison between FE, CONTACT, and Hertz analysis. (Adapted from Telliskivi T. and Olofsson U., 2001, Proceedings of the IMechE Part F, Journal of Rail and Rapid Transit, Vol. 215, pp. 65–72.)
the contacting bodies. In the flange contact particularly, the contact radius can be as small as 10 mm, which means this assumption can be invalid. Numerical solvers have been developed for calculating area and stress for contacts approximated to Hertzian ellipses, such as FASTSIM (Kalker, 1982) and non-Hertzian contacts, such as CONTACT (Kalker, 1990). CONTACT, however, requires a large amount of computing resource and again is limited by the half-space assumption. Finite element modeling carried out by Telliskivi and Olofsson (2001), including the plastic deformation and actual wheel and rail profiles showed good correlation in terms of contact area and stress with Hertz and CONTACT for the railhead contact, but considerably different stress and area for the rail gauge corner flange contact (as shown in Figure 22.6, where Case 1 represents the gauge corner contact and Case 2 represents the head contact), due to the limiting half-space assumption in Hertz and CONTACT analysis. Recently an innovative ultrasonic technique has been used to study the wheel/rail contact (Marshall et al., 2004). Figure 22.7 shows contact pressures derived from ultrasonic scans compared with numerical calculations using actual roughness profiles and Hertz analysis. There is good global geometric correlation between the ultrasonic results and the numerical model. The degree and fragmentation of the ultrasonic and numerical contacts are qualitatively similar. But on a local
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Handbook of Lubrication and Tribology
(a)
(b)
2200 2000 1800
1400 1200 (c)
(d)
1000 800
Contact pressure (MPa)
1600
600 400 200 0
FIGURE 22.7 Contact pressure maps for a load of: 80 kN (a) ultrasonic measurement; (b) Hertzian; (c) elastic model; (d) elastic–plastic model. (Adapted from Marshall, M.B., Lewis, R., and Dwyer-Joyce, R.S., 2004, Ultrasonic characterisation of a wheel/rail contact, Proceedings of the 30th Leeds-Lyon Symposium on Tribology, Elsevier Triboloby Series No. 43, pp. 151–158.)
level the ultrasonic results and numerical solutions differ. This is likely due to difficulty in aligning the surfaces to the same orientation in both the experiment and model. The elastic case assumes no localized yielding at the contact; this results in predicted contact pressures in excess of yield for the contacting surfaces. The pressures determined for the elastic case are far in excess of those measured ultrasonically. However, the elastic–plastic case shows similar peak pressures to the experiment. The experimental and elastic–plastic numerical model peak pressures are in excess of the Hertz solution, this is due to the reduced contact conformity attributable to roughening. Hertzian theory dictates that in static loading the maximum compressive stress is at the surface and the maximum shear stress is below the surface (at a depth of 0.78a, where a is the contact half width). When tractive force is applied at the surface the shear stress increases and the position of the maximum value moves closer to the surface. Because of the rolling/sliding behavior of a wheel on a rail, a cyclic build-up of plastic deformation occurs beneath the material surfaces. It is this behavior that leads to rolling contact fatigue and wear occurring. Figure 22.8 shows a shakedown map, which illustrates the relationship between friction in the wheel/rail contact and the load carrying capacity of the contact. It shows the limits of material behavior in terms of nondimensional contact pressure, p0 /k as a function of friction coefficient, µ (=T /N ), where p0 is the normal contact pressure, k is the shear yield strength, T is the tractive force, and N is the normal load. At relatively low friction coefficients, cumulative plastic flow occurs subsurface. For friction coefficients above about 0.3, plastic flow is greatest on the surface. The worst position in terms of damage to the material is in the ratchetting region, where strain is accumulated until the ductility of the material is exceeded and it is lost as wear debris or a crack is initiated.
22.2.4 Wheel and Rail Profiles Wheel and rail profiles are designed to optimize their performance for their given application. Performance is generally assessed in terms of (Zakharov, 2001): resistance to damage mechanisms (wear, fatigue, etc.), minimization of noise, and maximization of vehicle stability.
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6 Alternating plasticity (plastic shakedown)
5
Load factor, p0 / k
4
Incremental growth (ratcheting)
Elastic shakedown 3
2 Elastic 1 Subsurface
Surface
0.2
0.4
0.6
Friction coefficient, m
FIGURE 22.8
Shakedown map. TABLE 22.1 Chemical Composition (weight %) of Rail Steels Elements C Mn Si S P Cr V Ni Mo
USA, Canada, Brazil
Australia
Europe (UIC60)
0.72–0.82 0.80–1.10 0.10–0.60 0.037 max. 0.035 max. 0.25–0.50 0.03 max. 0.25 max. 0.10 max.
0.72–0.82 0.80–1.25 0.15–0.58 0.025 max. 0.025 max. — — — —
0.60–0.82 0.80–1.30 0.30–0.90 0.025 max. 0.025 max. 0.80–1.30 — — —
Source: Adapted from Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA.
In terms of contact stress, this means that stresses need to be kept low and contact points need to be spread across the wheel tread and railhead to reduce concentrated wear occurring. This, however, is difficult to achieve. A piece of track will see many different wheel profiles in a given length of time, which will result in many different contact positions and stresses. The main aim though is to design the profiles such that regrinding intervals are as long as possible, to reduce costs.
22.3 Wheel and Rail Materials Rail and wheel steels are metallurgically very similar. Both utilize high carbon (0.65 to 0.82%) and have a pearlitic microstructure. Some typical chemical compositions of rail steels from around the world are shown in Table 22.1. There are many different specifications of wheel steels, the main difference between them being carbon content. Choice varies according to final use, which may be freight, passenger, or locomotive. Some typical examples are shown in Table 22.2.
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Handbook of Lubrication and Tribology TABLE 22.2 Steels Elements C Mn Si S P Cu V Ni+Cr+Mo
Chemical Composition (weight %) of Wheel
USA, Canada, Brazil (Class C)
Australia
Europe (R7T)
0.67–0.77 0.60–0.85 0.15 max. 0.05 max. 0.05 max. — — —
0.67–0.77 0.60–1.00 0.15 max. 0.035 max. 0.04 max. — — —
0.51 0.77 0.35 0.009 0.009 0.14 0.001 0.20
Source: Adapted from Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA.
TABLE 22.3
Mechanical Properties of High Strength Rail Steels
Property Yield strength (MPa) (min.) Tensile strength (MPa) (min.) Elongation % (min.) Brinell surface hardness
USA, Canada, Brazil
Europe (UIC60)
758 1172 10 340–390
640 1080 9 320–360
Source: Adapted from Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA.
The mechanical properties of rail steels of most interest are: yield strength, which dictates the plastic flow and work hardening characteristics of the material; tensile and fatigue strength, which is an indication of the material’s resistance to fatigue and hardness, which can indicate how the material will resist wear. Typical rail mechanical properties are shown in Table 22.3. While the majority of rail steel is currently pearlitic, bainitic rail steels have been developed to try and improve damage resistance of track (Kalousek et al., 1985; Ueda et al., 1997; Singh et al., 2001). Whereas these offer greater wear resistance than pearlitic steels, their resistance to rolling contact fatigue is lower (Jiang, 1999; Yokoyama, 2002; Sawley, 2003). Work is, however, continuing on their development with a view to overcoming this problem.
22.4 Wheel and Rail Damage Mechanisms A number of damage mechanisms exist for both rail and wheels. The most significant are wear, plastic deformation, and rolling contact fatigue (RCF). These individually can cause problems, but due to their close interaction, measures introduced to reduce one may increase another. In this section, each mechanism is looked at in detail and then the interaction of wear and RCF is outlined.
22.4.1 Wear and Plastic Deformation A number of different techniques have been used for studying wear of railway wheel and rail steels. Field measurements have been used in the past to study the causes of wheel and rail wear (Dearden, 1960). A large amount of data has also been gathered from simulated field experiments carried out on specially built test tracks (Steele, 1982). Laboratory methods used range from full-scale laboratory experiments
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(McEwen and Harvey, 1985) and scaled-down tests (Kumar, 1984) to bench tests using a twin disc setup (Beagley, 1976; Bolton et al., 1982; Bolton and Clayton, 1984; Krause, 1986; Garnham, 1992; and DwyerJoyce, 2004a). The twin disc approach has been used more than most because it offers greater control over experimental variables as well as the ability to test a wide range of materials at lower cost. Twin disc testing carried out to study the wear behavior of railway wheel and rail steels has led to the identification of a number of wear regimes (Beagley, 1976; Bolton et al., 1982; Kumar et al., 1984; Bolton and Clayton, 1984; Lewis and Dwyer-Joyce, 2004a). Early tests demonstrated that two wear regimes existed (Beagley, 1976; Bolton et al., 1982). These were designated mild and severe. Later work led to the identification of a further regime, designated catastrophic wear (Bolton and Clayton, 1984; Lewis et al., 2004). Figure 22.9 shows results of twin disc testing of R8T wheel steel against UIC60 900A rail steel (data from Lewis and Dwyer-Joyce, 2004a). The results are plotted in terms of wear rate (µg material lost/ m rolled/mm2 contact area) against an index based on work done in the contact, T γ /A, where T is tractive force (normal load divided by friction coefficient), γ is slip, and A is contact area. This is a useful way to plot the data as it allows comparison of different test geometries. The curve is typical of those for rail steels (Lewis and Olofsson, 2004), as shown in Figure 22.10. As can be seen wear rates are gradually reducing with time as new and more wear resistant rail materials have been introduced. Wheel tread and railhead wear are thought to fall within the mild regime and wheel flange and rail gauge corner wear in the severe or catastrophic regime. This has been verified with field measurements carried out by Olofsson and Nilsson (2002). As a result of twin disc tests (Lewis and Dwyer-Joyce, 2004a), the regimes have been characterized in terms of wear rate and wear mechanism. At low T γ /A, in the mild wear regime, oxidative wear occurs on both wheel and rail discs. The disc surfaces turned a rusty brown color. Closer examination of the wear surfaces revealed abrasive score marks and evidence of the oxide layer breaking away from the surface (see Figure 22.11[a]). This ties in with observations made in the field that on straight track where low slip occurs on the high rail, oxidative wear is prevalent generating magnetite (Fe3 O4 ) (Broster et al., 1974) and in full-scale test-rig results, where reddish oxide film appeared for low slip conditions (McEwen and Harvey, 1985). As T γ /A was increased, the wear mechanism of the wheel discs altered. The wheel disc appeared to be wearing by a ratcheting process (deformation followed by crack growth and subsequent material removal). Closer examination of the wheel disc surfaces revealed that this was the case (see Figure 22.11[b]).
3,000
Wear rate (mg / m / mm2)
2,500 2,000 Catastrophic
1,500 1,000 Mild
Severe 500 0 0
20
40 Tg /A
60
80
100
120
(N / mm2)
FIGURE 22.9 Wear rates and regimes for R8T wheel steel tested against UIC60 900A rail steel. (Adapted from Lewis, R. and Dwyer-Joyce, R.S., 2004b, Wear mechanisms and transitions in railway wheel steels, Proceedings of the IMechE, Part J: Journal of Engineering Tribology, Vol. 218, pp. 467–478.)
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Handbook of Lubrication and Tribology
Wear rate (mg / m / mm2)
100,000
BS11 vs. Class D (Bolton and Clayton, 1984)
10,000 1,000
Standard carbon rail (Danks, 1987)
100 10
UIC60 900A vs R8T (Lewis and Olofsson, 2004)
1 0.1 0.01 0
50
100
150
200
Tg /A (N / mm2)
FIGURE 22.10
Rail steel wear rates.
Region of delamination of oxide film (a)
Abrasive score marks (b)
FIGURE 22.11 Wheel disc surface run at (a) T γ /A = 0.21 and (b) T γ /A = 4.1. (Adapted from Lewis, R. and Dwyer-Joyce, R.S., 2004b, Wear mechanisms and transitions in railway wheel steels, Proceedings of the IMechE, Part J, Journal of Engineering Tribology, Vol. 218, pp. 467–478.)
Figure 22.12[a] shows a section through a wheel disc, run at low T γ /A, parallel to the rolling direction. At the surface the oxide layer is just visible. There is a very small amount of deformation just below the wear surface of the disc. At higher levels of T γ /A, observation of the subsurface morphologies revealed that a larger amount of plastic deformation was occurring below the wheel disc wear surface (see Figure 22.12[b]) and crack formation just below the surface was visible, which was leading to thin slivers of material breaking away from the surface. The slivers have a similar thickness to the oxide layer and could indicate a severe oxidative wear mechanism occurring, where fracture occurs between the oxide layer and the metal. As T γ /A was increased further far greater cracking was visible at the wear surface and some of these cracks were seen to alter direction from running parallel to the wear surface and turning up to turning down into the material causing larger chunks of material to break away (see Figure 22.12[c]).
22.4.2 Wear Transitions While rail steel wear regimes have been defined well in terms of wear rate, metallographic features, and wear debris, there was not a great understanding of what mechanisms are leading to the changes in wear rate that occur (see Figure 22.9). In order to further understand the wear mechanisms, the transitions
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Industrial Lubrication Practice
22-11
(a) Oxide layer
Extent of subsurface deformation
20 mm
Rolling direction
(b)
40 mm Crack formation (c)
40 mm
More cracking and larger chunks of material breaking away
FIGURE 22.12 Sections parallel to rolling direction through wheel disc run at (a) T γ /A = 0.21; (b) T γ /A = 16.6; and (c) T γ /A = 28.3. (Adapted from Lewis, R. and Dwyer-Joyce, R.S., 2004b, Wear mechanisms and transitions in railway wheel steels, Proceedings of the IMechE, Part J: Journal of Engineering Tribology, Vol. 218, pp. 467–478.)
between these regimes were studied in more detail (Lewis and Dwyer-Joyce, 2004a). It was proposed that the first transition is associated with the onset of fully sliding contact conditions and the second is a result of surface temperature effects. Figure 22.13 shows friction measurements taken during UIC60 900A vs. R8T wear tests carried out at 1500 MPa plotted against slip, that is, a creep curve. As would be expected the friction reaches a threshold. This transition represents the change from partial slip in the disc interface to full slip conditions. Also shown is the Carter creep curve for an assumed limiting friction of 0.5. This model creep curve is based on smooth elastic cylinders in contact (Carter, 1926). The wear data (also shown in Figure 22.13) follows a similar pattern indicating that at the point of transition from partial slip to full slip a wear transition also occurs. After the full slip condition has been reached, increasing the magnitude of slip has no effect. Calculations were carried out to determine temperatures at a twin disc contact for the UIC60 900A rail steel vs. R8T wheel steel (Lewis and Dwyer-Joyce, 2004a; Lewis and Olofsson, 2004). The results, shown in Figure 22.14, indicate that the transition from severe to catastrophic wear occurs around 200◦ C. These temperatures correspond with those causing a drop in the yield strength of carbon manganese steels similar to rail steels (British Steel Makers Creep Committee, 1973).
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Handbook of Lubrication and Tribology 10,000
1
Wear rate (mg / m / mm2)
Carter creep curve 0.8
1,000
0.6 m
100 0.4 10
Wear rate
0.2
Measured m 1
0 0
5
10
15
20
Slip (%)
FIGURE 22.13 Friction vs. slip in the twin disc contact (for tests carried out at 1500 MPa). (Adapted from Lewis, R. and Dwyer-Joyce, R.S., 2004b, Wear mechanisms and transitions in railway wheel steels, Proceedings of the IMechE, Part J: Journal of Engineering Tribology, Vol. 218, pp. 467–478.) 400
3500 3000
300
Total max. temp.
2500
250 2000 200 1500
150 Total avg. temp.
100 50
1000 Wear transition
Wear rate
0 0
20
40
60
80
100
Wear rate (mg/m/mm2)
Temperature (8C)
350
500 0 120
Tg/A (N/mm2)
FIGURE 22.14 material.
Twin disc contact temperatures and wear coefficients for UIC60 900A rail material vs. R8T wheel
22.4.3 Wear Mapping Although using the T γ /A method for plotting wear rate data enables wear transitions to be identified easily and comparisons of different material combinations to be made, it does not help in fully understanding how the individual contributions of different parameters such as contact pressure and slip affect wear rate. Lewis and Olofsson (2004) proposed a wear mapping method that would give a more complete analysis of the effect of individual parameters. It is based on the technique developed by Lim and Ashby (1987) for mapping sliding wear mechanisms. Wear coefficients were calculated from the rail steel wear data using Archard’s equation (Archard, 1953): Vh K = (22.3) Ns where K is the wear coefficient, V is the wear volume, N is the normal load, s is the sliding distance, and h is the material hardness. Wear coefficients were then plotted against contact pressure and sliding speed in the contact. Two types of plots were constructed; contour maps and 3D point graphs (see Figure 22.15). Obviously, the accuracy of the contour map is limited by the amount of data available. The accompanying 3D graphs
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Industrial Lubrication Practice Severe
40
1000 900 800
hic transition
1100
)
–4
Catastrophic
Severe – catast rop
Contact pressure (MPa)
1300 1200
35
30.63 -- 35.00 26.25 -- 30.63 21.88 -- 26.25 17.50 -- 21.88 13.13 -- 17.50 8.750 -- 13.13 4.375 -- 8.750 0 -- 4.375
Wear coefficient (×10
Mild 1400
22-13
30 25 20 15 10 5
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
0.8 0.6 1200
Con
tac
700 0.0
1.0
0 1400 0.4 1000
t pr
ess
ure
800 (MP a)
0.2 0.0 600
Sliding velocity (m/sec)
c)
se
m/
y(
it loc
e gv
din
Sli
FIGURE 22.15 Wear coefficient maps for UIC60 900A rail material vs. R7 wheel material data from Olofsson and Telliskivi (2003). (Adapted from Lewis, R. and Olofsson, U., 2004, Mapping rail wear regimes and transitions, Wear, Vol. 257, No. 7–9, pp. 721–729.) Railhead/wheel tread
Rail gauge / wheel flange
Contact pressure (MPa)
2500
2000
Severe — catastrophic transition
1500
UIC60 900A vs. R7 wear map
Mild Catastrophic
1000
500
Severe
0 0.0
0.2
0.4
0.6
0.8
1.0
Sliding velocity (m/sec)
FIGURE 22.16 UIC60 900A rail steel wear map plotted over wheel/rail contact conditions derived from GENSYS simulations. (Adapted from Lewis, R. and Olofsson, U., 2004, Mapping rail wear regimes and transitions, Wear, Vol. 257, No. 7–9, pp. 721–729.)
give an indication of where data is lacking on a particular map. Wear transitions were marked on the contour plots. To study how the wear regimes identified above fit in with the wheel/rail contact conditions shown in Figure 22.15 the wear map of UIC60 900A rail steel vs. R7 wheel steel has been overlaid, as shown in Figure 22.16 (Lewis and Olofsson, 2004). This indicates that the railhead/wheel tread contact will experience mild to severe wear and the rail gauge/wheel flange contact will experience severe to catastrophic wear. This backs up previous suppositions regarding the wear regimes into which the railhead/wheel tread and rail gauge/wheel flange contacts fall.
22.4.4 Rolling Contact Fatigue Fatigue failures for rails and wheels can be categorized into surface and subsurface failures. Surface fatigue in rails can lead to head checking or squat formation and subsurface fatigue can result in shelling or the formation of tache ovale.
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Handbook of Lubrication and Tribology
Crack initiation and propagation by ratcheting
Crack driven by bending and residual stresses
Contact stresses dominate
(da/dn)
Crack driven by ratcheting
Crack driven by contact stress field Bending stress dominates crack propagation
Crack length
FIGURE 22.17 Phases of crack growth in rail. (Adapted from Kapoor, A., Fletcher, D.I., and Franklin, F.J., 2003, Proceedings of the 29th Leeds-Lyon Symposium on Tribology, pp. 331–340.)
Motion
Fluid filled crack
FIGURE 22.18
Crack opening driven by fluid pressurization.
Head checks (shallow cracks at the rail surface) normally occur on the rail gauge corner on curves. They result from accumulation of plastic strain increments (ratcheting), which eventually exhaust the ductility of the surface material, at which point cracks can initiate. The critical contact conditions for this to occur are high load and friction (see ratcheting region in the shakedown map shown in Figure 22.8). Plotting surface crack growth rates (da/dn) against crack length reveals that there are a number of phases present as shown in Figure 22.17 (Kapoor et al., 2003). After initiation, crack growth is driven by ratcheting in the plastically deformed layer. As the crack becomes longer and deeper crack growth is driven by the stress field due to the repeated contact loading. Finally the crack turns downward and growth is driven by bending stresses in the rail. If the crack reaches a critical crack length at this stage fast fracture can occur resulting in a rail break. Water and lubricants trapped in a crack increase the speed of propagation (Bower, 1988; Bogdanski et al., 1996) as shown in Figure 22.18. This is because liquids are trapped in the rail cracks as the wheel passes over causing pressurization, which increases the crack growth rate. Work has been undertaken to study the effect of rail surface indentations (Gao, 2000) and contaminants such as ballast dust and sand applied to increase adhesion in the wheel/rail contact (Grieve et al., 2002; Lewis, 2004) on rail fatigue life. Tests were carried out under various states of lubrication. These indicated that in dry or water lubricated conditions surface indentations had negligible effect on the fatigue life and that the dents were removed by plastic flow of surrounding material. For oil lubricated conditions, however, surface damage acted as a fatigue initiation site (Gao, 2000). For the contacts contaminated with ballast dust, wear rates were enhanced. The particles were seen to embed in the softer wheel material and abrade the harder rail surface (Grieve et al., 2002). Tests with sand showed very high wear rates. In addition to abrading the surfaces the sand initiated a low cycle fatigue process, which resulted in the removal of fatigue spalls from the material surface (Lewis and Dwyer-Joyce, 2004b).
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Industrial Lubrication Practice (a)
22-15 (b) Fatigue crack Point of propagation initiation
Initial wheel surface
Fatigue crack
Final fracture
FIGURE 22.19
RCF damage: (a) surface fatigue; (b) crack morphology for a subsurface induced fatigue failure.
Squats occur on straight track on the surface of the railhead. They appear as darkened areas on the rail. They consist of two cracks, one in the direction of travel and the other in the opposite direction, which is much longer. They can initiate as a result of ratcheting and fluid pressurization and also from white etching layers (WELs). WELs result from modification of the microstructure of the rail surface material from pearlite to martensite (Pyzalla et al., 2001). They not only have been noted to have a hardness of up to 1200 Hv (Feller, 1991), but are also extremely brittle and crack initiation is therefore more likely to occur within a WEL. WELs normally occur as a result of high temperatures, which may result from wheel skids, for example. Shelling in rails occurs at the rail gauge corner in curves and is a subsurface initiated defect (Grassie, 1997). Elliptical shell-like cracks propagate parallel to the rail surface and in many cases cause material to spall away. Tache ovale are defects which develop about 10 to 15 mm below the surface of the railhead from cavities caused by the presence of hydrogen (Grassie, 1997). The cavities may be present in the parent rail material or can arise during welding if it is carried out poorly. If the crack becomes sufficiently large, transverse fracture of the rail can occur. In surface initiated fatigue of railway wheels, fatigue cracks result from excessive plastic flow of the surface material. This will cause crack initiation due to ratcheting or low cycle fatigue of the surface material on both. Once initiated, the cracks typically grow into the wheel material to a maximum depth of 5 mm. Final fracture occurs as the cracks branch toward the wheel tread. The typical appearance of surface initiated fatigue failure is shown in Figure 22.19(a). Surface initiated cracks are normally not a safety issue. However, they are the most common type of fatigue damage in wheels. They are costly in requiring reprofiling of the wheel and causing unplanned maintenance. In the case of subsurface fatigue, cracks will initiate several millimeters below the surface. They continue to grow under the surface and final fracture will normally occur due to branching toward the surface. Such a failure will lead to a large piece of the wheel surface breaking away, as shown in Figure 22.19(b).
22.4.5 Interaction of Wear and Fatigue If the effect of wear on crack propagation is considered (as shown in Figure 22.20[a]), the crack is truncated as more material is removed. If crack truncation rate is laid over the crack growth rates the net crack growth rate can be determined. In Figure 22.20(b), two different wear rates are considered. If the wear rate (and hence crack truncation rate) is high, it is likely that most cracks will be worn away before progressing beyond Stage A (cracks driven by ratcheting). If, however, the wear rate is low, cracks will initiate and progress along Stage A. They may stabilize at Point 1, where growth rate equals truncation rate. These curves will vary considerably with contact conditions and as such the crack may be carried from Point 1 to Point 2 and then continue propagating. The crack reaches a length at which the growth
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Handbook of Lubrication and Tribology (a)
Material removed by
Truncation of crack = depth of material
Crack at angle u
(b)
(c) Life line due to fatigue
Crack truncation due to high wear rate
A
C Life
(da/dn)
B
1 2
3 Crack truncation due to low wear rate Crack length
Life line due to wear Fatigue failure unsafe
Failure by wear safe
Material removal rate (by grinding or wear)
FIGURE 22.20 Interaction of wear and fatigue: (a) crack truncation by wear; (b) crack growth rate vs. crack length; (c) rail life vs. material removal rate.
rate declines and it stabilizes at Point 3. If the wear rate drops below the intersection of Stages B and C, the crack can move to Stage C, which may lead to a dangerous conclusion. Figure 22.20(c) shows the effect of material removal rate on rail life. This may be reduced by using harder rail or by lubrication. Grinding rail increases the removal rate. Clearly the aim should be to have a removal rate at the optimum point, where the wear and fatigue lines cross. This, however, will be very difficult because of variations in the wheel/rail contact conditions, but with good interface management systems in place a good life can be attained.
22.4.6 Modeling Damage Mechanisms A number of design tools have been created to predict damage mechanisms in the wheel/rail contact (Fries and Dávila, 1987; Pearce and Sherratt, 1991; Zobory, 1997; Jendel, 2000a; Lewis et al., 2004). Typically wheel profile evolution prediction tools incorporate dynamics modeling to predict global contact parameters to assess contact position, load, and slip; local contact modeling based on, for example, FASTSIM or CONTACT, to calculate slip and load distributions within the contact and a semiempirical wear model based on coefficients derived from rolling/sliding or sliding wear tests (as shown in Figure 22.21). Examples of these tools include the wheel durability predictor based on ADAMSRail multibody dynamics software, developed during the EC funded HIPERWheel project (Lewis et al., 2004) and the wheel wear assessment code developed by Jendel (2000). The wear model in the HIPERWheel work was based on the T γ approach and that used by Jendel involves wear coefficients from pin-on-disc testing. The HIPERWheel predictor also included a fatigue assessment tool called FIERCE (Fatigue Impact Evaluator for Rolling Contact Environments), created by Eckberg (2002). The three mechanisms of fatigue initiation: surface, subsurface, and deep defect, are considered separately and quantified by three fatigue indices (details in Eckberg [2002]). The fatigue indices evaluated by FIERCE are represented as histograms. In addition, the surface fatigue impact may also be represented in the form of work-points plotted in a shakedown diagram. Interaction of fatigue and wear is not, however, considered. The work initiated by Jendel, has been taken forward and wear predictions now include the effects of natural lubrication and braking (Enblom, 2004).
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Industrial Lubrication Practice
22-17
Wheel and rail profiles
Multi-body dynamics model
Local contact analysis
Wear model
Change in wheel profile
FIGURE 22.21 Wheel profile evolution prediction tool schematic.
The two main drawbacks in some of these analysis tools are in the contact mechanics method used and the approach to wear modeling. Telliskivi and Olofsson (2001, 2004) studied different contact mechanics methods used in wear simulations and concluded that they significantly affected the accuracy. They also developed a new method with increased accuracy and no loss of calculation speed. Wear models are typically semiempirical and based on the results of laboratory bench tests. Ideally a model is required that is reliant on material properties alone. It is apparent at present that the detail and accuracy of the dynamic modeling far exceeds that of the wear modeling. An alternative approach is that of the Whole Life Rail Model developed by Kapoor et al. (2003). Contacts mechanics predictions are used to determine the plastic flow caused by each wheel passage. This is then used to determine the ratcheting rate. Crack initiation is considered to take place when the strain in a material element exceeds a critical value. Wear takes place when a region (or group of elements) loses integrity. This approach has the advantage that it does not rely on empirical data for life predictions and can implement varying materials properties in the rail section.
22.5 Friction Modification Friction modifiers are applied to the wheel/rail contact to generate required coefficients of friction. These may act to increase or decrease friction depending on the situation. A decrease in friction may be needed where wear rates are high, for example in low radius curves, and an increase may be desired where adhesion loss is prevalent, this would be used in traction and braking. Friction modifiers are divided in to three categories (Kalousec and Magel, 1999): 1. Low coefficient friction modifiers (lubricants) are used to give friction coefficients less than 0.2 at the wheel flange/gauge corner interface. 2. High friction modifiers with intermediate friction coefficients of 0.2 to 0.4 are used in wheel tread/ rail top applications. 3. Very high friction modifiers (friction enhancers) are used to increase adhesion for both traction and braking. Friction modifiers are classified according to their influence after full slip conditions have been reached in the wheel/rail contact, as shown in Figure 22.22 (Eadie et al., 2000). If friction increases after the saturation point the modifiers have positive friction properties, if friction reduces the modifier has negative friction properties. Positive friction modifiers can be described as high positive friction (HPF) or very high positive friction (VHPF), depending on the rate of increase in friction.
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Handbook of Lubrication and Tribology
Positive friction Saturation — full slip
Tractive force (= mN)
Neutral friction
Partial slip Negative friction Free rolling Creep
FIGURE 22.22 Behavior of friction modifiers. (Adapted from Eadie, D.T., Kalousek, J., and Chiddick, K.C., 2000, Proceedings of the 5th International Conference on Contact Mechanics and Wear of Rail/Wheel Systems, Tokyo, pp. 36–41.)
22.5.1 Increasing Friction Friction (or adhesion) loss has a large impact on safety and performance of railway networks. Poor adhesion in braking is a safety issue as it leads to extended stopping distances. If a train experiences poor adhesion in traction when pulling away from a station and a delay is enforced the train operator will incur costs. Similar delays will occur if a train passes over areas of poor adhesion while in service. Work carried out to investigate the causes of adhesion loss using both laboratory and field tests has identified the major causes as being: water (from rainfall or dew), humidity, leaves, wear debris, and oil contamination (Collins, 1972; Broster et al., 1974; Beagley, 1975; Beagley et al., 1975a, 1975b). Work carried out more recently has reemphasized the effect of the problems outlined above and identified further causes of adhesion loss, such as frost and mud deposited on rails by automobile wheels passing over level crossings (Logston, 1980; Nagase, 1989; Jenks, 1997). Most of the work cited above was carried out at relatively low speed. Work on adhesion issues related to high speed lines using both full-scale roller rigs and field measurements has shown that adhesion decreases with train velocity and wheel/rail contact force (Ohyama, 1991; Zhang et al., 2002). The most commonly used friction modifier used on railway networks worldwide to combat adhesion loss is sand. Sanding is used to improve adhesion in both braking and traction. In braking it is used to ensure that the train stops in as short a distance as possible. It usually occurs automatically when the train driver selects emergency braking. Sanding in traction, however, is a manual process. The train driver must determine when to apply the sand and how long the application should last. The sand is supplied from a hopper mounted under the train. Compressed air is used to blow the sand out of a nozzle attached to the bogie and directed at the wheel/rail contact region. This is quite a rudimentary solution and can cause problems to infrastructure. Wheel and rail wear rates increase severely when sand is applied (Kumar et al., 1986; Jenks, 1997; Lewis, 2004b) and there are problems associated with sand build-up at track sites where application is quite frequent. Very high positive friction modifiers to enhance the coefficient of friction from 0.4 to 0.6 are available, but are really only in the development stage. There are a number of different products available, but most involve a solid stick of material that is applied directly to the wheel tread.
22.5.2 Reducing Friction High friction coefficients are most prevalent at the wheel flange/rail gauge corner contact, particularly in curves. Load and slip conditions are also high, which means that high energy consumption and noise generation occurs and wear and rolling contact fatigue are more likely to occur at these sites.
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Industrial Lubrication Practice
22-19
Lubrication is commonly used on curved track to reduce the impact of high loads and slips and reduce friction. This offers a number of benefits including wear reduction (Clayton et al., 1989; Alp et al., 1996; Zhao et al., 1997) and energy saving (Reiff, 1999). The two main methods used to apply lubricant to the wheel/rail contact are wayside lubricators and on-board lubricators. Wayside lubricators are mounted next to the track and apply lubricant to the rail gauge corner. These are of three types: mechanical, hydraulic, and electronic. On-board lubricators apply grease or solid lubricant or spray oil on to the wheel flange, which is then transferred to the gauge corner of the rail. Complex control systems are used in the application process to avoid the application of lubricant at inappropriate locations. Mobile Lubricators, which are essentially railway vehicles designed to apply lubricant to the gauge corner of the track, are also sometimes used. Mechanical wayside lubricators are operated when the wheel makes contact with a mechanical plunger. This operates a pump, which supplies lubricant from a reservoir to a distribution unit. The wheel passes the distribution unit picking up the lubricant and spreading it along the track. Electronic lubricators work slightly differently. They use sensors to detect the approach of a train and activate electric pumps to deliver the lubricant. They are inherently more reliable than mechanical or hydraulic lubricators and can also be adjusted remotely from the track. On-board lubricators supply lubricant to the wheel flange/rail gauge corner. Most apply lubricant to the wheel flange, which is then spread along the rail as the wheel progresses. Some, however, deposit lubricant directly on the rail. Grease or oil spray systems are used, which employ complex control strategies using sensors measuring vehicle speed and track curvature to govern lubricant application. Solid stick lubricators are also available, in which a stick of lubricant is spring loaded against the wheel flange. On-board systems have a number of advantages over wayside lubricators (Sinclair, 2004). There is a reduced safety risk to staff during installation, inspection, and maintenance is easier as it can be carried out in more controlled conditions. The rail will continue to receive some friction control protection in the event of the failure of an individual on-board lubricator. However, despite these advantages, at problem tracks and sites, wayside lubricators are still a necessity. Problems exist with lubrication systems. These are related to both technical and human issues (Thelen and Lovette, 1996). The main technical problems with wayside lubricators are related to mechanical issues such as blocked applicator openings, leaking holes, ineffective pumps, and trigger mechanisms. Poor choice of lubricant can also lead to poor functioning of a lubricator. Human related problems can result from the technical issues. If over lubrication occurs and lubricant migrates onto the rail top, adhesion loss can occur. Train drivers may then be tempted to apply sand to compensate and increase friction, however, this will lead to increased wear and could cause the applicators to become blocked. The thought that application of lubricant will lead to wheel slip can also lead train drivers to switch off on-board lubrication systems. Poor wayside lubrication can lead to potentially serious problems including wheel slip and loss of braking and poor train handling (Roney, 2001). Other issues may be prevention of ultrasonic flaw detection, wastage of lubricant, and high lateral forces in curves and subsequent increase in wear. The key characteristics required of a lubricant are (Roney, 2001): lubricity or the ability of the lubricant to reduce friction; retentivity or the measure of time over which the lubricant retains its lubricity; and pumpability or how easily the lubricant can be applied to the track. The temperature is an issue as some track locations will experience a wide range across which some lubricants may not maintain their pumpability. Some networks use different lubricants in the winter and summer for this reason. The contact temperature is also important as flash temperatures can be as high as 600◦ C to 800◦ C, these can lead to the lubricant in the contact being burned up. Correct positioning of a wayside lubricator is critical to providing effective lubrication. Each site will require something different, which makes this task quite complex. Controlled field testing has been used to assess the reliability and efficiency of wayside lubricators based on a number of factors related to the lubricant including: waste prevention, burn up, distance covered, washing off by rain or snow, and migration to the rail top. This data and factors related to the track, such as length of curve, gradient, and applicator configuration, and traffic, including direction, types of bogie, axle loads, and speeds,
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Handbook of Lubrication and Tribology Angle a Load Grinding stone
FIGURE 22.23
Facet
Rail grinding.
have been combined to develop criteria and a model for positioning wayside lubricators (de Koker, 1994). Ultimately, however, the most critical element in preserving effective lubrication is maintenance. Once in place wayside lubricators need regular maintenance to prevent the problems outlined occurring.
22.6 Maintenance of Wheels and Rails In order to maintain the required contact conditions between wheel and rail, both are subjected to reprofiling programs. Rail reprofiling is achieved using a grinding process. The objectives of rail grinding are to keep control of the wheel/rail contact conditions (Roney, 2001) to: • maintain a balance between damage mechanisms to negate the need for premature rail replacement • facilitate the desired steering and dynamic stability of railway vehicles • reduce high dynamic loads and track vibrations Rail grinding is performed by specially designed railway vehicles using rotating grinding stones (as shown in Figure 22.23). The flat side of the stones is used, unlike machining process where the edge is used. The material removal process is dependent on the abrasive stone and the load applied to it, as well as the grinding speed and angle of the stone. Rail grinding can serve a number of purposes, as defined by Cooper (1993): 1. Preparative. Cleaning mill scale or surface defects introduced during laying of the rail to ensure a good start to service for the rail. 2. Preventative. Removing layers of fatigued metal before cracking leads to serious damage (see section on Interaction of Wear and Fatigue). 3. Curative. Recovering rail damages during wheel skids. These types of grinding will obviously require different levels of material removal. Wheel reprofiling has the same objectives as rail grinding. It is carried out during routine maintenance. Wheels undergo visual and ultrasonic inspection for damage at set intervals and at the same time regrinding is undertaken depending on the severity of the wear observed or other damage due to RCF or wheel flats etc. Reprofiling is undertaken using special lathes mounted beneath the track. Standards are set for levels of wear. These are related to measurements taken from datums on the flange tip and the inside of the wheel of tread wear and flange thickness reduction, as shown in Figure 22.24.
22.7 Conclusions The wheel/rail interface can appear deceptively simple, but is in fact a complex mixture of many interacting phenomena, which are affected by a multitude of variables many of which are uncontrollable. Management
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Industrial Lubrication Practice
22-21 W
Z Y X
13
70
W X Y Z
Flange thickness (new) Flange thickness (old) Flange height (new) Flange height (old)
Typical values (P8 profile) (mm) 29 24 30 36.5
FIGURE 22.24 Wheel profile limits.
of the wheel/rail interface is therefore a costly business, but is clearly essential for successful and safe rail infrastructure operation. Modeling of the phenomena occurring on wheels and rails leading to degradation and damage is very hard. Most predictive tools are subsequently based on empirical data, which limits their usefulness. The wheel/rail interface is an open system so there are many unpredictable variables and contact conditions vary at virtually every contact point. This means that it would be extremely time consuming to compute wear or fatigue predictions over an entire network. However, there are now models available to predict wear and fatigue of both wheels and rails, which will aid maintenance scheduling. Some railway networks have been slow to take on this new technology, but now at least are starting to appreciate the importance of the role that research in this area can play in improving the management of the wheel/rail interface.
References Alp, A., Erdemir, A., and Kumar, S., 1996, Energy and wear analysis in lubricated sliding contact, Wear, Vol. 191, pp. 261–264. Archard, J.F., 1953, Contact and rubbing of flat surfaces, Journal of Applied Physics, Vol. 24, pp. 981–988. Beagley, T.M. and Pritchard, C., 1975, Wheel/rail adhesion — the overriding influence of water, Wear, Vol. 35, pp. 299–313. Beagley, T.M., McEwen, I.J., and Pritchard, C., 1975a, Wheel/rail adhesion — the influence of railhead debris, Wear, Vol. 33, pp. 141–152. Beagley, T.M., McEwen, I.J., and Pritchard, C., 1975b, Wheel/rail adhesion — boundary lubrication by oily fluids, Wear, Vol. 33, pp. 77–88. Beagley, T.M., 1976, Severe wear of rolling/sliding contacts, Wear, Vol. 36, pp. 317–335. Bogdanski, S., Olzak, M., and Stupnicki, J., 1996, Influence of liquid integration on propagation of rail rolling contact fatigue cracks, Proceedings of the 2nd Mini Conference on Contact Mechanics and Wear of Wheel/Rail Systems, Budapest, pp. 134–143. Bolton, P.J., Clayton, P., and McEwen, I.J., 1982, Wear of rail and tyre steels under rolling/sliding conditions, ASLE Transactions, Vol. 25, pp. 17–24.
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Bolton, P.J. and Clayton, P., 1984, Rolling–sliding wear damage in rail and tyre steels, Wear, Vol. 93, pp. 145–165. Bower, A.F., 1988, The influence of crack face friction and trapped fluid on surface initiated rolling contact fatigue cracks, Transactions of the ASME, Journal of Tribology, Vol. 110, pp. 704–711. British Steel Makers Creep Committee, 1973, BSCC High Temperature Data, The Iron and Steel Institute for the BSCC, London. Broster, M., Pritchard, C., and Smith, D.A., 1974, Wheel–rail adhesion: it’s relation to rail contamination on British railways, Wear, Vol. 29, pp. 309–321. Cannon, D.F. and Pradier, H., 1996, Rail rolling contact fatigue – research by the European Rail Research Institute, Wear, Vol. 191, pp. 1–13. Carter, F.W., 1926, On the action of a locomotive driving wheel, Proceedings of the Royal Society, Vol. A112, pp. 151–157. Clayton, P., Danks, D., and Steele, R.K., 1989, Laboratory assessment of lubricants for wheel/rail applications, Lubrication Engineering, Vol. 45, pp. 501–506. Collins, A.H. and Pritchard, C., 1972, Recent research on adhesion, Railway Engineering Journal, Vol. 1, pp. 19–29. Cooper, J., 1993, Rail flaw detection: a particular challenge, Proceedings of the 5th International Heavy Haul Conference, Beijing, China. Danks, D. and Clayton, P., 1987, Comparison of the wear process for eutectoid rail steels: field and laboratory tests, Wear, Vol. 120, pp. 233–250. Dearden, J., 1960, The wear of steel rails and tyres in railway service, Wear, Vol. 3, pp. 43–49. de Koker, J., 1994, Development of a formula to place rail lubricators, Proceedings of the 5th International Tribology Conference. Eadie, D.T., Kalousek, J., and Chiddick, K.C., 2000, The role of high positive friction (HPF) modifier in the control of short pitch corrugation and related phenomena, Proceedings of the 5th International Conference on Contact Mechanics and Wear of Rail/Wheel Systems, Tokyo, pp. 36–41. Ekberg, A., Kabo, E., and Andersson, H., 2002, An engineering model for prediction of rolling contact fatigue of railway wheels, Fatigue and Fracture of Engineering Materials and Structures, Vol. 25, pp. 899–916. Enblom, R. and Berg, M., 2004, Wheel wear modelling including disc braking and contact environment, accepted for presentation at the 14th International Wheelset Congress, Florida, 17th–21st October 2004. Feller, H.G. and Walf, K., 1991, Surface analysis of corrugated rail treads, Wear, Vol. 144, pp. 153–161. Fries, R.H. and Dávila, C.G., 1987, Wheel wear predictions for tangent track running, Transactions of the ASME, Journal of Dynamics Systems, Measurement, and Control, Vol. 109, pp. 397–404. Garnham, J.E. and Beynon, J.H., 1992, Dry rolling–sliding wear of bainitic and pearlitic steels, Wear, Vol. 57, pp. 81–109. Gao, N. and Dwyer-Joyce, R.S., 2000, The effects of surface defects on the fatigue life of water and oil lubricated contacts, Proceedings of the IMechE Part J, Journal of Engineering Tribology, Vol. 214, pp. 611–626. Grassie, S. and Kalousek, J., 1997, Rolling contact fatigue of rails: characteristics, causes and treatments, Proceedings of the 6th IHHA Conference, Capetown. Grieve, D.G., Dwyer-Joyce, R.S., and Beynon, J.H., 2001, Abrasive wear of railway track by solid contaminants, Proceedings of the IMechE Part F, Journal of Rail and Rapid Transport, Vol. 215, pp. 193–205. Jendel, T., 2000a, Prediction of wheel profile wear — comparisons with field measurements, Proceedings of the International Conference on Contact Mechanics and Wear of Rail/Wheel Systems, Tokyo, Japan, 25–28 July, pp. 117–124. Jendel, T., 2000b, Prediction of Wheel Profile Wear — Methodology and Verification, Licentiate Thesis, TRITA-FKT 2000:9, Royal Institute of Technology, Stockholm, Sweden.
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Jenks, C.W., 1997, Improved Methods for Increasing Wheel/Rail Adhesion in the Presence of Natural Contaminants, Transit Co-operative Research Program, Research Results Digest, No. 17. Jiang, Y.Y. and Sehitoglu, H., 1999, A model for rolling contact failure, Wear, Vol. 224, pp. 38–49. Johnson K.L., 1985, Contact Mechanics, Cambridge University Press. Kalker J.J., 1990, Three-Dimensional Elastic Bodies in Rolling Contact, Kluwer, Dordrecht. Kalker, J.J., 1982, Fast algorithm for the simplified theory of rolling contact, Vehicle System Dynamics, Vol. 11, pp. 1–13. Kalousek, J., Fegredo, D.M., and Laufer, E.E., 1985, The wear-resistance and worn metallography of pearlite, bainite and tempered martensite rail steel microstructures of high hardness, Wear, Vol. 105, pp. 199–222. Kalousek, J. and Magel, E., 1997, Optimising the wheel/rail system, Railway and Track Structure, January. Kalousek, J. and Magel, E., 1999, Modifying and managing friction, Railway and Track Structures, May. Kapoor, A., Fletcher, D.I., and Franklin, F.J., 2003, The role of wear in enhancing rail life, Proceedings of the 29th Leeds-Lyon Symposium on Tribology, pp. 331–340. Krause, H. and Poll, G., 1986, Wear of wheel–rail surfaces, Wear, Vol. 113, pp. 103–122. Kumar, S., Krishnamoorthy, P.K., and Prasanna Rao, D.L., 1986, Wheel–rail wear and adhesion with and without sand for a North American locomotive, Journal of Engineering for Industry, Transactions of the ASME, Vol. 108, pp. 141–147. Kumar, S. and Rao, D.L.P., 1984, Wheel–rail contact wear, work, and lateral force for zero angle of attack — a laboratory study, Transactions of the ASME, Journal of Dynamic Systems, Measurement, and Control, Vol. 106, pp. 319–326. Lewis, R. and Olofsson, U., 2004, Mapping rail wear regimes and transitions, Wear, Vol. 257, No. 7–9, pp. 721–729. Lewis, R. and Dwyer-Joyce, R.S., 2004a, Wheel–rail wear and surface damage caused by adhesion sanding, Proceedings of the 30th Leeds-Lyon Symposium on Tribology, Elsevier Tribology Series No. 43, pp. 731–741. Lewis, R. and Dwyer-Joyce, R.S., 2004b, Wear mechanisms and transitions in railway wheel steels, Proceedings of the IMechE, Part J, Journal of Engineering Tribology, Vol. 218, pp. 467–478. Lewis, R., Dwyer-Joyce, R.S., Bruni, S., Ekberg, A., Cavalletti, M., and Bel Knani, K., 2004, A new CAE procedure for railway wheel tribological design, Proceedings of the 14th International Wheelset Congres, Florida, 17–21 October 2004. Lim, S.C. and Ashby, M.F., 1987, Wear mechanism maps, Acta Metallica, Vol. 35, pp. 1–24. Logston, C.F. and Itami, G.S., 1980, Locomotive friction-creep studies, Transactions of the ASME, Journal of Engineering for Industry, Vol. 102, pp. 275–281. Marshall, M.B., Lewis, R., and Dwyer-Joyce, R.S., 2004, Ultrasonic characterisation of a wheel/rail contact, Proceedings of the 30th Leeds-Lyon Symposium on Tribology, Elsevier Tribology Series No. 43, pp. 151–158. McEwen, I.J. and Harvey, R.F., 1985, Full-scale wheel-on-rail testing: comparisons with service wear and a developing theoretical predictive model, Lubrication Engineering, Vol. 41, pp. 80–88. Nagase, K., 1989, A study of adhesion between the rails and running wheels on main lines: results of investigations by slipping adhesion test bogie, Proceedings of the IMechE Part F, Journal of Rail and Rapid Transit, Vol. 203, pp. 33–43. Ohyama, T., 1991, Tribological studies on adhesion phenomena between wheel and rail at high speeds, Wear, Vol. 144, pp. 263–275. Olofsson, O. and Nilsson, R., 2002, Surface cracks and wear of rail: a full scale test and laboratory study, Proceedings of the IMechE Part F, Journal of Rail and Rapid Transit, Vol. 216, pp. 249–264. Olofsson, U. and Telliskivi, T., 2003, Wear, friction and plastic deformation of two rail steels — full-scale test and laboratory study, Wear, Vol. 254, pp. 80–93. Pearce, T.G. and Sherratt, N.D., 1991, Prediction of wheel profile wear, Wear, Vol. 144, pp. 343–351. Pyzalla, A., Wang, L., Wild, E., and Wroblewski, T., 2001, Changes in microstructure, texture and residual stresses on the surface of a rail resulting from friction and wear, Wear, Vol. 251, pp. 901–907.
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Reiff, R. and Creggor, D., 1999, Systems approach to best practice for wheel and rail friction control, International Heavy Haul Conference. Roney, M.D., 2001, Maintaining optimal wheel and rail performance, in Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA. Sawley, K. and Kristan, J., 2003, Development of bainitic rail steels with potential resistance to rolling contact fatigue, Fatigue and Fracture of Engineering Materials and Structures, Vol. 26, pp. 1019–1029. Sinclair, J., 2004, Friction modifiers, in Vehicle Track Interaction: Identifying and Implementing Solutions, IMechE Seminar, 17th February. Singh, U.P., Roy, B., Jha, S., and Bhattacharyya, S.K., 2001, Microstructure and mechanical properties of as rolled high strength bainitic rail steels, Materials Science and Technology, Vol. 17, pp. 33–38. Stanca, M., Stefanini, A., and Gallo, R., 2001, Development of an integrated design methodology for a new generation of high performance rail wheelsets, Proceedings of the 16th European MDI User Conference, Berchtesgaden, Germany, 14–15 November. Steele, R.K., 1982, Observations of in-service wear of railroad wheels and rails under conditions of widely varying lubrication, ASLE Transactions, Vol. 25, pp. 400–409. Telliskivi, T. and Olofsson, U., 2001, Contact mechanics analysis of measured wheel-rail profiles using the finite element method, Proceedings of the IMechE Part F, Journal of Rail and Rapid Transit, Vol. 215, pp. 65–72. Telliskivi, T. and Olofsson, U., 2004, Wheel–rail wear simulation, Wear, Vol. 257, pp. 1145–1153. Thelen, G. and Lovette, M., 1996, A parametric study of the lubrication transport mechanism at the rail–wheel interface, Wear, Vol. 191, pp. 113–120. Tournay, H., 2001, Supporting technologies vehicle track interaction, in Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA. Ueda, M., Uchino, K., Kageyama, H., Motohiro, K., and Kobayashi, A., 1997, Development of bainitic steel rail with excellent surface damage resistance, Proceedings of the 6th IHHA Conference, Capetown. Yokoyama, H., Mitao, S., Yamamoto, S., and Fujikake, M., 2002, Effect of the angle of attack on flaking behaviour in pearlitic and bainitic steel rails, Wear, Vol. 253, pp. 60–67. Zakharov, S., 2001, Wheel/rail performance, in Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA. Zhao, X.Z., Zhu, B.L., and Wang, C.Y., 1997, Laboratory assessment of lubricants for wheel/rail lubrication, Journal of Materials Science and Technology, Vol. 13, pp. 57–60. Zhang, W., Chen, J., Wu, X., and Jin, X., 2002, Wheel/rail adhesion and analysis by using full scale roller rig, Wear, Vol. 253, pp. 82–88. Zobory, I., 1997, Prediction of wheel/rail profile wear, Vehicle System Dynamics, Vol. 28, pp. 221–259.
© 2006 by Taylor & Francis Group, LLC
23 Lubrication in the Timber and Paper Industries 23.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 23.2 Lubrication in the Timber Industry . . . . . . . . . . . . . . . . . .
23-1 23-2
Timberlands Operations • Felling Equipment • Harvester, Skidder, and Forwarder Lubrication • Chainsaw Lubrication
23.3 Lubrication in the Pulp and Paper Industry . . . . . . . . .
23-5
Wood Yard Operations • Sawmill Equipment Lubrication • Pulp Mill Operations • Pulp Mill Equipment Lubrication • Paper Mill Operations — Wet End Equipment • Paper Mill Operations — Dry End Equipment • Paper Machine Lubrication — Wet End • Paper Machine Lubrication — Dry End • Paper Machine Oil Filterability
23.4 Lubricant Selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Paul W. Michael Milwaukee School of Engineering, Fluid Power Institute
23-14
Bearing Lubricant Selection • Gear Lubricant Selection • Synthetic Gear Lubricants • Hydraulic Fluid Selection • Biodegradable Lubricants • Vegetable Oil Based Lubricants — HETG • Synthetic Ester Based Lubricants — HEES • Lubrication Management Strategies • Oil Analysis
References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
23-21
23.1 Introduction Timber harvesting and papermaking push the limits of lubrication technology. In timber harvesting, rugged terrain, extreme temperatures, and the sheer mass of the crop place unparalleled strain upon equipment. In papermaking, the economics of a global market mandate large capital investments, high productivity, and extraordinary equipment reliability. As a result, these industries demand topflight lubricants and lube management practices. Both applications are frequently in close proximity to precious natural resources — wilderness and fresh water. Consequently lubricants for these applications are increasingly formulated to maximize lubricant longevity and minimize environmental impact. This chapter reviews features of lubricant selection, application, and management that are to some extent unique to industries that largely rely on forests for their prime raw material.
23-1
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Handbook of Lubrication and Tribology
23.2 Lubrication in the Timber Industry 23.2.1 Timberlands Operations Timber harvesting systems are characterized by the manner in which wood is moved from the stump to the roadside. The three major types of harvesting methods are tree-length, whole-tree, and shortwood methods [1]. Geographic, economic, regulatory, and silvicultural factors generally dictate the optimum system for timber harvesting. The tree-length harvesting method involves felling, delimbing, and topping trees in the woods and skidding the tree lengths to the landing. The tree lengths are bucked into logs or sticks at the landing. In some instances, the tree lengths are hauled to a mill site for processing. The whole-tree harvesting method involves transporting the entire felled tree to the landing for processing. One advantage of the whole-tree method is that it can utilize the tree’s entire biomass. The shortwood harvesting method involves the conversion of trees into desired length products at the stump, either by hand with chainsaws or by using a mechanized processor which fells, delimbs, and bucks the tree into sawlogs, pulpwood sticks, or other products. The individual pieces are then transported to the landing with a forwarder. The shortwood method is used for thinning a stand or single-tree selection silviculture because it causes less forest damage [2]. Silviculture has features in common with traditional agriculture, particularly with the increasing reliance upon commercial forests and plantation crops such as Eucalyptus. However, the vast majority of wood that is harvested is derived from natural hardwood and softwood forests. Additionally, harvesting timber when the ground is frozen is often desirable because it reduces the impact of heavy equipment on surface vegetation. Thus, timber harvesting equipment must be able to fell, process, and transport a heavy crop through rugged terrain at temperatures that are well below those encountered in most agriculture applications.
23.2.2 Felling Equipment Manual logging is labor intensive and hazardous since it requires workers to operate chainsaws in close proximity to falling trees. In order to reduce labor and insurance costs, the timber industry increasingly relies upon mechanized harvesting equipment such as the feller buncher [3]. A feller buncher is similar to a front-end loader or excavator (see Figure 23.1). It is a rubber tired or tracked vehicle with an articulating extendable arm onto which a felling head is attached. The felling head typically consists of hydraulic grappling devices and a disc saw. Feller bunchers may also be equipped with harvester heads that delimb, debark, and cut trees to length. When felling a tree, the operator positions the felling head at the base of the trunk and hydraulic grappling arms wrap around the tree as the saw removes the tree from the stump (see Figure 23.2). The machine then takes the severed tree (or trees) and lowers it to the ground where it may be loaded into a forwarder or skidded to the roadside. Feller bunchers are advantageous in that they are able to cut trees closer to the ground than chainsaws. However, they may not be suitable for use in steep terrain or for harvesting some large diameter trees. Felled trees generally are transported to a roadside landing area by skidders or forwarders. A skidder is a rubber tired or tracked vehicle that tows cut trees to the landing for transport or processing. Classified as either cable, clambunk, or grapple skidders (see Figure 23.3), their names are derived from the method used to attach the logs to the skidder. They come in a variety of sizes but all have four-wheel drive, good speed, high ground clearance, and diesel power as basic features. Wheeled skidders are used in a wide range of conditions. Track skidders have a bulldozer like undercarriage and are suitable for use in very steep terrain. Skidders are also equipped with an articulating blade which makes them suitable for moving debris that may impede travel through the woods. A forwarder has a front section like a skidder, but instead of a cable winch, has a hydraulic knuckleboom log loader behind the cab and a log deck above the rear axle (see Figure 23.4) [4]. Forwarders are used with conventional chainsaw felling and feller harvesters. Forwarders have a bogie axle that float with the rougher terrain which reduces ground disturbance and counteracts the effect of high center of gravity.
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Lubrication in the Timber and Paper Industries
23-3
FIGURE 23.1 Feller buncher — timber harvesting machine.
FIGURE 23.2 Feller buncher — timber harvesting machine, courtesy TimberPro.
A feller forwarder is a feller buncher with a bunk to the rear of the operator into which the felled trees are lowered and carried to the next tree to be felled. The process is repeated until the bunk is full. The machine then forwards the trees to the landing and unloads them.
23.2.3 Harvester, Skidder, and Forwarder Lubrication Lubrication of timber harvesting machines is similar to that of off-highway construction equipment. Diesel engine oil, hypoid gear lube, transmission oil, hydraulic fluid, and grease are the primary lubricants. The engine oil requirements for this class of equipment are similar to that of construction and off-highway diesel engines. The same is generally true of gear lube and transmission fluid requirements. However, the demands upon lubricants used in timber harvesting equipment are more severe than encountered in many construction applications. Not only are the payloads and temperature ranges extreme, but also construction applications often involve intermittent duty (position and hold) whereas timber harvesting
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Handbook of Lubrication and Tribology
FIGURE 23.3
Grapple skidder — timber harvesting machine.
FIGURE 23.4
Forwarder — timber harvesting machine, courtesy TimberPro.
tends to be continuous duty. These factors create a need for shear stable lubricants with exceptional oxidation resistance and effective antiwear chemistry. Table 23.1 provides a list of lubricant recommendations for timber harvesting equipment.
23.2.4 Chainsaw Lubrication While automated timber harvesting equipment is employed throughout much of Europe and North America, chainsaws remain the most popular tool in the timber harvesting industry. In regions where capital investment is limited, two-cycle chainsaws may be the only power tool available to loggers. Two-cycle engines are used in chainsaw applications because of their light weight, high speed, and ability to be operated at any orientation. They utilize two-cycle engine oils that are diluted at a 50 : 1 fuel to
© 2006 by Taylor & Francis Group, LLC
Lubrication in the Timber and Paper Industries TABLE 23.1
23-5
Lubrication of Timber Harvesting Equipment
Component Engine Powershift transmission Hydrostatic pump drive Hydrostatic transmission Swing drive Axle differentials (May include wet brakes) Bogie axle Boom hydraulics Harvester heads and grapples Chain auto-lube system
Lubricant
Notes (h = hour)
Diesel engine oil 10W30 or 15W40 Dexron III or C-4 fluid SAE 10 GL-5 Hypoid gear oil 75W90 Multipurpose tractor fluid or multigrade hydraulic oil GL-5 Hypoid gear oil or multigrade hydraulic oil Multipurpose tractor fluid or 75W90 gear oil GL-5 hypoid gear oil 75W90 Multigrade hydraulic fluid HV 32 to HV 68 EP 2 grease 5% Moly Bar and chain oil
Check oil level daily Sample interval: 250 h Check oil level weekly Sample interval: 1000 h Check oil level weekly Sample interval: 1000 h Check oil level weekly Sample interval: 1000 h Check oil level weekly Sample interval: 1000 h Check oil level weekly Sample interval: 2000 h Check oil level weekly Sample interval: 2000 h Check oil level weekly Sample interval: 2000 h Grease daily Check oil level daily
Source: Schierschmidt, B., personal communication, 2004. Lambert, J., personal communication, 2005.
oil ratio. This 50 : 1 premix is aspirated into the crankcase prior to entering the combustion chamber and serves as both a fuel and a lubricant. Two-cycle engines run very hot, which causes pistons to expand, thus decreasing the piston-cylinder clearance. This loss of clearance increases engine friction and the possibility of scuffing. In order to prevent engine seizure, two-cycle oils incorporate lubricity additives that reduce sliding friction on the cylinder wall. Two-cycle oils must also burn cleanly. Clean burning oils reduce visible smoke exhaust by protecting engines against exhaust port blocking, combustion chamber deposits, ring sticking, and wear. In order to reduce the amount of noncombustible materials entering the combustion chamber, two-cycle oils are formulated with low-ash or ashless additive systems. A typical two-cycle oil contains the following components: • • • •
Synthetic esters for lubricity Solvent for miscibility and low-temperature fluidity Polybutene for lubricity and smoke reduction Ashless antiwear and detergency additives
Bar and chain oils are also used in chainsaw applications. Typically these products are SAE 10 or SAE 30 lubricants that contain tackifiers to enhance adhesion. They reduce groove and rail wear on bars by minimizing drive link friction. Since bar and chain oils are “total loss” lubricants, that is they are used once and then released into the environment, increasingly biodegradable lubricants are used in this application [5].
23.3 Lubrication in the Pulp and Paper Industry The pulp and paper industry presents a distinctive environment for lubricants and maintenance engineers. Intense global competition and environmental regulation create an economic environment that favors large-scale, highly capitalized operations. In order to generate a satisfactory return on investment (ROI) in this economic climate, it is necessary to maintain high levels of equipment reliability and productivity.
© 2006 by Taylor & Francis Group, LLC
23-6
Handbook of Lubrication and Tribology Veneer logs Plywood mill
Wood residues
Pulpwood
Particle board mill
Sawlogs Sawmill
Pulpwood
Chemical pulping
Pulpwood
FIGURE 23.5
Wood residues
Paper mill
Groundwood mill
Integrated wood-yard operations.
23.3.1 Wood Yard Operations Wood yard operations in lumber and pulp mills have a number of similarities. In both industries, logs are unloaded from semi-trailers or railcars and transported to log decks by modified end loaders or knuckle-boom cranes. If necessary, tree length logs are cut to length on slasher decks. Saw logs are then conveyed to debarking in order to produce clean lumber or pulp. Rosser-head or ring debarkers are used in sawmill applications while drum debarkers are used in pulp mills. Drum debarkers are not used in sawmill applications because they cause the log ends to splinter, which reduces their value as lumber. In pulp mill operations debarked wood is chipped, washed, and screened. In sawmill operations debarked wood is sawed, edged, trimmed, and kiln dried. Sawmill trimmings are processed through a chipper for use in pulp mills. In both applications, bark is usually ground into small pieces by refuse hogs and used as boiler fuel. In order to maximize economic utilization of timber resources, lumber and pulping operations may be integrated as depicted in Figure 23.5.
23.3.2 Sawmill Equipment Lubrication Sawmills utilize a vast assortment of kickers, turners, roller decks, and chain conveyors. Chain oil, gear lube, and grease are the primary lubricants used in this equipment. In order to avoid staining the wood, light colored oils are preferred for the lubrication of roller-chain conveyors. Optimizing solid-wood yields is particularly important in sawmill operations where laser sensors and hydraulic setworks are used to maximize the quantity and value of lumber extracted from logs. Band saws are increasingly used in these applications because they have a narrow kerf that produces about half the sawdust generated by a circular saw. While sawdust is a useful fuel source for the boilers that heat lumber kilns, solid wood has a much higher economic value. With the periodic exception of kiln fan bearings, sawmill operating temperatures are not particularly high. However, low temperatures, moisture, dirt, and sawdust are endemic, as is shock-loading due
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Lubrication in the Timber and Paper Industries
23-7
to the constant pounding of the logs that are processed. These conditions mandate frequent grease application, use of synthetic gear oils, and generous lubrication of roller-chain conveyors. Table 23.2 provides lubrication recommendations and guidelines for sawmill equipment.
23.3.3 Pulp Mill Operations Pulp is the key raw material for papermaking. Pulp is made by mechanically or chemically separating the fibers in wood or other cellulosic materials from nonfibrous material. Mechanical pulping converts wood into pulp by shearing the fibers in logs or chips with grinders and disk refiners. Chemical pulping converts wood chips into pulp by cooking them in an aqueous solution at high temperatures and pressures. The two principal chemical pulping methods are the (alkaline) kraft process and the (acidic) sulfite process. In the kraft process wood chips are cooked in a solution of sodium hydroxide (NaOH) and sodium sulfide (Na2 S). In the sulfite process wood chips are cooked in a solution of sulfurous acid (H2 SO3 ) and bisulfite ion (HSO− 3 ). Both processes chemically degrade the lignin that binds the fibers within the wood. Due to advantages in chemical recovery and fiber quality, the kraft process has gained larger market acceptance. In an integrated mill, pulp is usually stored as a medium consistency stock prior to forming, pressing, and steam-cylinder drying on the paper machine. For nonintegrated pulp and fiber operations, pulp must be dewatered to decrease transportation costs. Dewatering is usually performed on a lapping press, which utilizes a sheet forming wet end, a press section, and a drying section similar to a paper machine [6]. The laps that are produced by this machine are stacked and bailed by a hydraulic layboy for economical transportation by truck or rail.
23.3.4 Pulp Mill Equipment Lubrication Lubrication of pulp mill equipment combines demanding elements of off-road construction and industrial equipment lubrication. Heavy loads, wide ranging temperatures, and contamination challenge the effectiveness of lubricants. High capital equipment costs and global competition also make equipment reliability and productivity a priority. Thus effective lubricants and lubrication management systems are critical to the economic success of mill yard operations. Use of synthetic lubricants and consolidation of lubricants are often helpful means of improving the efficiency of mill yard equipment lubrication. Table 23.3 provides a list of lube recommendations for pulp mill equipment.
23.3.5 Paper Mill Operations — Wet End Equipment Papermaking furnish is prepared in the Approach System of the paper machine. The machine chest, fan pump, headbox, and slice are the key components of the approach system. Here, pulp is diluted, mixed, and screened prior to being pumped into the headbox by the fan pump. Pressurized stock from the fan pump is discharged from the headbox through the slice opening and deposited on the forming fabric or wire. The wire is a continuous plastic mesh belt that carries the high-water content stock across a series of forming boards on the Fourdrinier. Stock is dewatered on the Fourdrinier by hydraulic pressure gradients created by these forming boards or foils. There also is a series of vacuum boxes on the Fourdrinier that assist in water removal. The forming section of the paper machine, which includes the Fourdrinier (Figure 23.6), incorporates several types of rolls: • The breast roll, located below the slice opening, where stock is deposited onto the forming fabric or wire • The dandy roll, positioned near the end of the Fourdrinier, which restructures fibers within the web and reduces the water content at the top of the sheet • The suction couch roll, at the end of the Fourdrinier, which transfers the formed sheet into the press section of the paper machine • Wire turning, guide, and tensioning rolls that control forming fabric travel
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23-8 TABLE 23.2
Handbook of Lubrication and Tribology Lubrication of Sawmill Equipment
Machine Log decks — kickers, loaders, and turners
Rosser head debarker
Component
Circle saw
Lithium complex grease NLGI 2 HM 32 HM 32
Chain drive gearbox
Synthetic ISO 220
Bull wheel bearings
Grease daily
Chains Chain drive gearbox
Lithium complex grease NLGI 2 Lithium complex grease NLGI 2 HM 32 Synthetic ISO 220
Hydraulic system
HM 32
Chains and air cylinders Chain drive gearbox
HM 32 Synthetic ISO 220
Arbor bearings
Lithium complex grease NLGI 2 HM 32
Check weekly Sample quarterly Hand oil weekly Inspect weekly Sample annually Grease weekly
Hydraulic turn, set, and feed
Band saw
Grease weekly Hand oil weekly Check weekly Sample quarterly Inspect weekly Sample annually
Grease daily Hand oil weekly Inspect weekly Sample annually
Check weekly Sample quarterly Hand oil weekly Grease weekly
Carriage knees, slides, and pinion Cable drum and idler bearings
HM 32 Lithium complex grease NLGI 2
Band saw wheel bearings
Hydraulic turn, set, and feed
Lithium complex grease NLGI 2 Lithium complex grease NLGI 2 HM 32
Saw guides and slides Wheel yoke screw drives
Spindle oil ISO VG 10 Synthetic ISO 220
Airline oiler Yoke bushing oil cup
HM 32 HM 32
Arbor bearings
Lithium complex grease NLGI 2 Spindle oil ISO VG 10
Grease weekly
Press and feed roll bearings
Gang saw
Notes
Conveyor and hour glass roll bearings Chains and air cylinders Log turner hydraulic system
Head wheel bearings
Ring debarker
Lubricant
Blade guide auto lube system
Clean and repack annually Grease weekly Check weekly Sample quarterly Check weekly Inspect weekly Sample annually Check weekly Fill weekly
Check weekly
Chipper
Main bearings
Lithium complex grease NLGI 2
Clean and repack at 6 month interval
Chip shaker screen
Eccentric bearings
Lithium complex grease NLGI 2 Lithium complex grease NLGI 2
Grease weekly
Knuckle bushings
Grease weekly
Bark hog
Main bearings
Lithium complex grease NLGI 2
Grease monthly
Kiln
Blower bearing fittings
Lithium complex grease NLGI 2 Synthetic ISO 220
Grease monthly
Blower bearing oil cups Source: Cornell, S., personal communication, 2005. Leja, R., personal communication, 2005.
© 2006 by Taylor & Francis Group, LLC
Fill weekly
Lubrication in the Timber and Paper Industries TABLE 23.3
23-9
Lubrication of Pulp Mill Equipment
Machine Debarker feed and deicing deck Drum debarker
Component
Lubricant
Notes
Hydraulic system
HV 46
Motor Gear coupling Fluid coupling
Grease Grease HM 68
Chain gear drive Chain Trunnions — grease lubed Trunnions — oil lubed
Synthetic ISO 220 HM 68 Lithium complex #2 Syn 460
Hydraulic discharge gate
HM 68
Chipper
Synchronous motor Bearings
Grease Grease
Motor bearing grease Lithium complex NLGI #2
Belt and screw chip conveyors
Motor PIV coupling Gear coupling Reducer
Grease Grease Grease Synthetic ISO 220
Chains Screw bearings Rollers Tail head
HM 68 Grease Grease Grease
Apply motor bearing grease annually Monthly Apply coupling grease annually Inspect weekly 6-month sample interval Inspect weekly Monthly Monthly Monthly
Falk drive
Synthetic ISO 220
Chains Screen bearings
HM 68 Grease
Pump bearings
HM 68
Gear coupling Reducer
Grease Synthetic ISO 220
Agitator bearings
Grease
Main drive shaft Plate adjust gear case
Grease Synthetic ISO 460
Rotary and disc chip screens
Digestors and agitators
Refiner
Washer
Lime kiln
Lube pump bearings
Grease
Motor bearings Coupling Gear case
Grease Grease Synthetic ISO 460
Shaft
Grease
Motor Coupling Reducer
Grease Grease Synthetic ISO 460
Drive gears Trunnions — grease lubed Trunnions — oil lubed
Open gear compound Grease Synthetic ISO 460
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Inspect weekly 3-month sample interval Apply motor bearing grease annually Apply coupling grease annually Inspect weekly 6-month sample interval Inspect weekly 3-month sample interval Automatic drip or brush oiler Grease monthly Inspect weekly 6-month sample interval Inspect weekly 6-month sample interval
Inspect weekly 6-month sample interval Check oil level weekly Lithium complex NLGI #2 Inspect daily 6-month drain interval Apply coupling grease annually Inspect weekly 6-month sample interval Monthly Monthly Inspect weekly 6-month sample interval Monthly Apply motor bearing grease annually Apply coupling grease annually Inspect weekly 6-month sample interval Monthly Apply motor bearing grease annually Apply coupling grease annually Inspect weekly 3-month sample interval Inspect weekly Grease weekly Inspect weekly 6-month sample interval
23-10 TABLE 23.3
Handbook of Lubrication and Tribology Continued
Machine Boiler
Component
Lubricant
Notes
Motor bearings
HL 32
Fluid coupling
HL 32
Feed water pump
HL 32
Steam turbine
Reservoir
HL 32
Inspect weekly 3-month sample interval
Stock and chemical pumps
Motor bearings Coupling Reducer Pump bearings
Grease Grease Synthetic ISO 460 HM 68
Apply motor bearing grease annually Apply coupling grease annually Inspect weekly 6-month sample interval Inspect weekly 6-month drain interval
Whitewater pumps
HM 68
Vacuum pumps
HM 68
Double wire press drive reducer Dancer roll Hydraulic felt tensioner
Synthetic ISO 220 Grease HM 46
Inspect weekly 3-month sample interval Inspect weekly 6-month sample interval Inspect weekly 6-month sample interval Grease quarterly Inspect weekly 6-month sample interval
Convey, lift, and discharge reducers Up-ender and scissors lift hydraulics Shafts and rolls
Synthetic ISO 220 HM 46
Lapping press
Pulp layboy
Grease
Inspect weekly 6-month sample interval Inspect weekly 6-month sample interval Inspect weekly 6-month sample interval
Inspect weekly 6-month sample interval Inspect weekly 6-month sample interval Grease quarterly
Source: Sobczak, T. and Prescher, S., personal communication, 2004.
Paper is transferred from the suction couch roll to the suction pickup roll as it enters the Press Section of the machine. In the press section of the machine, the paper web is conveyed through a series of two-roll nips by a synthetic felt. Pressing the paper removes water from the sheet and consolidates the web. Since it is more economical to remove water from paper by pressing than evaporation, press section design and operation is carefully optimized for maximum water removal. A myriad of specialized rolls are used to dewater paper in the press section including: • • • •
Perforated suction press rolls with suction boxes that draw water through the felt Grooved and blind-drilled rolls that provide a path for water to escape the nip Swimming rolls that utilize oil pressure to compensate for roll deflection Nipco rolls with individual hydrostatic pistons that allow precise crown control in different zones of the roll • Extended nip and other stationary shoe presses that are used to increase nip dwell times • Granite and covered steel rolls that serve as sealing surfaces when mated with controlled crown and perforated rolls • Felt turning, guide, and tensioning rolls that control press fabric travel
23.3.6 Paper Mill Operations — Dry End Equipment After press section operations, paper is transferred to the Dryer Section of the machine where it is carried by a dryer felt over a series of rotating steam-heated cylinders or dryer rolls. Typically, the dryer section of
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Lubrication in the Timber and Paper Industries
23-11 Dryer section
Headbox
FIGURE 23.6
Fourdrinier
Calender
Press
Reel
Typical paper machine layout, adapted from Smook courtesy Angus Wilde Publications [6].
the machine contains 40 to 100 dryer rolls that are 1 to 2 m in diameter. Water is removed from the paper in a two-step process as it is heated by contact with the rotating cylinder and flashes off as steam in the open draw or pocket between upper and bottom cylinders [7]. Pocket ventilators are utilized to hasten evaporation by supplying hot dry air through the felt rolls or an exterior duct. In the manufacture of lightweight paper such as tissue a single large dryer cylinder, known as a Yankee dryer, is employed. Yankee dryers vary in size from 4 to 7 m in diameter. A pressure roll transfers the sheet to the Yankee dryer surface. As the thin sheet rotates around the Yankee, it quickly dries and then is scraped off by a metal blade known as a doctor. Following the last dryer, paper calendaring and reeling operations are performed. The calendar is a set of rolls that smooth the paper surface to improve printability. In traditional hard-nip machines the paper is compressed to a uniform thickness by two hard rolls. In super calendars, mating a hard roll with a soft roll forms the nip. Calendaring can also be performed off-machine, although it is less common. The finishing section contains the reel and winder. The reel collects the paper coming off the paper machine. The winder is used to cut the paper into smaller rolls that are then shipped to their customers or gets converted onsite into the end product.
23.3.7 Paper Machine Lubrication — Wet End The wet end of a paper machine operates in a deluge of water and fiber. It is also subject to frequent wash-downs with high-pressure aqueous cleaners. These conditions present a significant challenge in terms of bearing lubrication. Upper press rolls are particularly problematic because of the position of the load zone [8]. Since grease can provide a physical barrier that inhibits water ingression, it is often the preferred lubricant for wet-end bearings [9]. Lubricating grease is a solid to semi-fluid substance formed by mixing a gelling agent or soap thickener into a lubricating fluid. The thickening agent keeps the lubricant in place, while a combination of heat, mechanical shear, and capillary action transfers lubricating fluid to bearing surfaces [10]. Typically NLGI #2 calcium sulfonate or lithium complex grease with good water washout and corrosion resistance is used for wet-end bearing lubrication. Since water contamination may cause grease to de-gel and soften in consistency, evaluating the roll-stability of grease after addition of water is advisable [11]. As shown in Table 23.4, wet-end bearings may also utilize oil. While grease offers the benefit of sealing out contaminants, oil is more effective at contamination removal. In high speed machines, lubricating oil systems with coolers, filters, and vacuum dehydrators are often employed. This is necessary because contamination removal is critical in bearings that rotate at high peripheral velocity [12].
23.3.8 Paper Machine Lubrication — Dry End Whereas some older machines with grease lubricated journal bearings still produce paper and liner board, oil lubricated spherical roller bearings are used in the dryer sections of modern machines. Since high-pressure steam passes through the journal of a dryer drum, bearing temperatures are quite high,
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23-12
Handbook of Lubrication and Tribology TABLE 23.4
Lubrication of Paper Machine Equipment
Machine
Component
Lubricant
Gould pump
Sight glass oiler
HM 68
Gorman Rupp pump
Sight glass oiler
HM 68
Sun Flo pump
Reservoir
Fan pump
Sight glass oiler
Synthetic 5W-30 HM 68
Pulper agitators
Reducer
PMO 220
Headbox slice
Reducer
Fourdrinier wire turning roll Couch roll
Drive gearbox
Synthetic ISO 220 Synthetic 1ISO 220
Drive gearbox
Synthetic ISO 220
Bearings grease lubricated Bearings oil lubricated
Grease
Wet end Wet end
PMO 220
Wet end press section Profile roll
Drives
Synthetic ISO 220
Hydraulic system
HM 100
Dryer bearings
Lube system
PMO 220 or PMO 320
Dryer planetary or hypoid drives
Lube system
PMO 150, 220, or 320
Size press Reel
Worm reducer Hydraulic system
Synthetic ISO 460 HM 46
Lineshaft turbine
Circulation system
HL68
Doctor oscillators Nash vacuum Pump
Reducer Reservoir
Synthetic ISO 460 HM 100
Hood lifts Disc filter Exhaust fans
Reducer Worm reducer Reducer Fan bearings
Exhaust fans
Motor bearings
Synthetic ISO 460 Synthetic ISO 460 Synthetic ISO 220 GC/LB Grease Grease
Transfer pumps
Falk reducer
PMO 220
Winder/unwinder
Hydraulic system
HM 46
Controlled crown roll
Hydraulic system
HM 46
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Notes Inspect daily 6-month drain interval Inspect daily 6-month drain interval Inspect daily 6-month drain interval Inspect daily 6-month drain interval Inspect weekly 6-month sample interval 6-month oil sample interval Inspect daily 12-month sample interval Inspect daily 12-month sample interval Grease monthly with a tacky lithium complex NLGI 2 Inspect sight glasses for oil flow and clarity daily Monthly oil sample interval Inspect daily 12-month sample interval Inspect daily 6-month oil sample interval Inspect sight glasses for oil flow and clarity daily Monthly oil sample interval Inspect sight glasses for oil flow and clarity daily Monthly oil sample interval 12-month drain interval Inspect daily 12-month sample interval Inspect weekly 6-month sample interval 12-month drain interval Inspect weekly 6-month oil sample interval 12-month drain interval 12-month drain interval 6-month oil sample interval Grease quarterly Apply motor bearing grease annually Inspect weekly 6-month oil sample interval Inspect daily 6-month oil sample interval Inspect daily 6-month oil sample interval
Lubrication in the Timber and Paper Industries TABLE 23.4
23-13
Continued
Machine
Component
Controlled crown roll
Circulating system
Reel
Drum drive
Calendar
Nip loading hydraulics
Lubricant
Notes
Synthetic ISO 220 Synthetic ISO 220 HM 68
6-month oil sample interval 6-month oil sample interval Inspect daily 6-month oil sample interval
Source: Ahrens, J.F. and Schlaefer, P., personal communication, 2005.
TABLE 23.5 Pall Filterability Index Test Criteria for Paper Machine Oils Volume passed, ml
Filter length of service
4000 2000 1000
Excellent Acceptable Short
especially when the journals are not insulated [13]. Hooding of the dryer section, which increases bearing exposure to radiant heat from the dryer cans, further exacerbates conditions for dryer bearings [14]. Consequently paper machine oils must be formulated to resist high temperature degradation and associated bearing deposits. Dryer bearing lubricants must be formulated to tolerate water contamination because steam and condensate leaks are a familiar occurrence. In order to prevent bearing corrosion and fatigue as well as avoid an unscheduled machine shutdown for high-water levels, it is common practice to routinely drain water from the bottom of paper machine oil reservoirs. Use of vacuum dehydrators is also a common strategy for coping with water contamination [15]. Since rapid water separation reduces the demands upon vacuum dehydrators and decreases the likelihood of filter by-pass due to additive precipitation, demulsibility is an important characteristic of paper machine oil.
23.3.9 Paper Machine Oil Filterability The filterability characteristic of a fluid can be defined as its ability to pass through a filter without giving rise to undue pressure drop. Excessive pressure drop is undesirable because it can lead to abbreviated filter life. Normally, flow degradation occurs over a period of time as filters accumulate dirt, sludge, and wear debris. When filters become blocked by additives that precipitate out of oil as a result of a chemical reaction with water or other liquid contaminants, filter usage and replacement costs can skyrocket. Several test methods have been developed for evaluating the filterability of lube oils. Filterability tests generally consist of filtering a specified quantity of fluid through a standard medium. The results are typically reported in terms of a ratio between flow rates with and without water in an attempt to compensate for the effect of viscosity on filterability [16]. However, filterability tests designed with hydraulic fluids in mind are generally not well suited for high-viscosity paper machine oils. For instance, ISO 13357-1 : 2002, Petroleum Products — Determination of the filterability of lubricating oils, is limited to a viscosity of up to ISO VG 100 [17]. The Pall Filterability Index for Paper Machine Oils test was developed to evaluate high-viscosity oils [18]. In this test 4000 ml of oil is heated to 145◦ F for 1 h and then pumped through a 3-µm nylon filter membrane. The amount of fluid that can be filtered before reaching a terminal pressure drop of 25 psi is measured. The test is performed with and without the addition of water. By measuring the total volume of fluid filtered rather than the time necessary to filter a specific volume of oil, the effect of fluid viscosity is reduced. The Pall Filterability Index rating’s relationship to filter life is shown in Table 23.5.
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Handbook of Lubrication and Tribology
23.4 Lubricant Selection 23.4.1 Bearing Lubricant Selection Roller bearings and journal bearings are lubricated by different mechanisms. In a journal bearing, a wedge of oil is drawn into the load zone as the shaft rotates. Hydrodynamic pressure created by the lubricant as it is drawn between these converging surfaces produces an oil film barrier that supports the bearing load. Since the journal and bearing surfaces conform to each other, the bearing load is distributed through the lubricant film over a relatively large area. In an antifriction or roller bearing, oil adsorbed on converging surfaces is also drawn into the bearing load zone by hydrodynamic forces. However, the contact surfaces of a bearing roller and race are nonconformal and the load is concentrated at the point or line where they make contact. The high pressure created within these contacts cause oil to undergo an instantaneous exponential viscosity increase. At the same time, surfaces elastically deform under load creating a Hertzian contact zone. As a result, the viscous lubricant is trapped within the bearing load zone and prevents surface contact. This enables a very thin elastohydrodynamic film to support the high loads generated within an antifriction bearing. Pressure distribution within an EHD contact is illustrated in Figure 23.7 [19]. The minimum film thickness, hm , occurs at the outlet of the conjunction zone. The value of hm , in a roller bearing can be determined using Equation 23.1 [20]. Bearing film thicknesses are normally in the 0.5 µm range [21]. This equation shows that film thickness is most sensitive to surface velocity (u), viscosity (µ0 ), and the viscosity–pressure coefficient (α) of the lubricant. The influence of the elastic modulus (E) and load (w) is very small since an increase in load merely increases the size of the Hertzian contact zone. As with conventional hydrodynamic lubrication, the lubricant’s viscosity must be high enough to resist being squeezed out of the converging zone. At the same time the entrainment velocity must be sufficient to continuously replenish the oil film. Generally the EHD film thickness should be 1.5 to 4 times the composite surface roughness to prevent contact of surface asperities. Failure to generate an adequate EHD film will result in a reduced bearing fatigue life [22].
hm = 2.65
(µ0 u)0.7 α 0.54 R 0.43 E 0.03 w 0.13
(23.1)
hm = film thickness at the rear constriction µ0 = viscosity at atmospheric pressure α = pressure–viscosity coefficient u = velocity defined as u = 12 (u1 + u2 ) where u1 and u2 are the individual velocities of the moving surfaces R = the radius of equivalent cylinder w = load per unit width E = elastic modulus if equivalent cylinder (flat surface assumed completely rigid) A variety of simplified methods have been developed for determining the minimum viscosity requirements for antifriction bearings [23,24]. An approximation of the oil viscosity required to achieve acceptable oil film thickness at the bearing operating temperature can be obtained by solving Equations 23.2 or 23.3, where n is the rotational speed in rpm and dm is the mean bearing diameter in mm [25]. The mean bearing diameter equals the average diameter of the bearing OD and ID (dm = (D + d)/2). where
4500 1000 1/3 v1 = √ n ndm 4500 v1 = √ ndm
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for n < 1000 r/min
for n ≥ 1000 r/min
(23.2) (23.3)
Lubrication in the Timber and Paper Industries
23-15
Hertzian pressure
Pressure Inlet region (Pumps film up)
hm Hertzian region (Rides it)
Outlet region (discharges it)
FIGURE 23.7 Pressure distribution within an EHD contact [18].
When applying these equations to paper machine dryer bearings, commonly the viscosity specified by the machine manufacturer is lower than the optimum level for full-film EHD lubrication. This is because process and design modifications that increase machine productivity also tend to cause higher bearing temperatures. When these conditions occur it is desirable to increase oil viscosity to provide enhanced bearing life. Whether or not this is feasible often depends upon the capacity of return lines to drain the oil back to the machine reservoir.
23.4.2 Gear Lubricant Selection The pulp and paper industry’s use of industrial gear drives is extensive. A recent survey of a single North American pulp mill identified 816 individual industrial gear drives [26]. In a nearby paper mill, 335 gear drives were counted among the support equipment for a single paper machine [27]. The majority of the gears used in these applications are enclosed gear drives of the spur, helical, and worm configuration. The contact mechanics of a spur, helical, or bevel gear creates a combination of sliding and rolling friction. When a gear tooth enters the mesh, contacting teeth slide toward each other with a slight rolling action. Sliding diminishes to pure rolling as pitch lines intersect, after which the contacting gear surfaces slide away from each other. Because of the complexity of this interaction, gears can operate under three different modes of lubrication: boundary, mixed, and elastohydrodynamic. Equation 23.1 may be used as a generalized form for calculation of the EHD film thickness in a line contact. In order to extend this equation to gear applications it is necessary to incorporate the mesh angle and transmission ratio, which adds to the complexity of the calculation [28]. An estimate of the required viscosity can be made based upon pitch line velocity using Equation 23.4 [29]. υ40 =
7000 (V )0.5
(23.4)
υ40 = lubricant kinematic viscosity at 40◦ C, cSt. V = operating pitch line velocity, ft/min. V = 0.262dn. d = operating pitch diameter of pinion, in., n = pinion speed, rpm. At low pitch line velocities, it is difficult to generate an effective EHD film. When low speeds are combined with high loads, boundary lubrication prevails. In boundary lubrication, the average film thickness is less than the surface roughness and asperity contact occurs [30]. As a result, friction and wear are dominated by contact surface properties, rather than bulk fluid properties. Extreme pressure (EP) additives are frequently incorporated in gear lubricants for the purpose of reducing friction and wear where
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Handbook of Lubrication and Tribology
TABLE 23.6 Properties of Synthetic and EP Gear Lubricants Property
Test procedure
EP gear oil
Synthetic gear oil
Viscosity index
ISO 2909 ASTM D2270
90 minimum
120 minimum
Oxidation stability
ASTM D2893
<6% Viscosity increase
<6% Viscosity increase
Rust protection
ISO 7120 ASTM D665b
No rust
No rust
Corrosion protection
ISO 2160 ASTM D130
1b
1b
Foam suppression
ASTM D892 Sequence I, 24C Sequence II, 93C Sequence III, 24C
75/10 75/10 75/10
75/10 75/10 75/10
2%
1%
1.0 ml 80.0 ml
2.0 ml 60.0 ml
60 # OK load
Not required
Fail stage >12
Fail stage >10
25 µ wet or dry
25 µ wet or dry
Demulsibility
Load carrying property
Filterability
ASTM D2711 MOD Max. percent water in the oil after 5-h test Max. cuff after centrifuging Min. total free water collected ASTM D2782 Timken test DIN 51 354 FZG test None
under boundary conditions. Typically sulfur and phosphorus compounds serve as EP additives in mineral oil based gear oil formulations. These additives modify the gear tooth surface by forming a tenacious reaction film that prevents scuffing, metal adhesion, and wear [31]. In 1994 the AGMA 250.04 Standard for Industrial Gear Lubrication was replaced by the ANSI/AGMA 9005-D94 American National Standard for Industrial Gear Lubrication [32]. This standard provides lubricant classifications and maintenance guidelines for industrial gearing. At the time of standard development, specifications for EP oils were upgraded to reflect advancements in additive technology. In addition, specifications for synthetic lubricants were incorporated in the new standard. These standards are comparable. However, the load carrying requirements are more demanding for EP mineral oil lubricants than synthetics (Table 23.6).
23.4.3 Synthetic Gear Lubricants Use of synthetic gear lubricants is popular in pulp and paper applications because they can reduce maintenance costs and improve equipment reliability. Synthetic gear lubricants are more oxidation resistant and thermally stable than mineral oil gear lubricants. They also maintain a stable viscosity over wider temperature range. Consequently synthetic lubricants provide the opportunity for extended drain intervals and generate more effective hydrodynamic and elastohydrodynamic lubricating films. Improved energy efficiency may also be realized through the use of synthetic lubricants, especially in worm drive applications (Table 23.7). Relative to mineral-derived lubricants, synthetics have a greater film thickness at high temperatures due to their higher viscosity index [33]. In rolling element bearings this can result in a fourfold increase in bearing life [34]. Polyalphaolefin (PAO) and polyalkyleneglycol (PAG) are the most popular synthetic base stocks for formulating gear lubricants. PAG is synthesized by polymerizing ethylene oxide with propylene oxide. PAO is synthesized by polymerizing C-10 or C-12 alphaolefins. Because these molecules are synthetically derived, they are devoid of wax molecules that hinder low-temperature fluidity and sulfur compounds that compromise oxidation stability [35].
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TABLE 23.7 Comparison of Efficiency of Lubricants in Worm Gear Applications Lubricant ISO 460 Mineral oil ISO 460 Polyalphaolefin ISO 460 Polyalkylene glycol
Viscosity index
Percent efficiency (%)
95 >150 >220
60 68 78
Source: Lauer, D.A., Gear Solutions, August 2004, pp. 21–30.
TABLE 23.8 Typical Test and Service Life of Solvent Refined and Hydrocracked Base Stocks Typical test and service life (h = hours, yr = years, m = months)
Turbine oil, TOST life Steam turbine, service life Hydraulic oil, TOST life EP gear oil, service life
Solvent-refined group I
Hydrocracked group II and III
4000 h 10 yr 2000 h 6m
18000 h 25 yr 6000 h 1 yr
Some of the benefits of synthetic lubricants can be obtained by using hydrocracked base stocks. As can be seen from Table 23.8, the oxidation life of lubricants formulated with these base stocks is at least double that of solvent refined lubricating oils [36]. Extreme Pressure (EP) gear oils exhibit less of a benefit from the use of hydrocracked base stocks because sulfurized EP additives catalyze oil oxidation. For this reason most synthetic gear oils formulated for industrial applications do not incorporate sulfurized EP additives.
23.4.4 Hydraulic Fluid Selection Viscosity is one of the most important criteria in the selection of a hydraulic fluid. A hydraulic fluid that is too low in viscosity will cause low volumetric efficiency, fluid overheating, and increased pump wear. A fluid that is too high in viscosity will cause poor mechanical efficiency, difficulty in starting and wear due to insufficient oil flow. Selecting the proper viscosity fluid requires an understanding of the viscosity requirements of the hydraulic components as well as the operating temperature range. Hydraulic component manufacturers were surveyed regarding the fluid viscosity requirements of their pumps and motors [37]. The majority of equipment was found to provide satisfactory performance with an operating viscosity range of 13 to 860 cSt. Based upon this viscosity range, a Temperature Operating Window (TOW) chart was developed for straight grade hydraulic fluids shown in Figure 23.8. When selecting a hydraulic fluid using TOW criteria, determine the lowest ambient temperature at start up and the highest temperature in use. Any fluid that has a TOW that encompasses this range may be used in the application. Since most of the hydraulic systems within a paper mill incorporate coolers and thermostats, operating temperatures are relatively stable. Consequently it is often possible to consolidate within a paper mill to a single VG 46 or ISO VG 68 hydraulic fluid. In forestry and wood yard applications, the operating temperatures of hydraulic systems are wideranging. Often the use of straight-grade oil is only feasible if seasonal oil changes are performed [38]. Unfortunately it is difficult to thoroughly drain a hydraulic system on a timber harvester or a log loader because upwards of half the fluid can remain trapped within lines and cylinders. As a result the seasonal oil change strategy often yields a mixture of viscosities that is inadequate at low and high temperatures. Multigrade hydraulic fluids may be used in forestry and wood yard operations to eliminate the need for seasonal oil changes. Multigrade oils contain a viscosity index improver that enhances a fluid’s resistance to viscosity change due to temperature. As a result, multigrade fluids have good low-temperature pumpability
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Handbook of Lubrication and Tribology
110
230
100
212 94
90 80
Temperature°C
60
140
64 55
50 40 30
122 104
44
86
32
20
68 +10
10
+4 –2
0
–30 –40
14
–15 –23
–20
–4
–33 10
50 32
–8
–10
FIGURE 23.8
158
73
Temperature°F
70
194 176
84
–22 15
22
32
46
68
100
–40
TOW for 13 to 860 mm2 /sec (cSt), 100 VI hydraulic fluid.
and good high temperature lubricity. Superior viscosity stability of multigrade oils also reduces mechanical energy losses during start-up and volumetric efficiency losses at high temperatures, which can result in enhanced productivity and fuel savings [39]. Selecting the optimum multigrade oil for these applications may be done using an ASTM Viscosity– Temperature Chart using the following procedure: 1. 2. 3. 4. 5. 6.
Determine the minimum and maximum viscosity requirements for system pumps and motors. Estimate the lowest and highest anticipated fluid temperatures in operation. Plot the highest recommended viscosity at the lowest anticipated temperature. Plot the lowest recommended viscosity at the highest anticipated temperature. Draw a line connecting these points. The viscosity requirements of the fluid are defined by the intersection of the line at 40 and 100◦ C.
23.4.5 Biodegradable Lubricants Biodegradable lubricants are increasingly used in forestry applications because of concern over the possibility of lubricants contaminating the environment, especially waterways. Most biodegradable lubricants exhibit two key environmental characteristics: virtual non-toxicity to aquatic life and aerobic biodegradability. In addition, oils that are derived from soybean, rapeseed, sunflower, and other plants are a renewable resource. Since these plants utilize carbon dioxide during photosynthesis as shown in Figure 23.9, they also do not increase green-house gasses in the atmosphere [40]. Organizations such as the Organization for Economic Co-operation and Development (OECD), the Co-ordinating European Council (CEC), and the U.S. Environmental Protection Agency (EPA) have developed standard test methods to determine the toxicity and biodegradability of substances. ASTM has also developed a Guide for Assessing Biodegradability of Hydraulic Fluids (ASTM D6006) and a Classification of Hydraulic Fluids for Environmental Impact (ASTM D6046) based on the above organizations’ methods. Utilizing the methodology from these organizations, standard classifications and performance
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Sustainable lubricant life cycle Used oil
Seed oil
Oilcake
Oil change
Combustion
Animal feed
Seed oil processing
Oxygen
Photosynthesis Carbon dioxide
Straw
Seed oil plant
Soil
Water
Nutrients
FIGURE 23.9
Sustainable lubricant life cycle, adapted from H.F. Eichenberger [40].
TABLE 23.9 ISO Environmental Hydraulic Fluid Classifications Symbol
Classification
Commercial designation
HETG HEES HEPG HEPR
Vegetable oil types Synthetic ester types Polyglycol types Polyalphaolefin types
Vegetable oils and natural esters Polyol esters, neopentylglycols, synthetic adipate esters Polyglycols Polyalphaolefins (PAO) or synthetic hydrocarbons (SHC)
requirements for environmental fluids have also been established by the International Standards Organization (ISO). ISO environmental hydraulic fluid classifications are described in Table 23.9.
23.4.6 Vegetable Oil Based Lubricants — HETG Type HETG fluids are based on naturally occurring vegetable oils. Rapeseed (canola), soybean, and sunflower oils are the most popular vegetable oil base stocks. Vegetable oils are triglyceride esters. Like all esters, vegetable oils are composed of a molecular combination of an alcohol and carboxylic acids. In vegetable oils, glycerin is the alcohol and the carboxylic acids are genetically determined combination of saturated and unsaturated fatty acids. Many of the plants grown for oil seed production have been bred or genetically engineered for enhanced oxidation stability and low-temperature fluidity [41]. The goal of such breeding is to maximize the monounsaturated fat content of the oil seed. Polyunsaturated fats are undesirable in lubricant applications because they readily oxidize and crosslink at high temperatures. Saturated fats are undesirable because they have a high melting point. (Vegetable shortening is solid at room temperature because it contains a high percentage of saturated fats.) A comparison of fatty acids is presented in Table 23.10. The characteristics of these fatty acids directionally correspond to characteristics of their triglycerides. Since vegetable oils are composed of a mixture of triglycerides they exhibit a wide range of melting points and oxidation resistance.
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23-20
Handbook of Lubrication and Tribology TABLE 23.10 Acids
Relative Rates of Oxidation and Melting Points of Fatty
Fatty acid
Class
Chemical abbreviation
Relative oxidation rate
Melting point, 0◦ C
Stearic Oleic Linoleic Linolenic
Saturated Monounsaturated Polyunsaturated Polyunsaturated
C18 C18 : 1 C18 : 2 C18 : 3
0.6 6 64 100
69.6 14 −5 −11
Source: Adapted from Honory, L., Handbook of Hydraulic Fluid Technology, Totten, G.E., ed, Marcel Dekker, New York, p. 1133.
TABLE 23.11
Properties of Biodegradable Synthetic Esters
Ester di-2-Ethylhexylazelate di- Isotridecyladipate NPG dioleate NPG diisostearate TMP trioleate TMP triisostearate PE tetraoleate
100◦ C, cSt 2.9 5.3 6.0 7.7 9.6 12.3 12.4
Viscosity at 40◦ C, cSt −40◦ C, c P 10.3 26.7 24.3 41.2 47.5 82.6 65.6
810 21,700 — 1,500 — 4,300 —
Biodegradability, CEC L-33-A-94 >95% >90% >90% — > 90% > 90% > 90%
Source: Cognis Tech data sheet #6D-7/2000, July 2000. Tribology Data Handbook, Booser, p. 46. 1997.
HETG fluids biodegrade rapidly, exhibit excellent natural lubricity, and have a high-viscosity index. However, all natural oils contain polyunsaturated and saturated fatty acids that limit their effectiveness to an approximate range of −20 to 65◦ C. Since timber harvesting equipment can operate at temperatures well beyond these limits, HETG oils are of limited utility in woodland applications.
23.4.7 Synthetic Ester Based Lubricants — HEES Raw materials for the production of synthetic esters may be derived from natural or petrochemical sources. Because of the array of alcohols and fatty acids available for production of synthetic esters, HEES fluids have a broader range of physical properties than HETG oils. As with HETG oils, the oxidation stability of an HEES is mainly determined by the degree of saturation of the fatty-acid moiety. Saturated adipic, 2-ethylhexanoic and stearic acids are commonly used in ester synthesis. Monounsaturated oleic acid is also used in ester synthesis because torsion created by the double bond lowers the melting point and enhances low-temperature fluidity. Further enhancements in oxidation stability are made possible by selecting a polyglycol with a substituted β carbon such as trimethylolpropane, pentaerythritol, or neopentylglycol. The properties of common synthetic esters are listed in Table 23.11. HEES fluids biodegrade rapidly, exhibit excellent lubricity, possess good oxidation resistance, are low in volatility, and have a high-viscosity index. However, the ester linkage that provides a site for microbes to begin biodegradation is also vulnerable to hydrolysis. In hydrolysis, an ester is cleaved into its parent carboxylic acid and alcohol as a result of a chemical reaction with water. Since water solubility in synthetic esters is high in comparison to mineral oil, degradation due to hydrolysis is an important consideration when using these fluids [43]. For satisfactory fluid life the upper limit for in-service water content in a synthetic ester is 0.2% [44]. The performance of HEES lubricants is more than satisfactory as long as this level of water contamination is not exceeded [45].
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23.4.8 Lubrication Management Strategies Lubrication and equipment maintenance not only have an impact upon productivity in the pulp and paper industries but they are also significant factors in determining product quality. Many maintenance professionals in the pulp and paper industry have adopted concepts that are popular in Total Quality Management (TQM) programs. These concepts are sometimes described as Reliability Centered Maintenance or Total Productive Maintenance (TPM). The goal of the TPM program is to increase production while increasing employee morale and job satisfaction. TPM focuses on equipment maintenance as a strategic element part of a business’ competitiveness. Maintenance is not regarded as an inconvenience or simply an overhead. Rather, equipment maintenance is incorporated into the production schedule in order to minimize unscheduled downtime. At the same time employees on the shop floor are encouraged to take initiative in equipment lubrication, maintenance, and process improvement [46]. TPM can have a significant impact on productivity and profitability. At a large North American paper mill where downtime is valued at $10,000/h, implementation of TPM in combination with enhanced sealing technology leads to a reduction from 22 to 23 h of lost time per month to 14 h per month [47]. Employee training and management commitment are requisite elements of implementing TPM.
23.4.9 Oil Analysis Whereas employee empowerment and enhanced cooperation between production and maintenance departments are crucial to the success of TPM, oil analysis, thermography, vibration analysis, and computer assisted lubrication scheduling have become valuable technological tools. Often thermography, vibration, and oil analysis are viewed as early warning systems to predict equipment problems. While this early alert system can reduce unscheduled downtime, these tools may also be used to: • • • •
Extend oil change intervals Identify the root cause of problems Prescribe corrective action Assess the effectiveness of the remedy
One of the most effective ways to maximize oil life is to use oil analysis to scientifically determine when oil should be changed. Tests such as viscosity, infrared, and spectrochemical analysis can be used to identify the point at which oil begins to degrade, necessitating a scheduled oil change. Contamination is measured using automated laser particle counters and Karl Fisher titration. Many industrial lubricants are formulated with base oils that have been catalytically processed to remove aromatic and sulfurized chemical impurities. Removal of these compounds can double the useful life of a lubricant. The most common problem identified in oil analysis is contamination with dirt or moisture. Degradation due to oil oxidation or additive depletion is relatively rare, although sometimes degradation can be severe enough to warrant an oil change. More often than not, the lubricants used in the pulp and paper industry can be refurbished while in service through auxiliary side-stream filtration or the use of a vacuum dehydrator. Thus changing oil based upon standard time intervals can amount to an unnecessary expense.
References [1] Stenzel, G., Walbridge, T.A., and Pearce, J.K., Logging and Pulpwood Production, 2nd ed., WileyInterscience Publication, New York, 1985, chap. 6. [2] Rushton, T., Brown, S., and McGrath, T., Impact of tree-length versus short-wood harvesting systems on natural regeneration, Report FOR 2003-3, No 70, Nova Scotia Department of Natural Resources, 2003. [3] Culhane, J., personal communication, 2004. [4] Mettler, D., Tech Update: Forwarders, Logging and Sawmilling Journal, 35(3), 50–54, 2004.
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[5] Carnes, K., Offroad hydraulic fluids: beyond biodegradability, Tribology and Lubrication Technology, 60, 32, 2004. [6] Smook, G.A., Handbook of Pulp and Paper Technologists, Angus Wilde, Vancouver, 1992, chap. 9. [7] Smook, G.A., Handbook of Pulp and Paper Technologists, Angus Wilde, Vancouver, 1992, chap. 17. [8] Hink, R., Best practice lubrication for paper machines — why bad things happen to good bearings, Machinery Lubrication, 200501, 60, 2005. [9] Hink, R., Best practice lubrication for paper machines, Machinery Lubrication, 200403, 60, 2004. [10] Michael, P., What is grease? Plant Services Magazine, Volume 13, No. 10, October 1992. [11] Mistry, A., Performance of lubricating greases in the presence of water, NLGI Spokesman, 68, 8, 2005. [12] Duchowski, J.K., Messerschmitt, A.P., Needham, T., and Collins, K.G., Improvements in equipment reliability and machine performance as a result of contamination control and monitoring procedures at a major pulp and paper facility, Tribology and Lubrication Technology, 60, 56, 2004. [13] Bergling, G., Effect of insulation and lubrication on the operational reliability and temperature of rolling bearings for steam-heated papermill cylinders, Ball Bearing Journal, 233, 1, 1989. [14] Burns, B.L., Carmichael, J.D., and Bogenholm, C.A., Paper machine bearing lubrication problems and solutions — a case history, Lubrication Engineering, 33, 173, 1976. [15] Williamson, M., Options for water removal, Practicing Oil Analysis, July–August 2003, pp. 48–53. [16] Givens, W.A. and Michael, P.W., Hydraulic fluids, in Fuels and Lubricants Handbook, Totten, G.E., Westbrook, S.R., and Shah, R.J., eds, ASTM, West Conshohocken, PA, 2003, chap. 13. [17] ISO 13357, Petroleum products — determination of the filterability of lubricating oils — part 1: procedure for oils in the presence of water. [18] Day, M., Filterability testing of paper machine oils, Machinery Lubrication Magazine, November 2001. [19] Wedeven, L.D., What is EHD? Lubrication Engineering, 31, 291, 1975. [20] Dowson, D., “Elastohydrodynamics,” Proceedings of the Institution of Mechanical Engineers., 182, Part 3A, 151–167, 1967–68. [21] Snyder, D.R., Selecting rolling element bearings for modern applications, Tribology and Lubrication Technology, 60, 28, 2004. [22] Cheng, H.S., Elastohydrodynamics and failure prediction, Lubrication Science, 2, 133, 1990. [23] Jackson, A., A simple EHL film thickness equation for rolling element bearings, ASLE Transactions, 24, 147, 1980. [24] SKF Bearing Installation and Maintenance Guide, p. 65, 2001. [25] Bergling, G., Effect of insulation and lubrication on the operational reliability and temperature of rolling bearings for steam-heated papermill cylinders, Ball Bearing Journal, 233, 1, 1989. [26] Ahrens, J.F., Unpublished data, 2004. [27] Schlaefer, P.J., Unpublished data, 2005. [28] Bala, V., Gear lubricants, in Fuels and Lubricants Handbook, Totten, G.E., Westbrook, S.R., and Shah, R.J., eds, ASTM, West Conshohocken, PA, 2003, chap. 16. [29] Errichello, R., Lubrication of gears — part 4, Lubrication Engineering, 46, 231,1990. [30] Hsu, S.M., Boundary lubrication: current understanding, Tribology Letters, 3, 1, 1997. [31] Tysoe, W.T. and Kotvis, P.V., Surface chemistry of extreme-pressure lubricants additives, in Surface Modifications and Mechanism, Totten, G.E. and Liang, H., eds, Marcel Dekker, New York, 2004, chap. 10. [32] ANSI/AGMA 9005-94: Industrial Gear Lubrication, AGMA, Alexandria, VA, August 1994. [33] Moore, L.D. et al., PAO based synthetic lubricants in industrial applications, Lubrication Engineering, 59, 23, 2003. [34] Siebert, H. and Mann, U., Gear Solutions, Volume 2, No. 3, March 2004, pp. 20–27. [35] Grega, G., Making synthetics work, Lubes ‘N Greases, 10, 22, 2004. [36] Khonsari, M. and Booser, E.R., New lubes last longer, Compoundings, 55, 21, 2005.
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[37] Michael, P.W., Herzog, S.N., and Marougy, T.E., Fluid viscosity selection criteria for hydraulic pumps and motors, Proceedings of the 48th National Conference on Fluid Power, NFPA, Milwaukee, 2000, 313. [38] Placek, D., Study examines multigrade fluids for forestry equipment, Hydraulics and Pneumatics, 54(3), 39, 2001. [39] Herzog, S.N. et al., Effect of operation time on oil viscosity and pump efficiency, NCFP I05-9.3, IFPE March, 2005, Las Vegas, NV, USA. [40] Eichenberger, H.F., Biodegradable hydraulic lubricant: an overview of current developments in central Europe, 42nd Earthmoving Industry Conference, Peoria, SAE 910962, 1991. [41] Maelor Davies, H. and Flider, F., Designer oils, Chemtech, April 1994, pp. 33–37. [42] Buenemann, T.F. et al., Non-aqueous synthetic lubricants, in Fuels and Lubricants Handbook, Totten, G.E., Westbrook, S.R., and Shah, R.J., eds, ASTM, West Conshohocken, PA, 2003, chap. 10. [43] Kempermann, C. and Murrenhoff, H., Reduction of Water Content in Biodegradable and Other Hydraulic Fluids, 1998 Earthmoving Conference, Peoria, SAE 981497, 1998. [44] Gere, R.A. and Hazelton, R.A., Polyol ester fluids, in Handbook of Hydraulic Fluid Technology, Totten, G.E., ed., Marcel Dekker, New York, 2000, chap. 19. [45] Young, K.J. et al., Environmental standards for biodegradable hydraulic fluids and correlation of laboratory and field performance, Lubricants for Off-Highway Applications, SAE SP-1553 (2000-012543), pp. 15–21. [46] Venkatesh, J., An introduction to total productive maintenance, Plant Engineering, February 2005. [47] Nooyen, J.G., personal communication, 2004.
© 2006 by Taylor & Francis Group, LLC
24 Textile Fibers/Fabrics 24.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 24.2 Friction and Lubrication . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
24-1 24-2
Fundamental Concepts • Testing Methods • Experimental Observations
24.3 Textile Processing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
24-11
Spinning Process • Texturing Process • Knitting Process • Weaving Process
24.4 Fiber Finishes and Components . . . . . . . . . . . . . . . . . . . . . .
Paul D. Seemuth Tribology Consulting International LLC
24-16
Formulated Finish Package • Lubricants • Surfactants (Emulsifiers) • Antistats • Other Finish Components • Coning Oils
Acknowledgments. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
24-26 24-27
24.1 Introduction The introduction of synthetic polymers has brought forth a wide range of applications to the modern world. Along with the various uses of these systems, tribology, the science of friction and wear, has had a pronounced effect on the knowledge of the properties achieved and the end-use possibilities for these materials. A major application of synthetic polymers influencing the world’s everyday life is textile fibers and related textile constructions. Textile materials are made from a variety of polymer types. The widely known textile related polymers are polyesters, polyamides, polyolefins, polyurethanes, and aramids. Fiber manufacturers produce a variety of chemical compositional forms within each class and introduce unique additives for special use characteristics. One fundamental challenge faces all fiber producers. These materials experience physical contact with numerous surfaces as well as rub against themselves. In contact with other fibers, other metal, or plastic surfaces, resistance to movement across the surface is noted and defined as friction. In a typical processing step, the number of surfaces and their nature, metal, hot, cold, smooth, rough, act as a destabilizing combination of friction forces. Therefore, the frictional forces imposed on these soft “plastic-type” materials frequently result in a number of product defects. To produce a quality product, friction needs controlling in a way to give uniform stress loads during use. A uniform and balanced friction load during processing eliminates most defects, that is, broken filaments, nonuniform fiber structure, static, and finally package formation. Unlike their naturally lubricated fiber counterparts, wool, cotton etc., the synthetic fibers require lubricants to control friction that is crucial to a formation of a quality fiber and its subsequent use. Therefore, application and extension of the principles learned in other business, that is, natural fiber and metalworking, were crucial for these new materials. 24-1
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
During the early 1900s, textile lubrication, fiber finishes that are combinations of a multitude of different chemical components, became key to the successful commercialization of the new textile fiber business. However, the unique character of synthetic polymers needed more than the old lubrication fluids of the past. Maintaining good yarn color, meeting processing demands, and imparting a soft nature to the fibers themselves demand new lubricants and auxiliaries that are suited to the production conditions and final use of these fibers. To overcome many processing issues, fiber finishes are normally applied as water-based emulsions. These emulsions can be used at concentrations from 0.1% active to >50% (typically 5–20%), dependent on the fiber type, process, and processing speed. Typical loading on a fiber is from 0.05% (staple) to 3–5% for polyurethane systems. Further, the finish must combine a number of physical properties, good “uniform” wetting on fiber surface and associated fiber bundle (wicking), low aerosol generation (environmental, health, and safety), static dissipation character, and finally good overall balanced frictional properties. These parameters are only a small set of the physical characteristics that the fiber finish must incorporate for use on the fibers. Improper cohesive forces may result in poor staple yarn spinning. It may result in a package of yarn being too hard or soft resulting in performance problems during knitting or weaving. “Sluffing,” multiple package yarn layers being removed at the same time, and tension plucks, high friction forces between the fibers themselves or filaments trapped underneath each other encountered during yarn removal from package, are common quality problems associated to poorly formulated finish systems. The list of finish characteristics is long and complex and many features are discussed later. Today’s concept of the finish must be an integral system component of the product. One will discover during work on fibers that the surface lubricant system is just as crucial as the underlying “solid” substrate for the final product, be it a garment to wear or a life saving bullet proof vest.
24.2 Friction and Lubrication 24.2.1 Fundamental Concepts Friction is defined as the resistance of movement of one body moving against another. This concept covers many contributing features of friction, apparent contact area, relative speed, sliding, and rolling cases. From Leonardo da Vinci’s work (1452–1519), the first law of friction evolved three key points: 1. Friction is independent of the area of the solids. 2. Friction is directly proportional to the load that is normal in the direction of movement. 3. Frictional force is independent of the speed. Therefore, the first law is expressed in Equation (24.1) as F = µW
where F = Friction and W = Load
(24.1)
This amounts to the coefficient of friction, µ being a constant, independent of load (W ). This general assumption works for most situations in first approximations though the frictional properties exhibited by real dynamic systems show variations from these simple laws. Work in the later 1800s and 1900s refined the knowledge of the frictional forces that are in play during the contact of two surfaces, normally one stationary relative to the second body. Second and Third Laws of Friction have been developed and studied and reviewed extensively [1]. Various authors [2–8] have described fiber friction. This frictional behavior is illustrated by Stribeck’s curve shown in Figure 24.1. The forces experienced by a fiber fall into four different regions. Of primary importance to most textile processes are the boundary and hydrodynamic regions. Both of these regions affect most of the properties of the yarn though the extent is variable over a considerable range. Examining Figure 24.1, boundary lubrication (boundary frictional forces) diminishes as the speed of the yarn accelerates. Similarly, hydrodynamic frictional forces increase though they do reach a temporary stable value plateau
© 2006 by Taylor & Francis Group, LLC
Textile Fibers/Fabrics
24-3 Basic friction regions
Boundary region
Coefficient of friction
Semi-boundary region
Hydrodynamic region
Solid-like region
[(Thread line speed) (Lub viscosity)] Pressure
FIGURE 24.1 Stribeck’s curve — general representation of frictional behavior of lubricated textile yarns (updated from Reference 21, p. 5, courtesy of Marcel Dekker, Inc. and Reference 17, p. 117, courtesy of Kao Corporation).
Low inter-molecular shear forces allowing lubricant to shear with itself prior to interfacial shear
Absorption of boundary lubricant to substrate by intermolecular forces (Van der Waals)
FIGURE 24.2 Representation of sheared lubricant layer under boundary conditions.
as the speed increases. The boundary and hydrodynamic region overlap at a minimum and this “semiboundary region,” a combination of both regions, operates in ill-defined manners. The main two regions are characterized and actually controlled independently using proper component selection by a finish formulator. Boundary lubrication (Fy−y or Fy−m ) is managed by application of an intramolecularly low shear material on the fiber surface [9–12]. This low shear material has a reasonable high affinity to stay “fixed” onto the polymer surface and it functions by shearing layers of itself from one another rather than shearing from the surface. Figure 24.2 best illustrates this. Under boundary conditions, the textile polymer is in close intimate contact with the other surface, be it another filament or an external contact point. During movement of the fiber, the lubricant film adheres to the fiber, the lubricant itself shears intramolecularly which accounts for low friction forces under highpressure and low speed conditions. Furthermore, this effect is necessary for damage prevention of the delicate fiber surface and uniform distribution of the stress load across each filament in a fiber bundle. The semi-boundary regions (stage two) are the less understood of the friction regions. Here, the relative speed change of the two surfaces introduces multiple interactions of both boundary and dynamic friction parameters and reaches a minimum value at this stage [2]. The third stage is the dynamic region (Fh ). It is here that textile processing requires considerable attention. The two bodies are in a high degree of relative motion, one normally stationary to the moving
© 2006 by Taylor & Francis Group, LLC
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Low shear strength at fiber–lubricant interface allowing fiber to ride (surf) on the lubricant “pool”
FIGURE 24.3
Representation of sheared lubricant layer under hydrodynamic conditions.
body, or fiber. This region is highly sensitive to lubricant viscosities, its inherent resistance to shear [13,14]. This is illustrated in Figure 24.3. In this region, the fiber glides on a “pool” of lubricant across the other surface. The shear point is at the fiber/lubricant interface, as the fiber/lubricant attractive forces are minimal. For the fiber finish formulator, the combinations of components serving multiple functions must be factored together to understand the final dynamic frictional values obtained. In the fourth region, hydrodynamic friction at very high speeds, work by Kao Corporation [17] in Japan suggests that the frictional forces start to decrease in what is termed the solid-like region. For most new textile processes, these domain speeds are now being explored further. Several hypotheses are presented to account for the decrease in friction. Friction decreases with speed for these possible reasons: 1. Speed results in slippage between the lubricant layer and the friction contact vs. lubricant shear. 2. Dynamic mechanical properties of the lubricant vs. the steady-state lubricant viscosity controls friction [14]. 3. High temperatures produced at extreme speeds further reduce the apparent lubricant viscosities, a controller of frictional forces.
24.2.2 Testing Methods Evaluation of frictional properties continues to receive considerable attention. The major work of the 1950s to 1990s focused on the macro scale for friction measurements [2–15]. The past 15 years has brought more focus on the microscopic as well as the nanoscale properties of frictional behavior [19]. All these areas are fundamentally crucial to the understanding of behavior; however the textile industry relies heavily on the macroscale effects. These effects are readily measurable and relate well to the daily processing behaviors. With the dependence on the practical aspects of friction for everyday processing, it is crucial to understand the available tools and testing protocols allowing rapid and consistent testing of the lubricants, surfactants, and other additives that are addressed later in this chapter. Evaluation of the fiber’s frictional performance containing a lubrication package requires the finish application onto a substrate of interest. Lawson-Hemphill provides a commercial yarn re-winding unit, a Precision Lab Winder, and the simple inclusion of a finish pump or syringe pump allows an effective adaptation for accurate finish application to a substrate. The best choice of substrate is specifically the polymer material that is under examination, be it polyester, polyamide, aramid, polyolefin, or some other type. There is always the dilemma for the formulator to have access to a suitable material containing no surface lubricants or other add-ons. However, frictional assessment within a fiber substrate set using a working process control provides an excellent relationship to another substrate set. Studies have been done which
© 2006 by Taylor & Francis Group, LLC
Textile Fibers/Fabrics
24-5
0.7 0.6 0.5
Fh
0.4 0.3 0.2 0.1 0 PET
Nylon
Cellulose acetate
Poly acrylonitrile
None
Hexadecane
Butylstearate
Oleic acid
Ethyl stearate
Octadeyl alcohol
Stearic acid
Octadecylamine
Solid stearic acid
Solid octadecylamine
FIGURE 24.4 Frictional response of different polymer surfaces (adapted from Reference 10, p. 1143, American Chemical Society).
illustrate this point. Schick studied the effect of substrate on friction. Though he was examining other attributes and these must be taken into account for the relationship to be binding, the effect is translatable [9]. The study shows that each polymer substrate has a different frictional response (Figure 24.4). Considering this, a single available control fiber material, using a control lubricant package, allows formulation of the systems to meet effectively the demands to a polymer type and its processing. Boundary (Fy−y ) and hydrodynamic (Fy−m or Fh ) friction are evaluated a number of ways. Commercial equipment for running frictional analyses can be obtained that is highly accurate with good precision. During the past years, many of the original frictometers have been discontinued or updated with modern computer data analysis features. Rothschild, Zurich Switzerland, markets the more notable of the newer instruments, upgraded from its long-term line of instruments devoted to the textile industry. LawsonHemphill Inc. provides a new design frictometer that allows improved pre-tension control during frictional testing. Both frictometer systems offer an excellent speed range for hydrodynamic (Fh ) evaluation up to 1000 m/m or 550 m/m respectively (Figure 24.5 and Figure 24.6). A schematic drawing of the Rothschild unit is shown in Figure 24.7. Friction is calculated from the input tension T1 and the resulting output tension T2 . These values give the coefficient of friction using the following Equation (24.2) [2,20–22]. T2 /T1 = eθ µ
(24.2)
Written in a more useful form, the coefficient of friction is expressed in Equation (24.3): µ=
ln(T2 /T1 ) θ
(24.3)
where θ is the wrap angle expressed in radians [3,23,24]. Graphical representations of the friction coefficient are best expressed by plotting the output tension, T2 , against increasing speed. One important reason for this is the calculation for µ will give the same
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
FIGURE 24.5
Rothschild frictometer R-2088 (photo courtesy of Rothschild Instruments, Zurich, Switzerland).
FIGURE 24.6
Lawson-Hemphill constant tension transport (photo courtesy of Lawson-Hemphill, Inc.).
number over a range of T2 values (at a given T1 input level). This is especially true under high T1 inputs. The false range of calculated µ masks the real frictional differences of experimental lubricant systems, which, if one translates the in-lab data to an operational process, will lead to inconsistencies in the results expected.
© 2006 by Taylor & Francis Group, LLC
Textile Fibers/Fabrics
24-7 T1 input tension head
T2 output tension head
u u expressed as radians for 1708
u u expressed as radians for 808
FIGURE 24.7
Thread line diagram Rothschild frictometer R-2088.
Using these calculations and graphical representations, friction performance can be evaluated for a set of lubricant systems. Additionally, the Rothschild and Lawson-Hemphill equipment allows modified setups so that different friction surfaces, process temperature environments, and contact angles are available in the thread path. The friction pin surface can be readily changed to a variety of chrome, ceramic, and other surface types along with modifications of surface roughness that will give even better experimental basis for textile operations and associated process behaviors. This provides detailed analyses of a finish lubricant’s properties under a variety of processing conditions. These frictometers have the capability of low speed operations for analysis of the boundary friction region. However, most of the boundary friction test methods, usually a three-twist configuration to the friction test surface, are limited to the types of yarns being studied [25–28]. POY (Partially Oriented Yarns) will undergo stretching during testing giving variability in friction values. FDY (Fully Drawn Yarns) provide the best overall stability under this three-twist configuration. In the early 1960s, a highly reliable method was developed by workers at DuPont [2,12]. With many improvements over the years, the primary equipment for this testing still is custom-made due to the nature of the testing, the importance placed on understanding this friction region, and the unique friction relationships crucial to the process. A typical unit is shown in Figure 24.8. The Capstan method uses a rotating mandrel containing a wound tube of yarn to move at speed ranges of from 0.0016 to 32 cm/sec. Over the mandrel is draped a single strand of fiber (from the same yarn/lubricant test source) and is attached to a strain gauge and held under constant T1 . The mandrel rotates and the frictional forces are recorded on a strip chart recorder or are recorded digitally. Using this method, the phenomenon of stick slip is easily observed. Stick slip is a characteristic of the overlapping nature of boundary, semi-boundary and hydrodynamic lubrication regions under slow speed conditions as seen in Figure 24.1. The effect that arises is a combination of several chemical and physical factors: 1. Shear strength of the fluid film 2. Adhesion forces between fibers (at dry contact regions resultant for nonuniformities in film thickness) 3. Van der Waals interactions (any interactions between uncharged molecules) These effects diminish as these forces are overcome and completely dampened with increasing speed. Figure 24.9 illustrates this. A good balance of the stick-slip friction magnitude will enhance the ability of the yarn to form a good quality yarn package and assist in uniform delivery of yarn off the package.
© 2006 by Taylor & Francis Group, LLC
24-8
Capstan frictometer (photo courtesy of INVISTA, Inc.).
Fm
FIGURE 24.8
Handbook of Lubrication and Tribology
10–3
10–1
10 Increasing speed
102
FIGURE 24.9 Typical stick-slip phenomena representation (reprinted from Reference 7, p. 105, courtesy of Textile Research Journal).
In the evaluation of yarns, there are other parameters that one needs to access. Static is a principle factor, along with friction, to control in many processes. Finish systems will have varying degrees of effectiveness to dissipate static charge that is built-up as fiber runs over guide surfaces and rubs against itself. Various methods have been outlined for assessing static charge dissipation with a fiber finish. One method was outlined by Schick [29]. The yarns are wound onto skeins and a voltage of 160 V was applied. The time is then measured for a drop to half value of applied voltage. Other methods have been applied for measurement of antistatic properties. Usually these results will be given as resistivity or as log Rp . In Figure 24.10, the general region for good static control in knitting and staple process is below log Rp of 9. A mid range of 9 to 11 will handle processes such as texturing. Once the value starts to rise above 11, static generation problems become a severe concern. While this serves as a guideline, measurement using the induced voltage charge and relaxation time to 12 values is an excellent basis. Under hydrodynamic
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Textile Fibers/Fabrics
24-9 14 13 12
Log Rp
11 10 9 8 7 6 5 excellent
good
fair
poor
Degree of antistat protection
FIGURE 24.10
General relationship of log R vs. static protection performance.
conditions, the Rothschild R-2088 unit has the capability to measure static charge generated at the friction pin surface using an electrostatic voltmeter. The Lawson-Hemphill CTT units have also adapted the frictometer in various ways to measure the electrical charge generated at the guide surface. As the charges are small, typically pico-ohms, consistent measurements require that care be taken to control the equipment’s environment and conditions of testing. Running the yarn at various speeds gives an accurate picture of the finish’s ability to dissipate charge under hydrodynamic conditions. A well-defined relationship between hydrodynamic friction, charge dissipation, and finish properties is achieved by collecting data from 50 to 1000 m/m.
24.2.3 Experimental Observations Numerous parameters influence the frictional response of a fiber bundle as it is moved over guide surfaces and the frictional effects are noted from the time the polymer exits the spinneret for fiber formation until the final fabric or garment is packaged for use. Several of these have been covered — speed of the yarn during process, process temperature, contact area between yarn and guide, tension or pressure at contact point which is a function of the contact angle influenced by guide diameter. Others include denier of fiber, finish viscosity and related physical properties, yarn luster (different surface roughness due to delusterant inclusion), and guide roughness [7,15,16]. These contributors play major roles in the overall requirements for lubrication balance during processing of synthetic fibers. Others that contribute to processibility are relative humidity and environmental temperatures that affect static propensity, moisture regain, and plasticization of a fiber’s surface (especially with nylon) [30–33]. Surface roughness is a major outside influence that affects the frictional forces experienced by a fiber [7]. While the focus is normally directed to the fiber system, that is, polymer and lubricant (finish), the roughness of the guide surface has as much influence as the changing luster of the fiber itself. As noted earlier, for a selected group of polymer types, PET friction was lowest of the group studied (see Figure 24.4). In the same way, within a polymer type, the hydrodynamic frictional forces are lowered for a fully dull yarn vs. a bright yarn. This is due to the decrease in surface contact and proposed theory on disruption of the fluid film between the guide and the fiber [1,4,6]. Reversing the situation by changing the roughness of the guide gives a similar effect as illustrated in Figure 24.11 [2,12,18]. With a given polymer type, increase guide roughness decreases the coefficient of friction quite dramatically. A drop of 10 to 60+% can be seen. When this is translated into a textile operation, guide roughness or multiple guides of different roughness along the running thread path can alter the fiber’s performance. It is in this world of nonuniform guides that the balance of frictional forces through lubrication is most demanding.
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24-10
Handbook of Lubrication and Tribology 4.6 micro inch Chrome
Coefficient of friction
8 micro inch Chrome 40 micro inch Chrome 40 micro inch Alsimag
90 micro inch nodular Chrome
Speed in yards/min
FIGURE 24.11 Frictional effect of guide roughness modification (data courtesy of INVISTA Inc.). (From Olsen, J.S., Frictional behavior of textile yarns, Textile Res. J., 39, 31, 1969).
Viscosity cps at 258C
120 100 80 60 40 20 0 0.5
0.6
0.7
0.8
0.9
1
1.1
1.2
1.3
1.4
Coefficient of friction
FIGURE 24.12 Relationship of viscosity to hydrodynamic friction for 25 commercial fiber finishes (Seemuth unpublished data, courtesy of INVISTA Inc.).
Viscosity of the lubricant package is the other major influence on frictional forces [13]. As noted in Figure 24.1, friction forces are directly relatable to the viscosity of a finish. While there are extenuating cases, first principles and numerous studies have shown a good relationship of hydrodynamic friction to viscosity measured using a Brookfield viscometer. A good example of this relationship is shown in Figure 24.12. A study of the friction vs. FOY on 40-13 dull nylon further illustrates the viscosity effect. Friction starts to reach a plateau at FOY levels of 0.8 to 1.0% (Figure 24.13). This is attributable to the increased shear stress of the total fluid film volume separating the surfaces, that is, a leveling out of the shear forces where lubricant‘s viscosity reaches an equilibrium value [13,20,22]. Once the plateau is reached, additional lubricant levels afford no additional value. One other premise in this explanation of the frictional curve profile is that the wetting of the lubricant on the surface forms a continuous fluid film in accord with the thermodynamics of wetting [34–37]. If the film was nonuniform, the friction forces would be governed by fiber surface properties and discontinuities of the fluid’s film. With modern textile operation reaching speeds rapidly approaching 6000 m/m, effects of temperature play an important consideration in the balance of lubrication properties. Viscosity decreases as the temperature increases. While there are limits to the viscosity value reached, temperature of the process can alter
© 2006 by Taylor & Francis Group, LLC
Textile Fibers/Fabrics
24-11 0.8
0.75 0.7 0.65 Fh
0.6 0.55 0.5 0.45 0.4 0.35 0.3 0.1
0.24
0.37
0.47
0.69
1
1.55
2
% FOY
FIGURE 24.13 Hydrodynamic friction vs. FOY on 40-13 dull nylon (data courtesy of INVISTA Inc.). (From Olsen, J.S., Frictional behavior of textile yarns, Textile Res. J., 39, 31, 1969.)
the performance. This temperature effect is also dependent on the molecular weight and thermal stability of the lubricant. Increasing temperature using low molecular weight lubrication of low thermal stability will result in changing composition and lubrication character as the lubricant volatilizes and decomposes. With the high speeds and increasing temperatures of processing, thermally stable lubricants are finding increasing use for textiles. These stable lubricants lower the rate of change of the compositional make up of the lubricant package and stabilize the frictional forces on the yarn during process transient. This effect is the result on less percentage loss of the lubricant, less decomposition of lubricant, and retention of the associated frictional properties. It is evident that the friction regions, especially boundary and hydrodynamic, are distinct for a synthetic fiber running against itself or an external surface. It should be further pointed out that while there is overlap of the regions, detailed work on the primary regions, boundary and hydrodynamic, provide most of the crucial information for controlling the friction balance necessary for a textile fiber to perform to it’s full capability during processing. While various references and researchers [2,23,29] provide slightly diverse ranges for the different regions, a general guide is presented. Boundary friction is attributed to speeds under 0.1 m/min. The semi-boundary region range is from 0.1 to 5.5–10 m/min. The hydrodynamic region is speeds exceeding the 5.5–10 m/min. While I do not understate or underestimate the importance of the underlying polymer structure or the external guides and processing conditions, textile lubrication researchers must realize that the fiber lubricant is an integral part of a widely interactive system, dependent on many parameters and independent on almost none. Therefore, the lubrication properties must match the process, be compatible with the substrate, and exhibit all the correct physical attributes necessary for optimum processibility [38]. Examining the delicate balance needed for the lubricant system, several key parameters surface that influence the overall lubrication quality of the fiber product. Dominating the boundary lubrication region are wetting and geometry of the surfaces in contact. Viscosity, wetting, and thermal stability are crucial parameters for hydrodynamic lubrication at high speeds.
24.3 Textile Processing 24.3.1 Spinning Process Polymer systems are produced by three major routes; melt, dry, and wet spinning [39–42]. All these routes are extensively covered in literature and will only be briefly mentioned for completeness. Thermoplastic polymers are normally melt spun and produced as multifilament bundles. Most nylons, polyester, polylactides [43–45] and polyolefins are pre-polymerized, then formed in chips used for later re-melt
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spinning or using a continuous process are produced directly into the filament form. Dry spinning is almost a misnomer as the polymer is dissolved in a solvent which flash evaporates upon exiting the spinneret. The polyurethanes fibers, Globespan® [46] and Lycra® [47] spandex fibers are examples of this type of process for fiber formation. Wet spinning is the most complex of the spinning techniques. Some wet spinning process requires the polymer solution to be kept above or below ambient temperature and thus a heat exchanger is used. The spinnerets are just above or immersed in tanks containing the coagulation medium into which the filaments are extruded. Wet spinning spinnerets can have up to 2000 holes. Wet-spun polymer types that are available include aramids, like Kevlar® [48] and Nomex® [48], viscose and acrylics. To produce the initial continuous fiber bundle, the polymer mix is pushed through a multi-holed spinneret having unique hole designs. Routinely, filaments are produced in round, trilobal or octalobal forms. Along with this common commodity set, modern spinneret cutting techniques can make a seemingly endless variety of shapes and filament sizes in recent years. Into the mid 1980s, deniers of >1.5 dpf were the most commonly produced. With the advent of more critical consumer demands, unique fibers were born, such as Nylon Supplex® [47], Coolmax® [47], MicroMattique® [47], Polarguard® [49], and Meryl® [50] Microfibre and Nateo. These are <1.3 dpf fibers offering cotton like appearance, cooling effects, warmth, and other aesthetic properties. The number of holes in the spinneret determines the number of the filaments being produced. With modern tooling methods, that is, laser cutting, a single spinneret of a 5-in. diameter can have as many as 132 or more holes aligned in a pattern to allow optimum quenching of the fibers upon exiting the spinneret. Each fiber type, a POY or FDY, staple or industrial will undergo different processes on the spinning equipment. For multifilament continuous fibers, POY fibers experience little or no heating. FDY products will be drawn to approximately <30% remaining elongation with moderate temperature heating, likely not exceeding 190◦ C to form the product’s internal structure. Industrial yarns will undergo high temperature drawing (>200◦ C) allowing for a highly crystalline structure and associated high strength. Once the filaments are formed, they are brought together, drawn, possibly entangled and when any other required processing steps are completed, are wound to produce a circular package. Staple fibers are processed similarly though they are drawn, heat-set, crimped, and then cut into small lengths, commonly from 0.5 to 1.5 in. in length. Smaller cut lengths of 2 to 5 mm produce products, referred to as flock, normally treated further for use under electrostatic processing conditions. To manage all these different processes, special lubrication packages are uniquely developed to meet the needs of spinning, drawing, and crimping [51].
24.3.2 Texturing Process For the last 35 years, twisting of POY, adding bulk to their appearance, has developed into a major industry for preparing a multitude of fabrics, most notable in the hosiery markets. Early twisting machinery accomplished this by means of a heated rotating pin though the yarn possessed only a small degree of the characteristics that the latest machines can produce. In the intervening years, speeds for these texturing units have progressed from running in the 50–150 m/min range to capabilities exceeding 1500 m/min. However, the typical speed ranges of 800–1200 m/min are routine and the once plastic-like appearance of the multifilament fibers is transformed into the softer feel and looks common to spun yarns. Referred now as False Twist Texturing (FTT), the process twists the yarn tangentially to the direction of travel with the number of turns per in. that can exceed 80. The term false-twist refers to the fact that the yarn, after the twist unit, exits in an untwisted form where the bulk induced by the twist is noted. The twisting is accomplished by a grouping of rotating discs made of various materials, alumina coated metal, ceramic, or varying hardness of polyurethane. A Barmag disc twist unit is shown in Figure 24.14. As the filaments move through the texturing process, the disc grabs the yarn and inserts the twist. When the linear speed is compared against the angular rotation velocity, insertion of twist is a boundary/semiboundary frictional region event. A Barmag AFT FTT machine is shown in Figure 24.15. The path the yarn travels is complex and sees many different conditions, especially temperature effects and friction regions as illustrated in Figure 24.16.
© 2006 by Taylor & Francis Group, LLC
Textile Fibers/Fabrics
24-13
FIGURE 24.14 Twist texturing disc stack (courtesy of Barmag-Saurer Group, Saurer Inc. USA).
To give the permanency of the false-twist appearance, the yarn is heated near the glass transition point to allow the twist to be set into the memory of the polymer structure. These heater temperatures are typically from 190 to 230◦ C. However, the new High Temperature Short Heater (HTSH) machines, like Barmag’s AFK, have heater systems with operating temperatures up to 550◦ C, allowing required heat transfer into the yarn at high linear speeds over short distances. This HTSH technology cuts energy consumption, allows increased speed without excessive heater lengths, and minimizes machinery height profiles for operator ergonomic considerations. With these factors on speed, twist insertion needs, temperature condition etc., lubrication becomes crucial to FFT textile processing. Three key areas are of concern: Hydrodynamic friction for ease of processing at high linear speeds; boundary lubrication for protection of filaments and insertion of twist; and finally, the chemical and physical properties of the lubricants to manage the friction properties and, in some cases, survive excessive temperature extremes without losing their ability to work. As these various frictional regions interplay with the overall performance, the selection of the finish package now has entered into a two-process requirement. First, the spinning process incorporating the application of the finish must be robust so that the yarn can be packaged with straight sidewalls at the high speeds used, surpassing linear velocities of 5500 m/min. Now the spin finish must perform in a multiple of ways. The finish must be compatible with the yarn, maintain a uniform film thickness along the entire length, and not affect the physical properties of the yarn from the spin day to the use date of a customer. Boundary friction forces must be such that the withdrawal of the yarn from the pack is smooth and without tension spikes, especially as the yarn transverses from the package’s front end to the back and returns to the package’s take-off end. Next, hydrodynamic forces must be acceptable throughout the path of transient and is of a level of friction that no excessive strain is added to the yarn than already being
© 2006 by Taylor & Francis Group, LLC
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FIGURE 24.15
Handbook of Lubrication and Tribology
Barmag AFK-2 (courtesy of Barmag-Saurer Group, Saurer Inc. USA).
Cross section (V-HTI)
System diagram
Cooling plate Tension monitor
Heater
Twist insertion Twist trap
W2
W1
Interlace jet
Creel supply
W3 Second heater
Take up
W4
FIGURE 24.16 Profile diagram and schematic thread path of Barmag AFK-2 (courtesy of Barmag-Saurer Group, Saurer Inc. USA).
© 2006 by Taylor & Francis Group, LLC
Textile Fibers/Fabrics
24-15
introduced by the process itself. In the drawing and twist zone, between the twist unit and the top of the heater, the finish formulation must have minimal volatility to prevent dripping of oil from the enclosed hot heater zone and keep decompositional build-up within the heater, due to oxidative degradation, to low levels. Within this zone the inserted twist translates along the length of the filament bundle from the twist unit to the top of the heater where there is inserted a device called a twist-stop. This transit length can exceed 2.5 m and boundary lubrication must provide uniform filament twisting, migration within the bundle for uniform stress across all filaments, and should minimize damage by abrasion against the guide and heater surfaces. It is in this region that the finish frictional properties are subject to the most severe criteria. Later we will discuss the finish formulation’s chemical characteristics to manage the mentioned needs for good texturing performance.
24.3.3 Knitting Process Knitting fashions an array of fabrics for use in garment manufacture. In most use, a fully drawn yarn is handled in one of two forms. The yarn is knitted directly from the package using a circular knitting machine. This knit style is used of hosiery and socks, jersey or other rib knits going into shirts and athletic garments. A typical machine for circular knitting can be viewed on the Internet [52]. The Santoni circular knitting machine has multiple packages and a computerized control system allowing garment manufacture without seams. This state of the art knitting technology provides individual custom garment manufacture if desired. The second way an FDY is used for knitting involves the removal of the yarn from individual packages onto a large beam, called warping. The warping process will take as many as 1600 individual thread lines (packages) and form a beam to be used in raschel or tricot knitting. Once the beam is formed, the beams are put into a group of 4 to 8 to feed a knitting machine that will form a flat fabric. This process, like circular knitting, may incorporate beams of other fiber types, especially spandex, to produce a variety of fabric, with and without stretch. The FDY finish is commonly a combination of two separate formulations. The first finish is present mainly to allow the drawing and lubrication needs during the spinning of the fiber. Most FDY spinning machines allow a secondary finish to be applied that is directed solely for the performance of the fiber during circular knitting or beaming which then is tricot or raschel knitted. Unlike the texturing finish systems, knitting processes will generate large amounts of static and thus antistatic character of the finish is an added feature that must be incorporated along with the frictional needs. As in texturing, the finish must provide superior boundary lubrication primarily for filament protection. For beaming and in all the knitting process, the yarns are subjected to considerable guide contacts at both high and low speeds. Abrasion is the most common problem that a finish must be designed to handle especially with the critical beaming process that subjects the yarn to multiple guides, sharp edges, and fine wire guides near the actual beam that cause filament damage and poor quality fabrics. Many resources are available to study the complexity of the various knitting process. It should be apparent that with the set-ups of the knitting operations, the need for package-to-package uniformity during yarn removal and consistent frictional properties between packages is essential.
24.3.4 Weaving Process Weaving and raschel/tricot knitting have many similar features. In most cases, FDY are warped and weaving done from sets of beams. Many of the finish characteristics are similar though there are some new features that must be recognized. In more than half the weaving operations, the warped beams are treated with a sizing material to provide improved weaving. The finish therefore must be chemically compatible with the size, allowing proper curing of the size prior to weaving and ease of removal after. Another added feature of the weaving process is that there are two distinct yarn forms being used. The warp yarn and the second are the fill yarns. The fill yarn must have more features incorporated for overall performance. Weaving fill insertion can be accomplished using mechanical, water or air transport of the yarn across
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
FIGURE 24.17 Weaving fill performance diagnostic unit (photo courtesy of Department of Textile Engineering, Auburn University).
the “weaving shed.” Mechanical filling, rapier looms, rely on the ability of physical transport of the yarn by a machinery part. With water and air weaving, transport of the yarn is accomplished by jets of water or air. Therefore, in water jet weaving (WJW), fill yarns must be water friendly and not react with the water to form deposits in the weaving shed. In air jet weaving (AJW), the filling yarn must accommodate aerodynamic properties with good cohesiveness to hold the yarn bundle together and accept being “grabbed” by the air and moved across the shed. A common test that has been developed in INVISTA Inc. labs is called “Yarn Time of Flight.” Using a set-up approximating a weaving fill process, the ability of finishes to transport the yarn across the shed can be evaluated. Similarly, researchers at Auburn University, under the auspices of the National Textile Center, built an instrumented unit to study the air friendliness dynamics of yarns and finishes for air jet weaving (Figure 24.17) [53–61].
24.4 Fiber Finishes and Components 24.4.1 Formulated Finish Package A fiber finish is formulated from one to multiple components based on the process and end-use needs. As I have mentioned, the fiber finish must endure the rigors of the spinning process in all its complexities and then continue to manage the tribologic needs of the fiber during processing. The final form could be a knitted to woven fabric, a hose or sock, a spun yarn mixed with another synthetic or natural yarn, to exotic fabrics used for bulletproof vest, geotextile road construction, awnings, and a vast array of other possibilities. This finish will take on a nature of its own, unique to the process it will endure. There are many systems, custom and commercial, that a fiber manufacturer can draw upon. Application of the finish onto the fiber either as neat oil or via a 1 to 20%+ oil-in-water emulsion will dictate the overall complexity of the formulation. Neat oils will typically be much higher in lubricant vs. the emulsion system that requires surfactants to generate stable oil-in-water emulsions. The principal finish properties will be based on the process, frictional, and chemical needs [62,63]. Generally, a fiber finish will have the following components; primary lubricant package, surfactants
© 2006 by Taylor & Francis Group, LLC
Textile Fibers/Fabrics
24-17
(also termed emulsifiers), wetting agents, antioxidants, antistats and auxiliaries, components for special needs [18]. Compositionally, the lubricant will comprise the larger overall percentage of the formulation. These lubricants will be selected primarily to manage both the boundary and hydrodynamic frictional needs. It is important to remember that the other components may contribute to the overall frictional forces when applied to the surface of fiber. The extent of contribution will be dependent on factors such as percentage contained and region of its function. With the description of the various processes, all the other effects that a finish might add to in a contributory or negative way must be considered. The patent literature is full of examples though one needs to be aware that the competitiveness of the fiber business restricts the disclosure of most of the real formulations used in today’s industry. Suffice it to say, for each fiber type, the science and art in formulation of a finish is crucial to success.
24.4.2 Lubricants Lubricants comprise the essence of the finish formulation. These materials control both hydrodynamic and boundary friction. The choices of lubricants are as varied as the imagination can dream [64]. Lubricant use has been described since the dawn of man. Most of the original lubricants served a simple function: to lessen the hydrodynamic friction between two surfaces. Natural oils and waxes were used to lubricate cartwheels and axels. As steam, then gasoline engines replaced the horse drawn cart, the need for improved lubrication at increasingly higher speeds was required. Textiles were no exception [65,66]. Spinning into yarns and knitting or weaving of cotton and wool was at very low speeds and these fibers have natural oils available for handling the friction. The advent of synthetic polymers for textile applications necessitated the expansion and discovery of new and varied lubricant types. As is the nature of textile processing, the selection of the type of fiber lubricant has become highly dependent on many factors including the frictional properties needed for the processing. Typically, the types of lubricants that will be employed are from the following classes, hydrocarbon oil, synthetic and natural esters, alcohol initiated ethylene oxide (EO) and propylene oxide (PO) copolymers, designated as EO–PO copolymers, silicones and alcohol or acid ethoxylates, capped and uncapped. Modifications within these classes are normally for specific needs, i.e. high thermal stability, special friction response under stress or temperature, polymer compatibility, or environmental concerns. Esters and mineral oils were the primary lubricants for most of the fiber finish formulations into the early 1980s. With the 1980s introduction of EO–PO copolymer lubricants, the fiber lubricant chemistries divided into several primary classes based on the polymer types, EO–PO systems for polyesters, mineral oil and ester lubricants for the amides and aramids, silicones for polyurethanes, and finally, the ethoxylates for polypropylene. While there is considerable overlap in some cases, this lubricant/polymer combination set is a basic starting point for the formulator. One of the most widely used classes of lubricants is esters. These may be naturally derived esters, that is, coconut oil, or those prepared synthetically from acids and alcohols derived from natural sources, animal, vegetable, coal, oil, and gas. This wide feedstock variety allows the textile formulator to structure the finish package specifically for a fiber type, processing steps and final end uses. Some typical examples of the frictional property refinements based on minor structural changes in an ester are shown in Figures 24.18 to 24.21. In frictional studies at INVISTA’s labs, esters prepared from a series of alcohols (mono, branched, and polyols) and various fatty acids (mixed, adipic, and pelargonic) were obtained from several ester suppliers and applied at fixed FOY on 70-34 nylon FDY. Studying frictional properties using an input T1 of 10 g, note the range of frictional values that these ester sets can produce. The different frictional properties are due to several key parameters, molecular weight, structural branching, and viscosity. Figure 24.18 shows the effect of molecular weight. Figure 24.19 provides the comparison if the ester is made with mixtures of straight and branched fatty acids. Figure 24.20 demonstrates a fine-tuning of frictional balance with another series of acids and finally, Figure 24.21 shows the similar effect when the alcohol moiety is modified and the acid remains constant. All these examples point out the diversity of frictional values that is available to the finish formulator. With this information, a formulator can blend the appropriate ester or a combination of esters to achieve
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
Coefficient of friction
1.52 1.42 Mw = 430 Mw = 540
1.32
Mw = 584 Mw = 512
1.22 1.12 1.02 100
250
500
Speed m/min
FIGURE 24.18 INVISTA Inc.).
Friction properties influenced by molecular weight (Seemuth unpublished data, courtesy of
44 40
T2
36 32 28 24 100 ypm170wrap
300 ypm170wrap
500 ypm170wrap
700 ypm170wrap
900 ypm170wrap
Speed nc8, nc10
2-EH
iC8
iC10
iC13
FIGURE 24.19 Frictional properties influenced by varying the acid portion for pentaethryitol tetra-esters (Seemuth unpublished data, courtesy of INVISTA Inc.).
the right balance of friction for a given polymer and process. Given the wide range of permutations possible, a normal basic formulation approach is using only one primary lubricant, while using combinations thereof can offer unique transitional properties between hydrodynamic and boundary lubrication regimes, offers interesting solutions, and should not be ignored. The other important second series of lubricants are the EO–PO systems [18]. These materials allow essentially water-soluble type finish systems to be developed whereas ester based packages tend to be microemulsions (see following section). While the benefits of using these types of lubricants are many, the most utility is on polymer systems that are hydrophobic and where the finish chemistry is exposed to hot surfaces. The EO–PO systems tend to leave little or no residue due to the rapid unzipping of the molecule during autooxidation [67]. However, as these EO-PO materials are hydrophilic, use on polymers that are water sensitive can result in adverse aging effects, swelling, oligomer migration to surface and yellowing. Table 24.1 shows some basic EO–PO systems that are used in textile fiber processing.
© 2006 by Taylor & Francis Group, LLC
Textile Fibers/Fabrics
24-19
44
40
T2
36
32
28
24 100 ypm 170wrap
300 ypm 170wrap
500 ypm 170wrap
700 ypm 170wrap
900 ypm 170wrap
Speed nc8, nc10
2-EH
iC8
iC10
iC13
FIGURE 24.20 Frictional properties influenced by varying the alcohol portion for an adipic acid di-ester (Seemuth unpublished data, courtesy of INVISTA Inc.).
46
42
T2
38
34
30
26 100 ypm170wrap
300 ypm170wrap
500 ypm170wrap
700 ypm170wrap
900 ypm170wrap
Speed NPG
TMP
PE
DiPE
Poly PE
FIGURE 24.21 Frictional properties of branched polyhydridic alcohol and fixed acid moiety (Seemuth unpublished data, courtesy of INVISTA Inc.).
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology TABLE 24.1 Typical Properties of Ucon® 50-HB and 75-H Lubricants Ucon® lubricant 50-HB-55 50-HB-100 50-HB-170 50-HB-260 50-HB-400 50-HB-660 50-HB-2000 50-HB-3520 50-HB-5100 75-HB-4500 75-HB-1400 75-HB-9500 75-HB-90,000
Average molecular weight
Viscosity at 40◦ C, cSt
Flash point, ◦ F, COC
Pour point, ◦ F, (ASTM)
270 520 750 970 1,230 1,590 2,660 3,380 3,930 980 2,470 6,950 15,000
8.4 19.1 33.8 52.0 80.0 132 398 700 1,105 90.9 282 1,800 17,850
200 365 400 460 480 445 440 445 450 465 520 510 536
−85 −60 −45 −40 −36 −30 −25 −20 −20 5 40 40 40
In general application, the most used EO–PO lubricants are in the range of 200–1200 cSt (10–800 at 40◦ C cSt) and are the random copolymers. This chemistry set is best represented by the Ucon® 50 HB or 75 HB series with the following empirical formula: RO–(CH2 CH2 O)x –(CH2 CH(CH3 )O)y –H where x and y are equal for the 50 HB series and are in a 75:25 ratio for the 75 HB series. Typical preparation of these lubricants is by simultaneously charging both ethylene and propylene oxide to the reaction vessel, initiating reaction with the EO and PO units associating in random order on the starting alcohol or acid moiety. While there has been considerable use of these EO–PO molecular weight ranges, aerosol toxicity issues have been shown for materials like Ucon® 50 HB-5100. The study shows a strong relationship of toxicity to molecular weight and is mostly associated with the random copolymers [68,69]. At present, all fiber finish formulation should be excluding use of the materials associated with the reported toxicity issues. The second major class are the block EO–PO copolymers. Like the previous systems, these can be represented in the following way. RO–(CH2 CH2 O)x –(CH2 CH(CH3 )O)y –(CH2 CH2 O)z –H where R can be hydrogen or a glycol or butanol initiated end group. In this preparation, each oxide type is charged into the reactor separately, allowed to completely react, then the second oxide type is introduced and reacted. In this way, the EO units are all linked together as are the PO units. The two hydrophilic EO units surrounded the hydrophobic PO units. For this type of blocked copolymer, additional frictional levels can be obtained as the molar ratio of x, y, z is varied. In Table 24.2, a series of the blocked systems and associated properties are provided. For the block copolymers, the toxicity is low and their use as replacements in the molecular weight range of the 50 HB-5100 materials is widely performed. These materials add good emulsification, wetting, and defoaming action to a finish formulation. The reverse block copolymers are known and used with good results. In this case, the PO unit blocks are at either end with the hydrophilic EO unit in the middle. This arrangement provide good properties like lower gelling propensity, defoaming action, and less swelling of polyurethane polymers, commonly used for many textile machinery parts, relative to the normal blocked EO–PO–EO copolymer. A small sampling of the reverse block copolymers is provided in Table 24.3.
© 2006 by Taylor & Francis Group, LLC
Textile Fibers/Fabrics
24-21 TABLE 24.2 Typical Properties of Pluronic® and Ethox® EO–PO–EO Lubricants Product code L31 L35 L42 L44 L61 L62 L63 L64 L81 L92 L101 L121 L122
Average molecular weight
Viscosity at 25◦ C, cPs
Pour point, (ASTM)
1,100 1,900 1,630 1,850 2,000 2,500 2,650 2,900 2,750 3,650 3,800 4,400 5,000
175 375 280 310 325 450 490 850 475 700 800 1,200 1,750
−32 7 −26 −1 −29 −4 10 16 −37 7 −23 5 20
TABLE 24.3 Typical Properties of Reverse Pluronic® and Ethox® PO–EO–PO Type Lubricants Product code
Average molecular weight
Viscosity at 25◦ C, cPs
Pour point, (ASTM)
1,950 4,550 1,800 3,350 1,900 2,150 2,700 4,250 8,550 3,250 3,300 4,150 7,000
440 400b 340 950 300 450 680 370a 2,600b 660 850 300a 1,440
−32 7 −26 −1 −29 −4 10 16 −37 7 −23 5 20
10R5 10R8 12R3 22R4 Ethox 17R1 17R2 Ethox 25R2 25R5 Ethox 25R8 Ethox 31R1 31R2 Ethox 31R4 Ethox 90R4
a 60◦ C, cPs; b 77◦ C, cPs.
24.4.3 Surfactants (Emulsifiers) Surfactants, or by another term emulsifiers, are the second major part of any finish formulation. This group of materials serves two purposes. Primarily used to generate a finish package that can be dispersed into water to form a stable emulsion for application in fiber manufacture [70–73], these materials also bring along both hydrodynamic and boundary lubrication properties. The EO–PO copolymers are an eminent example of lubricant as well as emulsifier. These materials function in both capacities. Along with the easily recognized cross function of these EO–PO copolymers, there are many additional classes of surfactants whose role, while essentially to assist in emulsion stabilization, contribute to the overall frictional properties that the final formulation will impart on the fiber during processing. In this section, I will outline three important chemical types primarily from the emulsification view through their frictional contributions that are present in the final system. These classes are nonionic surfactants represented by alcohol or acid ethoxylated materials, capped or uncapped, polyhydritic alcohols retaining OH functionality, like sorbitol or sorbitan esters, random and blocked EO–PO systems where R > 5 carbon units as represented in the Ucon® and Pluronic® systems, and finally the anionic or cationic surfactants.
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Handbook of Lubrication and Tribology
Surfactancy
High
R=Alcohol
R9 = Acid
Low Increasing moles of ethylene oxide
FIGURE 24.22 oxide (EO).
General trend of surfactancy effectiveness of ethoxylated alcohols and acids vs. moles of ethylene
Each of these systems adds unique properties to the finishes. These materials may function as the primary emulsifier, co-emulsifier, or coupling agent. The first surfactant class is the nonionics [74–76]. These systems are prepared from an acid or alcohol unit, the hydrophobic part, and ethylene oxide forming the hydrophilic unit. These nonionic surfactants are widely used as the primary or the secondary (coupling) emulsifier systems for a fiber finish though their effectiveness does vary based on structure, hydrophobe [77] and EO unit length. Figure 24.22 illustrates the effectiveness of the acid vs. alcohol initiated ethoxylated materials. As the number of moles of EO increases relative to the molecular weight of the hydrophobic group, the ability to emulsify the hydrophobic oil also increases. In this regard, there is an EO content level where the surfactant transitions from a liquid to a solid and the ability to formulate with these materials may be hindered by oil and surfactant solubility issues. In these cases, the use of several compatible surfactants, one functioning as a so called co-emulsifier, will be necessary to provide a stable oil formulation as well as being able to give stable emulsions of the primary lubricant and other materials. In most cases, the molar range of EO finding most use for the mono alcohols or acids is 5 to 15 EO units. This is common for alcohol or acids of 6 to 20 carbon atoms. For polyhydritic alcohols, meaning having more than one alcohol moiety, the range of EO units will range from 0 to 200 EO units. This class of surfactants can best be represented by R–(C)a (O)x -(EO)y –Rz where R is hydrogen or alkyl, a is 3 to 6, x is 2 to 6, y is 0 to 200, R is H or alkyl, and z is equal to (x − 1). Several commonly used surfactants of this class are the glycerol monoester, glycerol mono-oleate (GMO) or glycerol mono-isostearate (GMiSt) and ethoxylated sorbitol mono or the multi-esterified esters. These nonionic surfactants from the multifunctional alcohol class give improvements in thermal stability due to their higher molecular weights, have a wider liquid state range per EO content than the simple straight chain emulsifiers, and exhibit good compatibility with most lubricants. These materials are widely available from many suppliers in many variations for the finish formulator. The second emulsifier class is the alcohol or acid PO–EO systems [78,79]. The availability of these is limited and used for special processing cases. While offering decent emulsification properties, the ratio of EO and PO will be the key parameter as to the extent of utility without excessive use of co-emulsifiers. Figure 24.23 represents the typical effect of emulsification potential for the various structural possibilities. The blocked systems with the EO unit at the end are the best surfactants of the series. This arrangement of structure allows the hydrophobic alkyl and PO units to align with the hydrophobic lubricants and the hydrophilic EO units to the corresponding hydrophilic materials, the other emulsifiers for oil compatibility or water at the surface of the micellar emulsion structure. One key property that these materials can provide is the lowering of chemical absorption (chemical migration) into the soft polyurethane or rubber machine parts. Whereas the simple ethoxylated materials have a high absorptive character to the typical PU and rubber machine parts, substitution of these EO–PO surfactants, though slightly lessening the overall surfactant package emulsifying potential, greatly lowers the damage related to the softening
© 2006 by Taylor & Francis Group, LLC
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Emulsification potential
Textile Fibers/Fabrics
-EO
PO R-
O
O-P
R-E
om
Rand
% ethylene oxide [fixed PO content]
FIGURE 24.23 General trend of surfactancy effectiveness for alcohol or acid EO–PO systems.
360
Visocsity in cps
340 320 300 280 260 240 220 200 0
1.5
3
5
10
Moles of PO for capping
FIGURE 24.24 PO end-capping viscosity effect on blocked alcohol PO–EO surfactants (data courtesy of Milliken and Co.).
of the PU and rubber machine parts, abrasion, wear, and altered physical properties. As can be expected, the increased molecular weights of these types of materials will raise the overall viscosity of the finish package and consequently, the frictional properties on the polymer surface. An additional enhancement available to the finish formulator is the ability to modify the structure of these systems altering the viscosity of the individual materials. While this route has several limitations, the effect can be dramatic and useful when improved compatibility with oil formulation and machine parts is a critical parameter to success. This effect is illustrated in Figure 24.24 as well as the limitation. Taking a base alcohol PO–EO system, simple capping with an additional PO molar ratio of 1.5, a 20 to 50 cSt unit or more dropped in viscosity can be obtained. However, this effect is quickly lost with increased PO addition. While this example is larger than most, it is a valuable synthetic modification tool to assist in maintaining frictional properties at improved compatibility with other materials. A similar effect also appears with other simple ethoxylated materials and the multifunctional alcohols. Furthermore, the PO unit preserves the emulsification and frictional properties of the related and widely used MPEG esters. These PO capped esters exhibit none of the incompatibility with certain polymer types nor exhibit some of the topical numbing effects known for many methyl capped EO alcohol systems. The last classes of surfactants are the anionic [80–82] and cationic [83–85] materials. Again, a wide variety of materials is available and several will be mentioned in the discussion on antistatic components.
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Handbook of Lubrication and Tribology
These materials were the first of the classes of surfactants used by finish formulators. Many were vegetable oil based and while use has decreased with time, they still form an integral part of the formulator’s chemical toolkit. Several of the most common are the sulfated peanut and vegetable oils. These provide good compatibility with various mineral oils and continue to be used in many cases. As process speeds have increased along with needs for low finish oil color, these materials are slowly being replaced in favor of the aforementioned nonionic surfactants where necessary. One widely used anionic surfactant is sodium di-octylsulfosuccinate, commonly referred to as DOSS. This material is extensively studied for its surfactant and wetting properties [86]. In many finish systems, the presence of DOSS allows for improvements in wetting, lower emulsion, and oil surface tensions, without the need of the expensive fluorochemicals or silicones. With its moderately high molecular weight, thermal stability, anionic stabilization of emulsion micelles, and low cost, this material continues as a mainstay surfactant for finishes.
24.4.4 Antistats Antistats incorporated into a finish systems manage and minimize electrical charge generation during processing. These materials introduce hygroscopicity along with the surfactants to the lubricant system. With these materials present, the nonconductive polymer surface possesses the functionality allowing discharge of generated electrical static into the surrounding environment, mainly into the air’s moisture. Both the surfactants and the antistats contribute to the overall antistatic properties of a fiber. The typical surfactants are able to assist in charge removal from the polymer’s surface though their function relies solely on the lone electron pairs of the oxygen atoms of the EO units. As these materials have limited dissipation propensity, many formulations, especially those used for knitting and weaving, incorporate the stronger anionic or cationic classes of antistats. These highly charged materials, provide rapid and effective elimination of static charge. Typical classes of materials are shown in Figure 24.25. Of these, the phosphates and quaternary amines comprise the most widely utilized materials. These two types of chemical systems allow good compatibility in most finish packages. Most widely used are the phosphates with the ammonium salts being used more for processing conditions where the control of processing humidity is less likely, that is, older plants or global facilities in undeveloped countries. Formulation with these materials is a challenge. Dependent on the nature and size of the R groups, solubility in primary finish lubricants can be tricky. Most ammonium systems formulate well though their overall effectiveness is lower than the corresponding phosphates. The phosphates, on the other hand, are most effective, not as the free acid, but as the corresponding salt though oil compatibility becomes an issue. The potassium cation is the preferred counter ion for the phosphate anion. The cation–anion pair provides a larger sized ionic sphere associated with improved static elimination. When the sodium salt, a small and closely associated ion pair with anions, is tested against the corresponding K+ salt, the latter is generally more effective to dissipate charge. Furthermore, as the MW of either system increases, the charge dissipation effectiveness decreases. This is the effect of charge density vs. molecular size. Using many of the very hydrophobic lubricants, the larger R groups afford better oil compatibility than the smaller phosphate salts. The R group choice is also dependent on the fiber type and structure. Low orientation, high elongation fibers will absorb the lower molecular weight though more highly effective antistats. • Anionic R2P(O)O– RCO2– RS+O3–
• Cationic R4N+X–
• Amphoteric • Nonionics RO-(CH2CH2O)x-H R-(PO)x-(EO)y-H
FIGURE 24.25
Chemical classes used for textile antistats.
© 2006 by Taylor & Francis Group, LLC
R3N+CH2CO2–
Textile Fibers/Fabrics
24-25
13 12
Log Rp
11 10 9 8 7 6 250 280 325 355 365 380 390 410 415 455 490 530 Antistat molecular weight Phosphate at 22% Rh
Quat Ammonium at 22% Rh
Phosphate at 47% Rh
FIGURE 24.26 Effect of relative humidity on antistat effectiveness (Seemuth unpublished data, courtesy of INVISTA Inc.).
With Low Orientation Yarns (LOY) and many nylon staple products, phosphates based on alkyl groups
24.4.5 Other Finish Components The preceding sections have examined the primary constituents of most fiber finish formulations, lubricants, surfactants, and antistats. Several other components are used commonly in finish formulations though use is at low levels and it is uncommon to see much impact on frictional properties. Slade [18], Becher [86], and Seemuth [92–94] cover the broad extent of finish components, individual types, and selected examples, best. A brief mention will cover the remaining fiber finish components.
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Cohesive agents allow for temporary bundle integrity during processing. This is important for staple processing steps, roving, and spinning. These materials are normally higher viscosity substances and though normally in low levels, can influence friction. Therefore, a compromise between the level of cohesion and increased friction forces must be established. Antioxidants are widely used though levels less than 1% are normal, offering low chances for impact on overall lubrication properties. Typical antioxidants are organic di- or tri-alkoxy phosphites, di-sulfides, and tert-butylated phenols and cresols. Another common component is an antimicrobial or biocide. These systems are generally incorporated into finishes that will be emulsified in water. These are present as many of the fiber finish components are good nutrient sources for bacteria. The final systems that may be found in a fiber finish are: 1. pH buffers, either inorganic salts or fatty acid soaps 2. Antifoam materials specifically chosen to eliminate foam formation detrimental to processing with a finish package 3. Fluorochemical [95] or silicone wetting [96] agents In most instances, these materials will be used sparingly only to enhance a specific property of the finish. It should be recognized that fatty acid soaps add emulsification capabilities, assist lubrication in the boundary region, and modifies thermal deposition character and pH control properties [97,98]. Therefore, these fatty acid soaps may be present at levels that frictional properties may be altered due to its presence.
24.4.6 Coning Oils These materials are also blends of finish-type components though their typical compositions are prepared with limited utility in mind. In many cases, coning oils are needed and applied to the fiber to supplement the frictional properties of the fiber finish that may be lost during earlier processing steps. Examining a coning oil formulation, one will discover that a major portion of the system will be a low viscosity, inexpensive ester or mineral oil. These are present for control of the hydrodynamic frictional forces encountered during knitting and some in weaving. Other minor components will be antistats and some surfactants for scouring of the oils from the fabric prior to heat setting. While coning oils may need to meet all the requirements of a fiber finish, their formulations demand compatibility with the next steps — scouring, heat setting, and dyeing. In addition, coning oils tend to be used at higher levels so low toxicity and anti-sling properties will be additionally important. To determine the properties of the coning oils, one need only maintain the same fundamental needs for each process stage and the frictional, chemical, and physical needs, as does the finish formulator. In the last several years, the greatest challenges presented to a formulator of fiber finishes, those specially applied to the fiber during manufacture, has been to design the system in such a way as to eliminate the needs for coning oil applications. This challenge has been met and is obtainable except for FTT processing where, in the early processing stages, low FOY levels and high(>170◦ C) and sometime very high temperatures (>300◦ C) volatilizes off much of the finish systems, forcing applications of additional lubricants to facilitate the next processing stage.
Acknowledgments I wish to acknowledge E.I. du Pont de Nemours and Co., Inc. and INVISTA Inc. for providing me the opportunity to work in the synthetic fiber field and develop the science of Fiber Tribology. I further wish to acknowledge several mentors who have guided my study and work in the field of tribology, polymer and finish sciences, and toxicology especially Brian Briscoe, Imperial College, London; Karl Jacobs, Georgia Institute of Technology; Euan McClelland, M. Godwin Jones and Robert Phillips, DuPont retirees; and Gerald Kennedy and John Gannon, DuPont. I gratefully thank my other colleagues around the world and all the contributors to this chapter.
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References [1] Hutchings, I.M., Tribology: Friction and Wear of Engineering Materials, Edward Arnold, London, 1992, p. 23. [2] Olsen, J.S., Frictional behavior of textile yarns, Textile Res. J., 39, 31, 1969. [3] Howell, H.G., Meiszkis, K.W., and Tabor, D., Friction in Textiles, Butterworths, London, 1959. [4] Cameron, A., Basic Lubrication Theory, Ellis Horwood, London, 1981. [5] Baird, M.E. and Meiszkis, K.W., J. Textile Inst., 46, 112, 1955. [6] Schick, M.J., Friction and lubrication of synthetic fibers, in Surface Characteristics of Fibers and Textiles, Schick, M.J. (Ed.), Marcel Dekker, New York, 1975, pp. 2, 3, 5. [7] Schick, M.J., Friction and lubrication of synthetic fibers part I: effect of guide surface roughness and speed on fiber friction, Textile Res. J., 43, 193, 1973. [8] Schick, M.J., Friction and lubrication of synthetic fibers part II: two component systems, Textile Res. J., 43, 198, 1973. [9] Bowden, F.P. and Tabor, D., The Friction and Lubrication of Solids, Clarendon Press, Oxford, Part 1, 1950 and Part II, 1964. [10] Ford, Jr., T., Adsorption and boundary friction on polymer surfaces, J. Phys. Chem., 66, 1136, 1962. [11] Ford, Jr., T., Adsorption and boundary friction on polymer surfaces, J. Phys. Chem., 44, 1136, 1962. [12] Ford, Jr., T., and Olsen, J.S., Boundary friction of textile yarns, Textile Res. J., 31, 1007, 1961. [13] Schick, M.J., Friction and lubrication of synthetic fibers Part IV: effect of fiber material and lubricant viscosity and concentration, Textile Res. J., 43, 342, 1973. [14] Park, H., Seefried, Jr., C.G., and Bryant, G.M., Relation of lubricant structure to frictional properties — polyoxyalkylene monoether lubricants on filament yarns, Textile Res. J., 44, 692, 1974. [15] Schick, M.J., Friction and lubrication of synthetic fibers Part V: effect of fiber luster, guide material, charge and critical surface tension of fibers on fiber friction, Textile Res. J., 44, 758, 1974. [16] Schick, M.J., Friction and lubrication of synthetic fibers Part III: effect of guide temperature, loop size, pretension, denier, and moisture regain on fiber friction, Textile Res. J., 43, 254, 1973. [17] Kao Corporation, Surfactants — A Comprehensive Guide, Tokyo, 1983, p. 107. [18] Slade, P.E., Handbook of Fiber Finish Technology, Marcel Dekker, New York, 1998. [19] Riedo, E., Levy, F., Brune, H., Li, Z.-Q., Kawazoe, Y., and Zhang, S.B., Kinetics of capillary condensation in nanoscopic sliding friction, Phys. Rev. Lett., 88, 185505, 2002. [20] Hansen, W.W. and Tabor, D., Hydrodynamic factors in the friction of fibers and yarns, Textile Res. J., 27, 300–306, 1959. [21] Chapman, J.A., Pascoe, M.W., and Tabor, D., The Friction and Wear of Fibers, Conference on Fiber Friction, Ghent, Belgium, September 1954, pp. P3–P19. [22] Lyne, D.G., The dynamic friction between cellulose acetate yarn and a cylindrical metal surface, Textile Res. J., 46, 112, 1955. [23] Rubenstein, C., The friction and lubrication of yarns, J. Textile Inst., 49, T42, 1958. [24] Rubenstein, C., General theory of the friction of solids, Proc. Phys. Soc., B69, 921, 1956. [25] Rothschild R-2088 Manual and Operating Instructions. [26] ASTM Method D 3412-89 (1991), Standard test method for coefficient of friction, yarn to yarn, Annual Book of ASTM Methods, Volume 07.02, ASTM, Philadelphia, pp. 11–15. [27] ASTM Method D 3103 (1991), Standard test method for coefficient of friction, yarn to solid surfaces, Annual Book of ASTM Methods, Volume 07.01, ASTM, Philadelphia, pp. 848–854. [28] Lawson-Hemphill CTT Operations Manual. [29] Schick, M.J., Friction and lubrication of synthetic fibers, in Surface Characteristics of Fibers and Textiles, Schick, M.J. (Ed.), Marcel Dekker Inc, New York, 1975, p. 11. [30] Prevorsek, D.C. and Sharma, R.K., Fiber–fiber coefficient of friction: effects of modulus and tan δ, J. Appl. Polym. Sci., 23, 173, 1979.
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[31] Pastore, C.M. and Kiekens, P. (Eds), Surface Characteristics of Fibers and Textiles, Volume 94, Surfactant Science Series, Marcel Dekker, New York, 2000. [32] Schlatter, C. and Demas, H.J., Friction studies on Caprolan® yarn, Textile Res. J., 32, 87, 1962. [33] Steinbuch, R. Th., How the crystallization of nylon affects processing and properties, Mod. Plastics, 42, 137, 1964. [34] Berg, J.C., Wettability, Volume 49, Surfactant Science Series, Marcel Dekker, New York, 1993. [35] Myers, D., Surfaces, Interfaces and Colloids, VCH, New York, 1991. [36] Rosen, M.J., Surfactants and Interfacial Phenomena, Second Edition, John Wiley and Sons, New York, 1989. [37] Birdi, K.S., Handbook of Surface and Colloid Chemistry, CRC Press, New York, 1997. [38] Seemuth, P.D. and Potter, J.F., Lubricated Fluoropolymer Yarn, U.S. Patent 6,764,762, July 2004. [39] Doufas, A.K., McHugh, A.J., Miller C. et al., Simulation of melt spinning including flow-induced crystallization — Part II. Quantitative comparisons with industrial spin line data, J. Non-Newton. Fluid, 92, 81–103, 2000. [40] Doufas, A.K. and McHugh, A.J., Simulation of melt spinning including flow-induced crystallization. Part III. Quantitative comparisons with PET spin line data, J. Rheol., 45, 403–420, 2001. [41] Ziabicki, A., Fundamentals of Fiber Formation, Wiley, New York, 1976. [42] Coleman, M.M. and Painter, P.C., Fundamentals of Polymer Science, Second Edition, CRC Press, 1998. [43] Schmack, G., Tandler, B., Optiz, G. et al., High-speed melt spinning of various grades of polylactides, J. Appl. Polym. Sci., 91, 800–806, 2004. [44] Seemuth, P.D. and Anderson, M.B., Bioabsorbable Filaments and Their Production, U.S. Patent 5,288,516, August 1994. [45] Agrawal, A.K. and Bhalla, R., Advances in the production of poly (lactic acid) fibers. A review, J. Macromol. Sci-Polym., R C43, 479–503, 2003. [46] Registered trademark of Bayer. [47] Registered trademark of INVISTA Inc. [48] Registered trademark of E.I. du Pont de Nemours and Co., Inc. [49] Registered trademark of KoSa Industries Inc. [50] Registered trademark of Nylstar®, a joint venture of Rhodia and Snia. [51] Rusznak, I., Handbook of Fiber Science and Technology: Chemical Processing of Fibers and Fabrics, Volume I, Part A, Lewin, A.S. and Sello, S.B. (Eds), Marcel Dekker, New York, 1984. [52] See www.santoni.com for examples of seamless circular knitting machines. [53] Adanur, S. and Mohamed, M.H., Weft insertion on air-jet looms: velocity measurements and influence of yarn structure. Part I: experimental system and computer interface, J. Text. Inst., 2, 297, 1988. [54] Adanur, S. and Mohamed, M.H., Weft insertion on air-jet looms: velocity measurements and influence of yarn structure. Part II: effects of system parameters and yarn structure, J. Text. Inst., 2, 316, 1988. [55] Adanur, S. and Mohamed, M.H., Analysis of yarn motion in single nozzle air-jet filling insertion Part I: theoretical models for yarn motion, J. Text. Inst., 83, 45, 1992. [56] Adanur, S. and Mohamed, M.H., Analysis of yarn motion in single nozzle air-jet filling insertion Part II: experimental validation of the theoretical models and statistical analysis, J. Text. Inst., 83, 56, 1992. [57] Salama, M., Adanur, S., and Mohamed, M.H., Mechanics of a single nozzle air-jet filling insertion system. Part III: yarn insertion through tubes, Textile Res. J., 57, 44, 1987. [58] Adanur, S. and Mohamed, M.H., Analysis of yarn tension in air-jet filling insertion, Textile Res. J., 61, 259, 1991. [59] Adanur, S. and Mohamed, M.H., Analysis of air flow in air-jet filling insertion, Textile Res. J., 61, 253, 1991.
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[60] Bakhtiyarov, S. and Adanur, S., Analysis of air flow in single nozzle air-jet filling insertion: corrugated channel model, Textile Res. J., 66, 401, 1996. [61] Bakhtiyarov, S. and Adanur, S., Airflow over wavy yarn in air-jet filling insertion, Math. Comput. Appl., 4, 1, 1999. [62] Blau, P.J. and Blau, Peter J., Friction Science and Technology, Dekker/CRC Press, New York, 1995. [63] Ikada, Y. and Uyama, Y., Lubricating Polymer Surfaces, CRC Press, 1998. [64] Lansdown, Ar.R., Lubrication and Lubrication Selection, Third Edition, American Society of Mechanical Engineers, 2003. [65] Mang, T. and Dresel, W. (Eds), Lubricants and Lubrication, Wiley-VCH, New York, 2001. [66] Klamann, D., Lubricants and Related Products: Synthesis, Properties, Applications, International Standards, John Wiley and Sons Inc., New York, 1984. [67] Streitwieser, A. and Heathcock, C.H., Introduction to Organic Chemistry, MacMillan, New York, pp. 270, 400, 510, 863, 1985. [68] Hoffman, G.M., Newton, P.E., Birnbaum, H.A., and Kennedy, Jr., G.L., Acute aerosol inhalation studies in several animal species of ethylene oxide/propylene oxide copolymer (Ucon 50-HB-5100), Drug Chem. Toxicol., 14, 243, 1991. [69] Ulrich, C.R., Geil, R.G., Tyler, T.R., and Kennedy Jr., G.L., Two-week aerosol inhalation study in rats of ethylene oxide/propylene oxide copolymers, Drug Chem. Toxicol., 15, 15, 1992. [70] Sjöblom, J. (Ed.), Emulsions and Emulsion Stability, Volume 61, Surfactant Science Series, Marcel Dekker, New York, 1996. [71] Rosano, H.L. and Clausse, M. (Eds), Microemulsion Systems, Volume 24, Surfactant Science Series, Marcel Dekker, New York, 1987. [72] Solans, C. and Kunieda, H. (Eds), Industrial Applications of Microemulsions, Volume 66, Surfactant Science Series, Marcel Dekker, New York, 1996. [73] Datyner, A. (Ed.), Surfactants in Textile Processing, Volume 14, Surfactant Science Series, Marcel Dekker, New York, 1983. [74] Schick, M.J. (Ed.), Nonionic Surfactants, Volume 1, Surfactant Science Series, Marcel Dekker, New York, 1966. [75] Schick, M.J. (Ed.), Nonionic Surfactants: Physical Chemistry, Volume 23, Surfactant Science Series, Marcel Dekker, New York, 1966. [76] van Os, N.M. (Ed.), Nonionic Surfactants: Organic Chemistry, Volume 72, Surfactant Science Series, Marcel Dekker, New York, 1997. [77] Lin, I.J., Friend, J.P., and Zimmels, Y., J. Colloid Interface Sci., 45, 378, 1973. [78] Nace, V.M. (Ed.), Nonionic Surfactants: Polyoxyalkylene Block Copolymers, Volume 60, Surfactant Science Series, Marcel Dekker, New York, 1996. [79] Schmolka, I.R., J. Am. Oil Chem. Soc., 59, 322, 1982. [80] Gloxhuber, C. and Künstler, K. (Eds), Anionic Surfactants: Biochemistry, Toxicology, Dermatology, Second Edition, Revised and Expanded, Volume 43, Surfactant Science Series, Marcel Dekker, New York, 1992. [81] Stache, H.W. (Ed.), Anionic Surfactants: Organic Chemistry, Volume 56, Surfactant Science Series, Marcel Dekker, New York, 1995. [82] Cross, J. (Ed.), Anionic Surfactants: Analytical Chemistry, Second Edition, Revised and Expanded, Volume 73, Surfactant Science Series, Marcel Dekker, New York, 1998. [83] Richmond, J.M. (Ed.), Cationic Surfactants: Organic Chemistry, Volume 34, Surfactant Science Series, Marcel Dekker, New York, 1990. [84] Rubingh, D.N. and Holland, P.M. (Eds), Cationic Surfactants: Physical Chemistry, Volume 37, Surfactant Science Series, Marcel Dekker, New York, 1990. [85] Cross, J. and Singer, E.J. (Eds), Cationic Surfactants: Analytical and Biological Evaluation, Volume 53, Surfactant Science Series, Marcel Dekker, New York, 1990. [86] Becher, P. (Ed.), Encyclopedia of Emulsion Technology, Volume 1, 1983, Volume 2, 1985, Volume 3, 1988, and Volume 4, 1996, Marcel Dekker, New York.
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[87] Hearle, J.W., Moisture and electrical properties, in Moisture in Textiles, Hearle, J.W. and Peters, R.H. (Eds), Butterworths, London, 1960. [88] Grindstaff, T., Improved staple processing performance through air change control, Textile Res. J., 55, 266, 1985. [89] Shenai, V.A., Technology of Textile Processing, Volume V, Sevak Publications, Bombay, 1976. [90] Anderson, N., Peak, R., and Moyse, J.A., Spin Finish with Anti-Static Agent, U.S. Patent 4,294,709, 1981. [91] Schonfeldt, N., Surface Active Ethylene Oxide Adducts, Pergamon Press, Oxford, 1969. [92] Seemuth, P.D., Aqueous Emulsion Finishes for Spandex Fibers Containing Polydimethyl Siloxane and Ethoxylated Long-chain Alkanol, U.S. Patent 4,999,120, 1991. [93] Seemuth, P.D., Chemical, Physical and Structure-Property Relationships of Finish Finishes and Components, 198th National ACS Meeting, August 1989. [94] Seemuth, P.D., Fundamental Interactions of Wetting Phenomena and Finish Compositions, Wetting Fundamental Conference, Charlotte, NC, June 1995. [95] Kissa, E. (Ed.), Fluorinated Surfactants: Synthesis, Properties, Applications, Volume 50, Surfactant Science Series, Marcel Dekker, New York, 1993. [96] Hill, R.A. (Ed.), Silicone Surfactants, Volume 86, Surfactant Science Series, Marcel Dekker, New York, 1999. [97] Patterson, H.T. and Proffitt, Jr., T.J., Fatty acids in textiles, in Fatty Acids in Industry, Johnson, R.W. and Fritz, E. (Eds), Marcel Dekker, New York, 1989, chapter 19. [98] Finch, N.L., Lemley, J.D., and Proffitt, Jr., T.J., High Temperature Resistant Textile Fiber Finish Composition, U.S. Patent 3,544,462, 1970.
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25 Food-Grade Lubricants and the Food Processing Industry 25.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25.2 The Food Processing Industry . . . . . . . . . . . . . . . . . . . . . . . . 25.3 Current Registration Practices . . . . . . . . . . . . . . . . . . . . . . . .
25-1 25-2 25-2
U.S. Regulations Prior to 1998 • Changes in Food-Grade Lubrication Standards After 1998 • Third-Party Certifications
25.4 Challenges Facing Food-Grade Lubricants . . . . . . . . . . 25.5 Food-Grade Lubricants Defined by Category . . . . . . . 25.6 Approved Lubricant Formulations in H1 Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
25-4 25-4 25-5
Acceptable Food-Grade Basestocks • Acceptable Food-Grade Additives and Thickeners
James C. Fitch, Sabrin Gebarin, and Martin Williamson Noria Corporation
25.7 Selecting What Machines Require Food-Grade Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25.8 Selecting an H1 Food-Grade Supplier [13] . . . . . . . . . . 25.9 Global Trends [2] . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25.10 Religious Organizations Influence in Food-Grade Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25.11 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
25-12 25-12 25-14 25-17 25-17 25-17
25.1 Introduction The food processing industry presents unique challenges to lubricant formulation engineers, lubricant marketers, plant lubrication engineers, equipment designers, and builders. While it is never desirable for lubricants to be allowed to contaminate raw materials, work-in-progress, or finished product, the consequences of a lubricant contaminated product is rarely more acute than in the food processing industry. As such, lubricants used in this industry have requirements, protocols, and performance expectations that go well beyond typical industrial lubricants.
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This chapter provides a general overview of the unique challenges associated with food-grade lubricants including recent revisions of the regulatory environment. The terminology commonly used by suppliers and clients of food-grade products will also be defined and discussed. So too, will be machine applications common to certain sectors of the food processing industry and their unique requirements for food-grade lubricants.
25.2 The Food Processing Industry Food-grade lubricants are significant in scope and application when you consider the size of the food processing industry. In the United States, food manufacturers represent a significant percentage of total manufacturing. According to the 1997 U.S. Census, approximately 485 billion dollars in sales revenue was generated (about the same amount generated in the transportation manufacturing industry). This represents almost 13% of all manufacturing in the United States. In 1997, approximately 28,000 manufacturing facilities employed 1.6 billion employees and produced 233 billion dollars in goods [1].
25.3 Current Registration Practices Historically, the two U.S. government agencies primarily involved in food processing were the United States Department of Agriculture (USDA) which regulates meat, poultry, and plants and the United States Food and Drug Administration (FDA) which monitors other food and pharmaceutical manufacturing operations.
25.3.1 U.S. Regulations Prior to 1998 Prior to 1998, approval and compliance of food-grade lubricants was the responsibility of the USDA. The Food Safety and Inspection Services (FSIS), headed by the USDA, reviewed the formulations of maintenance and operating chemicals. FSIS required meat and poultry facilities to use only nonfood compounds that were pre-approved by the USDA authorization program. However, these programs spread to other food market sectors such as fisheries and retail food operations [2]. To gain USDA approval, lubricant manufacturers had to prove that all the ingredients in the formulation were allowable substances. Allowable substances, in this instance, are those listed by the FDA in accordance with the Guidelines of Security Code of Federal Regulations (CFR) Title 21, §178.3570. This did not include lubricant testing. Rather, the approval was based primarily on a review of the formulation ingredients of the lubricant [2].
25.3.2 Changes in Food-Grade Lubrication Standards After 1998 Starting February 1998, FSIS significantly altered their program by implementing a system established by Hazard Analysis and Critical Control Point (HACCP) requiring the manufacturer to assess risk at each point in the operation where contamination might occur. The National Aeronautics and Space Administration (NASA) originally developed the HACCP system in the 1960s to prevent astronauts from receiving any food-borne illnesses. It established measures like minimum cooking temperatures for each control point, instituted procedures to monitor these measures, and also provides corrective actions if critical limits are not met [3]. In essence, the manufacturer became responsible for reviewing and approving the chemical compositions of lubricants to decide whether they were safe or not as food-grade lubricants.
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FIGURE 25.1 NSF’s search engine of approved lubricants in H1, H2, and H3 applications.
25.3.3 Third-Party Certifications In response to the change in the approval process, several commercial organizations developed external certification programs. Three such organizations were the National Sanitation Foundation (NSF), Underwriters Laboratory (UL), and a joint effort by three recognized industry professional associations: The National Lubricating Grease Institute (NLGI), The European Lubricating Grease Institute (ELGI), and the European Hygienic Equipment Design Group (EHEDG). NSF has developed a lubricant evaluation program that essentially mirrors the FSIS program by evaluating the candidate lubricant formulations to verify compliance with the various FDA CFR guidelines. Each component in the formulation is submitted to NSF by the lubricant manufacturer along with other supporting documentation. This is then reviewed to verify it is within the FDA list of permitted substances [4]. NSF’s website provides food processing manufacturers with a continually updated list of approved lubricants at www.nsfwhitebook.org (Figure 25.1). Underwriters Laboratory is another organization that began third-party certification of food-grade lubricants but no longer is doing so. While they have not been as active as NSF in the area of food-grade lubricants, in the past, UL has organized several informational meetings inviting lubricant and chemical manufacturers to attend [5]. The NLGI/ELGI/EHEDG Joint Food-grade Lubricants Working Group has been active in drafting an authorization program for food-grade lubricants. This group’s program is also based on the former USDA/FSIS authorization program and CFR policies. Their plan is to develop a DIN (the German Institute for Standardization) standard in Germany and use the DIN standard to later develop an ISO (International Organization for Standardization) standard [5]. Not all countries use third-party certifications. Canada, New Zealand, Australia, and Japan are some of the countries that federally regulate food-grade lubricants [1]. However, the Canadian Food Inspection Agency (CFIA) is working on a food-grade lubricants approval system, and NSF will help with the
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CFIA review process. Also, the Australian Quarantine Inspection Service has approved approximately 50 food-grade lubricants based on NSF registration [6].
25.4 Challenges Facing Food-Grade Lubricants Agricultural and animal substances go through a number of processes in a manufacturing plant such as cleansing, sterilizing, blending, mixing, cooking, freezing, cutting, packaging, canning, and bottling. Large-scale food processing requires machinery such as pumps, mixers, tanks, hoses, and pipes, chain drives, and conveyor belts. Machinery used in food processing facilities face many of the same tribological and lubrication challenges found in other nonfood processing plants. In that sense, lubricants must offer similar protection of internal surfaces to control friction, wear, corrosion, heat, and deposits. They must also offer good pumpability, oxidation stability, hydrolytic stability, and thermal stability where the application so requires. Many of the raw materials used to formulate lubricants that effectively address these challenges in conventional industrial applications are not permissible in food applications for safety reasons. In addition, certain applications within the food and drug manufacturing facilities demand that lubricants resist degradation and impaired performance when in contact with food products, certain process chemicals, water (including steam), and bacteria. They must also exhibit neutral behavior toward plastics and elastomers and have the ability to dissolve sugars. In general, these lubricants must comply with food/health and safety regulations, as well as be physiologically inert, tasteless, odorless, and internationally approved [7]. Lubricants in many food processing plants can be subjected to ingression and contend with an assortment of environmental contaminants. For instance, a corn-milling environment generates significant dust. Although not as hard as silica-based terrain dust, it still presents a problem for filtration. A meat plant requires stringent steam cleaning at all times, so the risk of water contamination is high. Water contamination in gear oils routinely exceeds 15% in some plants. Another aspect of lubrication contamination that poses unique risk to food-grade lubricants is the growth of microorganisms such as bacteria, yeast, and fungi. While these can also be challenging to conventional industrial lubricants, the opportunity for microbial contamination in the food-production industry is considerably greater.
25.5 Food-Grade Lubricants Defined by Category Food-grade lubricants are either compounded or uncompounded products acceptable for use in meat, poultry, and other food processing equipment, applications, and plants. The lubricant types in food-grade applications are broken into categories based on the likelihood that they will contact food. The original food-grade designations H1, H2, and H3 were created by the USDA. The approval and registration of a new lubricant into one of these categories depends on the ingredients used in the formulation. The three designations are described here [2]. H1 lubricants are food-grade lubricants used in food-processing environments where there is some possibility of incidental food contact. Lubricant formulations must be composed of one or more approved basestocks, additives, and thickeners (if grease) listed in 21 CFR 178.3750. Only the minimum amount of lubricant required should be used on the equipment. H2 lubricants are lubricants used on equipment and machine parts in locations where there is no possibility that the lubricant or lubricated surface contacts food. Because there is not risk of contacting food, H2 lubricants do not have a defined list of acceptable ingredients. They cannot, however, contain intentionally heavy metals such as antimony, arsenic, cadmium, lead, mercury or selenium. Also, the ingredients must not include substances that are carcinogens, mutagens, teratogens, or mineral acids [4].
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Food-Grade Lubricants and the Food Processing Industry TABLE 25.1
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H-3 Soluble Oil Approved Lubricants
Lubricant type
Regulations they must meet
Edible oils (corn oils, cottonseed oil, soybean oil) Certain mineral oils Generally recognized as safe (GRAS)
21 CFR 172.860 21 CFR 172.878 21 CFR 182 or 21 CFR 184
H3 lubricants, also known as soluble or edible oil, are used to clean and prevent rust on hooks, trolleys, and similar equipment. Equipment applied with H3 lubricants should be cleaned by washing or wiping the surface before putting the equipment in service. These lubricants can only consist of ingredients as shown in Table 25.1 [4]. Deciding whether there is a possibility of contact is tough, and many have erred on the side of safety with respect to selecting H1 over H2.
25.6 Approved Lubricant Formulations in H1 Lubricants As previously mentioned, the USDA/FSIS approvals are based on the various FDA Codes in Title 21 that dictate approval for ingredients used in lubricants that may have incidental contact with food. These are mentioned as follows: • 21.CFR 178.3570 — allowed ingredients for the manufacture of H1 lubricants • 21.CFR 178.3620 — white mineral oil as a component of nonfood articles intended for use in contact with food • 21.CFR 172.878 — USP mineral oil for direct contact with food • 21 CFR 172.882 — synthetic isoparaffinic hydrocarbons • 21.CFR 182 — substances generally recognized as safe Based on the Title 21 FDA regulations noted, the following is paragraphs discuss the allowable basestocks, additives, and thickeners in food-grade lubricants.
25.6.1 Acceptable Food-Grade Basestocks Depending on whether the food-grade lubricant is H1 or H2, the list of approved basestocks will vary. H2 lubricant basestock guidelines are less restrictive and consequently allow a broader variety of basestocks. Many products used in industrial (nonfood) plants are also used in food plants for H2 applications. H1 lubricants are much more limited since they are designed to allow for accidental exposure with the processed foods. The approved H1 lubricant basestocks can be either mineral or synthetic. 25.6.1.1 Petroleum-Based Lubricants Mineral oils used in H1 food-grade lubricants are either technical white mineral or USP-type white mineral oils. White oils start as normal paraffinic petroleum stocks and are processed into pure branched paraffin stocks, stripped free of the majority of aromatic hydrocarbons, sulfur, and nitrogen contaminants. They are highly refined and are colorless, tasteless, odorless, and nonstaining. Technical white oils meet the regulations specified in 21 CFR 178.3620. Based on the American Society for Testing Materials (ASTM) method D156-82, “Standard Test Method for Saybolt Color of Petroleum Products (Saybolt Chromometer Method),” the Saybolt color must be a minimum of 20 to be considered a technical white oil [8]. USP mineral oils are the purest of all white mineral oils, and are the most oxidatively stable [5]. Historically, white mineral oils were first listed in the United States Pharmacopoeia (USP) in 1926. Later, a paper on the general principles of white oil manufacturing was written in 1935 followed by other papers [1].
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
TABLE 25.2 Approved Substance for H1 Lubricants per 21CFR 178.3570 Substance
Limitations
Aluminum stearoyl benzoyl hydroxide
For use only as a thickening agent in mineral oil lubricants at a level not to exceed 10% by weight of the mineral oil
N , N -bis(2-ethylhexyl)-armethyl-1H-benzotriazole-1methanamine (CAS Reg. No.94270-86-7) BHA BHT
For use as a copper deactivator at a level not to exceed 0.1% by weight of the lubricant
[alpha]-Butyl-omega- hydroxypoly(oxyethylene) poly(oxypropylene) produced by random condensation of a 1 : 1 mixture by weight of ethylene oxide and propylene oxide with butanol; minimum mol wt 1500; Chemical Abstracts Service Registry No. 9038-95-3 [alpha]-Butyl-omega- hydroxypoly (oxypropylene); minimum mol wt 1,500; Chemical Abstracts Service Registry No. 9003-13-8 Castor oil
Addition to food not to exceed 10 ppm
Castor oil, dehydrated
Addition to food not to exceed 10 ppm
Castor oil, partially dehydrated
Addition to food not to exceed 10 ppm
Dialkyldimethylammonium aluminum silicate (CAS Reg. No. 68953-58-2), weight 1, 6hexanediol (CAS Reg. No. by weight of the mineral oil. 629-11-8), where the alkyl groups are derived from hydrogenated tallow fatty acids (C14 –C18 ) and where the aluminum silicate is derived from bentonite Dimethylpolysiloxane (viscosity greater than 300 cSt)
For use only as a gelling agent in mineral oil lubricants at a which may contain up to 7% by level not to exceed 15%
Addition to food not to exceed 10 ppm
Addition to food not to exceed 10 ppm
Addition to food not to exceed 1 ppm
Di (n-octyl) phosphite (CAS Reg. No. 1809-14-9)
For use only as an extreme pressure-antiwear adjuvant at a level not to exceed 0.5% by weight of the lubricant
Disodium decanedioate (CAS Reg. No. 1726514-4)
For use only:
Disodium EDTA (CAS Reg. No. 139-33-3)
For use only as a chelating agent and sequestrant at a level not to exceed 0.06% by weight of lubricant at final use dilution
© 2006 by Taylor & Francis Group, LLC
1. As a corrosion inhibitor or rust preventative in mineral oil–bentonite lubricants at a level not to exceed 2% by weight of the grease 2. As a corrosion inhibitor or rust preventative only in greases at a level not to exceed 2% by weight of the grease
Food-Grade Lubricants and the Food Processing Industry
TABLE 25.2
25-7
Continued
Substance Ethoxylated resin phosphate ester mixture consisting of the following compounds: 1. Poly(methylene-p-tert-butyl phenoxy)poly(oxyethylene) mixture of dihydrogen phosphate and monohydrogen phosphate esters (0 to 40% of the mixture). The resin is formed by condensation of 1 mol of p-tertbutylphenol with 2 to 4 mols of formaldehyde and subsequent ethoxylation with 4 to 12 mols of ethylene oxide 2. Poly(methylene-p-nonylphenoxy) poly(oxyethylene) mixture of dihydrogen phosphate and monohydrogen phosphate esters (0 to 40% of the mixture). The resin is formed by condensation of 1 mol of p-nonylphenol with 2 to 4 mols of formaldehyde and subsequent ethoxylation with 4 to 12 mols of ethylene oxide 3. n-Tridecyl alcohol mixture of dihydrogen phosphate and monohydrogen phosphate esters (40 to 80% of the mixture; CAS Reg. No. 56831-62-0)
Limitations For use only as a surfactant to improve lubricity in lubricating fluids complying with this section at a level not to exceed 5% by weight of the lubricating fluid
Fatty acids derived from animal or vegetable sources and the hydrogenated forms of such fatty acids 2-(8-Heptadecenyl)-4,5-dihydro-1H-imidazole-1ethanol (CAS Reg. No. 95-38-5)
For use at levels not to exceed 0.5% by weight of the lubricant
Hexamethylenebis(3,5-di-tert-butyl-4-hydroxy hydrocinnamate) (CAS Reg. No.35074-77-2)
For use as an antioxidant at levels not to exceed 0.5% by weight of the lubricant
[alpha]-Hydro-omega-hydroxypoly (oxyethylene) poly(oxypropylene) produced by random condensation of mixtures of ethylene oxide and propylene oxide containing 25 to 75% by weight of ethylene oxide; minimum mol wt 1,500; Chemical Abstracts Service Registry No. 9003-11-6 12-Hydroxystearic acid
Addition to food not to exceed 10 ppm
Isopropyloleate
For use only as an adjuvant (to improve lubricity) in mineral oil lubricants
Magnesium ricinoleate
For use only as an adjuvant in mineral oil lubricants at a level not to exceed 10% by weight of the mineral oil
Mineral oil
Addition to food not to exceed 10 ppm
N-Methyl-N-(1-oxo-9-octadecenyl) glycine (CAS Reg. No. 110-25-8)
For use as a corrosion inhibitor at levels not to exceed 0.5% by weight of the lubricant
N-phenylbenzenamine, reaction products with 2, 4, 4-trimethylpentene (CAS Reg. No. 68411-46-1)
For use only as an antioxidant at levels not to exceed 0.5% by weight of the lubricant
Petrolatum
Complying with Sec. 178.3700. Addition to food not to exceed 10 ppm
Phenyl-[alpha]-and/or phenyl [beta]-naphthylamine
For use only, singly or in combination, as antioxidant in mineral oil lubricants at a level not to exceed a total of 1% by weight of the mineral oil (Continued)
© 2006 by Taylor & Francis Group, LLC
25-8
Handbook of Lubrication and Tribology
TABLE 25.2
Continued
Substance
Limitations
Phosphoric acid, monohexyl and dihexyl esters, compounds with tetramethylnonylamines and C11−14 alkylamines Phosphoric acid, monoisooctyl and diisooctyl esters, reacted with tert-alkyl and (C12 –C14 ) primary amines (CAS Reg. No.68187-67-7)
For use only as an adjuvant at levels not to exceed 0.5% by weight of the lubricant
Phosphorothioic acid, O, O, O-triphenyl ester, tertbutyl derivatives (CAS Reg. No. 192268-65-8)
For use only as an extreme pressure-antiwear adjuvant at a level not to exceed 0.5% by weight of the lubricant
Polyurea, having a nitrogen content of 9 to 14% based on the dry polyurea weight, produced by reacting tolylene diisocyanate with tall oil fatty acid(C16 and C18 ) amine and ethylene diamine in a 2 : 2 : 1 molar ratio Polybutene (minimum average mol wt 80,000)
For use only as an adjuvant in percent level not to exceed 10 mineral oil lubricants at a by weight of the mineral oil
For use only as a corrosion inhibitor or rust preventative in lubricants at a level not to exceed 0.5% by weight of the lubricant
Addition to food not to exceed 10 ppm
Polybutene, hydrogenated; complying with the identity prescribed under Sec. 178.3740 Polyethylene
Addition to food not to exceed 10 ppm
Polyisobutylene (average mol wt 35,000–140,000 (Flory))
For use only as a thickening agent in mineral oil lubricants
Sodium nitrite
Use only as a rust preventive in mineral oil lubricants at a level not to exceed 3% by weight of the mineral oil
Tetrakis[methylene(3,5-di-tert-butyl-4hydroxyhydro-cinnamate)] methane (CAS Reg. No. 6683-19-8) Thiodiethylenebis (3,5-di-tert-butyl-4- hydroxyhydrocinnamate) (CAS Reg. No. 41484-35-9)
For use only as an antioxidant in lubricants at a level not to exceed 0.5% by weight of the lubricant
Tri[2(or 4)-C9−−10 -branched alkylphenyl] phosphorothioate (CAS Reg. No. 126019-82-7)
Triphenyl phosphorothionate (CAS Reg. No. 597-82-0) Tris(2,4-di-tert-butylphenyl)phosphite (CAS Reg. NO. 31570-04-4) Thiodiethylenebis(3,5-di-tert-butyl-4-hydroxyhydro-cinnamate) (CAS Reg. No. 41484-35-9) Zinc sulfide
Addition to food not to exceed 10 ppm
For use as an antioxidant at levels not to exceed 0.5% by weight of the lubricant For use only as an extreme pressure-antiwear adjuvant at levels not to exceed 0.5% by weight of the lubricant For use as an adjuvant in lubricants herein listed at a level not to exceed 0.5% by weight of the lubricant For use only as a stabilizer at levels not to exceed 0.5% by weight of the lubricant For use as an antioxidant at levels not to exceed 0.5% by weight of the lubricant For use at levels not to exceed 10% by weight of the lubricant
Source: 21 CFR 3570 — Lubricants with incidental food contact. Retrieved online at www.access.gpo.gov/ nara/cfr.index.asp.
25.6.1.2 Synthetic Lubricants Synthetic H1 lubricants are mainly polyalphaolefins (PAO). They were first introduced in 1981 by Gulf Research and Development Company [1]. Compared to white mineral oils, they have significantly greater oxidation stability and greater range of operating temperatures. Another H1 synthetic lubricant used is Polyalkylene glycols (PAG). These lubricants are more increasingly used in high temperature applications.
© 2006 by Taylor & Francis Group, LLC
Food-Grade Lubricants and the Food Processing Industry TABLE 25.3
25-9
Bakery and Confectionery Products
Flour pumps/blowers/fluidizers Greased bearings NLGI No. 2 Grease Oiled bearings ISO R&O or AW Oil Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray
Blenders Greased bearings Enclosed gears
Blanchers — initial cooking Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray
Shapers/rounders/moulders/elongators Geared bearings NLGI No. 2 Grease Enclosed gears ISO 460 Gear Oil Chains NLGI No. 2 Grease Spray
Dough mixers Greased bearings Oiled bearings Enclosed gears Chains Hydraulics
NLGI No. 2 Grease ISO R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil
Panners/depanners Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 220 or 460 Gear Oil NLGI No. 2 Grease Spray
Sifters/separators Greased bearings Chains Air Line Lube
NLGI No. 2 Grease NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil
Proofers/coolers Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 220 or 460 Gear Oil NLGI No. 2 Grease Spray
Dividers/portioners Greased bearings Oiled bearings Enclosed gears Chains
NLGI No. 2 Grease ISO R&O or AW Oil ISO 460 Gear Oil NLGI No. 2 Grease Spray
Pan-tray washers Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 220 Gear Oil
Ovens/fryer/roasters/cookers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 460 Gear Oil Chains NLGI No. 2 Grease Spray Baggers/packagers Greased bearings Oiled bearings Open gears Enclosed gears Chains
NLGI No. 2 Grease ISO 46 R&O or AW Oil NLGI No. 2 Grease Spray ISO 220 or 460 Gear Oil NLGI No. 2 Grease Spray
Pan-tray stackers/unstackers Greased bearings NLGI No. 2 Grease Chains NLGI No. 2 Grease Spray Hydraulics ISO 46 R&O or AW Oil Air Line Lube ISO 46 R&O or AW Oil
Wrappers Greased bearings Oiled bearings Open gears Enclosed gears Chains
NLGI No. 2 Grease ISO 46 R&O or AW Oil NLGI No. 2 Grease Spray ISO 220 or 460 Gear Oil NLGI No. 2 Grease Spray
Liquor mills Greased bearings Enclosed gears
NLGI No. 2 Grease ISO 220 Gear Oil
Dust collectors Greased bearings Enclosed gears
NLGI No. 2 Grease ISO 100 R&O or AW Oil
Depositors/applicators Greased bearings Oiled bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 46 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray
Conveyers Greased bearings Oiled bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 460 Gear Oil NLGI No. 2 Grease Spray
Extruders Greased bearings Oiled bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 46 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray
Pumps Greased bearings
NLGI No. 2 Grease
Slicers Greased bearings Enclosed gears Chains Air Line Lube
NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray or ISO 100 R&O or AW Oil
Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With permission.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
TABLE 25.4
Beverages
Depalletizers/palletizers Greased bearings Enclosed gears Chains Hydraulics Air Line Lubes
NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil ISO 46 R&O or AW Oil
Bottle-can uncaser/packer/case wrapper Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains/open gears NLGI No. 2 Grease Spray Air Line Lube ISO 46 R&O or AW Oil
Bottle-can washers/rinsers Greased bearings Enclosed gears Chains Open gears
NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray NLGI No. 2 Grease Spray
Bottle-can fillers Greased bearings Oiled bearings Enclosed gears
NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 460 Gear Oil
Pasteurizers Greased bearings Enclosed gears Hydraulics Chains
NLGI No. 2 Grease ISO 460 Gear Oil ISO 46 R&O or AW Oil NLGI No. 2 Grease Spray
Bottle cappers Greased bearings Enclosed gears Thread roller
NLGI No. 2 Grease ISO 460 Gear Oil ISO 100 R&O or AW Oil
Conveyors Greased bearings Oiled bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 460 Gear Oil NLGI No. 2 Grease Spray
Can closers Greased bearings Enclosed gears Chains Open gears
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray NLGI No. 2 Grease Spray
Pumps Greased bearings
NLGI No. 2 Grease
Bottle labelers Greased bearings/cams Oiled bearings Chains Open gears Enclosed gears
NLGI No 2 Grease ISO 100 R&O or AW Oil NLGI No. 2 Grease Spray NLGI No. 2 Grease Spray ISO 220 Gear Oil
Syrup mixers/pumps/fillers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 460 Gear Oil Open gears NLGI No. 2 Grease Spray
Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With permission.
Dimethylpolysiloxane (silicones) with a viscosity greater than 300 cSt [9] is also permitted for H1 lubricants. Sanction letters for the use of silicone fluids as defoaming agents show up as early as 1953. Silicones were not approved until soon after a petition filed by General Electric in 1965 [1]. Silicones have even higher thermal and oxidation stability than PAO and PAG base oils. 25.6.1.3 Differences Among Basestocks Although synthetics are more expensive than mineral oils, tests performed on H1 PAO and white mineral oils on drive chains show that the useful life of PAOs is almost twice that of white oils. Testing has shown PAG base oils have a service life five times longer than white mineral oils [7]. In addition to longer service life, there is evidence that synthetic H1 oils do a better job of protecting metal surfaces from corrosion and wear and withstand greater temperature extremes required around freezers or ovens.
25.6.2 Acceptable Food-Grade Additives and Thickeners Often basestocks are not able to meet the severe demands required in food processing work environments. To improve the performance characteristics of base oils, additives are blended into the formulation. The types of antioxidants, corrosion inhibitors, antiwear, extreme pressure additives, and concentration are limited by 21 CFR 178.3570.
© 2006 by Taylor & Francis Group, LLC
Food-Grade Lubricants and the Food Processing Industry TABLE 25.5
25-11
Canned, Preserved, and Frozen Fruits and Vegetables
Peelers/pitters/huskers Greased bearings/slides Oiled bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 46 R&O or AW Oil ISO 460 Gear Oil NLGI No. 2 Grease Spray
Centrifuges Greased bearings Oiled bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 100 R&O or AW Oil NLGI No. 2 Grease Spray
Snippers Greased bearings/slides Oiled bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 46 R&O or AW Oil ISO 460 Gear Oil NLGI No. 2 Grease Spray
Presses Geared bearings Enclosed gears Chains Hydraulics
NLGI No. 2 Grease ISO 100 R&O or AW Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil
Graders/food washers Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray
Cookers/coolers Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 100 R&O or AW Oil NLGI No. 2 Grease Spray
Blanchers — initial cooking Greased bearings NLGI No. 2 Grease Enclosed gears ISO 460 Gear Oil Chains NLGI No. 2 Grease Spray
Freezing tunnels Greased bearings Hydraulics Chains
NLGI No. 2 Grease ISO 46 R&O or AW Oil NLGI No. 2 Grease Spray
Grinders/blenders/finishers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray
Vacuum filters/evaporators Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray
Hammer mill Greased bearings Enclosed gears
NLGI No. 2 Grease ISO 100 R&O or AW Oil
Homogenizers/pasteurizers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 100 R&O or AW Oil or ISO 460 Gear Oil Fluid drive ISO 46 R&O or AW Oil Hydraulics ISO 68 R&O or AW Oil
Baggers/sealers Greased bearings Oiled bearings Enclosed gears Chains/open gears Air Line Lubes
NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray ISO 100 R&O or AW Oil
Depalletizers/palletizers Greased bearings Enclosed gears Chains Hydraulics Air Line Lubes
NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil ISO 46 R&O or AW Oil
Can-jar washers Greased bearings Enclosed gears Chains/open gears Can-jar filters Greased bearings/valves Oiled bearings/valves Enclosed gears
Can closers Greased bearings Enclosed gears Chains/open gears
NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray
Labelers Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray
Can-jar packers/casers/uncasers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains/open gears NLGI No. 2 Grease Spray Air Line Lube ISO 46 R&O or AW Oil NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 460 Gear Oil
Conveyers Greased bearings Oiled bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 460 Gear Oil NLGI No. 2 Grease Spray (Continued)
© 2006 by Taylor & Francis Group, LLC
25-12
Handbook of Lubrication and Tribology TABLE 25.5
Continued
Can-jar coolers Greased bearings Enclosed gears Chains/open gears
NLGI No. 2 Grease ISO 100 R&O or AW Oil NLGI No. 2 Grease Spray
Jar cappers Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray
Pumps Greased bearings
NLGI No. 2 Grease
Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With permission.
Greases are lubricating oils that have a thickening agent added to the formulation. Approved grease thickeners include aluminum stearate, aluminum complex, organo clay, and polyurea [10]. Aluminum complex is the most common H1 food-grade grease thickener. They can withstand high temperatures and are water resistant, which are important properties for food processing applications. Greases with calcium sulfonate thickeners have not been explored for approval by the USDA or FDA, but has been approved in Canada for incidental contact [11]. The list of approved base oils, additives, and thickeners for H1 incidental contact with food is available in Table 25.2.
25.7 Selecting What Machines Require Food-Grade Lubricants Selecting whether to use an H1 or H2 lubricant can be challenging. As previously mentioned, H1 lubricants are permitted where incidental contact might be possible, whereas an H2 lubricant is only permitted where there is no possible contact with the food product. For example, a lubricant used on a conveyor system running over a food line must be an H1 category oil, while a conveyor system running underneath a food line may use either an H1 or H2 lubricant. Because H1 lubricants are limited by types of additives and in the past only used mineral oil basestocks, H1 lubricants in certain instances provided less protection and shorter lubricant life. Now that synthetics are used, some H1 lubricant performance can exceed nonfood-grade lubricants. This is highly significant in allowing consolidation and avoiding accidental cross-contamination of H1 and H2 oils, and contamination of H2 oils with food [7]. Tables 25.3 through 25.9 are designed as a quick reference for some food processing applications generic for several types of industries [12]. The specific application should be checked to verify the lubricant grade or viscosity. The tables do not identify whether to use an H1 or H2 lubricant. It is ultimately the food-processing plant’s decision to determine whether an H1 is required or if an H2 lubricant is allowable.
25.8 Selecting an H1 Food-Grade Supplier [13] Finding the right lubricant supplier is as important as selecting the right lubricant. It is important to find a food-grade lubricant supplier that understands specific applications and requirements. Also, a supplier can serve as a part of the maintenance department, to help educate staff on lubrication maintenance and provide training to get the most performance and service life possible out of the lubricant. Other important qualities of a lubricant supplier are product consolidation, oil analysis, on-time delivery, speedy response to questions, and ability to tailor products to client needs.
© 2006 by Taylor & Francis Group, LLC
Food-Grade Lubricants and the Food Processing Industry TABLE 25.6
25-13
Dairy Products
Separators/clarifiers Enclosed gears
ISO 100 R&O or AW Oil or ISO 220 Gear Oil
Labelers Greased bearings Enclosed gears
NLGI No. 2 Grease ISO 220 Gear Oil
Homogenizers/pasteurizers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 100 R&O or AW Oil or ISO 220 Gear Oil Hydraulics ISO 46 R&O or AW Oil
Casers/packers/stackers/destackers Geared bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains/open gears NLGI No. 2 Grease Spray Guides NLGI No. 2 Grease Spray Hydraulics ISO 46 R&O or AW Oil Air Line Lube ISO 46 R&O or AW Oil
Tank-vat agitators Greased bearings Enclosed gears
NLGI No. 2 Grease ISO 220 Gear Oil
Cheese fillers/presses Greased bearings Oiled bearings Slides Cams
NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 100 R&O or AW Oil NLGI No. 2 Grease Spray
Fillers/cappers Geared bearings Enclosed gears Chains/open gears Air Line Lube
NLGI No. 0 or 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil
Butter churns/boats Greased bearings Enclosed gears Hydraulics Rear leg
NLGI No. 2 Grease ISO 220 Gear Oil ISO 46 R&O or AW Oil ISO 68 Turbine Oil
Packagers Geared bearings Enclosed gears Chains/open gears Air Line Lube
NLGI No. 0 or 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil
Centrifuges Oiled bearings
ISO 100 R&O or AW Oil
Dryers Greased bearings Enclosed gears
NLGI No. 2 Grease ISO 220 Gear Oil
Fruit feeders Hydraulics
ISO 46 R&O or AW Oil
Powder baggers/bag closers Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 Gear Oil
Ice cream freezer Greased bearings Oiled bearings Enclosed gears Chains Valves
NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil
Mixers/hammers/mills/vibrators Greased bearings NLGI No. 2 Grease Enclosed gears ISO 100 R&O or AW Oil
Ice cream fillers Greased bearings Slides Oiled bearings Clutches
NLGI No. 2 Grease NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 68 Turbine Oil
Dust collectors Greased bearings Enclosed gears
NLGI No. 2 Grease ISO 100 R&O or AW Oil
Conveyers Greased bearings Oiled bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray
Liquifiers Greased bearings
NLGI No. 2 Grease
Pumps Greased bearings
NLGI No. 2 Grease
Thermutators Greased bearings
NLGI No. 2 Grease
Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With permission.
© 2006 by Taylor & Francis Group, LLC
25-14
Handbook of Lubrication and Tribology
TABLE 25.7
Fat and Oil Products
Expellers Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray
Centrifuges Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray
Bean cleaners/shakers/dehullers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray
Mixers Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray
Dryers/condensers/coolers/toasters Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chain/open gears NLGI No. 2 Grease Spray
Dust collectors Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray
Crackers/grinders/hammer mills Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray
Conveyors Greased bearings Oiled bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray
Flakers Greased bearings Oiled bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray
Pumps Greased bearings
NLGI No. 2 Grease
Solvent extractors Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray
Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With permission.
25.9 Global Trends [2] USDA H1 and H2 still stand as a recognized approval for food and drug suitability. In fact, many lubricant manufacturers still aspire to the USDA H1 and H2 categories and approval process, and supply certification from their boards of directors to guarantee that claim. However, efforts chaired by Klüber Lubricants of Germany led to the creation of a new standard, DIN V 0010517, 2000-08 (Food-grade Lubricants — Definitions and Requirements). This standard has since been approved at a higher DIN level. This German standard has been submitted by DIN as a draft to ISO in Geneva. It may take several years from the date the application is accepted for an international standard to be released. NSF has evolved globally to succeed the USDA. NSF International, The Public Health and Safety Company™ , has been committed to public health, safety, and protection of the environment for more than 55 years. NSF has earned the Collaborating Center designations by the World Health Organization (WHO) for both food safety and for drinking water safety and treatment. It is conceived and administered as a public service organization serving as an independent and neutral body to resolve issues between regulatory bodies, business, industry, and the public. The DIN standard V 0010517, 2000-08 has also been adopted by ELGI and NLGI as their guideline.
© 2006 by Taylor & Francis Group, LLC
Food-Grade Lubricants and the Food Processing Industry TABLE 25.8
25-15
Grain Mill Products
Milling separatos/degerminators Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray
Tempering bins Greased bearings Enclosed gears
Shakers Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray
Grinding mills/hammer mills/comminuters/crumblizers Geared bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Adjusting screws NLGI No. 2 Grease Spray
Washers/cleaners Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray
Dryers Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray
Evaporators Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray
Ovens/cookers Greased bearings Enclosed gears
NLGI No. 2 Grease ISO 220 or 460 Gear Oil
Chains
NLGI No. 2 Grease Spray
NLGI No. 2 Grease ISO 220 Gear Oil
Sifters Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Loaf molders/extruders Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray Hydraulics ISO 46 R&O or AW Oil Dough mixers Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray Hydraulics ISO 46 R&O or AW Oil Centrifuges/filters/oil extractors Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil or ISO 220 Gear Oil Enclosed gears ISO 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray
Package fillers/baggers/bag closers Greased bearings NLGI No. 2 Grease Oiled bearings ISO 46 R&O or AW Oil Enclosed gears ISO 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray
Dust collectors Greased bearings
Mixers/blenders Greased bearings Enclosed gears Chains
Conveyors Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray
Pellet mills — feed processing Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 100 R&O or AW Oil ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray Hydraulics ISO 46 R&O or AW Oil Coolers Greased bearings Enclosed gears Chains
Pumps Greased bearings
NLGI No. 2 Grease
NLGI No. 2 Grease
NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray
Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With Permission.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology TABLE 25.9
Meat, Seafood, and Poultry
Parts washers/scalders Greased bearings NLGI No. 2 Grease
Mixers/mincers Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 Gear Oil Chains/open gears NLGI No. 2 Grease Spray Air Line Lube ISO 46 R&O or AW Oil
Feather pickers
Meat saws/meat and bacon slicers/peelers/skinners/ chippers/venters Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 460 Gear Oil Chains NLGI No. 2 Grease Spray Air Line Lube ISO 46 R&O or AW Oil
Greased bearings
NLGI No. 2 Grease
Smoke houses/ovens Greased bearings NLGI No. 2 Grease Linkage NLGI No. 2 Grease Spray Open gears NLGI No. 2 Grease Spray
Sausage linkers/frank machines/patty machines Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 100 R&O or AW oil or 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray
Cookers Greased bearings Linkage
NLGI No. 2 Grease NLGI No. 2 Grease Spray
Stuffers Greased bearings Oiled bearings Enclosed gears
Open gears
NLGI No. 2 Grease Spray
Chains
NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 100 R&O or AW Oil or 220 or 460 Gear Oil NLGI No. 2 Grease Spray
Grinders/disintegrators Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 Gear Oil Chains/open gears NLGI No. 2 Grease Spray Air Line Lube ISO 46 R&O or AW Oil
Centrifuges/separators/dryers/filters Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 or 460 Gear Oil Air Line Lube ISO 46 R&O or AW Oil
Pickling injectors Greased bearings Enclosed gears Chains Hydraulics
Graders/deboners Greased bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 220 or 460 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil
Can washers, fillers and closers/labelers/packers/ wrappers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray Hydraulics ISO 46 R&O or AW Oil Air Line Lube ISO 46 R&O or AW Oil
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray
Freezers Greased bearings Enclosed gears Chains Hydraulics
NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil
Rehangers/neck breakers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 or 460 Gear Oil Hydraulics ISO 46 R&O or AW Oil
Conveyors Greased bearings Oiled bearings Enclosed gears Chains
NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray
Eviscerators/remove entrails Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 or 460 Gear Oil Hydraulics ISO 46 R&O or AW Oil
Pumps Greased bearings
NLGI No. 2 Grease
© 2006 by Taylor & Francis Group, LLC
Food-Grade Lubricants and the Food Processing Industry
25-17
TABLE 25.9 Continued Gizzard machines/lung pullers/neck skin cutters/venters Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 or 460 Gear Oil Hydraulics ISO 46 R&O or AW Oil Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With permission.
25.10 Religious Organizations Influence in Food-Grade Lubricants The Muslim and Jewish religions further restrict the formulation of food-grade lubricants. Today, there are approximately 14 million Jews and 1.3 billion Muslims worldwide [14]. Both religions have rules covering aspects of food processing. “Kosher for Pareve,” or Kosher, is the term used to describe Jewish dietary laws. Kosher law is approved by several rabbinic orders. In the United States, the Orthodox Union and the Organized Kashrus Laboratories are major approval organizations active in the approval of food-grade lubricants. Kosher law outlaws the use of pork and pork by-products. Kosher law also prohibits any mixing of meats and dairy and eggs. Any equipment must be properly cleaned and left idle for 24 h before and after making kosher foods [1]. Under Islamic law, “Halal,” meaning lawful or permitted in Arabic, laws are imposed on their food products. In the United States, the Islamic Food and Nutrition Council of America issues Halal Certificates. Similar to Kosher laws, Halal foods exclude the use of pork and pork by-products. Also, Halal excludes the use of alcohol in its products, which potentially limits some of the additives used in food-grade lubricants [1].
25.11 Conclusions The food and beverage processing industries with respect to food-grade lubricants has changed dramatically within the last five years. Understanding the differences between H1, H2, and H3 lubricants and making the proper lubricant selection is critical to food safety and machine reliability. As an additional source, NSF’s website provides lubricant requirements for food-grade products and gives a free access listing of certified food-grade lubricants on their website at www.nsfwhitebook.org.
References [1] Raab, Michael J., Food-grade lubricants: a new world order, NLGI Spokesman, 66, 2, 2002. [2] Williamson, M., Understanding food-grade lubricants. Machinery Lubrication Magazine, 64, 2003. [3] Hodson, D., Food-grade lubricants reduce contamination threats for food and beverage manufacturers. Machinery Lubrication Magazine, 24, 2004. [4] NSF International Registration Guidelines (July 2003) version 3.3, Retrieved October 2004 from http://www.nsf.org/business/nonfood_compounds/guidelines.pdf. [5] Girard, J., The continuing evolution of food-grade lubricants. Machinery Lubrication Magazine, 20, 2002. [6] Email correspondence with Dr. Kenji Yano, program manager for NSF International Nonfood Compounds Registration Program. [7] Lauer, D.A., Special lubricants for the food-processing and pharmaceutical industries. Lubrication Excellence 2003 Conference Proceedings, 439, 2003. [8] 21 CFR 178.3620 — Technical White Mineral Oil as a Component of Nonfood Articles Intended For The Use of Contact with Food. Retrieved November 2004 at http://www.gpoaccess.gov/cfr/index.html.
© 2006 by Taylor & Francis Group, LLC
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[9] 21 CFR 178.3570 — Lubricants with Incidental Food Contact. Retrieved November 2004 at http://www.gpoaccess.gov/cfr/index.html. [10] Food-grade lubricants, Machinery Lubrication Seminar, Noria Corporation, 209, 2004. [11] Mackwood, W. and Muir, R., Calcium sulfonate complex grease: the next generation food machinery grease. NLGI Spokesman, 17, 2003. [12] Food Processing Industry. Lubricant Selector Guide. Lubrication Engineers, Inc. (no date) [13] Lesinski, D.J. and Raab, M.J., Brand insurance: using the right food lubricant to protect your company. Machinery Lubrication Magazine, 54, 2003. [14] Major Religions of the World Ranked by Number of Adherents, retrieved from Adherents website December 2004 at http://www.adherents.com/Religions_By_Adherents.html, last updated September 2002.
© 2006 by Taylor & Francis Group, LLC
26 Aviation Industry 26.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.2 Lubrication of Aviation Piston Engines . . . . . . . . . . . . . .
26-1 26-1
Types of Engines • Wet Sump Engines • Dry Sump Engines • Pressure Section • Scavenging Section • External Lubrication System • Supply Tanks • Oil Temperature Control Devices • Oil Pressure and Temperature Gauges • Types of Piston Engine Oil • Grades of Piston Engine Oil • Important Oil Properties • Oil Drain Intervals
26.3 Lubrication of Aviation Turbine Engines . . . . . . . . . . . .
H.A. Poitz and R.E. Yungk Air BP Lubricants
26-10
Turbine Engine Lubrication Systems • Turbine Engine Oils • Classification of Turbine Engine Oils • Important Oil Properties • Spectrographic Analysis (SOAP)
26.4 Airframe Lubrication . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
26-13 26-15
26.1 Introduction In 2003 the U.S. aviation industry consisted of approximately 219,000 active aircraft including those operated by the airlines, general aviation, helicopters, and both piston engine powered and turbine powered aircraft. The majority of this number (211,000) is piston engine powered. The amount of lubricants consumed annually by this aviation population consists of about 8 million gal of piston engine oil. A similar volume of turbine oil is also consumed, but some of this is used in marine and power generation equipment. The General Aviation and Regional/Commuter segments of the industry are expected to enjoy the most future growth. Most of this growth will be turbine engine powered. Factors for this turbine powered aircraft growth are the increases in fractional jet ownership and regional jets providing point-to-point service [1].
26.2 Lubrication of Aviation Piston Engines 26.2.1 Types of Engines There are basically two types of piston engines involved in general aviation: radial engines (Figure 26.1) and “so-called” flat engines or horizontally opposed engines (Figure 26.2). Most radial engines have been manufactured by Pratt & Whitney or Curtiss-Wright. The bulk of general aviation is powered by flat engines manufactured by either Textron Lycoming or Teledyne Continental Motors. 26-1
© 2006 by Taylor & Francis Group, LLC
26-2
FIGURE 26.1
Handbook of Lubrication and Tribology
Pratt & Whitney R-1830 Twin Wasp — a 1200 HP, 14 cylinder radial engine.
Various subtypes of these flat engines can be indicated by a letter prefix such as the letter) meaning opposed cylinders, VO — vertical opposed such as in helicopter use, IO — fuel injected cylinders opposed, TIO — turbo charged cylinders opposed, and R — radial cylinders not opposed. The subscript such as O-360 indicates not only that the cylinders are opposed, but also that the total displacement of the cylinders is 360 in.3 .
26.2.2 Wet Sump Engines Most low output engines used in the so-called “light” or “private” plane class employ the “wet sump” type of crankcase lubrication quite similar to that used in automotive. As indicated in Figure 26.3, the crankcase is the oil reservoir in the wet sump system. From this source the oil passes through an internal strainer into a gear pump which forces the oil under pressure into a passage leading to the propeller end of the camshaft. From here the oil flows through the hollow camshaft to the cam bearings. Each cam bearing is provided with an annular groove from which the oil passes through a drilled passage in the crankcase webs to the main bearings. Oil is introduced into the hollow crankshaft through the main bearings and is forced through passages in the crankshaft throws to the connecting rod bearings. Surplus oil escaping from the connecting rod bearings is thrown by centrifugal force onto adjacent parts where it serves to lubricate cylinder walls, piston rings, and pins as well as to cool the pistons as it strikes the underside of the piston heads. In some larger engines actual oil jets impinge on the under side of piston heads to assist in cooling this area and lubricating cylinder walls.
26.2.3 Dry Sump Engines The internal lubricating system of larger aircraft engines and especially radial engines is of the somewhat more complex dry sump type (Figure 26.4). Such systems can be discussed with greater simplicity by
© 2006 by Taylor & Francis Group, LLC
Aviation Industry
FIGURE 26.2
Lycoming AEIO-540 — a 300 HP, opposed 6 cylinder engine.
26-3
© 2006 by Taylor & Francis Group, LLC
26-4
Handbook of Lubrication and Tribology
Breather port Bearings Oil separator
Air
Engine lubrication points Pressure screen
Camshaft, rocker arms Piston rings Push rods, sockets
Oil pump Suction screen
Oil
FIGURE 26.3
Typical wet sump lubrication system on an aircraft engine.
Breather line
Oil separator Breather port
Engine lubrication points
Scavenge pump
Air
Oil cooler Oil filter
Oil tank Lubrication pump
Oil scavenge points
FIGURE 26.4
Oil
Typical dry sump lubrication system on an aircraft engine.
dividing them into pressure section, a scavenging section, and an external system. Pressure section includes the pressure pump, relief valve, check valve, strainers and/or filters, pressure gauge, and the necessary interconnecting pipes and passages. The scavenging section includes the collecting sumps, scavenging pumps, pump screens, and the necessary piping for interconnecting the various units. Finally, the external system includes the oil tanks, controls, and instrumentation.
26.2.4 Pressure Section In many modern engines, the pressure and scavenging pumps may be contained in the same housing and driven by the same shaft. The pressure pump inlet is connected in such a manner as to permit the pump to supply oil as long as any remains in the supply tank. The capacity of the pump is sufficient to furnish a surplus volume of oil at ample pressure to all pressure lubricated parts. Surplus oil from these parts lubricates adjacent parts by the splash system. To prevent excessive rise in pressure, a ball or piston-type relief valve is either built into the pump or connected at some point in the pressure system convenient for adjustment. In inverted or radial engines, where all or part of the cylinders operate in an inverted position, another spring-loaded check valve is often provided to prevent oil from flowing into the engine and accumulating in the combustion chambers of the lower cylinders when the engine is not operating.
© 2006 by Taylor & Francis Group, LLC
Aviation Industry
26-5
Oil strainers and filters are usually used to remove solid contaminants such as carbon, dust, dirt, and metal particles from the oil. The filters are commonly installed on the pressure side of the oil pump and may be any one of several types: non-cleanable depth media element, centrifugal, or stack metal disk. It is to be emphasized that no currently known oil filter will successfully remove or completely neutralize the various forms of liquid and oil soluble contaminants resulting from the deterioration of the lubricant itself. One important feature of many aviation engine lubrication systems is an “anti-sludge” tube to prevent the accumulation of sludge in the crankshaft throws from clogging the hole in the crankshaft through which the oil is supplied to the bearing. The anti-sludge tube is simply a small tube connected to the oil hole in the journal and projecting vertically into the large oil reservoir within the hollow crankshaft. The high speed at which the crankshaft rotates exerts considerable centrifugal force on the oil passing through the hollow shaft. As a result, sludge, which might otherwise stop the flow of oil through the oil hole to the bearing, is effectively thrown out and accumulates harmlessly in the cavity inside the crank pin. Drain plugs, located on the “cheeks” of the crankshaft throw, provide openings at both ends of the anti-sludge chamber for cleaning during overhaul. In addition to supplying oil to the engine, the internal lubrication system also provides lubrication for accessories such as the distribution shaft, tachometer drive, vacuum pump drive, supercharger mechanism, and propeller-reduction gears. Furthermore, where controllable pitch or contact speed hydraulic propellers are used, engine oil must be supplied both for their lubrication and their hydraulic actuation.
26.2.5 Scavenging Section Some engines of the larger sizes are provided with as many as three sumps and an equal number of scavenging pumps. In engines such as the twin row radial type, one sump is usually located at a low point in the nose section, another between the rocker arm boxes of the lower cylinder, and the third or main sump at the lowest point in the crankcase. These various sumps are provided to collect oil readily from the different positions in the engine regardless of its attitude. For example, in a “nose down” position most of the oil will flow to and must be scavenged from the main or engine sump, while some also must be collected and scavenged from the nose section. Meanwhile, excess oil from the rocker arm boxes is collected in the sump located between the rocker arm boxes of the lowest cylinder. In a single row radial engine usually one engine sump located at the lowest point in the crankcase collects all the return oil. In connection with the scavenging system, Figure 26.5 shows how excess oil is prevented from flowing into the lower cylinders of a radial engine. Sumps are always provided with drain plugs to permit complete drainage during oil change. The drain plugs are often provided with small magnets which collect metallic particles resulting from wear or breakage. Wire mesh screens may also be located in the lines. The scavenging pump of an aviation engine is very similar in construction to the pressure pump although it is usually considerably larger to keep the sump relatively free from oil under all conditions. The scavenging pump must first lift the oil from the sump and then force the oil through the restricted passages of the oil cooling unit to the top of the oil supply tank. Scavenging or return oil lines from the sump to the scavenging pump may be either externally located or built integrally with the crankcase section. A breather is provided on the upper part of the crankcase to relieve internal pressure resulting from high-speed piston action, blowby, and the high temperatures prevalent in aircraft engines. Such breathers are so constructed as to relieve the pressure without excessive oil loss. In some engines the breather unit may include a gravity operated valve which remains open when the engine is in a normal position and closes when the engine is inverted to prevent oil loss.
26.2.6 External Lubrication System For purposes of clarity the external may be broken down into three units: the supply tank or tanks and their related parts; the oil temperature controlling units; and the necessary pressure gauges, temperature gauges, and related acessories.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
FIGURE 26.5
26.2.7 Supply Tanks These tanks must be capable of withstanding internal test pressures of approximately 5 lb/in.3 without failure or leakage. The tank should be located as close as possible to the engine and in most cases at a sufficiently higher level to provide for gravity flow to the engine when it is in normal position. Capacity of the supply tanks will vary according to engine size, and regulations require a certain minimum oil capacity per gallon of fuel capacity. If the engine manufacturer’s specifications for oil capacity exceed the minimum governmental requirements, then the capacity must be no less than that recommended by the manufacturer. Regulations also require that oil tanks be properly vented and provided with an expansion space which cannot be unintentionally filled with oil. In preparation for starting in cold weather, viscosity of the oil may be lowered by means of electrical heaters lowered into the oil tank just prior to starting. Electric current from the hangar circuits may operate these auxiliary heating devices.
26.2.8 Oil Temperature Control Devices One of the most important units for control of oil temperature in the external lubrication system is the oil cooler, a heat exchanger similar to an automobile radiator. Air from the outside is the heat exchanging medium, operating in the same manner as air passing through the radiator of an automotive vehicle (Figure 26.6). Two types of control valves are used to determine whether or not the oil should be bypassed around or directed through the radiator core (1) the thermostatic type which is sensitive to change in temperature
© 2006 by Taylor & Francis Group, LLC
Aviation Industry
26-7
Oil in
Oil out
Air flow
FIGURE 26.6
and (2) the spring loaded or pressure type. When the oil returning from the engine is comparatively cool, the thermostatic valve opens and permits the oil to bypass the cooling surface and flow through the core jacket and then to the top of the oil supply tank. As the oil returning from the engine increases in temperature, the thermostatic valve, by means of a temperature sensitive element closes the bypass passage and requires the oil to flow through the cooling core. Thus by varying the amount of oil passing around and through the cooler, the thermostat maintains the required temperature. The spring loaded valve depends entirely upon the pressure in the oil leading from the scavenging pump. When the oil is cool and its viscosity is high, an excessive pressure is required to force it through the cooling passages. This pressure causes the spring loaded valve to open and allow the oil to pass directly into the tank without cooling. As the temperature of the oil increases due to engine operation, its viscosity decreases and less pressure is required to pump it. The valve then closes resulting in the oil being forced through the cooling core. In general, both the thermostatic and the pressure actuated types are calibrated to require passage of the oil through the cooling core when the returning oil reaches approximately 60◦ C.
26.2.9 Oil Pressure and Temperature Gauges To enable the pilot or flight engineer to monitor operation of the lubrication system, temperature and pressure gauges are provided. The pressure gauge usually indicates oil pressure at or near the main gallery line of the engine. The temperature of the oil is generally determined from thermocouples and is indicated on gauges located in the cockpit control panel.
26.2.10 Types of Piston Engine Oil The majority of aircraft piston engines in operation today were developed on a non-additive petroleumbased oil that has been in existence since the 1920s and is presently covered by SAE J1966 [2], formally Military Specification, MIL-L-6082. Little significant change has occurred in these specifications and usage of this type of oil is still sizable. Very few significant improvements in piston engine oil were made until 1958, when nonmetallic dispersants, antioxidants, and antiwear additives were introduced. This type of oil offered significant
© 2006 by Taylor & Francis Group, LLC
26-8
Handbook of Lubrication and Tribology TABLE 26.1 Military grade 1065 1080 1100 1120
Grades of Piston Engine Oil Viscosity range
Commercial grade
SAE grade
cSt at 100◦ C
SSU at 210◦ F
65 80 100 120
30 40 50 60
10.7–12.4 14.2–16.7 18.7–21.1 23.9–26.1
62–68 75–85 93–103 115–125
improvements in engine cleanliness, reduced wear in oil lubricated areas, and improvement in oil life in high temperature operation. Military Specification MIL-L-22851 was written around this type of oil, but this has been superceded by SAE J1899 [3]. This type now represents over 70% of aircraft piston engine oil sold. Although oils meeting this specification soon became incorrectly known as “detergent” oils, they were never intended to clean up dirty engines. Their correct designation is “dispersant” oils, inasmuch as they have the ability to disperse or keep in suspension products of combustion that blew by the piston rings, as well as products of oil degradation. By keeping these products in dispersed condition, they were removed from the engine when the oil was drained and this led to much cleaner engines. To take advantage of the performance of the newer dispersant oil, some engine models have been type-tested on this oil and restricted to its use. Some new engines have been designed and certified to only use dispersant oils.
26.2.11 Grades of Piston Engine Oil SAE Technical Committee TC-8 is the custodian of the two piston oil specifications used today. J1966 (nondispersant) and J1899 (dispersant) oils both include viscosity grades 65, 80, 100, and 120 as well as multigrade viscosities. The larger the grade number, the more viscous the oil. These commercial grade numbers are the approximate midpoint of the SSU viscosity range at 100◦ C. Early Pratt & Whitney and Wright Aeronautical Specifications, as well as those of the smaller engine manufacturers Avco-Lycoming and Continental Motors, only covered Grades 120 and 100. In later years, the smaller engine manufacturers expanded their specifications to include Grades 80 and 65, primarily for use in colder climates. Present-day designations have dropped the four-digit military grade numbers in favor of a two or three digit grade, Grade 1080 now becoming Grade 80. Table 26.1 shows the historical military grade, commercial grade, SAE grade (automotive oil), and viscosity ranges of the four grades of oil.
26.2.12 Important Oil Properties Lubricating oil in an aircraft engine must serve many functions in temperature ranging from −35◦ C to over 95◦ C, and at elevations from sea level to 6000 m. These functions include lubrication of cylinders, piston rings, valves, gears, and bearings to minimize wear; cooling of the engine hot areas; and performing as hydraulic oil for operation of variable pitch propellers as well as other hydraulically operated mechanisms. Because of these many functions, the following properties are usually defined in the many specifications covering aircraft oil: • Viscosity (ASTM D445) — helps provide a correct lubrication film and correct operating temperatures. Both pour point and viscosity play a major role in the crank-ability of the engine at cold temperatures. • Viscosity Index (ASTM D2270) — the higher the viscosity index, the better the oil resists thinning at the elevated temperature at which an aircraft oil performs. The oil must remain thick enough to provide an effective film between hot metal parts yet must be fluid enough to be readily circulated through the oil system.
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Aviation Industry
26-9
• Pour point (ASTM D97) — discussed previously and relates to the ability to crank the engine in cold weather. • Flash point (ASTM D92) is used to control flammability and volatility of the oil. • Sulfur content by x-ray spectroscopy (ASTM D2273) — sulfur not removed in the refining process can react with moisture and other products of combustion to form compounds highly corrosive to aircraft engine bearings and valve guides. Total sulfur is controlled by specifications and the copper strip corrosion test (ASTM D130) controls corrosivity by sulfur compounds. • Neutralization Number (ASTM D664) — acidic material in an oil may corrode bearings and valve guides and must be controlled. • Ash Content (ASTM D482) — low ash content is also desirable to preclude formation of excessive combustion chamber deposits that might turn into miniature preignition sites and cause extensive piston damage. The limits of ash content usually approach zero for aircraft engine oils. • Effect of an oil on metals (ASTM D130) — to predict the effect of the oil on bronze valve guides and other copper containing components. In the newer ashless dispersant oils the following additional properties are quite important: • Wear properties (ASTM D6709) — measures the ability of the oil to minimize wear as the result of the use of antiwear additives. Testing involves measuring wear on various parts of a small engine in prescribed test. • Antifoam (ASTM D892) — especially important at higher altitudes, because excessive foaming could result in pressure pump starvation and damage as well as foaming out of a tank vent system and partial loss of oil. Foaming is usually controlled by a small addition (5 ppm) of a commercial antifoam additive that modifies surface tension of the oil and permits rapid release of entrapped air. All candidate oils show satisfactory performance (wear, deposit control, oil condition control, dispersancy, etc.) in a 150 h endurance test run on a Textron Lycoming TIO-540-J2BD engine under a test program listed in SAE J1899 and J1966.
26.2.13 Oil Drain Intervals Oil drain intervals for aircraft piston engines are established by the engine manufactures. During the time when the major power sources of larger aircraft used by the airlines and daily utilization was high, many airlines operated their engines full overhaul life with no drains for the lube oil. This was possible because the clearances in the large radial engine were such that a large amount of oil escaped by the piston rings and valve stems and was burned by the exhaust system, accounting for very large oil add amounts and a virtual replacement of used oil with makeup or new oil every 40 to 50 h. Flat engines manufactured by Textron Lycoming or Teledyne Continental Motors are usually drained at 50-h intervals unless equipped with approved full-flow filters and then may be extended to 100-h drain intervals. Again, the maximum authorized drain period is specified by the engine manufacturer in the operator’s handbook. Engine manufacturers sometimes allow extension of oil drain periods under very well regulated conditions. For the average operator, however, it is imperative that they abide by the present oil drain periods as specified in the aircraft operating manual. Some factors that influence oil drain intervals can be itemized as follows: 1. Current engine age In a new engine or one that has very low oil consumption because of tight clearances, the oil remains in the crankcase in usage for a considerably longer time than in an engine that is toward the end of its overhaul, and has worn sufficiently to allow higher oil consumption and thus a much higher rate of new oil makeup. With this in mind, oil should be drained more frequently in a new or tight engine than in an older engine where new oil is added quite frequently. 2. Average monthly usage If an aircraft only flies 4 or 5 h on the weekend and then is idle the remainder of the week, a very severe service condition is established with regard to oil deterioration. The engine lands
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hot, cools down, and water then condenses in the crankcase and reacts with the combustion chamber blowby products to form highly corrosive acids. If these acids are not burned off by operating the engine again, they will tend to corrode all of the bearings and cause considerable degradation to the oil itself. Conversely, an operator that flies every day for long periods of time, for instance 10 h a day, causes much less severe stress on the oil than the “weekend pilot.” 3. Filtration While filtration was once considered to impose quite a weight penalty, recent studies have indicated that good full-flow filters can double the oil drain interval under normal circumstances. Engine manufacturers now issue service bulletins indicating approved filter kits with many types of oil, which can double the authorized oil drain period. 4. Type of oil The most desirable type of oil available at the present time, with regard to compatibility with filters and resistance to sludge and carbon deposit build-up is the ash-less dispersant oil. This type provides the maximum oil drain intervals permitted by the engine manufacturers.
26.3 Lubrication of Aviation Turbine Engines 26.3.1 Turbine Engine Lubrication Systems Turbine engines have lubrication systems very similar to the piston engine dry sump systems described earlier. Temperature can be much higher with compartment drain temperature in excess of 150◦ C. Because of the high heat loads inherent in a turbine engine due to both radiant heat and internal friction, lube systems employ both air and fuel heat exchanges for temperature control. Bearing compartments and gear box accessory (for fuel pumps, generators, air driven starters, oil pumps, etc.) drive shafts are sealed with carbon face seals, labyrinth seals, or hydraulic seals. Static seals
Overboard discharge Breather pressure valve Deoiler
Rear bearing compartment
Gearbox
Front bearing compartment
Mid bearing compartment
Scavenge pump
Scavenge pump
Scavenge pump
Scavenge pump
Chip detector
Chip detector
Chip detector
Chip detector
Deaerator Oil reservoir Main oil pump
FIGURE 26.7
Main oil filter
Typical lubrication system on a gas turbine engine.
© 2006 by Taylor & Francis Group, LLC
Oil coolers
Aviation Industry
26-11
are typically accomplished with orings or gaskets made from fluorocarbon elastomers although some nitrile, silicone, and fluorosilicone seals are employed in older designs. Because of the good thermo-oxidative stability of modern day turbine oils, a reasonable level of consumption (0.2 to 1 lb/h), and high utilization rates found in commercial applications, turbine oil is seldom drained on a periodic basis. The oil remains serviceable by topping off lost oil between overhaul periods.
26.3.2 Turbine Engine Oils Aircraft gas turbines developed prior to 1948 were lubricated satisfactorily with light mineral oils. As power requirement increased and engine cycle temperatures rose, however, mineral oils lacked the necessary high-temperature stability. Severe oxidation of the oil resulted in excessive oil thickening. Oxidation and thermal degradation also initiated a search for a new class of oils which resulted in the development of the synthetic aircraft turbine oils. The synthetics in common aircraft use are ester fluids which have several inherently desirable characteristics. One of these is good response of viscosity to temperature (viscosity index). Another is good thermo-oxidative stability, as evidenced by high bulk fluid temperatures (up to 200◦ C) and thermal decomposition temperatures in transient hot spots (up to 370◦ C). Starting with such base oils, these properties are further enhanced by the addition of antioxidants, metal deactivators, load carrying agents, etc. Esters used as lubricating base oils have three or four ester groups attached to a central carbon atom. The ester groups may contain a variety of carbon chain links ranging from four to eleven, depending upon the selection of the alcohol and acids used as a starting material. With the selection of the proper base oil and the incorporation of the necessary additive systems, fully formulated ester oils have several good properties such as (1) excellent deposit control, (2) good lowtemperature fluidity, (3) good oxidative and thermal stability, (4) good bearing and gear fatigue resistance, (5) good gear load-carrying ability, (6) good corrosion protection, and (7) good compatibility with all engine materials of construction, especially non-metallic seals. Among these, the major feature most important to engine operation is good deposit control. Many engines used in aircraft do not receive full overhauls until 20,000 to 40,000 h have elapsed, thus any deposits formed within the engine may cause lubrication system problems. Such problems are manifested by plugging of oil jets, resulting in oil starvation to bearings and gears; plugging of the breather system, resulting in excessive oil system breather compartment pressures, that result in higher than normal compartment temperature and severe oil leakage; plugging of scavenge systems, resulting in severe oil loss; and carbon shaft seals being rendered inoperative thereby allowing excessive air leakage into the oil system, resulting in excessively rapid oil degradation and deposit formation.
26.3.3 Classification of Turbine Engine Oils Most of the turbine engine oils used today are described in military specifications MIL-PRF-23699 [4] and MIL-PRF-7808 [5]. The oil described by MIL-PRF-23699 is a 5 cSt viscosity grade (at 100◦ C) and serves not only U.S. Army and Navy turbine applications but also most commercial and general aviation turbine applications worldwide. There are now three classes of oil described in MIL-PRF-23699: STD or Standard class, HTS or High Thermal Stability class, and C/I or Corrosion Inhibiting class. The HTS class of oil uses very stable ester basestocks and advanced antioxidant packages to achieve optimized thermo-oxidative stability for the purpose of minimizing deposits that limit time between overhauls (TBO). The C/I class of oil uses a corrosion inhibition package to minimize corrosion in military applications where operating environments are particularly corrosive. MIL-PRF-7808 consists of two viscosity grades, 3 and 4 cSt (at 100◦ C) and is utilized in applications requiring low temperature start capability as low as −54◦ C. These applications include U.S. Air Force fighter aircraft in cold climates as well as commercial aircraft auxiliary power units (APUs) that endure
© 2006 by Taylor & Francis Group, LLC
26-12
Handbook of Lubrication and Tribology TABLE 26.2
Grades and Classes of Turbine Engine Oil
Specification
Viscosity cSt at 100◦ C
MIL-PRF-7808, Grade 3 MIL-PRF-7808, Grade 4
3 4
MIL-PRF-23699, Class STD MIL-PRF-23699, Class HTS MIL-PRF-23699, Class C/I DOD-PRF-85734
5 5 5 5
Attributes Lowest temperature start capability Improved cleanliness/stability with good low temperature start capability Higher viscosity for improved system durability Best cleanliness/stability Corrosion Inhibition for salt environments High load carrying for helicopter gearboxes
cold temperatures at altitude. The more viscous MIL-PRF-7808, Grade 4 also has higher thermal stability requirements recognizing the need for more stable oils for advanced military aircraft that operate at increasingly hotter temperatures. So-called high load carrying oils are typically used for helicopter transmission gearboxes and some gas turbine systems with highly loaded gear sets. These 5 cSt grade oils are described in DOD-PRF-85734 [6] are similar to the MIL-PRF-23699 oils with an additional Extreme Pressure (EP) additive system that promotes better boundary film lubrication in high load situations. The forum for many turbine oil issues, evaluation method development and future specifications is an SAE Standards Development committee, E-34 (Propulsion Lubrication). E-34 published a specification, AS 5780 [7], for 5 cSt turbine oils used in commercial applications. Original equipment manufacturer (OEM) are using this specification as part of their regulated oil approval activities.
26.3.4 Important Oil Properties In addition to the properties important to piston engine oils, turbine engine oils are evaluated for other lubricant attributes appropriate to their base chemistry, engine materials of construction and engine time between overhauls. Lubricant compatibility (FED-STD-791, Method 3403) evaluates the hot aging miscibility of candidate oils with other oils likely to be encountered in service. This is important given the possible additive package diversity as allowed by performance (not material) specifications. An acid assay (FED-STD-791, Method 3500 (1)) is performed to control the acid distribution used in an oil formulation’s ester basestock. This ensures the oil chemistry is controlled on a batch-to-batch basis similar to the originally qualified formulation. Elastomer compatibility (e.g., Def Stan 05-50 (Part 61) Method 22) measures swell and deterioration of seal materials in contact with hot oil. Elastomers of interest include fluorocarbon, perfluorocarbon, fluorosilicone, silicone, and nitrile. This becomes important because the ester basestocks and antioxidant packages tend to be aggressive toward fluorocarbon, while extreme pressure additives are aggressive toward silicone and fluorosilicone. Oxidation and Corrosion Stability (FED-STD-791, Method 5308 mod 1 or ASTM D4636) evaluates bulk oil stability at temperatures up to 218◦ C as well as any breakdown products consequential corrosive impact on representative metallurgies. Thermal Stability and Corrosivity (FED-STD-791, Method 3411) has proven an effective quality control method for detecting undesirable contaminations from other non-aviation ester based products. Deposit control is measured in a full scale, heated bearing rig (FED-STD-791, Method 3410, severity rating of 1.5). This test is run for 100 h (HTS oils require 200 h) with evaluation of oil condition control and a cleanliness demerit rating. Emerging subscale tests are being developed and evaluated to determine control of liquid phase deposits (SAE ARP 5996 to simulate oil pressure pipes and jets), mixed phase deposits (GE Alcor High Temperature Deposition Test to simulate scavenge system pipes), and vapor phase deposits (U.S. Navy test to simulate
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Aviation Industry
26-13
breather system pipes). This activity is indicative of the importance of controlling deposits in engines with ever increasing time between overhauls. Load carrying or boundary film lubricating ability is measured by the Ryder Gear test (FED-STD-791, Method 6508 and/or SAE AIR 4978 Appendix E). This becomes especially import for the DOD-PRF-85734 oils that are used in highly loaded helicopter transmission gearboxes. Hydrolytic stability (Def Stan 05-50 (part 61) Method 6) is important to storage stability and the stability of oils in capped (not breathed) lubrication systems often found in engine accessories such as generators and air starters. Hydrolysis is a degradation chemical reaction between an ester and water at elevated temperatures that reverts the ester to acid and alcohol destroying oil properties. Corrosion Inhibition in MIL-PRF-23699, Class C/I oils is controlled by SAE ARP 4249 ball corrosion testing. Trace metals are controlled because new oil needs to provide a baseline for spectrographic oil analysis program diagnostic testing discussed later.
26.3.5 Spectrographic Analysis (SOAP) The parts per million (ppm) of several metals (e.g., Fe, Ag, Cr, Al, Mg, Ti, Mo, V, etc.) in a sample of used oil is usually determined by exposure of a sample to an emission spectrometer. Spectrographic oil analysis was first used in the 1940s by the railroads to determine bearing wear in diesel engines, and was very successful. The military studied its use in predicting power plant failure in aircraft engines and the practice is now in widespread use. Many analytical laboratories offer such service to airlines and operators of general aviation. Provided the proper techniques are utilized in obtaining oil samples and sufficient background information is provided to the analytical laboratory on the type of engine and oil history, spectrometric analysis is a very useful tool in determining the condition of the oil. Caution should be observed in the use of one random sample of oil to assess any engine condition. Each engine has its own normal wear pattern: only after observing such a pattern over several oil changes in many hours of engine operation will significant deviations from this pattern become meaningful. Most operators use a combination of engine parameter trending (oil consumption, temperature, pressure, etc.), periodic examination of magnetic chip collector, examination of oil filter/screen for wear debris, and spectrometric oil analysis to judge the condition of their engines. One cardinal rule is to immediately obtain a second sample of oil for analysis if an alarm is sent back from the laboratory as to potential engine problems as observed by excessive wear metals in the spectrometric analysis. Spectrometric oil analysis of oils from turbine-powered aircraft is somewhat more meaningful than similar analysis of oils from piston-powered aircraft because of the absence of combustion by-products in the turbine oil. These by-products may mask determinations of some wear metals. Magnetic chip collectors incorporated into the scavenge system also serve as a valuable diagnostic compliment for engine condition monitoring. These chip collectors are inspected by mechanics at regular intervals. The debris collected can be evaluated with scanning electron microscopes equipped with energy dispersive spectroscopy (SEM/EDS) to ascertain the material type and wear mechanism by which the debris was generated. Advanced lubrication system designs are beginning to incorporate oil and debris monitors that are integrated into the engine’s control system providing a more comprehensive Prognostic Health Monitoring (PHM) system.
26.4 Airframe Lubrication Selection of proper lubricants, frequency of lubrication, and points requiring lubrication are the responsibility of the airframe manufacturer. This is true for all aircraft, whether a Piper Cub, or a Boeing 747 aircraft. A subcomponent manufacturer, for example an engine manufacturer or a flap gearbox manufacturer,
© 2006 by Taylor & Francis Group, LLC
250
CARBURETOR & CABIN HEAT GUIDE
100
ELEVATOR PULLEYS THROTTLE LEVERS
100
LUBRICANT HOURS IDLER PULLEYS 250 (SEE CAUTION 3)
100 RUDDER HINGES
FLAP PULLEY STABILIZER ADJUSTMENT PULLEY (SEE CAUTION 3) 250
RUDDER & 100 ELEVATOR HORNS
BRAKE MASTER CYLINDERS
ELEVATOR HINGES 100 LEFT & RIGHT
50
CONTROL STICK BEARINGS TORQUE TUBE BEARINGS 100
ELEVATOR PULLEY
ENGINE OIL SUMP DRAIN AND REFILL (SEE NOTE 6, PAGE 3)
100 TAIL WHEEL BEARING
100
50
26-14
LUBRICATIONCHART
HOURS LUBRICANT FLAP PULLEYS
50
TAIL WHEEL SWIVEL
STABILIZER ADJUSTMENT 100 MECHANISM (SEE CAUTION 3)
ENGINE
100 GREASE FITTING ELEVATOR & STABILIZER PULLEYS, FLAP HINGE 100 BEARINGS, FLAP CRANK & PUSH ROD BEARINGS (SEE NOTE 4)
RUDDER PEDAL BEARINGS BRAKE PEDAL BEARINGS 100 FLAP HANDLE BEARINGS FLAP HANDLE RATCHET 100
AILERON HINGE BEARINGS AILERON HORN 100 AILERON PULLEYS LEFT & RIGHT (SEE NOTE 4)
SHOCK STRUT PIVOTS
100
LANDING GEAR WHEEL BEARINGS
100
AILERON PULLEY 100 RUDDER PULLEY LEFT & RIGHT
GEAR HINGES 100 LANDING LEFT & RIGHT
LEGEND MIL-G-23827
NOTES 1. CARBURETOR AIR FILTER — CLEAN PER MANUFACTURER’S INSTRUCTIONS ON FILTER BOX OR INSTRUCTIONS IN OWNER’S HANDBOOK. (UNDER ABNORMAL CONDITIONS, FILTER REQUIRES CLEANING MORE FREQUENTLY. REPLACE AS REQUIRED.) 2. LUBRICATION POINTS — WIPE ALL LUBRICATION POINTS CLEAN OF OLD GREASE, OIL, DIRT, ETC. BEFORE RELUBRICATING. 3. WHEEL BEARING REQUIRES CLEANING AND REPACKING AFTER EXPOSURE TO ABNORMAL QUANTITY OF WATER. 4. AILERON AND FLAP HINGES-HINGE BLOCKS WITH LUBRICATION HOLES IN THEIR UNDERSIDE MAY BE PRESSURE LUBRICATED WITH GREASE MIL-23827
MIL-L-7870 MIL-L-3545 MIL-H-5606
ENGINE
GREASE, AIRCRAFT AND INSTRUMENT, GEAR AND ACTUATOR SCREW OIL-GENERAL PURPOSE LOW TEMP. LUBRICATION GREASE-LUBRICATION HIGH TEMPERATURE HYDRAULIC FLUID (RED)
SAE 50 ABOVE 60 F AIR TEMP SAE 40 BETWEEN 30 F AND 90 F AIR TEMP LYCOMING 0-290-D2 & 0-320 ENG. SAE 30 BETWEEN 0 F AND 70 FAIR TEMP SAE 20 BELOW 10 F AIR TEMP SAE 20 BELOW 32 F AIR TEMP SAE 40 ABOVE 32 F AIR TEMP CONTINENTAL C90 ENGINE SEE LYCOMING SERVICE INSTRUCTIONS NO. 1014 FOR USE OF DETERGENT OIL
FIGURE 26.8 © 2006 by Taylor & Francis Group, LLC
CAUTIONS 1. DO NOT USE A HYDRAULIC FLUID WITH A CASTER OIL OR ESTER BASE. 2. DO NOT APPLY LUBRICANT TO RUBBER PARTS 3. TRIM CABLES — UNDER NO CIRCUMSTANCES SHOULD THE TRIM CABLES FROM THE COCKPIT TO THE REAR OF THE FUSELAGE BE LUBRICATED. (TO PREVENT SLIPPAGE) 4. CONTROL CABLES — WIPE CLEAN AT REGULAR INTERVALS BUT DO NOT LUBRICATE. UNDER SALT CONDITIONS OCCAOCCASIONAL LUBRICATION WITH MIL-L-7870 IS RECOMMENDED.
Handbook of Lubrication and Tribology
50
CARBURETOR AIR FILTER (SEE NOTE 1)
Aviation Industry
26-15
may also specify a certain oil or grease for their particular unit, but final responsibility for the correct lubricant rests with the airframe manufacturer. Each aircraft has a lubrication chart provided by the airframe manufacturer (e.g., Figure 26.8) [8] that clearly specifies the type of lubricant (usually by a military specification), the point of lubrication, and the frequency of lubrication. In the case of aircraft, this figure is usually expressed in hours of operation instead of miles. If an aircraft is operated infrequently, that is less than 20 h/month, the lubrication intervals are normally shortened by half of what is shown in the maintenance manual. By specifying lubricants by military specification (consensus OEM specifications are under development) the airframe manufacturer does not restrict the buyer to one brand name product and widens the availability of the lubricant geographically. One brand-name product is occasionally recommended on an exclusive basis because of superior performance, but airframe manufacturers try to avoid this situation. The industry forum for developing both OEM based specifications and test methods are SAE AMS M Aviation Greases committee.
References [1] FAA Aerospace Forecasts, Fiscal Years 2004–2015, U.S. Department of Transportation, Federal Aviation Administration, Office of Aviation Policy and Plans, March 2004. [2] SAE Specification J1966, Lubricating Oils, Aircraft Piston Engine (Non-Dispersant Mineral Oil), 2005. [3] SAE Specification J1899, Lubricating Oil, Aircraft Piston Engine (Ashless Dispersant), 2005. [4] MIL-PRF-23699 Performance Specification, Lubricating Oil, Aircraft Turbine Engine, Synthetic Base, NATO Code Number O-156, 1997. [5] MIL-PRF-7808 Performance Specification, Lubricating Oil, Aircraft Turbine Engine, Synthetic Base, 1997. [6] DOD-PRF-85734 Performance Specification, Lubricating Oil, Helicopter Transmission System, Synthetic Base, 2004. [7] SAE Aerospace Standard AS5780, “Core Requirement Specification for Aircraft Gas Turbine Engine Lubricants,” 2005. [8] Lubrication Chart for the Piper Super Cub Aircraft provided by Piper Aircraft Corporation, 2002.
© 2006 by Taylor & Francis Group, LLC
27 Lubrication for Space Applications 27.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27.2 Lubrication Regimes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
27-1 27-2
Hydrodynamic, EHL, Mixed, and Boundary Lubrication • Factors Influencing Boundary Film Formation
27.3 Liquid Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
27-3
Types of Liquid Lubricants • Liquid Lubricant Properties
27.4 Greases and Solid Lubricants . . . . . . . . . . . . . . . . . . . . . . . . .
27-13
Greases • Solid Lubricants
27.5 Mechanism Components and Re-Lubrication Mechanisms . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
27-17
Spacecraft Components • Re-Lubrication Mechanisms
27.6 Lubricant Testing and Analysis . . . . . . . . . . . . . . . . . . . . . . .
27-20
Types of Testing • Accelerated Testing • Facilities for Space Lubricant Testing
William R. Jones and Mark J. Jansen NASA Glenn Research Center Tribology and Surface Science Branch
27.7 Lubricant Selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
27-29
Relative Life and Wear Characteristics • General Mechanism Effects • Cleaning and Surface Preparation
27.8 Summary . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
27-33 27-33
27.1 Introduction Space tribology is a subset of the lubrication field dealing with reliable satellite and spacecraft performance. It encompasses the entire gamut of tribologic regimes, including elastohydrodynamic lubrication (EHL), parched EHL, transient EHL, boundary lubrication, and mixed lubrication. Historically, choices of space mechanism lubricants were based on space heritage rather than on the latest technology or best available materials. With the limited mission lives and minimal duty cycles of the early space program, this strategy was highly successful. As missions extended, other spacecraft components such as electronics, batteries, and computers, failed before lubricated mechanisms [1]. However, during the 1980s and 1990s, these ancillary components vastly improved and tribologic systems have become one main factor limiting spacecraft reliability and performance. Although tribologic components represent only a small fraction of the spacecraft’s cost, they are often single point failures that cripple or debilitate expensive spacecraft.
27-1
© 2006 by Taylor & Francis Group, LLC
27-2
Handbook of Lubrication and Tribology h~0.0025 – 1.25 mm
h~0.0025 mm
h~0.0025 – 0.025 mm h >0.25 mm Hydrodynamic
Elastohydrodynamic
Mixed
Boundary
Cofficient of friction
.150
.001 Viscosity × Velocity , ZN Load P
FIGURE 27.1 Coefficient of friction as a function of viscosity–velocity–load parameter. (From Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.)
The Galileo spacecraft is a classic example of single point tribologic failure affecting the entire mission. The Space Shuttle Atlantis launched Galileo in 1989, starting a six-year journey to Jupiter. A high gain antenna, used to transmit control and telemetry data to earth, was one of the craft’s most important components. The umbrella shaped antenna was stowed closed behind a sun shield. In 1991, after the craft’s pass near the sun, the antenna was deployed, but it only opened partially. Engineers concluded three of the 18 ribs in the antenna’s umbrella-like structure were stuck in place. Ground-based tests [2] showed titanium alignment pins, which had been lubricated with a bonded dry film lubricant, had galled and subsequently seized due to lack of lubricant, causing the antenna failure. In this case, engineers salvaged the mission using the low gain antenna combined with data transmission advancements. This chapter will discuss basic lubrication ideas that play a major role in space mechanisms, space lubricant types, details of the most common space lubricants, mechanism components, testing and test facilities, and factors affecting lubricant selections.
27.2 Lubrication Regimes Lubrication separates surfaces in relative motion by interposing a third body that has low shear resistance, thus preventing serious surface damage or wear. The third body can be a variety of different materials including adsorbed gases, reaction films, and liquid or solid lubricants.
27.2.1 Hydrodynamic, EHL, Mixed, and Boundary Lubrication Depending on the third body type and thickness, several different lubrication regimes can be identified and are depicted in the Stribeck curve (Figure 27.1). Stribeck performed a series of journal bearing experiments in the early 1900s measuring friction coefficients as a function of load, speed, and temperature [3]. Later, Hersey performed similar experiments and devised a plotting format based on a dimensionless parameter, ZN/P [4]. The Stribeck/Hersey curve plots friction coefficient as a function of viscosity (Z), velocity (N), and load (P).
© 2006 by Taylor & Francis Group, LLC
Lubrication for Space Applications
27-3
When the ZN/P value is high, surfaces are completely separated by a thick (>0.25 µm) lubricant film and occur at high speeds, high viscosities, and/or low loads. In this region, termed hydrodynamic lubrication, lubricant rheology determines the friction. As the ZN/P parameter decreases, the lubrication regime changes from hydrodynamic to elastohydrodynamic, then to mixed, and finally to boundary. The EHL regime occurs in nonconformal, concentrated contacts where high loads cause surfaces to elastically deform and pressure–viscosity effects to occur in the lubricant. Film thickness in this regime ranges from 0.025 to 1.250 µm. As ZN/P continues to decrease, film thickness also decreases and surface interactions start taking place. This regime, where both surface interactions and fluid film effects occur, is referred to as the mixed regime. Finally, at low ZN/P values, the boundary lubrication regime is entered, where surface interactions are the primary factor. The boundary lubrication regime is a highly complex arena involving metallurgy, surface topography, physical and chemical adsorption, corrosion, catalysis, and reaction kinetics [5, 6]. Formation of protective surface films, which minimize wear and surface damage, is the regime’s most important characteristic. Typically, space mechanisms operate in the EHL, mixed, or boundary lubrication regimes, with the boundary lubrication regime being the most severe.
27.2.2 Factors Influencing Boundary Film Formation Both lubricant and bearing surface chemistry govern film formation. Additional environmental factors, such as temperature, also influence the lubricant’s film forming ability. The lubricant’s physical properties determine the film’s effectiveness at minimizing wear. Properties affecting film formation include shear strength, thickness, surface adhesion, film cohesion, melting point or decomposition temperature, and bulk lubricant solubility. 27.2.2.1 Starved EHL An EHL subdivision, starved EHL, describes the situation occurring in ball bearings having a restricted oil supply, where pressure build-up in the contact inlet region is inhibited, resulting in a thinner film thickness than calculated by classical EHL theory [7, 8]. Starvation theory was first described by Wedeven [9] in the early 1970s. 27.2.2.2 Parched EHL In many space mechanisms, instrument bearings are lubricated with a minimal amount of oil. When no free bulk oil is available to form a meniscus, starvation theory cannot adequately describe lubricant behavior. Another EHL subdivision, parched elastohydrodynamics, describes this behavior [10,11]. Lubricant films in this regime are so thin that they are immobile outside the Hertzian contact zone. This regime is particularly important for space mechanisms because parched EHL bearings require the least driving torque and have the most precisely defined spin axis, making them an ideal choice for many applications. 27.2.2.3 Transient/Non-Steady State EHL For space mechanisms, transient or nonsteady state behavior is another important EHL area. In this area, load, speed, and contact geometry are not constant over time. Unlike steady state EHL behavior, nonsteady state behavior is not well understood. However, many practical machine elements, including rolling element bearings, gears, cams, and traction drives, operate under nonsteady state conditions. In particular, stepper motors, commonly used in many space mechanisms, operate in this state. This regime has been studied theoretically for line contacts [12–14] and experimentally for point contacts [15,16].
27.3 Liquid Lubricants For space applications, designers use both liquid and solid lubricants. Both have merits and deficiencies, which appear in Table 27.1 [17].
© 2006 by Taylor & Francis Group, LLC
27-4
Handbook of Lubrication and Tribology TABLE 27.1
Relative Merits of Solid and Liquid Space Lubricants
Dry lubricants Negligible vapor pressure Wide operating temperature Negligible surface migration Valid accelerated testing Short life in moist air Debris causes frictional noise Friction speed independent Life determined by lubricant wear Poor thermal characteristics Electrically conductive
Wet lubricants Finite vapor pressure Viscosity, creep, and vapor pressure all temperature dependent Sealing required Invalid accelerated testing Insensitive to air or vacuum Low frictional noise Friction speed dependent Life determined by lubricant degradation High thermal conductance Electrically insulating
Source: Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.
27.3.1 Types of Liquid Lubricants In the last three decades, space applications have used many different liquid lubricants, including mineral oils, silicones, esters, and perfluoropolyethers (PFPE). More recently, a synthetic hydrocarbon (Pennzane®) has been replacing many older lubricants. Each lubricant type will be discussed briefly but since the majority of current spacecraft use either a formulated Pennzane® or one of the PFPE lubricants, these two classes will be discussed in detail. 27.3.1.1 Mineral Oils This lubricant class consists of a complex mixture of naturally occurring hydrocarbons with a wide range of molecular weights. Examples include V-78, BP 110, Apiezon C, Andok C (Coray 100) [18], and the SRG series of super refined mineral oils, including KG-80 [19]. The super refined fluids have been highly processed to remove polar impurities, either by hydrogenation or percolation through bauxite. Refining makes them poorer neat lubricants, but greatly improves additive response. While Apiezon C is still commercially available, production of all others was discontinued many years ago. Nevertheless, some companies have stockpiled SRG oils and still use them to lubricate momentum and reaction wheel bearings. SRG oils have an estimated shelf life in excess of 20 years [19]. 27.3.1.2 Esters Esters, which are available in a wide viscosity range, are inherently good boundary lubricants. In the 1970s, British Petroleum developed a triester base lubricant which was laboratory tested but whose production stopped before it flew in space. Another ester, NPT-4 (neopentylpolyol ester), has been used in the past, but is no longer produced. Nye Lubricants also markets a series of low volatility neopentylpolyol esters (UC4 , UC7 , and UC9 ). 27.3.1.3 Silicones This fluid class was used early in the space program but silicones are poor boundary lubricants for steel on steel systems. Boundary lubrication comparisons of this fluid with a PFPE and a PAO have been reported [20] and Figure 27.2 shows relative lifetimes. Silicone performed poorly, degrading into an abrasive, polymerized product. Versilube F-50, a chloroarylalkylsiloxane, is an early example of this lubricant class. 27.3.1.4 Synthetic Hydrocarbons Two synthetic hydrocarbon groups are available today: polyalphaolefins (PAO) and multiply alkylated cyclopentanes (MACs).
© 2006 by Taylor & Francis Group, LLC
Lubrication for Space Applications
27-5
16 14
Relative life
12 10 8 6 4 2 0 Chloroarylalkylsiloxane
PFPE
PAO
Lubricant type
FIGURE 27.2 Screening test results (scanner and mechanism). (From Didziulis et al., “Lubrication,” NASA/TP-1999206988, 1999.) TABLE 27.2 Typical Properties for Three Commercial Polyalphaolefins
Viscosity at: 210◦ F, SUS 210◦ F, cSt 100◦ F, SUS 100◦ F, cSt 0◦ F, cSt Flash point Pour point Evaporation 6 12 h at 350◦ F Specific gravity at 25◦ C
Oil 132
Oil 182
39 3.9 92 18.7 350 440◦ F −85◦ F 2.2% 0.828
62.5 10.9 348 75.0 2700 465◦ F −60◦ F 2.0% 0.842
Oil 186 79.5 15.4 552 119 7600 480◦ F −55◦ F 1.9% 0.847
27.3.1.4.1 Polyalphaolefin (PAO) Polyalphaolefins are made by the oligomerization of linear α-olefins having six or more carbon atoms [21]. Nye Lubricants markets a number of PAOs for space applications and properties for three commercial PAOs appear in Table 27.2. A new synthetic hydrocarbon based on PAO chemistry has been developed [22]. 27.3.1.4.2 Multiply Alkylated Cyclopentanes Multiply alkylated cyclopentanes (MACs) make up the second hydrocarbon class. These materials are synthesized by reacting cyclopentadiene with various alcohols in the presence of a strong base [23,24]. The reaction products are hydrogenated to produce the final product, which is a mixture of di-, tri-, tetra-, or penta- alkylated cyclopentanes. Varying reaction conditions control the distribution. Originally, only one product, known as Pennzane® SHF-X2000 or Nye Synthetic Oil 2001A, was available for space applications, and is primarily the tri-2-octyldodecyl substituted cyclopentane [23]. However, Pennzane® SHF-X1000, a lower viscosity but higher volatility version, is now available [25]. SHF-X1000 is primarily a di-substituted cyclopentane. A variety of formulated versions for both oils are also available. Properties of SHF-X1000 and SHF-X2000 appear in Table 27.3. Recent experience with SHF-X2000 appears in Carré et al. [26]. A six-year life test of a CERES elevation bearing assembly using a Pennzane®/lead naphthenate formulation yielded excellent results [27]. Additional life test data for another Pennzane®/lead naphthenate formulation for the MODIS instrument appears in VanDyk et al. [28].
© 2006 by Taylor & Francis Group, LLC
27-6
Handbook of Lubrication and Tribology TABLE 27.3
Typical Properties of Two Pennzane® Fluids
Property Viscosity at 100◦ C (cSt) Viscosity at 40◦ C (cSt) Viscosity at −40◦ C (cSt) Viscosity index Flash point (◦ C) Fire point (◦ C) Pour point (◦ C) Specific gravity at 25◦ C Coefficient of thermal expansion (cc/cc/◦ C) Total weight loss, 24 h, 125◦ C, 10−5 Torr Refractive index at 20◦ C Vapor pressure at 125◦ C (Torr)
TABLE 27.4
SHF-X1000
SHF-X2000
9.4 60 N/A 131 290 N/A −52 0.85 N/A <0.4% 1.4682 6 × 10−8
14.6 108 80,500 137 300 330 −55 0.85 0.0008 <0.2% 1.4671 4 × 10−7
Physical Properties of Four Commercial PFPE Lubricants and Pennzane® SHF-X2000 Vapor pressure, Pascal
Lubricant
Average molecular weight
Viscosity at 200◦ C, cSt
Viscosity index
Pour point ◦ C
9500 3700 6250 8400 1000
255 230 800 500 330
355 113 134 210 137
−66 −40 −35 −53 −55
Fomblin™ Z-25 Krytox™ 143AB Krytox™ 143AC Demnum™ S-200 Pennzane® SHF-X2000
At 20◦ C 3.9 × 10−10 2.0 × 10−4 2.7 × 10−6 1.3 × 10−8 2.2 × 10−11
At 100◦ C 1.3 × 10−6 4.0 × 10−2 1.1 × 10−3 1.3 × 10−5 1.9 × 10−8
27.3.1.5 Perfluoropolyethers (PFPE) These fluids, which are designated as either PFPE or PFPAE, have been commercially available since the 1960s and 1970s in the form of a branched fluid (Krytox™) manufactured by DuPont [29], a linear fluid (Fomblin™ Z) (a similar product marketed as Brayco™ 815Z) [30], and a branched fluid (Fomblin™ Y) [31], both manufactured by Solvay Solexis. In Japan, Daikin [32] developed another linear fluid (Demnum™). Preparation and properties of these fluids appear in Synthetic Lubricants and HighPerformance Functional Fluids [33]. Some typical lubricant properties appear in Table 27.4. PFPE fluid’s high density, nearly twice that of non-PFPE fluids, is an advantage for EHL lubrication. For similar kinematic viscosities, PFPE fluids will yield almost twice the EHL film thicknesses compared to conventional fluids [34]. 27.3.1.5.1 PFPE Formulations Currently no space applications use liquid PFPEs with additives, although many PFPE soluble additives have been developed in recent years, including antiwear, anticorrosion, and antirust additives [35–43]. Several different additives exhibited antiwear behavior in a Krytox™ basestock in vacuum four-ball tests (Figure 27.3) [41]. 27.3.1.6 Silahydrocarbons Silahydrocarbons, another newer space lubricant, were first introduced by the Air Force Materials Laboratory [44,45]. These materials contain only silicon, carbon, and hydrogen and thus, do not exhibit the poor boundary lubricating ability seen with silicones, which contain oxygen. Additionally, these unimolecular materials have exceptionally low volatility and are available in a wide range of viscosities. Three silahydrocarbon types are available, based on the number of silicon atoms present in the molecule (i.e., tri, tetra, or penta) [46,47]. Kinematic viscosities as a function of temperature were measured for a series of synthesized silahydrocarbons (Figure 27.4) [48]. For comparison, the plot includes Pennzane® SHF-X2000 data. As can be seen,
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Lubrication for Space Applications
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1.0E-09 9.0E-10 8.0E-10
Wear volume, mm3
7.0E-10 6.0E-10 5.0E-10 4.0E-10 3.0E-10 2.0E-10
io
l
te Am
id
e/
th
lfi
to ike ß-
D
ph os Ph
Su
ne
e at
e at ph
Th
io
ph
os
hi nz ot Be
Ba
se
st
az ol
e
oc k
1.0E-10
Additive
FIGURE 27.3 Antiwear behavior of several additives in Krytox™ basestock using a vacuum four-ball tribometer.
the silahydrocarbon viscosity properties bracket the Pennzane® data. Table 27.5 lists the EHL properties of two silahydrocarbons [49]. Based on these values, silahydrocarbons will generate thicker EHL films than Pennzane® SHF-X2000 under the same conditions. Tribologic properties of several silahydrocarbons appear in Jones et al. [50].
27.3.2 Liquid Lubricant Properties Numerous reviews of liquid lubricants for space applications have been published [51–53]. Liquid lubricant data also appears in some handbooks [54–56]. Since most applications today use either PFPEs or Pennzane® (MAC) formulations, these two classes are presented in more detail. To function properly in a lubricated contact, a liquid lubricant has to possess certain physical and chemical properties. To be considered for space applications, the lubricant must have vacuum stability (i.e., low vapor pressure), low creep tendency, high viscosity index (i.e., wide liquid range), good EHL and boundary lubrication properties, and radiation and atomic oxygen resistance. In some applications, optical or infrared transparency is also important. 27.3.2.1 Volatility Although space mechanisms use labyrinth seals extensively, lubricant loss can still be a problem for longterm applications [57]. For a fixed temperature and outlet area, lubricant loss is directly proportional to vapor pressure. Compared to conventional lubricants with similar viscosities, PFPE fluids are particularly good candidates (Figure 27.5) [58]. Vapor pressure data for four commercial PFPE fluids and Pennzane® SHF-X2000 appear in Table 27.5. Additional vapor pressure data is found in Nguyen and Jones [59]. A theoretical model to predict evaporative oil losses in spacecraft mechanisms has been reported [60]. Besides base oil volatility, additive volatility must also be addressed [61].
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Handbook of Lubrication and Tribology Tri(94-95) Tri(89-118) Penta(95-49) Tri(88-103) Pennzane Tetra(91-65) Tetra(88-102)
170
Kinematic viscosity
150 130 110 90 70 50 30 10 75
125
175
225
Temperature(°F)
FIGURE 27.4 Kinematic viscosity as a function of temperature for a series of silahydrocarbons. (From Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.) TABLE 27.5 Pressure–Viscosity Coefficient (GPa−1 ) (α value) of Two Silahydrocarbons and Pennzane® Temperature (◦ C) 21 30 40
Pennzane®
Trisilahydrocarbon
Pentasilahydrocarbon
N/A 9.8±0.3 N/A
16±0.3 N/A 11±1
17±0.3 N/A 13.5±1
Source: Spikes, H.A., “Estimation of the Pressure–Viscosity Coefficients of Two Silahydrocarbon Fluids at Two Temperatures,” Imperial College (Tribology Section), Report TS024b/96, London, UK, 1996. Mineral oil
15 Years to lose 1.0 ml, per cm2 outlet
Z fluid (z-25) K fluid (143 AC)
12
PAO Ester
9
Mil-Std-1540 (71°C) 6
3
0 20
40
60
80
100
120
140
160
180
200
Temperature,°C
FIGURE 27.5 Relative evaporation rates of aerospace lubricants. (From Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.)
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Compound A B C D E F
1.8
ASTM slope
1.6
CF3(OC2F4)4OCF3 CF3(OC2F4)4OCF3 C2F5OC3F6OC2F5 C2F5OC4F8OC2F5 C2F5OC2F4OC2F5 C2F5(OC4F4)2OC2F5
1.4
1.2
1.0
0.8
0
1 2 3 4 Compound carbon to oxygen ratio, C/O
5
FIGURE 27.6 Viscosity–temperature slope as a function of carbon-to-oxygen ratio. (From Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.)
27.3.2.2 Creep A liquid lubricant’s tendency to creep, or migrate over bearing surfaces, is inversely related to its surface tension. Because PFPE fluids have unusually low surface tensions (γLV , 17 to 25 dynes/cm at 20◦ C), they are more prone to creep than conventional fluids such as hydrocarbons, esters, and silicones. Low surfaceenergy fluorocarbon barrier films, placed on bearing lands, are used to contain PFPE fluids within bearing raceways [62]. However, there is a tendency for PFPE fluids to dissolve the barrier films with prolonged contact [57], rendering them ineffective in preventing PFPE migration. Pennzane®-based lubricants have higher surface tensions and thus are less prone to creep and are easier to contain. Sharp corner geometries also will minimize creep losses. 27.3.2.3 Viscosity–Temperature Properties Historically, liquid lubricated space mechanisms have not been exposed to wide temperature ranges. However, there is a trend for newer mechanisms to operate at lower temperatures (−10 to −40◦ C) therefore, low pour point fluids that retain low vapor pressure and reasonable viscosities at temperatures between −40C and 75◦ C are needed. The viscosity–temperature slope of PFPE, un-branched fluids is directly related to the carbon-tooxygen (C/O) ratio in the polymer-repeating unit (Figure 27.6) [63]. Here, the ASTM slope is used for the correlation. High values of the ASTM slope indicate large viscosity changes with temperature. Additionally, branching, such as the trifluoromethyl pendant group in the Krytox™ fluids, causes deterioration in viscometric properties. Comparison of ASTM slopes for three commercial fluids appears in Figure 27.7. Here the low C/O ratio fluid Fomblin™ Z has the best viscometric properties. The Demnum™ fluid, with a C/O ratio of three, has intermediate properties, while the branched Krytox™ fluid, has the highest slope. 27.3.2.4 Elastohydrodynamic Properties Successful operation of continuously rotating, medium to high-speed bearings relies on the formation of an EHL film. The section titled Lubrication Regimes briefly discussed this regime and a more detailed
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Handbook of Lubrication and Tribology 1.0 K fluid
ASTM slope
.8
.6
D fluid
Z fluid
.4
.2
0 10
20
40 100 200 400 Kinematic viscosity, cSt at 20°C
1000
2000
FIGURE 27.7 Viscosity–temperature slope (ASTM D341-43) as a function of kinematic viscosity at 20◦ C for Krytox™ (K), Demnum™ (D), and Fomblin™ (Z) fluids. (From Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.)
discussion appears in Reference 64. Two lubricant physical properties influence EHL film formation: the absolute viscosity (µ) and the pressure–viscosity coefficient (α) [8]. Viscosity is influenced by molecular weight and chemical structure. Except for low molecular weight fluids, α values are only related to chemical structure [65]. Conventional, high-pressure viscometers [34,66,67] can directly measure pressure–viscosity coefficients or they can be measured indirectly by optical EHL experiments [68]. Conventional viscometry normally uses the Barus equation [69] for correlations: µp = µo eαp
(27.1)
where µp is absolute viscosity at pressure (p), µo is absolute viscosity at atmospheric pressure, α is a temperature dependent but pressure independent constant. This implies that a log µp vs. p plot should yield a straight line with a slope of α. Unfortunately, this simple relationship is seldom obeyed. The pressure–viscosity properties that are important in determining EHL film thickness occur in the contact inlet, where pressures are much lower than in the Hertzian region. Therefore, the slope of a secant drawn between atmospheric pressure and approximately 0.07 GPa is typically used for film thickness calculations. Some researchers [34] favor using a different pressure–viscosity parameter, the reciprocal asymptotic isoviscous pressure (α ∗ ) based on work by Roelands [70]. Table 27.6 lists pressure–viscosity coefficients (α ∗ ) for several lubricants at three temperatures (38, 99, and 149◦ C). Additional values of α ∗ as a function of temperature for Pennzane® X-2000, Nye 186 A and Fomblin™ Z-25 are found in the work of Nélias et al. [71]. Figure 27.8 shows α values for a branched PFPE (Krytox™ 143AB). Data obtained by conventional (low shear) pressure–viscosity measurements are denoted with open symbols, while indirect measurements from EHL experiments (effective α values) are shown with solid symbols. Good agreement exists between the different sources as well as between different measurement techniques. Figure 27.9 contains similar data an for un-branched PFPE (Fomblin™ Z-25) as a function of temperature. Here, definite data grouping exists, with effective α values being substantially lower than conventional measurements values. Two possibilities exist for this discrepancy. First, inlet heating can occur during the EHL measurements, leading to lower viscosities and lower film thicknesses which result in lower calculated α values. The second
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TABLE 27.6 Pressure–Viscosity Coefficients at Three Temperatures α ∗ , PA−1 × 108 for Several Lubricants Lubricant
38◦ C
99◦ C
149◦ C
1.3 1.8 1.8 2.5 3.1 4.2
1.0 1.5 1.5 1.5 1.7 3.2
0.85 1.1 1.3 (extrapolated) 1.3 0.94 3.0
Ester Synthetic paraffin Z fluid (Z-25) Naphthenic mineral oil Traction fluid K fluid (143AB)
Source: Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.
4.8 Westlake and Cameron Foord at al. NASA Jones et al. DuPont Alper et al. Solid symbols denote effective a values
a, (Pa)–1×108
4.4
4.0
3.6
3.2
2.8 20
40
60
80
100
120
140
160
Temperature, °C
FIGURE 27.8 Pressure–viscosity coefficients for PFPE Krytox™ 143AB as a function of temperature. (From Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.)
Bair and Winer 2.4
Wedeven, 1975 Corti and Savelli, 1993
a, (Pa)–1 × 108
2.0
Vergne and Reynaud, 1992 Cantow et al., 1987
1.6
Spikes et al., 1989 Tanaka et al., 1989 Solid symbols denote effective a values
1.2 0.8 0.4 0 0
20
40
60
80
100
120
140
Temperature, °C
FIGURE 27.9 Pressure–viscosity coefficients for PFPE Fomblin™ Z-25 as a function of temperature. (From Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.)
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TABLE 27.7 Measured Viscosities and Calculated Pressure Viscosity Coefficients (α Values) for Several Space Lubricants [73] Pennzane® 2001–Synth. Oil
PAO-186–Synth. Oil
40 80 100 120
88 19 12 8
90 21 13 8
40 80 100 120
11.0 9.5 7.0 7.0
12.5 9.0 7.0 5.0
Temp ◦ C
Pennzane®2001+5% Pb naphthenate
Pennzane®2001+3% Pb naphthenate
Viscosity, cP 37 10 6 4
98 21 12 8
96 21 12 8
αGPa−1 6.5 5.0 5.0 5.0
12.0 9.0 6.5 6.0
10.0 9.0 7.0 7.0
NPE UC-7–Ester
Source: Spikes, H.A., Imperial College (Tribology Section), Report TS037/97, London, UK, 1997.
possibility is a non-Newtonian shear thinning effect, which can occur with polymeric fluids. Shear rates in EHL inlets can range from 105 to 107 sec−1 [72]. However, the EHL measurements do represent actual film thicknesses that may be expected in practice. Effective α values for several non-PFPE space lubricants, including Pennzane® SHF-X2000 base fluid and some Pennzane® formulations, appear in Table 27.7 [73]. From EHL theory, the lubricant with the largest α value should yield the thickest film at room temperature, assuming approximately equal contact inlet absolute viscosities. However, in many applications, lubricants must perform over a wide temperature range. In these cases, the EHL inlet viscosity can be the overriding factor if the temperature coefficient of viscosity is high. This can cause a crossover in film thickness as a function of temperature for some PFPE fluids as shown by Spikes [65] in Figure 27.10. 27.3.2.5 Boundary Lubrication As described in the section titled Lubrication Regimes, in boundary lubrication, surfaces are not completely separated, resulting in surface asperity interactions. In this regime, the most important requirement is the formation of protective surface films that minimize wear, surface damage, and friction. The lubricant and the contacting surface chemistry govern film formation. Non-additive hydrocarbons, mineral oils, and esters react in a boundary contact, producing “friction polymer” [74]. Except for electrical contacts, this material usually has a short-term, beneficial effect, but does represent a lubricant loss mechanism. However, conventional lubricants are usually formulated with antiwear, anticorrosion, extreme pressure, or antioxidant additives to enhance their performance and stability. In contrast, PFPE boundary lubricant is a relatively inert, very pure fluid, which in past years contained no additives. If these fluids were truly inert, they would not provide any surface protection except for some local fluid film effects (micro-EHL) and wear debris removal. However, PFPE fluids do react with bearing surfaces, producing a series of corrosive products and a friction polymer, which, in turn, react with existing surface oxides to produce metallic fluorides [75–77]. The fluorides are effective, in-situ solid lubricants, which reduce friction and prevent catastrophic surface damage [75]. This is the current understanding of the PFPE’s lubricating mechanism. Unfortunately, the fluorides are also strong Lewis acids (electron acceptors) and readily attack and decompose PFPE molecules [77–79]. This causes the production of additional reactive species, which, in turn, produce more surface fluoride, resulting in an autocatalytic reaction that may cause abrupt failure of a boundary-lubricated contact. Therefore, the very reaction that allows pure PFPE fluid use in boundary contacts leads to an early destruction and accompanying contact failure shown in Figure 27.11(a). In contrast, a non-PFPE space lubricant, such as Pennzane® SHF-X2000, has a greater lubricated lifetime under boundary lubrication and a slower progression to failure, characterized by a gradually increasing
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Lubricant parameters (m0a)
103
102
Z 25 Y 25 101 Z 25 Y 04
100 20
40
60 80 100 120 140 Temperature, °C
FIGURE 27.10 Lubricant parameters for PFPE Y and Z fluids. (From Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.)
friction coefficient (Figure 27.11(b)). It should be noted that PFPE degradation rates are highly dependent on the local contact conditions (i.e., surface passivation degree, surface oxide type and thickness, surface contamination level, temperature, load, speed, etc.). Substituting standard bearing balls with ceramic or ceramic-coated balls can provide significant improvements in bearing lifetimes [80] when using a PFPE. Balls coated with TiC [81] have shown considerable promise in alleviating some lubricant degradation problems. By replacing 52100 steel balls with TiC coated balls, Gill et al. [82] have shown a nine-fold increase in bearing lifetime with a PFPE (Fomblin™ Z25). In accelerated life tests [83,84] using a spiral orbit tribometer, lubricant lifetimes with another PFPE (Krytox™ 143AC) were extended by factors of two to four, depending on the stress level. Hybrid bearings have also been used for dry, unlubricated applications [85,86].
27.4 Greases and Solid Lubricants 27.4.1 Greases Space mechanism designers have extensively used greases based on PFPEs with PTFE thickeners (Krytox™ 240 series and Braycote 600 series). More recently, hydrocarbon greases based on Pennzane® SHF-X2000 (marketed by Nye Lubricants under the name of Rheolube® 2000) are available. Performance of various Pennzane® based greases appear in work of Rai et al. [87]. Also, new PFPE grease formulations (commercially designated as Braycote® 700 and 701) incorporating a boundary additive have yielded significantly improved wear characteristics (Figure 27.12) [88]. Some PFPE based greases also contain sodium nitrite as a rust inhibitor. Studies have compared the lubricated lifetime of PFPE and MAC-based greases under boundary lubrication conditions [89–91]. As with the base oils, MAC-based greases have a longer lubricated life compared to PFPE-based greases. Relative lives are shown in Figure 27.13 [89].
27.4.2 Solid Lubricants Over the last 30 years, several solid lubricants have been used in space, including lamellar solids, soft metals, and polymers. Lamellar solids include transition metal dichalcogenides like molybdenum disulfide and
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Handbook of Lubrication and Tribology
Friction coefficient
(a) 0.3
0.2
0.1
0.0 0
1
2
3 4 Time, h
5
6
7
20
40
60 80 Time, h
100
120
140
Friction coefficient
(b) 0.3
0.2
0.1
0.0 0
FIGURE 27.11 Coefficient of friction as a function of time for (a) Krytox™ 143AC and (b) Pennzane® SHF-X2000 (spiral orbit rolling contact tribometer).
Mean wear rate(×10–9 mm3/mm)
2.0 1.8 1.6 1.4 1.2 1.0 0.8 0.6 0.4 0.2 0.0 143AC
S-200
Z-25
2001A (MAC)
2001 (MAC)
SiHC
PAO-100
Lubricant
FIGURE 27.12 Mean wear rates of steel lubricated with various PFPE formulated greases using a vacuum four-ball tribometer. (From Jones, D. et al., “An Additive to Improve the Wear Characteristics of Perfluoropolyether Based Greases,” NASA TM-1999-209064, 1999.)
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Normalized life (orbits/mg)
10000
1000
100
10
1 MAC
PFPE1
PFPE2
Grease base oil type
FIGURE 27.13 Relative life of three space greases using the Spiral Orbit Tribometer. (From Marchetti, M. et al., “Preliminary Evaluation of Greases for Space Mechanisms Using a Vacuum Spiral Orbit Tribometer,” NASA TM 2001-211157, 2001.)
tungsten disulfide. Soft metals include lead, gold, silver, and indium. Polyimides and polytetrafluoroethylene are polymeric materials that have lubricating properties. More recently, diamond like carbon (DLC) coatings are being investigated for space applications and discussed in the section titled Diamond Like Carbon Coatings. Unlike liquid lubricants, metals and lamellar solids are applied as thin films (less than a micron), preferably by ion plating [92–94] or sputtering [92,95,96]. Another application method involves bonded films [97,98], where lubricants are mixed with an organic binder and applied to the surface by spraying or dipping. The films, typically greater than 10µm thick, are cured at high temperature. Sometimes self-lubricating polymers and polymer composites [99,100] are used. Rolling element bearing cages and retainers or bushings are the main applications for solid lubricants. Today, ion-plated lead and various forms of sputter-deposited MoS2 are the most common solid lubricants used in space mechanisms. 27.4.2.1 Ion-Plated Lead In Europe, ion-plated lead is the choice solid lubricant for precision spacecraft bearings and normally is used in conjunction with a leaded bronze cage. Lead coatings and cages saw early success in cryogenic space applications [101]. Other successful implementations include GIOTTO [102], OLYMPUS [103], and GERB [104]. Although not as common in the United States, the combination was used in SABER encoder bearings. Extensive data has been published on speed, thickness, and substrate surface roughness effects [105]. One disadvantage of this lubricant is limited life in air, where significantly higher wear rates occur and produce copious amounts of lead oxide. This debris also causes torque noise. 27.4.2.2 Molybdenum Disulfide For many years, molybdenum disulfide (MoS2 ) has been successfully used in space applications [106–109]. In vacuum, these films display extremely low friction (0.01 or less). Optimized thin films (one micron or less) are deposited by sputtering. The film’s tribologic performance is extremely dependent on the sputtering conditions, which control the microstructure that, in turn, determines crystallinity, morphology, and composition [110]. For instance, the presence of oxygen in the sputtering environment can affect
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Handbook of Lubrication and Tribology 0.20
0.15
Humid air
Coefficient of friction
0.10
0.05
0.04 0.03 0.02
Ultra-low friction
Vacuum
0.015 0.01
0.005
Super-ultra low friction
FIGURE 27.14 Friction variation of sputtered MoS2 films. (From Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.)
both friction and wear life [111]. For more details about the sputtering process, see Spalvin’s work [95]. Additionally, substrate surface roughness has a pronounced effect on friction and wear. For steel bearing surfaces, optimum durability occurs at a nominal surface roughness of 0.2 µm (Ra ) [112]. Operating environment greatly affects the frictional behavior and life of MoS2 films. In ultrahigh vacuum, these films display ultra-low friction (less than 0.01), shown in Figure 27.14. Under normal vacuum conditions, the friction may range from 0.01 to 0.04 with exceptionally low wear and long endurance lives. When the same films were tested in humid air, initial friction coefficients were near 0.15 and life was severely limited [106,113]. Since many space mechanisms must be ground tested before launch, sometimes in room air, a great deal of research has been done to improve the film’s performance under atmospheric conditions. One method to improve atmospheric life has been to layer or co-deposit metals such as gold [114–116] with MoS2 films. In this work, gold inclusions doubled the film durability in dry nitrogen and tripled or quadrupled it in air. Co-deposition with other metals (chromium, cobalt, nickel, and tantalum) has also shown synergistic effects [117]. Ion implantation with silver has also been reported to be beneficial [118], but not when co-deposited [117]. Co-deposition with titanium also improves the properties of MoS2 films [119], making the films less sensitive to atmospheric water vapor than pure MoS2 films. Accelerated test results on some films appear in Fusaro and Siebert [120]. Studies show the problems occurring in moist air are associated with water molecule adsorption at edge sites on the MoS2 lattice [121]. Using a PTFE composite retainer is another method for enhancing MoS2 endurance in ball bearings [122]. In gimbal bearing life tests [109], advanced MoS2 films combined with PTFE-based retainers demonstrated lives in excess of 45 million cycles.
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27.4.2.3 Diamond Like Carbon Coatings Diamond like carbon coatings are being examined for space applications. Because of their vast diversity, DLCs pose unique challenges. Previous studies [123–128] show a wide range of friction coefficients in vacuum or dry nitrogen, due to different deposition methods, different alloying elements, and a variety of multilayer structures. Recent studies [129–131] show films high in hydrogen (>40%) display low friction and low wear. While current DLCs have much higher wear rates than MoS2 under vacuum, they show promise for future applications, particularly to provide a passivation layer on steel surfaces, which could extend PFPE lubricant lifetimes. An in-depth review of solid and liquid lubricants appears in Didziulis et al. [132].
27.5 Mechanism Components and Re-Lubrication Mechanisms Nearly all spacecraft systems contain mechanisms requiring lubrication. With spacecrafts’ ever-expanding exploration roll, including the Martian surface and the atmosphere of Jupiter’s moons, mechanism and lubricant demands are always growing. Each mission’s unique challenge is to match the lubricant with the component’s primary function and operating environment. Solar array drives, momentum, reaction, and filter wheels, tracking antennas, slip rings, scanning devices, sensors, rover wheels, robotic arms, antenna arrays, gearboxes, and actuators are some of the many components requiring lubrication. Each has unique hardware, mission requirements, and operating environment and therefore unique lubrication requirements. When selecting a lubricant for a mechanism component, many factors must be considered. One of the most fundamental criteria is the mechanism’s operating regime (boundary, mixed, or EHL). This will guide in proper lubricant and additive selection. Another important consideration is mechanism design life and duty cycle. Since a mechanism has a finite lubricant amount and re-lubrication is almost never possible, lubricant degradation (consumption) is one of the most common types of mechanism failure. Understanding how the selected lubricant will interact with the mechanism (such as surface chemistry’s relation to degradation rate, dewetting, creep, and lubricant evaporation paths) is vital to long life operation. The mechanism’s operating environment is another important design factor. For example, the lubricant used in a mechanism inside a satellite, which is exposed to mild temperatures (20 to 50◦ C) and a vacuum environment, will be very different from the lubricant used in a rover wheel, which might see a temperature range −135 to 30◦ C in a harsh, dusty, partial atmospheric environment.
27.5.1 Spacecraft Components 27.5.1.1 Electrical Contact Ring Assemblies Electrical Contact Ring Assemblies (ECRA) [133] are a good example of unique lubrication requirements. Excessive electrical noise, usually due to surface contamination, is the most common failure mechanism in ECRAs [134]. Therefore, electrical signal integrity, even when the lubricant begins to breakdown, is one primary lubricant selection criteria. Since most ECRAs operate at low speed, good boundary or mixed regime lubrication properties are required for long life. Proper selection and understanding of the lubricant and its degradation product’s electrical properties are vital to trouble-free ECRA operation. 27.5.1.2 Gyroscopes, Momentum, and Reaction Wheels Gyroscopes, which are used to measure changes in orientation, operate at high speeds, typically between 8,000 and 20,000 rpm with high accuracy. Fluctuations in the bearing reaction torque, noise, and excess heat generation can cause a null position loss in the gyroscope, making the bearings a vital gyroscope component. The ideal lubricant for a gyroscope provides a high level of wear protection, produces minimal friction, and has a low evaporation rate [20]. Because of the limited lubricant quantity, evaporation and degradation rates and creep properties must be well understood. Gyroscopes also contain another
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mechanism, the gimbal supports, which operate at low speed in the boundary regime only, presenting completely different lubricant selection criteria. Momentum wheels, which typically operate between 3,000 and 10,000 rpm, pose their own lubricant selection criteria. Historically, the lubricant is the cause of the majority of problems experienced by momentum wheels, with most wheel failures caused by inadequate lubrication, loss of lubricant, or lubricant degradation [20]. As higher speed, higher energy wheels are designed, lubricants will be subjected to higher operating temperatures and stresses, which can increase creep or degradation rates. To ease lubricant limitations, current design practices include use of improved synthetic lubricants, labyrinth seals and barrier coatings, lubricant impregnated retainers, or a lubricant re-supply system. Reaction wheels have similar design concepts as momentum wheels, but operate at lower speeds. The support bearings spend more time in the mixed lubrication regime. Therefore, reaction wheel lubricants must also have good boundary lubrication characteristics. Control moment gyroscopes (CMG) combine the aspects of the gyroscope and the momentum wheel to provide spacecraft attitude control. Therefore, considerations of both groups must be weighed when selecting a CMG lubricant [20]. 27.5.1.3 Sensors Many spacecraft use sensors containing rotating or dithering components supported by bearings. Proper lubricant selection is vital to ensure the sensor bearings are operating in the correct lubrication regime to fulfill mission life and environmental requirements. By adjusting the lubricant type and viscosity, the designer has control over the mechanism’s characteristics. An example of a rotating mechanism is a scanning horizon sensor, used for spacecraft orientation. The device operates at moderate operational speeds (400 to 1,600 RPM), constant, moderate temperature, and low loads, making lubricant selection easy. On the other hand, sensors using oscillatory motion place a high demand on the lubricant. Typically, the oscillation angle is small and the bearing only operates in the boundary lubrication regime. With the small oscillatory angle, no new lubricant is brought into the contact zones [135]. Between the operating regime and the lack of fresh lubricant, the lubricant is in a high demand, severe operating environment. Mechanism designers should consider an option where the bearing can rotate after operating a predetermined amount of time, bringing fresh lubricant into the contact region. 27.5.1.4 One Time Use Mechanisms Not all mechanisms require long-term lubrication. For example, satellite solar array or antenna deployment is a one-time operation. However, lubricant selection is also critical in these applications. If these mechanisms fail to deploy, the spacecraft’s functionality will be greatly reduced or totally lost. These mechanism types have unique lubrication needs because they only operate once, are low speed applications, may have long dormancy before use, and may be subject to harsh (external space) environments. Usually, solid lubricants are a good choice for these applications due to their low friction properties and ability to stay in the contact region. However, failure to fully understand the system, its dynamics, and the build-up, testing, and final operating environmental effects can lead to crippling results, as evidenced by the high gain antenna failure on the Galileo spacecraft (see Introduction to this chapter). 27.5.1.5 Actuators and Gearboxes Gearboxes and actuators are other mechanisms requiring lubrication. With these mechanism types, the lubricant may not be subject to many duty cycles. However, these cycles may involve high lubricant stresses with long, dormant periods between cycles. Often, the mechanisms may not be re-lubricated between missions or require long terrestrial storage times. Because examination and re-lubrication is often an expensive, complicated, and time consuming process, understanding of how the lubricant may react with the actuator components during storage and nonoperating times is important to ensure that mechanisms will function properly. 27.5.1.6 Planetary Rovers With advancements in robotics, computers, and communications, autonomous rovers are set to be the planetary explorers of the near future. Rovers have many components needing lubrication and have
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unique lubrication requirements. During space travel, the lubricant will be subject to low pressure and controlled temperature of the spacecraft, but once deployed, it will be exposed to harsh environments, wide temperature ranges, various gaseous atmospheres, and other environmental conditions including dust and solid contaminants. Rover mechanisms include robotic arms to deploy instruments and manipulate the environment, mast assemblies to hold cameras and viewing devices, solar arrays to provide power, antennas and communication equipment masks, and a mobility system consisting of wheels, legs, and other moving components. In addition to the rover, the associated landing craft also has many lubricated mechanisms. 27.5.1.7 Other Mechanisms Many other mechanisms that require lubrication are used in space. Some examples include solar array drives (SAD), screw drives, and many types of gears and transmission assemblies [136].
27.5.2 Re-Lubrication Mechanisms Because lubricant loss or degradation is a common reason for space mechanism failure, mechanism relubrication methods have been examined [53,137]. Since it is impossible to service most spacecraft after launch, in-situ and remote re-lubrication systems have been explored to extend mission lives. Although currently not widely used, many devices have been developed, including centrifugal oilers [138], positive commandable lubricators [139], wick feed systems [140], oozing flow lubricators [141,142], lubricant reservoirs, porous retainers [143], and remote controlled, in-situ lubrication mechanisms [144]. 27.5.2.1 Passive and Reservoir Re-Lubrication Many systems, such as centrifugal oilers, wick feed systems, and porous retainers are passive systems. In these systems, lubricant is either constantly fed into the contact region or drawn into the contact region by surface tension. These systems drawbacks are added complexity, additional space, additional weight, and lack of control. While porous retainers do not take up space, in some cases they have been shown to act as lubricant sponges rather than lubricant suppliers [18]. Lubricant reservoirs provide a re-supply mechanism. However, they are typically bulky and continuously supply lubricant to the contact region. This often leads to excessive lubricant in the bearing, which can be just as detrimental to the instrument as insufficient lubricant. A newer version of the reservoir concept, the oozing-flow lubricator, has been developed. This system has a reservoir, but the designer can specify a flow rate, which is controlled using a proprietary grooving system between the reservoir and the bearing. The groove geometry determines the flow rate, which is usually extremely low (micrograms per hour). This allows the designer to balance the degradation rate with the rate new oil is introduced into the contact region, thus avoiding an over or under lubrication situation. This system is described in more detail in other works [141,142]. All of the re-lubrication systems listed here are only applicable to liquid lubricants. 27.5.2.2 Active, In-Situ Re-Lubrication A newer re-lubrication concept is a small, in-situ, remotely controlled, lubricant reservoir. With this concept, the device is remotely activated when bearing torque increases, which evaporates a minimal lubricant charge into the contact zone, thus reducing friction and torque back to an acceptable level. The control can either come from a ground-based command or with a sensor/controller integrated in the mechanism. Additionally, both solid and liquid lubricants can be used with this system. The concept was first used to combat solid coating wear [145–147] and perform in-situ deposition of a solid film coating [148]. Both types of in-situ lubrication worked well and the concept extended to liquid lubricants [137]. All these experiments showed that in-situ lubrication worked in a laboratory environment. However, a viable solution for real systems had not been developed. Marchetti et al. [144] extended the concept to an attachment for an actual bearing and successfully demonstrated its functionality. In this system, a porous ring is impregnated with a liquid lubricant and attached to the edge of the bearing. The system can turn on and off a heating element on the top of the ring. When the bearing requires lubricant, the
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heater is activated and lubricant is exuded, which re-lubricates the contact region. This system is a low power, low volume system, and can be easily adapted to many different bearings. The system has been successfully demonstrated with both PFPE and MAC lubricant classes. While there are many re-lubrication systems available, the designer must balance the added weight, complexity, and power requirements with the mission life. Once again it is important to understand the whole mechanism, its function and design life, and how the selected lubricant will interact with the components in the system.
27.6 Lubricant Testing and Analysis With increasing demands and expanding operating environments of spacecraft mechanisms, new lubricants and additives are always under development. Since tribologic failure is a leading cause of spacecraft mechanism malfunctions, new lubricants must undergo extensive ground-based testing to ensure they will meet stringent mission requirements. In the past, lubricant selection has been based on “heritage,” or actual flight experience. In early space flight, this approach worked well because of short mission lives and minimal duty cycles. However, with today’s longer missions and improved components, new lubricant properties must be well understood to ensure they will perform properly with the system and the environment.
27.6.1 Types of Testing There are three levels of tribologic testing (1) tribometer, (2) component, and (3) system or mechanism. Each level has advantages and drawbacks are discussed in more detail in the following section.
27.6.1.1 Mechanism and System-Level Testing Historically, mechanisms were qualified with either system-level tests on actual flight hardware or duplication of flight system conditions [134]. However, these tests are expensive, time consuming, and, with long life requirements of modern missions, often cannot be completed before launch. Unless a mechanism fails early, the tests are useless for lubricant selection. Additionally, since tests are long and expensive only a few candidate lubricants can be tested. However, should an anomaly occur on orbit, system-level tests are useful for troubleshooting or to develop a working solution. Due to the limitations of this testing, component and tribometer level tests were developed.
27.6.1.2 Tribometer Level Testing Tribometer level testing involves measuring fundamental lubricant properties under various conditions. Tests are often short and the results quickly compared. This guides in lubricant selection for more extensive testing or application in an actual mechanism. Usually, properties measured include friction coefficient, wear rate, bearing material effects, and atmospheric effects. Typical test devices include pin-on-disk, ballon-disk, four-ball, block-on-ring, disc-on-disc, and several other geometries. Some newer tribometers, such as the Spiral Orbit Tribometer [149,150], can determine the lubricant consumption rate in addition to the previously mentioned effects. For space applications, the facilities operate at ultrahigh vacuum, but can also introduce various gaseous atmospheres. All the devices operate in the boundary or mixed lubrication regime. At this test level, several lubricants and conditions can be quickly evaluated at relatively low cost, making it particularly valuable for selecting the best candidate for an application. However, most of these devices, which operate in pure sliding, do not simulate the contact conditions seen in theultimate
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application — a major disadvantage. Some newer tribometers, the Spiral Orbit Tribometer, for example, provide a more realistic simulation of ball bearing contact conditions. 27.6.1.3 Component Testing Testing specific system components is a compromise between tribometer and system level tests. Typically, components include ball bearings, ball screws, slip rings, or gears. Obviously, component tests are more expensive and time consuming compared to tribometer tests. However, they are considered more reliable and can duplicate anomalies seen in specific components, either in flight or during system-level testing. Most component level tests use angular contact bearings, operate in vacuum, and are nonaccelerated, although sometimes limited lubrication, higher than normal loads, or elevated temperatures accelerate tests. Often, long-term component tests study liquid lubricants under boundary conditions. Component level tests are also applied to study lubrication regime transitions due to varying conditions and new solid/liquid lubricant combinations. 27.6.1.4 Life Prediction Ideally, all mechanism designs should incorporate generous margins, assuring that design lifetimes are always achieved. For low duty cycle operations, this is easily accomplished. However, this is almost impossible for long-lived missions involving millions of cycles. Two options exist that can provide confidence when selecting a lubricant for long life missions. First, the mechanism or component can be life tested before the final design. As mentioned previously, the cost and time involved usually eliminate this option. Second, looking at a “heritage” mechanism, which has operated successfully in space for the duration in question provides a database and allows for confidence that the mechanism will achieve its design life. However, these options are rarely possible. Statistical methods for fatigue life predictions do exist, but fatigue is not a failure mode in space applications. Since most liquid or grease lubricated bearing failures and anomalies occur from lubricant loss or tribologic consumption, a ball-pass or stress-cycle analysis can be used to estimate component life. 27.6.1.5 Lubricant Consumption Based on Ball-Pass/Stress-Cycle Analysis In the boundary lubrication regime, lubricant is continuously consumed within the Hertzian contact zone. Lubricant degrades into a nonlubricating, friction polymer or into gaseous fragments, which are lost to the surrounding vacuum. As long as fresh lubricant is entrained into the contact zone, consumption occurs. Thus, the consumption rate is directly related to the rotation rate. The Hertzian stress level also affects the consumption rate. Life generally decreases exponentially with increasing Hertz stress [84,151]. However, for mean Hertzian stresses over a narrow range (0.4 to 0.6 GPa), the following equation shows a method of normalizing the consumption rate: CDF =
Bc • Ph
(27.2)
where CDF is the Cumulative Degradation Factor (sometimes referred to as lubricant stress cycles; in units of psi-crossings), Bc is the number of times a given spot on the inner raceway is compressed (ball crossings or passes), and Ph is the mean Hertzian contact stress at the inner raceway [152,155]. Since this method normalizes data to units of psi-crossings, the CDF for any component or mechanism can be calculated and compared to existing life test data. This method was used to evaluate the AURA TES (Tropospheric Emission Spectrometer) Life Test Unit. The instrument’s design lifetime was 7 × 106 scans; however, after approximately 2.2 × 106 scans, motor current increased, indicating increased bearing torque. Motor current continued to increase, which eventually resulted in shutdown at approximately 5 million scans. Upon disassembly, it was discovered that the encoder bearings, which were lubricated with Braycote™ 815Z, had failed. CDF calculations were done and compared to CDF values from several other bearing life tests (Table 27.8). From these comparisons, it is obvious that the encoder bearing lifetime was in
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Handbook of Lubrication and Tribology TABLE 27.8 Measured Cumulative Degradation Factors for Several Bearing Life Tests Lubricated with 815Z
Bearing life test TES LTU Encoder Ball aerospace [152] Hughes (flooded) [155] Lockheed-Martin [158]
Mean Hertzian stress (MPa)
Cumulative degradation factor for initiation of lubricant degradation or bearing failure (×1012 ball crossing-psi)
503 448 648 751
2.0a (3.8)b 1.4c (2.0)b 2.8 2.2–8.7 (average 5.5)
a Start of motor current increases. b Test termination. c Initial torque increases.
3500
Normalized life (Orbits/mg)
3000
2500
2000
1500
1000
500
0 17-4PH
440C Steel type
FIGURE 27.15
Relative life of Pennzane® SHF-X2000 on two different bearing steels.
family, not a premature occurrence. Therefore, it was concluded there was a very low probability the TES instrument would reach the 7 million scan design life. 27.6.1.5.1 Other Factors Affecting Consumption Rate However, the consumption rate is also highly dependant on the lubricant and bearing surface chemistries. This is easily illustrated using Spiral Orbit Tribometry (discussed in detail in the section on Spiral Orbit Tribometer [SOT]). Figure 27.15 shows an example of this effect for a formulated Pennzane® in vacuum with two different bearing materials (440C and 17-4 PH stainless steels). Lubricant life is drastically reduced with the 17-4 PH material. Other surface chemistry effects were observed in Pepper and Kingsbury [153]. In this work, the detrimental effect of increasing chromium concentration on Fomblin™ Z-25 degradation is shown.
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27.6.2 Accelerated Testing Accelerated lubricant testing is required to rapidly screen several lubricants and additive packages. Because accelerated tests typically do not involve actual flight hardware, testing is significantly less expensive. Generally, test results cannot be extrapolated to predict the life of components that are lubricated similarly but operate under different conditions [58]. However, accelerated testing can rank lubricant’s life and performance relative to one another, providing an inexpensive way to screen several candidate lubricants. Once lubricants are screened using accelerated testing, the best candidates can continue to full-scale life testing. Test times are usually days to weeks rather than months to years. Test acceleration is typically achieved by subjecting the lubricant to a condition more extreme than the service requirements. Extreme conditions are produced by varying test parameters, including increasing speed, load, or temperature, adding contaminants, reducing lubricant quantity, or varying surface roughness [156]. When selecting which parameters to vary, understanding of how those parameters affect contact conditions and how the variations will relate to actual mechanisms is important. Ideally, accelerated tests should have as many parameters as close to an actual mechanism as possible. Because several different accelerated tests are required to measure various lubricant attributes, such as wear rates and boundary lubrication properties, a combination of tests is required to select lubricants for full mechanism testing. Solid and liquid lubricants have their own considerations when choosing a method for acceleration. 27.6.2.1 Accelerated Testing of Liquid Lubricants For liquid lubricants, changes in speed, temperature, contact stress, or available lubricant could change the film thickness substantially and subsequently change the test’s lubrication regime. As described earlier in this chapter, each lubrication regime has specific, different wear characteristics. Therefore, changing the lubrication regime could yield irrelevant results. However, the designer can change multiple parameters to maintain the proper lubrication regime. For example, if the speed is increased, the temperature can also be increased to try to maintain a constant lambda ratio (film thickness to composite surface roughness). When choosing what parameters to vary, it is important to understand how the variation will affect the contact and lubricant. For example, increasing the temperature might cause the lubricant to oxidize or react with contact surfaces differently than at the normal operating temperature. The designers must also consider if it is worth the effort to design a facility to maintain all the operating parameters. Ideally, the minimum number of parameters are varied. Additionally, accelerated testing with liquid lubricants does not account for time dependent parameters, including creep, loss of lubricant through evaporation and centrifugal forces, and in some cases lubricant degradation [58]. 27.6.2.2 Accelerated Testing of Solid Lubricants Accelerated life testing with a solid lubricant has fewer considerations. If a system uses only solid lubricants and operates at low speed, accelerated testing can be done simply by increasing the system speed. However, it is important to consider that the lubricant wear rate may be speed dependant and additional loads from inertial effects or component instability may be present [58]. However, as with liquid lubricants, the short-term tests do not account for long-term effects, such as exposure to various environments during storage, build-up, and operation. 27.6.2.3 Summary of Accelerated Testing Strengths and Weaknesses Even with the limitations of accelerated testing, it is a valuable tool for quickly screening and evaluating several lubricants. Lubricants can be ranked relative to each other and compared to “heritage” lubricants, providing vital information to help designers select lubricants or further tests. Additionally, additive performance can be quickly evaluated. With some modern accelerated testing equipment, such as the Spiral Orbit Tribometer, many parameters can be similar to actual mechanism conditions. Accelerated test limitations include chemistry effects, wetting, and lubrication regime changes with temperature, speed, and contact pressure, lack of standardized testing procedures, and coating fractures under high loads. Strengths and weaknesses are summarized in Table 27.9.
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Handbook of Lubrication and Tribology TABLE 27.9 Strengths and Weaknesses of Accelerated Life Testing Weaknesses Wetting condition with temperature Chemistry changes with temperature and pressure Oxidation changes with temperature and pressure Hydrodynamics region changes Wear/friction polymers changes Coating fracture under high load (solid lube) Nonstandardized Dynamic changes in cages and components Low confidence
Strengths Easy to monitor Used to enhance design Used to validate model Rapid baseline data generation Lower cost
Source: Murray, S. and Heshmat, H., NASA CR-198437, 1995.
27.6.3 Facilities for Space Lubricant Testing 27.6.3.1 Facilities for Accelerated Testing Accelerated test facilities for space mechanisms have some common features. To simulate the space environment, testing is done in a high or ultrahigh vacuum chamber. Because of this, space lubricant test facilities are usually costly, most are unique, and test apparatus size is usually limited. As mentioned earlier, accelerated test apparatus do not use flight hardware or mockups, but rather very simple hardware with minimal parts. This keeps test costs low and usually provides specimens that can be easily analyzed after test completion. Several different tribometers have been developed to qualify space lubricants. Because most space mechanism failures occur due to lubricant degradation rather than fatigue, degradation rates should be determined to help select lubricants. Several unique facilities measure one or more of the following: friction, substrate wear, lubricant degradation rates, lubricated lifetimes, or bearing system properties. 27.6.3.1.1 Spiral Orbit Tribometer (SOT) A novel vacuum tribometer exists at NASA Glenn Research Center and is based upon a simplified, retainerless thrust bearing with one ball and flat races (Figure 27.16). The SOT is used to measure lubricant degradation rates and friction in boundary-lubricated rolling/pivoting contact. One of the flat plates rotates and the ball, under load, is driven in a nearly circular orbit. However, the orbit is actually an opening spiral with the pitch directly related to the friction coefficient [150]. At the end of the orbit, the ball contacts the “guide plate,” returning it to the original orbit diameter. A force transducer, mounted in-line with the guide plate, measures the force exerted by the ball on the plate, which is directly related to the friction coefficient. For most of the orbit, the ball undergoes pure rolling with pivot or spin. However, when the ball is in contact with the guide plate, rolling, pivoting, and sliding occur. This is the most tribologically severe part of the orbit, termed the “scrub,” where the majority of the lubricant’s tribodegradation occurs. This is advantageous when doing post-test or in-situ analysis of degradation products (i.e., mass spectrometer signatures). Lifetime is controlled in two ways. Typically, the facility operates with a minute amount of liquid lubricant (∼50 µg). With this lubricant quantity, the entire amount is degraded, usually limiting test times to less than one or two days. Varying load, which changes stress and therefore changes degradation rates, is another method for controlling lifetimes. The facility can operate either in ultrahigh vacuum (10−9 Torr) or with various atmospheres. Loads between 45 and 450 N are used and, by varying the ball diameter, achieve mean Hertzian stress between 0.5 and 5.0 GPa. Ball rotational speeds between 4 and 100 rpm have been used. A full description of the tribometer and its kinematics can be found in works of Pepper and Kingsbury and that of Jones et al. [150,154,157]. The advantages of this facility over more traditional systems are as follows: the ball undergoes rolling, pivoting, and sliding contact — typical of a real bearing, short lifetimes due to finite lubricant quantity,
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(a)
(b)
Rotating plate (tilted up for clarity)
Rotation Spiral orbit
Ball orbit Guide plate Force transducer
Scrub Ball Spiral orbit Stationary plate
FIGURE 27.16
Spiral orbit rolling contact tribometer (overview and detail).
simple operation, and easy post-test sample analysis. Because the ball remains a sphere with no wear, posttest analysis can be extremely detailed. Additionally, there is excellent correlation between relative lifetimes obtained using the SOT and actual bearing tests. Relative lifetimes for several space lubricants appear in Figure 27.17 [159] and comparisons with actual bearing tests appear in the work of Bazinet et al. [158]. To date, the majority of SOT tests were done with liquid lubricant. However, the facility has also been used to study greases and thin solid films.
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Normalized lifetime
1000
100
10
5Z 81
Z2 5
ila
3A C 14
ta s
0 Pe n
PA O 10
ila Tr is
ila Te tra s
Pe n
nz
an
e 20
01
A
1
FIGURE 27.17 Relative lifetimes of several space lubricants using the SOT tribometer. (From Bazinet, D.G. et al., NASA CP-2004-212073, Galveston, TX, May 19–21, 2004.)
27.6.3.1.2 Vacuum Four-Ball Tribometer Figure 27.18 shows a tribometer developed at NASA Glenn Research Center. This tribometer is used to measure wear of materials. The facility is based upon a four-ball configuration and designed to test lubricants under pure sliding conditions. Three balls are fixed in a cup and immersed in lubricant while a fourth ball is mounted on a rotating shaft (Figure 27.19). The rotating ball is brought into contact with the stationary balls generating wear scars on the stationary balls. The test is stopped every hour to measure the wear scar diameters and calculate the wear volume. Using a special platform, the wear scars can be measured without removing the balls from the cup [160], allowing the test to resume exactly where it was stopped. A test takes four hours and, upon completion, wear volume is plotted as a function of sliding distance. The wear rate (wear as a function of time) is calculated from the slope of the line. Typical tests are run with 9.82 mm diameter balls, 100 rpm, room temperature, a vacuum level of at least 10−6 Torr, and an initial Hertzian stress of 3.5 GPa. However, as the wear scars develop and the contact area increases, the contact stress drops. This apparatus rapidly provides information about the wear-resisting capability of a lubricant or additive package. In general, lubricants exhibiting high wear rates correlate with lower bearing lifetimes. 27.6.3.1.3 Vacuum Pin-on-Disk Tribometer Figure 27.20 shows an ultrahigh vacuum pin-on-disk tribometer [161] for solid film coating tests. The pin-on-disk facility is used to measure friction and wear rates in sliding conditions. The apparatus is a simple setup using a stationary pin with a known end diameter loaded against a rotating disk coated with the desired lubricant. Loads, typically less than 1000 g, are applied by a dead weight system directly above the pin. The friction force displaces the pin tangentially. The tangential pin displacement is measured and the friction coefficient directly calculated. The rotating mechanism is enclosed in a vacuum chamber and tests are typically run at a vacuum level of at least 10−9 Torr. The rig can also be back-filled with a variety of gases to study their effects on the lubricants. 27.6.3.2 Facilities for Component Life Testing of Space Lubricants 27.6.3.2.1 Eccentric Bearing Test Apparatus Aerospace Corporation [134] developed an apparatus based upon a 22.1 mm diameter bearing. In this test, the bottom (rotating) raceway is flipped over and the flat side polished to a 0.25 µm finish. This configuration allows the lower raceway to be mounted with eccentricity. The intentional misalignment
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Drive motor Speed sensor Feedthrough Pump
Eddy current damper Flex pivot Load cell
Hall effect torque sensor
Pneumatic load system
FIGURE 27.18
Overview of four-ball vacuum tribometer.
accelerates the wear process and the flat surface operates at a higher stress level than the curved raceway. Both conditions induce more severe tribologic conditions. Additionally, posttest analysis is easier on a flat surface. Figure 27.21 shows the eccentric bearing tester. The lower raceway is mounted to a rotating shaft and rigidly supported. The upper raceway is placed inside an assembly containing a load cell. The load is induced into the bearing by compressing springs that push on the upper raceway fixture. The eccentricity can be altered from 0 to 3.06 mm and introduces a skidding element into the ball motion, greatly accelerating lubricant degradation. It also provides a wide wear track, making posttest surface analysis easier. A set of aluminum flexures connects the upper housing to the lower housing, allowing the upper housing to flex slightly. Deflection is measured through a proximity sensor mounted in the lower housing. From the deflection, operating torque is calculated. The entire assembly operates under vacuum of approximately 10−7 Torr. A similar test device exists at the European Space Tribology Laboratory [162]. 27.6.3.2.2 GOES Bearing Test Facility A vacuum, bearing test facility at NASA Glenn Research center is a good representation of component level testing. It was originally developed to study filter wheel bearings from the Geostationary Operational Environmental Satellite (GOES) when it was discovered that bearings onboard GOES missions were operating at a higher temperature than expected. Concerns about the bearing’s operating regime arose and it would have a severely shortened life due to lubricant degradation. This facility isolates a single bearing inside a vacuum chamber (Figure 27.22). An inline torque meter and load cell monitor bearing health. Cross-bearing resistance is also measured and used to determine the bearing’s operating regime. Load
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Rotating ball
Lubricant cup
Staionary balls
Flex pivot
FIGURE 27.19
Detail of four-ball vacuum tribometer.
Detail Load Friction force sensor
Ball of pin specimen Disk specimen
FIGURE 27.20
Ultra-high vacuum pin-on-disk tribometer.
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Proximity probe
Vibration surface
Total bearing Eccentricity Elastic flexture
Proximity probe
Reference surface
Torque sensor
FIGURE 27.21 Aerospace Corp.’s eccentric bearing test mechanism. (From Didziulis et al.,“Lubrication”, Chapter 15, NASA/TP-1999-206988, 1999.)
is applied through a dead weight system. The facility operates at a vacuum level of at least 10−6 Torr. This system’s advantage is the ability to isolate a single bearing, which is useful for studying bearing characteristics without the complexity of an entire system. Some tests on GOES filter wheel bearings are found in the works of Jansen et al. [163,164]. Because of the system’s simplicity and success with the GOES bearing investigations, the facility underwent many design modifications making it a more flexible test facility. The system can be adapted to accept almost any size bearing (up to 100 mm) and heating and cooling modifications were added. The facility’s operating range is −40 to 100◦ C. An advanced stepper motor system accommodates complex motion profiles in addition to rotary and oscillatory motion. The system has been used to investigate toroid issues, torque spikes, and lubricant regime transition temperatures of component bearings.
27.7 Lubricant Selection Lubricant selection for space applications is unique for several reasons. The lubricant must have extremely low vapor pressure, operate at a variety of temperatures, be a good boundary lubricant, and have a low consumption rate, which is defined as the rate at which the lubricant degrades into a nonlubricating
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Drive motor Slip ring assembly
Test section
In-line torque meter Test section
Test section and bearing
Load cell
FIGURE 27.22 GOES bearing test facility (overview and detail). (From Jansen, M.J., Jones, W.R., Jr., and Predmore, R.E., “Evaluation of Temperature and Material Combinations on Several Lubricants for Use in the Geostationary Operational Environmental Satellite (GOES) Mission Filter Wheel Bearings,” NASA TM 2001-211121, August 2001.)
material. Consumption rate is the most vital factor for space lubricants because virtually all bearing failures are due to lubricant degradation and not bearing fatigue or wear.
27.7.1 Relative Life and Wear Characteristics Various Pennzane® formulations have been compared to other hydrocarbons (PAOs) and PFPEs in eccentric bearing tests [26]. The data, shown in Figure 27.23, indicates that a Pennzane® formulated with antimony dialkyldithiocarbamate yielded a lifetime several times that of a PFPE (Krytox™ 143AB). Linear ball screw tests at ESTL (European Space Tribology Laboratory) [165] compared several lubricants. These included Fomblin™ Z-25 oil, Braycote™ 601 grease, Pennzane® SHF-X2000 oil, sputter coated MoS2 , ionplated lead, and Braycote™ 601 + ion-plated lead. The two PFPE lubricants failed rapidly. The Braycote 601 + ion-plated lead combination reached the full 2 million cycle test requirement; however, ion-plated lead alone failed at 400K cycles. The full test was completed by MoS2 , but some polishing of the contact zone was noted. The Pennzane® oil completed the test, but lead screw wear occurred and some dewetting was noted. Earlier work at ESTL [166] compared several solid and liquid lubricants’ performance in oscillating ball bearings. Three liquid/grease lubricants were tested with phenolic cages: Fomblin™ Z-25, Braycote™ 601, and Pennzane® SHF-X2000. Torque levels were measured over 10 million oscillations. For a low angle of oscillation (±0.5◦ ), Fomblin™ Z-25 yielded average torque levels 1.5 times that of Braycote™ and
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6
Relative wear life
5 4 3 2 1 0 PFPE
PAO + Sb MAC + Pb Lubricant
MAC + Sb
FIGURE 27.23 Eccentric bearing screening test results for PFPE, PAO, and MAC oils. (From Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.) 18 Wear rate × 10–9(mm3/mm)
16 14 12 10 8 6 4 2 0 GXL- GXL- GXL- GXL320A 320B 296 296B (Base) (Base)
GXL- GXL- Z-25 320A 320B (Liquid) (Base)
AISI52100 steel
440C stainless steel
FIGURE 27.24 Mean wear rates of various space lubricants using a vacuum four-ball tribometer. (From Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.)
approximately 3 times that of Pennzane®. At ±5◦ , mean torque measurements were similar for all three lubricants, but Pennzane® exhibited the lowest levels. Finally, at ±20◦ , the Z-25 yielded the highest torque, Braycote™ was intermediate, and Pennzane® was the lowest. Long-term angular contact bearing tests were also reported by Gill [167]. Pennzane® SHF-X2000 was tested at 200 and 1,400 rpm for 108 and 109 revolutions, respectively. The bearings did not fail, but did suffer from oil starvation due to evaporation losses. It was not clear if this was due to insufficient sealing or related to a lubricant batch problem. However, it was noted that similar tests with PFPE oils have never failed due to starvation by oil evaporation. A vacuum four-ball tribometer [168] has been used to rank various space lubricants according to wear rates (Figure 27.24), including three PFPEs, Pennzane® base fluid, a formulated Pennzane®, and two other unformulated fluids (a silahydrocarbon and a PAO). In general, higher wear rates represent lower lifetimes in space.
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27.7.2 General Mechanism Effects In space applications, many factors influence a mechanism’s lubricated life. Because lubricant degradation, not mechanical wear or fatigue, is the most common cause of failure, understanding how various design parameters affect the mechanism and its tests is critical. Selection of materials, stress levels, environment, and operational parameters (speed, load, etc.) all play a major role in the success of a lubricated space mechanism. 27.7.2.1 Stress Level Life Relation It has been shown that, under boundary lubrication, the lubricated lifetime is directly related to the operating stress level [84,151]. The life/stress relationship is exponential; therefore, a small change in stress level can have a significant effect on the lubricated life. Proper bearing sizing, resulting in reduced operating stress, is one way designers can easily increase mechanism lifetimes. 27.7.2.2 Environmental Effects Since some lubricants, such as MoS2 , are greatly affected by environment, the designer should carefully consider all of the mechanism’s operating atmospheres. Not only is the final operating environment important, but also the environments during build-up, storage, and run-in testing. Because some liquid lubricants or additives may have a high vapor pressure and evaporate quickly, tests, including accelerated tests, should be done as close to the final vacuum level as possible. The mechanism’s operating temperature is another important consideration. Relatively small temperature changes can greatly affect the lubrication regime and driving torque. Finally, exposure to other external environmental factors, such as radiation, can accelerate lubricant degradation and should be carefully considered before final lubricant selection is made. 27.7.2.3 Mechanism Speeds As with temperature, a mechanism’s operating speed can have a great effect on the operating regime. Small changes in speed can shift a bearing from the boundary regime into the mixed or even EHL regime. Because lifetime and lubrication mechanisms change greatly with the operating regime, this is an important consideration when designing component or system-level tests. When designing a test, both speed and temperature can be adjusted to try to maintain the proper lubrication regime. However, other heating effects such as evaporation and oxidation rates should be considered. 27.7.2.4 Retainers and Ball Separators Use of retainers or other ball separators, such as toroids, can greatly affect mechanism lifetimes. For example, if a porous retainer material is improperly selected or if it is not fully impregnated with lubricant before build-up it can act as a sponge taking lubricant from the mechanism during operation which can lead to contact starvation [18]. In addition, humidity-induced dimensional changes can occur in phenolic retainers [169]. Toroid design, implementation, and material interactions with the lubricants must also be carefully examined before final selection can be made. Improper toroid material selection can accelerate lubricant breakdown due to tribochemical reactions. Additionally, improper toroid application, such as placing them on every ball or using a bearing with too much space between balls, has led to torque anomalies, excessive toroid wear debris in the bearing, and total bearing seizure. When selecting retainer and toroid materials, the designer should consider how the material will interact with both the mechanism and the lubricant. 27.7.2.4.1 Retainerless Bearings One way to eliminate retainer problems is to remove it entirely. This is not a widely accepted practice due to concerns over ball-to-ball contact. By classical EHL theory, no film exists in ball-to-ball contact because the relative surface speed is zero. However, research has shown that a protective lubricant film does exist between balls [170,171] and classic EHL theory does not adequately describe this contact [7,8]. Several bearings have been operated for hundreds of hours without ball-to-ball failure [141,142]. Other
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advantages of eliminating the retainer are greater load distribution among more balls (when compared to the equivalent bearing with a retainer), resulting in a lower operating stress, elimination of possible cage instabilities, and elimination of a source for wear debris in the bearing. While not a standard practice, a designer should not rule out retainerless bearings if the ball separator is causing problems in a mechanism.
27.7.3 Cleaning and Surface Preparation Cleaning or surface treatments are an important part in the preparation of lubricated space mechanisms. Some space lubricants’ degradation rates can be particularly sensitive to changes in bearing surface chemistry [63]. Traditionally, space mechanisms were cleaned with trifluorotrichloroethane (CFC 113) or 1,1,1-trichloroethane (TCA), both ozone-depleting chemicals (ODC). During the 1990s, production of ODCs was ceased and several studies were conducted on alternate cleaning methods [172–176]. Although results varied, all the studies showed both positive and negative cleaning effects on the mechanism’s lubricated life. Although surface cleaning techniques do not have as great an effect on lifetime as the materials used or the mechanism design, it is still a factor that should be specified before build-up to ensure the longest lubricated life. Because many space lubricants are sensitive to surface chemistry, surface passivation is another technique used to extend lubricant life under boundary lubrication. With this technique, the bearing surfaces are passivated to become less chemically active, reducing the rate the lubricant reacts with the surface thus increasing lifetime. Several passivation methods exist, including chromic acid, tricresyl phosphate (TCP) presoak, high temperature chromic acid, and ultra-violet ozone. Many studies have been conducted to study passivation effects with various classes of lubricants [177–187]. While TCP treatment improved results with mineral oils and synthetic esters, surface passivation did not have a significant effect with PFPE lubricated systems [186,187].
27.8 Summary This chapter is to provide a current state-of-the-art review of the lubrication technology for space mechanisms. It is not intended to be an in-depth study. For more details, the readers are directed to the following sources: NASA Space Mechanisms Handbook, edited by Robert L. Fusaro, NASA/TP-1999-206988; Space Vehicle Mechanisms Elements of Successful Design, edited by Peter L. Conley, John Wiley & Sons, Inc, New York, 1998 and Space Tribology Handbook, edited by Emyr W. Roberts, European Space Tribology Laboratory, 2nd Edition, 1997.
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[8] Hamrock, B.T. and Dowson, D., Ball Bearing Lubrication: The Lubrication of Elliptical Contacts, Wiley, New York, 1981. [9] Wedeven, L.D., Evans, W., and Cameron, A., “Optical Analysis of Ball Bearing Starvation,” ASME J. Lubr. Tech., 98, 349, 1971. [10] Kingsbury, E., “Parched Elastohydrodynamic Lubrication,” ASME J. Tribol., 107, 229, 1985. [11] Jones, W.R., Prahl, J., Jansen, R., Schritz, B. et al., “Parched Elastohydrodynamic Lubrication: Instrumentation and Procedure,” Tribol. Trans., 37, 1, 13, 1994. [12] Wu, Y. and Yan, S., “A Full Numerical Solution for the Non-Steady State Elastohydrodynamic Problem in Nominal Line Contacts,” Proc. 12th Leeds-Lyon Symp. on Tribology, p. 291, 1986. [13] Ai, X. and Yu, H., “A Full Numerical Solution for General Transient Elastohydrodynamic Line Contacts and its Application,” Wear, 121, 143, 1988. [14] Hooke, C.J., “The Minimum Film Thickness in Lubricated Line Contacts During a Reversal of Entrainment-General Solution and the Development of a Design Chart,” Proc. Instn. Mech. Engrs., J208, 53, 1994. [15] Sugimura, J., Jones, W.R., Jr., and Spikes, H.A., “EHD Film Thickness in Non-Steady State Contacts,” Trans. ASME J. Tribol., 120, 442, 1998. [16] Glovnea, R.P. et al., “Film Thickness Measurements on Several Lubricants for Space Applications During Rapid Stop-Start Motion,” Proc. 2000 AIMETA Int. Trib. Conf., L’Aquila, Italy, p. 429, 2000. [17] Roberts, E.W. and Todd, M.J., “Space and Vacuum Tribology,” Wear, 136, 157, 1990. [18] Bertrand, P.A., “Oil Absorption into Cotton-Phenolic Material,” J. Mater. Res., 8, 7, 1749, 1993. [19] Dromgold, L.D. and Klaus, E.E., “The Physical and Chemical Characteristics of an Homologous Series of Instrument Oils,” Bearing Conference, Dartmouth College, Hanover, NH, 1968. [20] Kalogeras, C. and Tran, A., “Rotating Mechanisms,” in NASA Space Mechanisms Handbook, Fusaro, R. (Ed.), Chap. 8, 1999, pp. 97–112. [21] Shubkin, R., “Perfluoroalkylpolyethers,” in Synthetic Lubricants and High-Performance Functional Fluids, Shubkin, R. (Ed.), pp. 1–40, 1993. [22] Bollea, D., Jones, W., Marchetti, M., Jansen, M., Dube, M.J. et al., “A New Synthetic Hydrocarbon Liquid Lubricant for Space Applications,” Tribol. Lett., 15, 1, 2003. [23] Venier, C.G. and Casserly, E.W., “Multiply-Alkylated Cyclopentanes (MACs): A New Class of Synthesized Hydrocarbon Fluids,” Lubr. Eng., 47, 7, 586, 1991. [24] Venier, C.G. and Casserly, E.W., “Cycloaliphatics,” in Synthetic Lubricants and High Performance Functional Fluids, Shubkin, R. (Ed.), 1993. [25] Casserly, E., Jones, W., Marchetti, M., Jansen, M., Predmore, R., Venier, C. et al., “Tribological Properties of a Pennzane-Based Liquid Lubricant (Disubstituted Alkylated Cyclopentane) for Low Temperature Space Applications,” Proc. 36th Aerospace Mechanisms Symp., NASA CP-2002-211506, p. 331, 2002. [26] Kalogeras, C.G., Didziulis, S., Fleishaver, P., Baver, R., Carré, D.J. et al., “Recent Experience with Synthetic Hydrocarbon Lubricants for Spacecraft Applications,” Aerospace Report TR-95(5935)-3, 1995. [27] Miller, J., Jones, W., Rasmussen, K., Wheeler, D., Rana, M., Peri, F., Brown, P.L. et al., “The Clouds and the Earth’s Radiant Energy System Elevation Bearing Assembly Life Test,” Proc. 33rd Aerospace Mechanisms Symp., Pasadena, CA, NASA CP-1999-209259, p. 197, 1999. [28] Dietz, B., Street, K., Jones, W., Jansen, M., Dube, M., Sharma, R., Predmore, R., VanDyk, S.G. et al., “The Role of Bearing and Scan Mechanism Life Testing in Flight Qualification of the MODIS Instrument,” Tribotest J., 9, 2, 2002. [29] Gumprecht, W.H., “PR-143-A New Class of High-Temperature Fluids,” ASLE Trans., 9, 24, 1966. [30] Pasetri, A., Fontanelli, R., Bernardi, G., Sianesi, D. et al., “Perfluoropolyethers by Photooxidation of Fluoroolefines,” Chim. Ind., 55, 208, 1973. [31] Zamboni, V., Fontanelli, R., Binaghi, M., Sianesi, D. et al., “Perfluoropolyethers. Their Physical Properties and Behavior at High and Low Temperatures,” Wear, 18, 85, 1971.
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[32] Ohsaka, Y., “Recent Advances in Synthetic Lubricating Oils (8). Perfluorinated Polyethers,” Petrotech, 8, 9, 840, 1985. [33] Del Pesco, T.W., “Perfluoroalkylpolyethers,” in Synthetic Lubricants and High-Performance Functional Fluids, Shubkin, R. (Ed.), pp. 145–172, 1993. [34] Johnson, R., Winer, W., Sanborn, D., Jones, W.R., Jr. et al., “Pressure–Viscosity Measurements for Several Lubricants to 5.5 × 108 Newtons per Square Meter (8 × 104 psi) and 149C (300F),” ASLE Trans., 18, 4, 249, 1975. [35] Gschwender, L.J., Snyder, C.E., Jr., and Fultz, G.W., “Soluble Additives for Perfluoropolyalkylether Liquid Lubricants,” Lubr. Eng., 49, 9, 702–708, 1993. [36] Gschwender, L., Sharma, S., Liang, J., Fuhz, G., Helmick, L.S. et al., “The Effect of Humidity on the Wear Behavior of Bearing steels with Rf O (n-C3 F6 O)m Rf Perfluoropolyalkylether Fluids and Formulations,” Tribol. Trans., 40, 3, 393, 1997. [37] Jones, W.R., Jr., Ajayi, O.O., and Wedeven, L.D., “Enhancement of Perfluoropolyether Boundary Lubrication Performance,” NASA TM 107306, 1996. [38] Masuko, M., Takeshita, N., and Okabe, H.,“Evaluation of Anti-Wear Performance of PFPE-Soluble Additives under Sliding Contact in High Vacuum,” Tribol. Trans., 38, 3, 679, 1995. [39] Nakayama, K., Dekura, T., and Kobayashi, T., “Effect of Additives on Friction, Wear and Iron Fluoride Formation under Perfluoropolyether Fluid Lubrication in Vacuum and Various Atmospheres Containing Oxygen,” Wear, 192, 1768, 1996. [40] Sharma, S.K., Gschwender, L.J., and Snyder, C.E., Jr., “Development of a Soluble Lubricity Additive for Perfluoropolyalkylether Fluids,” J. Synth. Lubr., 7, 1, 15, 1990. [41] Jones, W.R., Herrera-Fierrro, P., Lin, T.Y., Kawa, M., Shogrin, B.A. et al., “Evaluation of BoundaryEnhancement Additives for Perfluoropolyethers,” Tribol. Trans., 42, 747, 1999. [42] Corti, C., Montagna, L., Savelli, P., Srinivasan, P. et al., “Soluble Additives for Perfluorinated Lubricants,” J. Synth. Lubr., 10, 143, 1993. [43] Williams, J.R., Feuchter, D.K., and Jones, W.R., Jr., “Development and Preliminary Evaluation of Aryl Ester Boundary Additives for Perfluoropolyethers,” NASA TM 106603, 1994. [44] Tamborski, C. et al., “Synthesis and Properties of Silahydrocarbons, A Class of Thermally Stable, Wide-Liquid Range Fluids,” Ind. Eng. Chem. Prod. Res. Dev., 22, 172, 1983. [45] Snyder, C., Gschwender, L., Randolph, B., Paciorek, K., Shih, J., Chen, G., Snyder, C.E., Jr. et al., “Research and Development of Low-Volatility Long Life Silahydrocarbon-Based Liquid Lubricants for Space,” Lubr. Eng., 48, 325, 1992. [46] Shih, J., Kratzer, R., Randolph, B., Snyder, C., Paciorek, K.J. et al., “Polysilahydrocarbon Synthetic Fluids. 1. Synthesis and Characterization of Trisilahydrocarbons,” I & EC Research, 29, 1855, 1990. [47] Shih, J., Kratzer, R., Randolph, B., Snyder, C., Paciorek, K.J. et al., “Polysilahydrocarbon Synthesis Fluids. 2. Synthesis and Characterization of Tetrasilahydrocarbons,” I & EC Research, 30, 2191, 1991. [48] Jones, W.R., Jr., Shogrin, B.A., and Jansen, M.J., “Research on Liquid Lubricants for Space Mechanisms,” J. Synth. Lubr., 17, 2, 109, 2000. [49] Spikes, H.A., “Estimation of the Pressure–Viscosity Coefficients of Two Silahydrocarbon Fluids at Two Temperatures,” Imperial College (Tribology Section), Report TS024b/96, London, UK, 1996. [50] Jansen, M., Gschwender, L., Snyder, C., Sharma, S., Predmore, R., Dube, M., Jones, W.R., Jr. et al., “The Tribological Properties of Several Silahydrocarbons for Use in Space Mechanisms,” J. Synth. Lubr., 20, 4, 305, 2004. [51] Fusaro, R.L. and Khonsari, M.M., “Liquid Lubrication in Space,” NASA TM 105198, 1991. [52] Stone, D. and Bessette, P., “Liquid Lubricants,” in Space Vehicle Mechanisms: Elements of Successful Design, Conley, P.L. (Ed.), Chap. 8, John Wiley & Sons, New York, 1998. [53] Zaretsky, E.V., “Liquid Lubrication in Space,” Tribol. Int., 23, 2, 75, and NASA Ref. Publ. 1240, 1990. [54] Fusaro, R.L. (Ed.), “NASA Space Mechanisms Handbook,” NASA/TP-206988, 1999. [55] Roberts, E.W. (Ed.), The Space Tribology Handbook, 2nd ed, publ. ESTL, AEA Technology, 1999.
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[56] McMurtrey, E.L., “Lubrication Handbook for the Space Industry,” NASA TM-86556, 1985. [57] Hilton, M.R. and Fleischauer, P.D., “Lubricants for High Vacuum Applications,” Aerospace Report TR-0091 (6945-03)-6, 1990. [58] Conley, P., “Space Vehicle Mechanisms: Elements of Successful Design,” John Wiley & Sons, New York, 1998. [59] Nguyen, Q.N. and Jones, W.R., Jr., “Volatility and Wear Characteristics of a Variety of Liquid Lubricants for Space Applications,” Tribol. Trans., 44, 4, 671, 2001. [60] Carré, D.J. and Bertrand, P.A., “A Model to Calculate Evaporative Oil Loss in Spacecraft Mechanisms,” Tribol. Trans., 42, 2, 282–288, 1999. [61] Carré, D.J. and Bertrand, P.A., “Modeling and Measurement of Aryl Phosphate Ester Vapor Pressures at 50◦ C,” Tribol. Trans., 42, 4, 777–782, 1999. [62] Kinzig, B.J. and Ravner, H., “Factors Contributing to the Properties of Fluoropolymer Barrier Films,” ASLE Trans., 21, 291, 1978. [63] Jones, W.R., Jr., “Properties of Perfluoropolyethers for Space Applications,” Tribol. Trans., 38, 3, 557, 1995. [64] Wedeven, L.D., “What is EHD?” Lubr. Eng., 31, 6, 291, 1975. [65] Spikes, H.A., Cann, P., and Caporiccio, G., “Elastohydrodynamic Film Thickness Measurements of Perfluoropolyether Fluids,” J. Synth. Lubr., 1, 1, 73, 1984. [66] Vergne, P. and Reynaud, P., “High Pressure Behavior of Space Liquid Lubricants,” Tribology 2000, Vol. 2, Bartz, W.J. (Ed.), Tech. Acad. Esslingen, Ostfildern, Germany, 1992. [67] Barrall II, E., Wolf, B., Geerissen, H., Cantow, M.J. et al., “Temperature and Pressure Dependence of the Viscosities of Perfluoropolyether Fluids,” J. Polym. Sci. Part B: Polym. Phys., 25, 603, 1987. [68] Johnson, G.J., Spikes, H., Caporiccio, G., Aderin, M. et al., “The Elastohydrodynamic Properties of Some Advanced Non Hydrocarbon-Based Lubricants,” Lubr. Eng., 48, 8, 633, 1992. [69] Barus, C., “Isothermals, Isopiestics and Isometrics Relative to Viscosity,” Am. J. Sci., 45, 87, 1893. [70] Roelands, C.H., “Correlational Aspects of the Viscosity–Temperature–Pressure Relationship of Lubricating Oils,” O.P. Books Program, University Microfilm, Ann Arbor, MI, 1966. [71] Nélias, D. et al., “Traction Behavior of Some Lubricants Used for Rolling Bearings in Spacecraft Applications: Experiments and Thermal Model Based on Primary Laboratory Data,” ASME Trans., 124, 72–81, 2002. [72] Foord, C.A., Hamman, W.C., and Cameron, A., “Evaluation of Lubricants Using Optical Elastohydrodynamics,” ASLE Trans., 11, 1, 31, 1968. [73] Spikes, H.A., “Film Formation and Friction Properties of Five Space Fluids,” Imperial College (Tribology Section), Report TS037/97, London, UK, 1997. [74] Lauer, J.L. and Jones, W.R., Jr., “Friction Polymers,” ASLE Spec. Publ. Sp-21, Tribology and Mechanics of Magnetic Storage Systems, III, 14–23, 1987. [75] Mori, S. and Morales, W., “Tribological Reactions of Perfluoroalkyl Polyether Oils with Stainless Steel under Ultrahigh Vacuum Conditions at Room Temperature,” Wear, 132, 111, 1989. [76] Carré, D.J., “Perfluoropolyalkylether Oil Degradation: Inference of FeF3 Formation on Steel Surfaces Under Boundary Conditions,” ASLE Trans., 29, 2, 121, 1986. [77] Herrera-Fierro, P., Jones, W.R., Jr., and Pepper, S.V., “Interfacial Chemistry of a Perfluoropolyether Lubricant Studied by X-Ray Photoelectron Spectroscopy and Temperature Desorption Spectroscopy,” J. Vac. Sci. Technol. A, 11, 2, 354, 1993. [78] Carré, D.J. and Markovitz, J.A., “The Reaction of Perfluoropolyalkylether Oil with FeF3 , AlF3 , and AlCl3 ,” STLE Trans., 28, 1, 40, 1985. [79] Kasai, P.H., “Perfluoropolyethers: Intramolecular Disproportionation,” Macromolecules, 25, 6791, 1992. [80] Carré, D.J., “The Use of Solid Ceramic and Ceramic Hard-Coated Components to Prolong the Performance of Perfluoropolyalkylether Lubricants,” Sur. Coat. Technol., 43/44, 609, 1990. [81] Boving, H.J., Heanni, W., and Hintermann, H.E., “Titanium Carbide Coatings for Aerospace Ball Bearings,” Proc. 22nd Aerospace Mechanisms Symp., Sunnyvale, CA, NASA CP-2506, p. 245, 1988.
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[82] Prie, W., Rowntree, R., Boving, H., Hintermann, H., Gill, S. et al., “In-Vacuum Performance of Fomblin Z25-Lubricated, 52100 Steel Bearings with TiC-Coated Balls,” Proc. Fifth European Space Mechanisms and Tribology Symp., Noordwijk, The Netherlands, ESA SP-334, p. 165, 1993. [83] Jansen, M., Helmick, L., Nguyen, Q., Wheeler, D., Boving, M., Jones, W.R., Jr. et al., “The Effect of Stress and TiC Coated Balls on Lubricant Lifetimes Using a Vacuum Ball-on-Plate Rolling Contact Tribometer,” Proc. 33rd Aerospace Mechanisms Symp., Pasadena, CA, NASA /CP-1999-209259, p. 237, 1999. [84] Pepper, S., Jansen, M., Nguyen, Q., Wheeler, D., Schröer, A., Jones, W.R., Jr. et al., “The Effect of Stress and TiC Coated Balls on Lifetime of a Perfluoropolyalkylether Using a Vacuum Rolling Contact Tribometer,” NASA TM 2000-209925, March 2000. [85] DiGesu, F. and Sicre, J., “Characterization of Lubrication Effect on Performances and Lifetime of Ceramic and Hybrid Ball Bearings,” Proc. 9th European Space Mech. and Tribology Symp., SP-480, Liège, September 19–21, 2001. [86] Popp, M. and Sternagel, R., “Hybrid Ceramic and All Ceramic Anti Friction Bearings,” Proc. 8th European Space Mech. and Tribology Symp., Toulouse, France, SP-438, p. 105, 1999. [87] Massey, M., Gschwender, L., Snyder, C., Zabinski, J., Sharma, S., Rai, A.K. et al., “Performance Evaluation of Some Pennzane®-Based Greases for Space Applications,” Proc. 33rd Aerospace Mechanisms Sym., Pasadena, CA, p. 213, 1999. [88] Fowzy, M., Landry, J., Jones, W., Shogrin, B., Nguyen, Q., Jones, D. et al., “An Additive to Improve the Wear Characteristics of Perfluoropolyether Based Greases,” NASA TM-1999-209064, 1999. [89] Jones, W., Street, K., Peppery, S., Jansen, M., Marchetti, M. et al., “Preliminary Evaluation of Greases for Space Mechanisms Using a Vacuum Spiral Orbit Tribometer,” NASA TM 2001-211157, 2001. [90] Marchetti, M., Jones, W.R., Jr., and Sicre, J., “Relative Lifetimes of MAPLUB® Greases for Space Applications,” NASA TM 2002-211875, 2002. [91] Jones, W., Street, K., Wheeler, D., Dixon, D., Jansen, M., Kimura, H., Marchetti, M. et al., “Tribological Performance of Some Pennzane-Based Greases for Vacuum Applications,” Tribol. Lett., 12, 4, 209, 2002. [92] Roberts, E.W., “Thin Solid Lubricant Films in Space,” Tribol. Int., 23, 2, 95–104, 1990. [93] Spalvins, T., “A Review of Recent Advances in Solid Film Lubrication,” J. Vac. Sci. Technol. A, 5, 2, 212, 1987. [94] Spalvins, T.,“The Improvement of Ion-Plated Ag and Au Film Adherence to Si3 N4 and SiC Surfaces for Increased Tribological Performance,” NASA TM-1998-207415, 1998. [95] Spalvins, T., “Lubrication with Sputtered MoS2 Films: Principles, Operation and Limitations,” J. Matl. Engr. Perf., 1, 347, 1992. [96] Fox, V., Hampshire, J., and Teer, D., “MoS2 /Metal Composite Coatings Deposited by CloseField Unbalanced Magnetron Sputtering: Tribological Properties and Industrial Uses,” Surf. Coat. Technol., 112, 118, 1999. [97] Fusaro, R.L., “Mechanisms of Lubrication and Wear of a Bonded Solid Lubricant Film,” ASLE Trans., 24, 2, 191, 1981. [98] Peebles, D., Ohlhavsen, J., Robinson, S., Sorroche, E., Dugger, M.T. et al., “Oxidation Effects on the Friction of Lubricants and Self-Lubricating Materials in the Enduring Stockpile,” Proc. 22nd Aging, Compatibility and Stockpile Stewardship Conf., Oak Ridge, TN, 1999. [99] Fusaro, R.L., “Self-Lubricating Polymer Composites and Polymer Transfer Film Lubrication for Space Applications,” Tribol. Int., 23, 2, 105, 1990. [100] Gardos, M.N., “Self Lubricating Composites for Extreme Lubricating Conditions,” in Friction and Wear of Polymer Composites, F. Klaus (Ed.), Elsevier Science Publishers, p. 397, 1986. [101] Brewe, D.E., Scibbe, H.W., and Wisander, D.W., “Performance of High-Speed Ball Bearings with Lead and Lead-Alloy-Plated Retainers in Liquid Hydrogen at 1.2 Million DN,” J. Lubr. Technol., p. 437, 1974.
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[102] Todd, M.J. and Parker, K., “Giotto’s Antenna De-Spin Mechanism: Its Lubrication and Thermal Vacuum Performance,” Proc. 21st Aerospace Mechanisms Symp., NASA CP 2470, 1987. [103] Sheppard, S., “Design and Development of an Advanced Solar Array Drive Mechanism,” Proc. 1st European Space Mechanisms and Tribology Symp., ESA-SP 196, p. 19, 1983. [104] Fabbrizzi, F., Sawyer, E., and Gill, S., “Life Test Development and Results for the Gerb Mirror De-Spinning Mechanism,” Proc. 33rd Aerospace Mechanisms Symp., Pasadena, CA, NASA/CP1999-209259, p. 221, 1999. [105] Arnell, R.D., “The Effects of Speed, Film Thickness and Substrate Surface Roughness on the Friction and Wear of Soft Metal Films in Ultra High Vacuum,” Thin Solid Films, 53, 333, 1978. [106] Lince, J.R. and Fleischauer, P.D., “Solid Lubricants,” in Space Vehicle Mechanisms: Elements of Successful Design, P.L. Conley (Ed.), Chap. 7, John Wiley & Sons, New York, 1998. [107] Hilton, M.R. and Fleischauer, P.D.,“Applications of Solid Lubricant Films in Spacecraft,” Aerospace Report TR-92(2935)-6, 1994. [108] Loewenthal, S.H. et al., “Evaluation of Ion-Sputtered Molybdenum Disulfide Bearings for Spacecraft Gimbals,” Tribol. Trans., 37, 3, 505, 1994. [109] Hopple, G.B. and Loewenthal, S.H., “Development, Testing and Characterization of MoS2 Thin Film Bearings,” Sur. Coat. Technol., 68/69, 398, 1994. [110] Hicton, M.R., Bauer, R., Fleishaver, P., Didziulis, S.V. et al., “Thrust Bearing Wear Life and Torque Tests of Sputter-Deposited MoS2 Films,” The Aerospace Corp., El Segundo, CA, Report No. TOR-92(2064)-1, 1992. [111] Lince, J.R., Hilton, M.R., and Bommannavar, A.S., “Oxygen Substitution in Sputter-Deposited MoS2 Films Studied by Extended X-ray Absorption Fine Structure, X-ray Photoelectron Spectroscopy and X-ray Diffraction,” Surf. Coat. Technol., 43/44, 640, 1990. [112] Roberts, E.W., Williams, B.J., and Ogilvy, J.A., “The Effect of Substrate Surface Roughness on the Friction and Wear of Sputtered MoS2 Films,” J. Phys. D: Appl. Phys., 25, A65, 1992. [113] Roberts, E.W., “The Lubricating Properties of Magnetron Sputtered MoS2 ,” European Space Agency, ESA-5615/83-NL-PP, 1987. [114] Lince, J.R., “Optimizing Co-Sputtered Nanocomposite Au/MoS2 Solid Lubricant Films for High or Low Contact Stress,” Tribol. Lett., in press. [115] Spalvins, T., “Frictional and Morphological Properties of Au-MoS2 Films Sputtered from a Compact Target,” Thin Solid Films, 118, 375, 1984. [116] Roberts, E.W. and Price, W.B., “Advances in Molybdenum Disulfide Film Technology for Space Applications,” Proc. Sixth European Space Mechanisms and Tribology Symp., Zurich, Switzerland, p. 273, 1995. [117] Stupp, B.C., “Synergistic Effects of Metals Co-Sputtered with MoS2 ,” Thin Solid Films, 84, 257, 1981. [118] Zhang, X., Yu, D., Wang, X., Liu, H. et al., “The Enhancement of Wear Life and Moisture Resistance of Sputtered MoS2 Films by Metal Ion Implantation,” Wear, 173, 145, 1994. [119] Fox, V., Teer, D., Hampshire, J., Renevier, N.M. et al., “Coating Characteristics and Tribological Properties of Sputter-Deposited MoS2 /Metal Composite Coatings Deposited by Closed Field Unbalanced Magnetron Sputter Ion Plating,” Proc. ICMCTF 99 Conf., San Diego, CA, Paper E1–5, 1999. [120] Fusaro, R.L. and Siebert, M., “Comparison of Several Different Sputtered Molybdenum Disulfide Coatings for Use in Space Applications,” Proc. 36th Aerospace Mechanisms Symp., NASA CP-2002211506, p. 305, 2002. [121] Fleishauer, P., “Effects of Crystallite Orientation on Environmental Stability and Lubrication Properties of Sputtered MoS2 Thin Films,” ASLE Trans., 27, 1, 82, 1984. [122] Suzuki, M. and Prat, P., “Synergism of an MoS2 Sputtered Film and a Transfer Film of PTFE Composite,” Wear, 995, 225–229, 1999. [123] Enke, K., Dimigen, H., and H`‘ubsch, H., Appl. Phys. Lett., 36, 4, 291, 1980. [124] Memming, R., Tolle, H.J., and Wierenga, P.E., Thin Solid Films, 143, 31, 1986.
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[125] Miyoshi, K., Pouch, J.J., and Alterovitz, S.A., Mater. Sci. Forum, 52–53, 645, 1989. [126] Heimberg, J.A. et al., “Superlow Friction Behavior of Diamond-Like Carbon Coatings: Time and Speed Effects,” Appl. Phys. Lett., 78, 17, 2449–2451, 2001. [127] Maillat, M. and Hintermann, H.E., Proc. Fifth European Space Mechanisms and Tribology Symp., ESTEC, Noordwijk, The Netherlands, ESA SP-334, April 1993. [128] Donnet, C., Surf. Coat. Technol., 80, 151, 1996. [129] Miyake, S. et al., ASLE Trans., 30, 1, 121, 1987. [130] Fontaine, J. et al., “Towards the Solid Lubrication of Space Mechanisms by Diamond-Like Carbon Coatings,” Proc. 8th European Space Mechanisms & Tribology Symp., Toulouse, France, SP-438, p. 297, 1999. [131] Vercammen, K. et al., “Study of RF PACVD Diamond Like Carbon Coatings Deposited at Low Bias for Vacuum Applications,” Proc. 9th European Space Mechanisms & Tribology Symp., Liege, Belgium, SP-480, p. 309, 2001. [132] Didziulis et al., “Lubrication,” NASA/TP-1999-206988, 1999. [133] Lince, J., “Electrical Contact Ring Assemblies,” in NASA Space Mechanisms Handbook, Chap. 9, Fusaro, R. (Ed.), 1999. [134] Hilton, M., Carré, D., Didzivlis, S., Fleischaver, P., Kalogeras, C. et al., “The Use of Screening Tests in Spacecraft Lubricant Evaluation,” Aerospace Corp. Report No. TR-93(3935)-6, 1993. [135] Postma, R.W., “Pointing Mechanisms,” in NASA Space Mechanisms Handbook, Chap. 9, Fusaro, R. (Ed.), 1999. [136] Sarafin, T. and Larson, W. (Eds), Spacecraft Structures and Mechanisms — From Concept to Launch, Microcosm, Inc., Torrance, CA, 1995. [137] Jansen, M.J., Jones, W.R., Jr., and Pepper, S.V., “Evaluation of an In-Situ, Liquid Lubrication System for Space Mechanisms Using a Vacuum Spiral Orbit Tribometer,” Proc. ASME/STLE Joint Conf., Cancun, Mexico, October 27–31, 2002. [138] Glassow, F.A., “Assurance of Lubricant Supply in Wet-lubricated Space Bearings,” Proc. 10th Aero. Mechanisms Symp., NASA CR-148515, p. 90, 1976. [139] James, G.E., “Positive Commandable Oiler for Satellite Bearing Lubrication,” Proc. 11th Aerospace Mechanisms Symp., NASA TM-79356, p. 87, 1977. [140] Loewenthal, S.H. et al., “Operating Characteristics of a 0.87 kW-hr Flywheel Energy Storage Module,” Proc. 20th Intersociety Energy Conv. Engr. Conf., 164, 1985. [141] Hanson, R.A., Jones, W., Predmore, R., Kingsbury, E. et al., “Cartridge Bearing System for Space Applications,” Proc. 33rd Aerospace Mechanisms Symp., Pasadena, CA, May 19–21, 1999. [142] Jones, W.R., Jr., Shogrin, B.A., and Kingsbury, E., “Long Term Performance of a Retainerless Bearing Cartridge with an Oozing Flow Lubricator for Spacecraft,” Proc. Int. Rolling Element Bearing Symp., 1997. [143] Marchetti, M. et al., “Lubricant Supply by Porous Reservoirs in Space Mechanisms,” Proc. 26th Leeds-Lyon Symp., Leeds, UK, p. 777, 1999. [144] Jones, W., Pepper, S., Jansen, M., Predmore, R., Marchetti, M. et al., “In-Situ, On-Demand Lubrication System for Space Mechanisms,” Tribol. Trans., 46, 3, 452, 2003. [145] Kato, K., Furuyama, H., and Mizumoto, M., “The Fundamental Properties of Tribo-Coating Films in Ultra-high Vacuum,” Proc. Japan International Tribology Conf., Nagoya, Japan, 261, 1990. [146] Kato, K. et al., “Basic Study of Lubrication by Tribo-Coating for Space Machines,” Tranp. Japan. Soc. Mech. Eng., 62, 600, 3237, 1996. [147] Adachi, K. and Kato, K., “Tribo-Coating in Ultra-high Vacuum,” Proc. 74th JSME Spring Ann. Meet., p. 420, 1997. [148] Adachi, K., Kato, K., and Oyamada, T., “Tribo-Coating for Friction Control in Ultra-high Vacuum,” Int. Symp. for High Perf. Tribol., 61, Taegu, Korea, May 29–30, 1998. [149] Pepper, S.V., Kingsbury, E., and Ebihara, B.T., “A Rolling Element Tribometer for the Study of Liquid Lubricants in Vacuum,” NASA TP 3629, 1996.
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[150] Pepper, S.V. and Kingsbury, E.P.,“Spiral Orbit Tribometry — Part I: Description of the Tribometer,” Tribol. Trans., 46, 1, 57, 2003. [151] Jones, W., Pepper, S., Wheeler, D., Schröer, A., Flühmann, F., Loenenthal, S., Shogrin, B., Jansen, M.J. et al., “The Effect of TiC Coated Balls and Stress on the Lubricant Lifetime of a Synthetic Hydrocarbon (Pennzane® 2001A) Using a Vacuum Spiral Orbit Tribometer,” Proc. Int. Tribol. Conf., p. 1433, Nagasaki, Japan, October 29–November 2, 2000. [152] Dayton, C.J. and Warden, R.M., “Effect of Alternate Solvent Diluents on PFPE Performance,” Proc. 35th Aerospace Mechanisms Symp., NASA CP-2001-209626, p. 25, 2001. [153] Pepper, S.V. and Kingsbury, E., “Destruction of Fomblin Z-25 by Different Bearing Metals,” Proc. 4th Int. Rolling Element Bearing Symp., 1997. [154] Pepper, S.V. and Kingsbury, E.P., “Spiral Orbit Tribometry — Part II: Evaluation of Three Liquid Lubricants in Vacuum,” Tribol. Trans., 46, 1, 65, 2003. [155] Conley, P. and Bohner, J.J., “Experience with Synthetic Fluorinated Fluid Lubricants,” Proc. 24th Aerospace Mechanisms Symp., NASA CP-3062, p. 213, 1990. [156] Murray, S. and Heshmat, H., “Accelerated Testing of Space Mechanisms,” NASA CR-198437, 1995. [157] Pepper, S., Jansen, M., Ngvyen, Q., Kingsbury, E., Loewenthal, S., Predmore, R., Jones, W.R., Jr. et al., “A New Apparatus to Evaluate Lubricants for Space Applications — The Spiral Orbit Tribometer (SOT),” Proc. Intern. Spring Fuels and Lube. Meet. and Exposition, Paris, France, June 19–22, 2000. [158] Bazinet, D.G. et al., “Life of Scanner Bearings with Four Space Liquid Lubricants,” Proc. 37th Aerospace Mechanisms Symp., NASA CP-2004-212073, Galveston, TX, May 19–21, 2004. [159] Jones, W., Predmore, R., Loewenthal, S., Jansen, M.J. et al., “Relative Lifetimes of Several Space Liquid Lubricants Using a Vacuum Spiral Orbit Tribometer (SOT),” in New Horizons for Tribology and Lubricants, Vol. 25, Expert Verlag, Germany, 2002. [160] Jones, W., Jansen R., Ebihara, B., Pepper, S., Helmik, L., Masuko, M. et al., “A Vacuum Four-Ball Tribometer to Evaluate Liquid Lubricants for Space Applications,” NASA TM-106264, 1994. [161] Iwaki, M., Gotoh, K., Obara, S., Imagawa, K., Miyoshi, K. et al., “Friction and Wear Properties of Selected Solid Lubricating Films,” NASA TM 209088, 1999. [162] Gill, S., Vine, M.K., and Rowntree, R.A., “BLAST, A New Lubricant Screening Tester for Space Oils and Greases,” Proc. Fifth European Space Mechanisms and Tribology Symp., Noordwijk, The Netherlands, ESA SP-334, p. 365, 1993. [163] Jones, W., Pepper, S., Jansen, M.J. et al., “EHL Transition Temperature Measurements on a Geostationary Operational Environmental Satellite (GOES) Filter Wheel Bearing,” NASA TM 2001-210670, January 2001. [164] Jansen, M.J., Jones, W.R., Jr., and Predmore, R.E., “Evaluation of Temperature and Material Combinations on Several Lubricants for Use in the Geostationary Operational Environmental Satellite (GOES) Mission Filter Wheel Bearings,” NASA TM 2001-211121, August 2001. [165] Gill, S., “A Comparison of the Performance of Linear Ballscrew Devices when Lubricated with Dry and Liquid Lubricants in Vacuum,” Proc. European Space Mechanisms and Tribology Symp., Noordwijk, The Netherlands, p. 101, 1997. [166] Gill, S.,“A Comparison of the Performance of Solid and Liquid Lubricants in Oscillating Spacecraft Ball Bearings,” Proc. 28th Aerospace Mechanisms Symp., Cleveland, OH, p. 229, 1994. [167] Gill, S. and Rowntree, R.A., “Interim Results From ESTL Studies on Static Adhesion and the Performance of Pennzane® SHF X-2000 in Ball Bearings,” Proc. Sixth European Space Mechanisms and Tribology Symp., Zurich, Switzerland, p. 279, 1995. [168] Poslowski, A., Shogrin, B., Herrera-Fierro, P., Jansen, M., Jones, W.R., Jr. et al., “Evaluation of Several Space Lubricants Using a Vacuum Four-Ball Tribometer,” NASA TM 208654, 1998. [169] Bertrand, P.A. and Sinsheimer, J.D.,“Humidity-Induced Dimensional Changes in Cotton-Phenolic Ball-Bearing Retainers,” Trans. ASME, 124, 3, 474–479, 2002.
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[170] Jones, W., Kingsbury, E., Jansen, M., Prahl, J., Shogrin, B.A. et al., “Experimental Determination of Load Carrying Capacity of Point Contact at Zero Entrainment Velocity,” NASA TM-1999-208848, 1999. [171] Jones, W., Jansen, M., Prahl, S., Thompson, P.M. et al., “The Effect of Sliding Speed on Film Thickness Land Pressure Supporting Ability of a Point Contact Under Zero Entrainment Velocity Conditions,” NASA TM-2000-210566, 2000. [172] Toddy, T., Predmore, R., Shogrin, B., Herrera-Fierro, R., Jones, W.R. et al., “The Effect of ODC-Free Cleaning Techniques on Bearing Lifetimes in the Parched Elastohydrodynamic Regime,” NASA TM 107322, 1996. [173] Hall, P.B. and Thom, R.L., “Adhesion Performance of Solid Film Lubricants on Substrates Cleaned with Environmentally Compliant Cleaners,” NASA CP-3349, pp. 703–707, March 1997. [174] Loewenthal, S.H., Jones, W.R., Jr., and Predmore, R. “Life of Pennzane® and 815Z-Lubricated Instrument Bearings Cleaned with Non-CFC Solvents,” NASA/TM-209392, 1999. [175] Jones, W., Wheeler, D., Keller, D., Jansen, M.J. et al., Evaluation of Non-Ozone-DepletingChemical Cleaning Methods for Space Mechanisms Using a Vacuum Spiral Orbit Rolling Contact Tribometer,” Lubr. Eng., 57, 10, 22, 2001. [176] Loewenthal, S.H. et al., “Instrument Bearing Life with Non-CFC Cleaners,” Proc. 4th Int. Rolling Elem. Symp., Orlando, FL, April 1997. [177] Singer, H.B., “The Effect of TCP Treatment on the Low-Speed Performance of Ball Bearings,” MIT/IL E-2317, September 1968. [178] Allen, S., “Effect of Surface Condition of Instrument Ball Bearings,” Proc. Int. Conf. Surf. Technol., 566–583, 1973. [179] Singer, H.B.,“A Potpourri of Ball Bearing Surface Chemistry Problems,” Draper Lab Report P-1474, 1982. [180] Murday, J.S. et al., “Surface Chemistry of Ball Bearing Steels I,” NRL Memorandum Report 3047, April, 1975. [181] Shafrin, E.G. and Murday, J.S., “Auger Compositional Analysis of Ball Bearing Steels Reacted with Tricresyl Phosphate,” ASLE Trans., 21, 4, 329–336, 1978. [182] Shafrin, E.G. and Murday, J.S., “Analytical Approach to Ball-Bearing Surface Chemistry,” J. Vac. Sci. Technol., 14, 1, Jan/Feb 1977. [183] Allen, S., “Final Report, Gyro Bearing Program, Tasks 1 to 6, NAS 9-3079,” MIT/DL R-586, June 1970. [184] Woodwell, R.G., Miller, L., and Anderson, C., “Alkaline Cleaning: Its Effects on Tricresyl Phosphate Coated Bearing Steels,” Non-Ozone Depleting Chemical Cleaning and Lubrication of Space Systems Mechanisms Comp., Pub. No S69-5762-1.0-1, September 26–27, 1994. [185] Hannan, C.H., “The Effect of Tricresyl Phosphate Coating on Reducing Torques in Low-Speed Bearings,” Proc. Dartmouth Bear. Conf., Dartmouth College, September 3–6, 1968. [186] Jones, W., Herrera-Fierro, P., Jansen, M., Shogrin, B.A. et al., “The Effects of Acid Passivation, Tricresyl Phosphate Presoak, and UV/Ozone Treatment on the Tribology of PerfluoropolyetherLubricated 440C Stainless Steel Couples,” J. Synth. Lubr., 19, 4, 283, January 2003. [187] Wedeven, L.D. and Goodell, A.J., “Evaluation of Passivated and TCP Treated 440C Surfaces with WA Machine,” Wedeven and Associates, Subcontract No. SS-274275-J, September 1991.
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28 Friction and Wear in Lubricated Sheet Metal Forming Processes List of Symbols . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28.2 Tribological System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
28-1 28-3 28-4
System Analysis • Operational Variables • Sheet Materials • Tool Materials • Lubrication in Sheet Metal Forming
28.3 Friction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
28-17
Lubrication Modes • Modeling Mixed Lubrication in Sheet Metal Forming • Calculation of the Stribeck Curve
28.4 Wear . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
28-22
Volumetric Wear of Forming Tools • Galling
Emile van der Heide TNO Industrial Technology
Dirk Jan Schipper University of Twente, Department of Mechanical Engineering Surface Technology and Tribology
28.5 Simulation of Tribological Contact Situations in Sheet Metal Forming . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
28-29
General Framework of Demands for Tribo Tests • Conventional Tribo Tests and Sheet Metal Forming • Simulation of Friction and Wear in Sheet Metal Forming
Acknowledgments. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
28-33 28-33
List of Symbols a, b B C d E Esheet Etool
geometrical features shown in Figure 28.6 contact length constant from the Nadai plastic deformation relation punch displacement reduced elastic modulus elastic modulus of the sheet elastic modulus of the tool
[m] [m] [Pa] [m] [Pa] [Pa] [Pa]
28-1
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28-2
FC Ff FH FN FT f fC H Hc HEI HEP HRI HRP Hsoft H , L, M h Kw k L M n pa pC pm pmax R R∗ Ra r s Tcr Tf V v v+ v dif α β γ˙ η0 η νtool νsheet σs τ τC τH τ0 φ ψ
Handbook of Lubrication and Tribology
load carried by the asperities friction force load carried by the hydrodynamic component normal force total normal load acting on a contact coefficient of friction coefficient of friction of a single asperity hardness dimensionless central separation elastic/isoviscous asymptote elastic/piezoviscous asymptote rigid/isoviscous asymptote rigid/piezoviscous asymptote hardness of the softest contact partner separation-, lubricant-, and load-number Equations (28.24) and (28.25) separation between two surfaces, film thickness coefficient of wear specific wear rate lubrication number bending moment density of the asperities load carried by the asperities asperity pressure mean contact pressure maximum contact pressure die radius, see Figure 28.4 combined radius (Equation [28.12]) center line average roughness punch radius, see Figure 28.4 sliding distance critical temperature flash temperature wear volume (sliding) velocity sum velocity sliding velocity viscosity–pressure coefficient (Barus) average summit radius shear rate viscosity at ambient pressure viscosity Poisson ratio of the tool Poisson ratio of the sheet standard deviation surface heights shear stress shear stress asperity contact shear stress hydrodynamic component Eyring shear stress probability height distribution of the asperity summits plasticity index
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[N] [N] [N] [N] [N] [-] [-] [Pa] [-] [-] [-] [-] [-] [Pa] [-] [m] [-] [m3 (N−1 m)−1 ] [-] [Nm] [m−2 ] [N] [Pa] [Pa] [Pa] [m] [m] [m] [m] [m] [◦ C] [◦ C] [m3 ] [m/sec] [m/sec] [m/sec] [Pa−1 ] [m] [sec−1 ] [Pa·sec] [Pa·sec] [-] [-] [m] [Pa] [Pa] [Pa] [Pa] [-] [-]
Friction and Wear in Lubricated SMF Processes
28-3
Abbreviations BA BL EBT EDT EG (E)HL EP FEM GI GA HSLA ML SEM SMF
Bright annealed Boundary lubrication Electron beam texturing Electro discharge texturing Electro galvanizing (Elasto)hydrodynamic lubrication Extreme pressure Finite element method Hot dip galvanizing Galvanealled High strength low alloy Mixed lubrication Scanning electron microscope Sheet metal forming
28.1 Introduction Metal forming processes are mechanical processes, designed to make products from metal sheet or billets without removal of the material. It is one of the major technologies to manufacture metal products such as kitchen sinks (deep drawing), profiles (extrusion), tubes (roll bending), or panel sides for automotive applications (drawing/stretching). Forming technology consists of a broad spectrum of processes, which can be classified roughly in two groups [1]: • Bulk or massive forming, where the work piece material undergoes large plastic deformation, resulting in an appreciable change in shape or cross section, usually without significant elastic recovery after deformation. Frequently used massive forming processes are extrusion, rolling, and forging. • Sheet metal forming (SMF), where sheet material is plastically deformed in two dimensions (the thickness of the sheet is more or less constant) with the possibility of significant elastic recovery or spring back. Figure 28.1 schematically shows the set up of three commonly applied SMF processes: deep drawing, air bending, and stretching. Each metal forming process has its own characteristic features in terms of tooling and material flow. A comprehensive review can be found in Reference 2. Common features of SMF processes, the focus of this chapter, are the use of initially flat sheet material cut into an optimized shape (the blank), a punch to transmit the energy needed for the mechanical work, a die that directs the material flow during the process, a blank holder, which controls undesired material flow and wrinkles, and draw beads that are used to restrict material flow. SMF processes are suited for
4
4
4
5 2
3
3 2
2 1
1
Deep drawing
Air bending
1
Stretching
FIGURE 28.1 Examples of sheet metal forming processes. Common features are (1) die, (2) blank, (3) blank holder, (4) punch, and (5) draw bead.
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28-4
Handbook of Lubrication and Tribology TABLE 28.1 Indication of Maximum Performance of Pneumatic, Mechanical, and Hydraulic Presses Type Force (kN) Strokes/minute Stroke length (mm) Stiffness in stroke direction
Pneumatic
Mechanical
Hydraulic
500 30 400 Poor
30.000 1200–2000 400 High
200.000 100 1200 Moderate
Source: Adapted from Bolt, P.J. et al., Materialen–vormgeven van dunne metaalplaat, Vereniging FME–CWM, Zoetemeer, The Netherlands, 1996.
mass production applications and the products are processed, fully automated, in large volumes with the help of industrial presses, see Table 28.1. Tribology — the science and technology of interacting surfaces in relative motion — plays an important role in the control of material flow during the process and in the control of the surface quality of the formed products. Poor lubrication conditions, for example, can easily result in local high friction, which in turn limits the desired material flow. In the worst case, fracture of formed products occurs, see Figure 28.2. Wear, the second tribological process in the work piece–tool interface, has a direct negative influence on the product’s dimensions and in the end will cause secondary wrinkles (pucker) to occur. Galling, a wear type that could occur during forming of metals, for example, stainless steel or aluminum (see Section 28.3.2), results in damaged tooling and in severely scratched products. The main objective of this chapter is to provide an introduction to the field of tribology in SMF. First, the relevant tribological system is described in Section 28.2. Details are given about the main system components showing the wealth of opportunities that exist in tool material and lubricant selection. Section 28.3 describes how friction forces in SMF processes can be predicted based on contact modeling in combination with hydrodynamic lubrication theory. Roughness is approached in a stochastic or deterministic way. Wear, and especially galling, is described in more detail in Section 28.4. Section 28.5 gives an introduction to wear and friction testing for SMF applications.
28.2 Tribological System 28.2.1 System Analysis Since friction and wear are system properties rather than material properties, it is of high importance to start with a system analysis [3,4]. The first step in such an analysis is the separation of a tribological contact situation by using a hypothetical system envelope. The contact situation separated by this envelope is regarded as a system, that is, a set of elements, interconnected by structure and function. A general presentation of tribological systems in (sheet) metal forming processes is depicted in Figure 28.3. The structure of the system consists of the interaction between the forming tool surface and the work piece material surface in the presence of a lubricant and surrounded by the environment. Tribological contacts in metal forming applications differ strongly from classical applications such as roller bearings, bushings, and gears, that is, with respect to the type of motion and the type of deformation. First of all, bulk plastic deformation of one of the contacting surfaces occurs: a state that is usually avoided in classical applications. In SMF processes only the tool material deforms elastically and is subjected to run-in. Second, the processed sheet material is always “fresh” in the tribo contact, since each pressing is done with a new blank. Therefore, running-in will occur only for one of the contacting surfaces, that is, the tool surface. The third common feature of tribological systems for forming processes is related to the microgeometry of the contacting surfaces. Figure 28.4 visualizes the roughness of a tool surface and a sheet surface, just before contact. It shows the interaction between a relatively smooth tool surface, with a relatively rough sheet surface. Furthermore, the sheet surface is relatively soft compared to the tool
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Friction and Wear in Lubricated SMF Processes
28-5
FRICTION Failure ty pe poor lubrication
WEAR die clearance to narrow due to galling
worn draw bead
die die radius clearance to large to large
Crack cylindrical part
Sec. wrinkle/pucker
Crack near bottom
Drawing marks
Asymmetrical flange
FIGURE 28.2 Possible failures as a result of poor control of friction and wear during axisymmetrix deep drawing. Adapted from Bolt, P.J. et al., Materialen — vormgeven van dunne metaalplaat, Vereniging FME–CWM, Zoetemeer, The Netherlands, 1996.
surface. Typically, the difference in hardness between the sheet material and the tool material is a factor of seven. The contact between two rough elastic surfaces can be simplified to the contact between a rough deforming surface with reduced elastic modulus E (Equation [28.1]) and a smooth rigid surface, see for example, Reference 5. This approach suits SMF-contacts well, considering the differences in roughness and hardness, between sheet and tool surfaces. 2 2 1 − νtool 1 − νsheet 2 = + E Etool Esheet
(28.1)
The surface asperities can deform elastically, plastically, and mixed elastic–plastic. The deformation mode can be estimated using a plasticity index, for example, the one developed by Greenwood and
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28-6
Handbook of Lubrication and Tribology
Operating variables
Structure of the tribo system
Forming tool surface
Lubricant Work piece surface
Environment
Changes in contact topography or surface composition loss
Mass loss wear data
Wear and friction characteristics
FIGURE 28.3
Tribological system in metal forming.
1.2 Tool surface 0.8
s [mm]
0.4 0.0 –0.4 –0.8 –1.2 0.00
Sheet surface
0.05
0.10
0.15
0.20
x [mm]
FIGURE 28.4 Roughness of a mirror finished tool surface compared to an AISI 304 2B sheet surface at the same magnification. Adapted from van der Heide, E., Lubricant failure in sheet metal forming processes, PhD Thesis, University of Twente, Enschede, The Netherlands, 2002.
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Friction and Wear in Lubricated SMF Processes TABLE 28.2
28-7
Plasticity Index for Two Sheet Materials σz [µm]
βm [µm]
H [GPa]
ψ[-]
2.5 0.50
5 80
0.9 1.8
1 >1
DC 06 AISI 304
Source: Roughness data extracted from de Rooij, M.B., Tribological aspects of unlubricated deepdrawing processes, PhD Thesis, University of Twente, 1998.
Williamson [6], see Equation (28.2). Roughness is represented in this equation by its probability asperity height distribution, with standard deviation σs . Furthermore, it is assumed that summits are equally shaped, that is: spherically with radius β. As a first approximation, it may be assumed that elastic contact occurs when ψ < 0.6 and plastic contact when ψ > 1. The intermediate phase between 0.6 and 1 corresponds with the mixed elastic–plastic regime. ψ=
E · 2 · Hsoft
σs β
(28.2)
Applying Equation (28.2) and using typical sheet material values from Table 28.2, it is found that the quotient of E over H is about 64 for stainless steel sheet and 128 for deep draw steel DC 06. Since the square root is at least about 0.05 (for stainless steel), plastic contact is likely to occur at the sheet asperity level. As run-in does not occur for the sheet material, a transition to an elastic regime is not likely to occur. The environment of the tribological contact is typically humid air at room temperature, although exceptions exist, that is, in hydro-forming and hot metal gas-forming applications.
28.2.2 Operational Variables The operating variables for tribological contacts in metal forming depend heavily on the actual application. Furthermore, differences exist between locations at the tool surface. The normal pressure at the blank holder–work piece interface, for example, is much lower than the normal pressure at the die radius, for a deep drawing setup, see Figure 28.5. A general overview of possible contact conditions in metal forming applications is given in Table 28.3. The presented values for the normal pressure can be much higher when considering critical contact situations at drawing radii or during forming of high(er) strength steels. As an example, the sliding velocity and normal pressure for a simple air bending set up are calculated. Tooling will increase in temperature from room temperature to about 40 to 70◦ C, as a result of the forming action. In some cases, forming is done at elevated temperature to make use of specific forming behavior of sheet materials. Indicative temperatures for warm forming and hot forming are 300–600◦ C and 800–1000◦ C, respectively. Let us take the example of air bending. Two tribological contact situations can be identified during the air bending process: the contact between the punch nose and the sheet and the contact between the sheet material and the die shoulder radius. The latter contact is of high importance since failure will induce fluctuating friction forces, which in turn will influence the amount of spring back that is generated after lifting the punch. Furthermore, galling can be initiated in this sliding contact. The operational conditions for the die shoulder–sheet contact has been estimated by ter Haar [7]. First, a time dependent function s(t ) based on the geometry of the process, see Figure 28.6, is derived, which describes the position of the contact on the sheet, relative to the original position for punch displacement, d = 0. s(t ) = lPD (t ) + r · α(t ) + b − a
© 2006 by Taylor & Francis Group, LLC
(28.3)
28-8
Handbook of Lubrication and Tribology F
FN
FN FN F FZ
F FZ FZ
FZ
Blank holder
Die
Forming part
FIGURE 28.5 Tribological contacts for a deep drawing process. Adapted from Netsch, T., Methode zur Emittlung von Reibmodellen fur die Blechumformung, PhD Thesis, Darmstadt University of Technology, Germany, Shaker-Verlag Aachen, Band 41, 1998.
TABLE 28.3
Operational Conditions in Metal Forming Processes Processes
Conditions Contact pressure ratio, p/Y a Contact pressure, p [MPa] Velocity, v [m/sec]
Sheet forming
Drawing ironing
Rolling rotary forming
Forging extrusion
0.1–1 1–100 10−3 –10−1
1–3 100–1000 10−2 –102
1–3 100–1000 10−2 –102
2–5 100–3000 10−3 –10−1
a Y : yield stress of the sheet material.
Source: Adapted from Wang, Z.G., J. Mater. Process. Technol., 151, 223–227, 2004.
with lPD (t ) =
(a − b)2 + (d(t ) − r − R)2 − (r + R)2
α(t ) = arc sin α(t ) = arc sin
r +R l ∗ (t ) r +R l ∗ (t )
l ∗ (t ) =
© 2006 by Taylor & Francis Group, LLC
− arc cos
+ arc cos
a−b l ∗ (t ) a−b l ∗ (t )
(28.4)
r + R > d(t )
(28.5) r + R ≤ d(t )
(a − b)2 + (d(t ) − r − R)2
(28.6)
Friction and Wear in Lubricated SMF Processes
28-9
d
Sheet
Punch a
Punch
r
Sheet
Lpd 2b
a
R
Die
Die
FIGURE 28.6 Two stages of air bending. Adapted from: ter Haar, R., Friction in sheet metal forming: the influence of (local) contact conditions and deformation, PhD Thesis, University of Twente, 1996.
The time derivative of s(t ) is the velocity of the displacement of the contact point, v(t ): v(t ) =
d [s(t )] dt
(28.7)
The sliding velocity can now be calculated by assuming a certain function for the punch displacement with time. The bending moment per unit width, M , needed for bending the sheet around the punch nose is estimated by ter Haar [7] based on the work of de Vin [8]: 2 M = r 2E 3
E C∗
3/(n+1)
(St /2)n+2 − r n+2 (E /C ∗ )(n+2)/(n−1) (n + 2)r n
+ 2C
∗
C∗ =
(n+1)/2 4 ·C 3
(28.8)
with (28.9)
This bending moment has to be supplied by the force FN at the die shoulder radius at a distance lPD . Hence: FN (t ) =
B·M lPD (t )
(28.10)
For reasons of simplicity, it is assumed that the contact between the die shoulder and the sheet material remains elastic during the forming process. The maximum contact pressure is found by applying the Hertzian theory for a strip contact: pmax =
FN /B · E π · R∗
(28.11)
Here, B represents the length of the line contact, E the combined modulus of elasticity (Equation [28.1]), and R ∗ the combined radius (Equation [28.13]). R∗ =
© 2006 by Taylor & Francis Group, LLC
R1 R2 R1 + R 2
(28.12)
28-10
Handbook of Lubrication and Tribology TABLE 28.4 Input for Calculation of Contact Pressure and Sliding Velocities During Air Bending Parameter
Value
Unit
Parameter
Value
Unit
dmax B a b r R tstroke
7.5 40 9.67 0.40 1.67 2.00 1
mm mm mm mm mm mm sec
st E1 = E2 υ1 = υ2 C n η —
1 2.1 × 1011 0.33 500 × 106 0.204 0.2 —
mm N/m2 — N/m2 — Nsec/m2 —
Source: Adapted from ter Haar, R., Friction in sheet metal forming: the influence of (local) contact conditions and deformation, PhD Thesis, University of Twente, 1996.
5
500 495
4
490
pmax [MPa]
480
2
475 1
470 465
v [mm/sec]
3
485
0
460
pmax [MPa]
455
v [mm/sec]
–1 –2
450 0
0.1
0.2
0.3
0.4
0.5 t [sec]
0.6
0.7
0.8
0.9
1
FIGURE 28.7 Calculated maximum Hertzian contact pressure and the sliding velocity for the die shoulder–sheet contact of the air bending process described Table 28.3.
The calculated values for the maximum contact pressure and sliding velocity as a function of time, for the geometry listed in Table 28.4, are shown in Figure 28.7. Applying the rule of the thumb pmax < 0.6 H with H = 900 GPa, Table 28.2 shows that the assumption of a Hertzian contact is valid for this particular configuration. In general, however, a finite element method calculation is needed to calculate the plastic component of the deformation.
28.2.3 Sheet Materials Sheet metal forming processes are used to produce a variety of products from automotive and aerospace applications to household and kitchen applications. Each product has to meet a specific set of demands with respect to, for example, corrosion resistance, toughness, strength, and visual appearance. Therefore, a range of sheet materials exists that varies in composition, forming behavior and surface quality. A summary of relevant properties for four groups of sheet materials, that is, low carbon steel, stainless steel, aluminum alloys, and copper alloys, is given in Reference 9, see Table 28.5. Low carbon steels represent the largest fraction of sheet material used nowadays in the forming industry. Automotive panels, for example, are typically made of 0.8 mm cold rolled deep drawing steel DC 06, alloyed low carbon steel with high formability. The nominal composition of this sheet material
© 2006 by Taylor & Francis Group, LLC
Friction and Wear in Lubricated SMF Processes TABLE 28.5
28-11
Survey of Representative Sheet Materials
Material
Typical example
Yield stress [MPa]
Tensile strength [MPa]
Brinell hardness [HB]
180 500–620
270–350 570–710
80 —
Carbon steel
DC 06 HSLA Docol 500 YP
Stainless steel
AISI 304/X5 CrNi 18 10 AISI 316/X5 CrNiMo 17 12 2 AISI 409/X2CrTi12 AISI 430/X6Cr17
210 220 220 260
520–720 520–670 380–560 430–630
150–190 150–190 150–180 150–190
Al-alloys
AA 1050A/Al99.5 (O) AA3103/AlMn1 (O) AA 5086/AlMg4 (O) AA 2024/AlCu4Mg1 (T4)
20 30 100 275
65 90 240 425
20 27 65 120
Cu and Cu-alloys
SW-Cu (F22) CuZn 30 (F27) CuSn8 (F33)
max. 140 max. 160 max. 190
220–260 270–350 330–380
40–65 55–85 65–95
Source: Adapted from Bolt, P.J. et al., Materialen — vormgeven van dunne metaalplaat, Vereniging FME–CWM, Zoetemeer, The Netherlands, 1996.
TABLE 28.6 Characteristic 2D Roughness Values, Bulk Hardness, Zinc Layer Thickness, and Hardness, Measured for Galvanized Steel
GI EDT EG GA
Ra [µm]
Rz [µm]
Rq [µm]
Rp [µm]
Rt [µm]
Bulk [HV0.2 ]
Layer [µm]
Layer hardness
0.9 1.2 0.9
4.3 7.0 7.1
1.0 1.5 1.2
1.9 3.6 2.5
5.0 10.1 8.1
81 77 73
14 8 10
46–57 HV0.005 32–39 HV0.001 236–287 HV0.005
is ≤0.02 wt% C, ≤0.25 wt% Mn, 0.020 wt% P, 0.02 wt% S, and ≤0.30 wt% Ti (EN 10130). The hardness of this sheet material is typically 70–80 HV, measured at the cross-section of the sheet material. For outer parts, this material is zinc coated to ensure corrosion protection. Zinc layers are deposited by electro galvanizing (EG), hot dip galvanizing (GI), or hot dip galvanizing with additional heat treatment of the surface (galvanealled, GA). The resulting layers range in thickness from 5 to 20 µm and in hardness from about 60 to 350 HV. Furthermore, pretextured rolls are used during the final cold rolling steps in the steel plant to create specific roughness textures such as electro discharge texturing (EDT) and electron beam texturing (EBT) at the sheet surface. Characteristic values for 2D roughness, hardness, and layer thickness are given in Table 28.6. Scanning electron microscope (SEM) images and 3D roughness measurements clearly show the surface quality and texture, see respectively, Figure 28.8 and Figure 28.9 for DC 06 GI EDT. The increasing demand for light and strong structures has led to the development of new types of carbon steel sheet materials, with high(er) strength, allowing the application of thinner cross-sections in mechanical designs. Typical yield strength of these materials ranges from 500 to 700 MPa, see for example the quality HSLA in Table 28.5. This development will further broaden the spectrum of varieties in sheet surfaces. Stainless steel is another important group of sheet materials especially for domestic and household appliances such as kitchen sinks or dish washers. The chemical composition that is used strongly depends on the application but consists typically of ferritic grades like AISI 430, or austenitic grades like AISI 304. The main difference between these grades is the alloying element Ni, see Table 28.7. Appearance is crucial in marketing of household equipment. Hence, several finishes are developed by the steel manufacturers. Commonly used finishes include bright annealed (denoted as BA), annealed pickled and skin-passed (denoted as 2B), and electro discharge textured (EDT). Typical 2D roughness data measured parallel to
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28-12
Handbook of Lubrication and Tribology
FIGURE 28.8
3D roughness measurements of galvanized steel quality GI EDT.
FIGURE 28.9
SEM image of galvanized steel quality GI EDT. TABLE 28.7 Nominal Composition of Frequently Used Stainless Steel Grades in wt% Type
AISI
C
Si
Mn
Cr
Ni
Ferritic Austenitic
430 304
0.05 0.04
0.35 0.5
0.4 1.1
16.5 18.2
— 8.7
the sheet’s rolling direction is given in Table 28.8, together with an indicative value for the hardness of these materials. Sheet metal forming of stainless steel is very sensitive to galling. This is partly caused by the relatively high hardness of the material and partly because of the poor thermal conductivity of the material. Both aspects result in high local temperature rises due to frictional heating, see Section 28.4.2. Furthermore, stainless steel can have a high surface roughness, which is strongly related to wear of forming tools, especially when those tools are made of rather soft aluminum bronze, see Section 28.4.1. As the selection of the sheet material is based on the application of the formed product, it is extremely difficult to change the selected choice based on tribological considerations. As such,
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Friction and Wear in Lubricated SMF Processes
28-13
TABLE 28.8
Characteristic Roughness (µm) and Hardness Values (HV)
Grade
Finish
Ra
Rq
Ry
Rz
Rt
Rp
HV1
AISI 304 AISI 430 AISI 430
2B BA EDT
0.28 0.04 2.33
0.39 0.05 3.01
2.99 0.44 17.56
2.46 0.31 13.65
3.07 0.45 18.42
0.61 0.15 5.25
172 157 165
sheet material usually represents a fixed boundary condition in tribological optimizations of SMF processes.
28.2.4 Tool Materials Current fast changes in product design have created a need for materials designed specifically for low or medium volume series. These tool materials are usually referred to as low-cost materials or soft tool materials and are easy to process into complex shapes, saving both developing time and costs of tooling for press operations. Typical examples that have strong links with rapid prototyping techniques are epoxy resins and aluminum. Furthermore, zinc alloys (e.g., ZnAl4 Cu3 ) and resin-impregnated laminated and condensed wood can be considered. The latter material is used successfully in low-volume aerospace applications. For more demanding loading conditions, typically cast iron–carbon alloys are selected. Gray cast iron GG 25 can still be considered a low-cost material. High-volume applications that exist in mass production of large automotive panels require at least ductile cast iron grade GGG 60, but preferably GGG 70L. Mechanical strength can also be created by alloying gray cast iron with chromium, molybdenum or vanadium, see Table 28.9. For smaller tools, one could also select cast steel instead of cast iron. Cast steel has a maximum carbon content of 0.45% and small amounts of alloying elements such as manganese, molybdenum, and chromium. Frequently selected grades include 16 MnCr5 and 42CrMo4. A further increase at scale of loading conditions requires the use of wrought iron carbon alloys, usually referred to as tool steel. These steels combine the advantages of “easy” machining in the soft annealed state with high wear resistance in the hardened and tempered state. Conventional tool steel, for example, X155CrMoV 12 1/WN 1.2379 in Table 28.9, consists of 1.5 wt% C and has chromium and vanadium as main alloying elements, respectively, about 12 and 1 wt%. By changing the carbide type, size, and amount in tool steels it is possible to optimize the balance between ductility and wear resistance. Carbide volumes above 25% require a manufacturing process that is based on powder metallurgy. Typical examples of these powder metallurgical steels are high tungsten and vanadium alloyed tool steels WN 1.3344 and Vanadis 10 [10]. Powder metallurgy is also used to produce a class of cemented carbides, which is frequently denoted by the industry as hard metals, consisting of tungsten carbide bonded together by cobalt. The grain size of the tungsten carbides and the cobalt–carbide content again determines the balance between ductility, needed for impact loading, and hardness, needed for wear resistance. Full ceramics, applied for decades in conventional tribological systems, where lubrication is restricted or where extreme wear resistance is required, can also be applied in SMF, for example, as die inserts. Especially Si3 N4 [11] and SiC [12] are promising materials for highly demanding forming applications, although the poor impact strength of the materials limits its application. The range of possibilities is further enlarged by the advanced state-of-the-art coating technology. Thin hard layers can be applied on sheet and tool surfaces to meet specific demands regarding wear and corrosion resistance in a cost-effective way. A comprehensive summary of general possibilities is given by Bushan and Gupta [13], an overview of results for metal forming applications is given in Reference 9. Proven technology, extremely important in the metal forming industry, consists of heat treatment by flame or induction hardening, nitriding, and hard chromium plating. Alternative “new” technologies include single layers created by physical or chemical vapor deposition (e.g., CrN or TiN) or advanced multilayer systems (e.g., laser hardening with additional multilayer based on carbon). Chemical diffusion technologies as chromizing and vanadizing are also able to create hard, wear resistant tool surfaces.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
TABLE 28.9
Tool Materials
Examples Soft tool materials
Cast iron
Aluminum bronze
Cast steel Tool steel
Cemented carbides Ceramics
Epoxy resin Laminated wood ZnAl4Cu3 GG 25 GGG 60 GG 25 CrMoV GGG 70L CuAlFe, for example, AMPCO 25 16MnCr5 42CrMo4 WN 1.2363 WN 1.2379 WN 1.3344 Vanadis 10 94% WC–6% Co 80%WC–20% Co SiC Si3 N4
Indicative surface hardness in service 20–45 HB
Application area Low volume forming, prototype series
150–300 HB
Large tools, for example, in automotive applications
160–360 HB
Tools for deep drawing, air bending, etc. with low–medium loading conditions. Tools for stainless steel forming Tools for deep drawing, air bending, etc. with demanding loading conditions Tools for deep drawing, air bending, etc. with demanding loading conditions
600–700 HV 700–840 HV
1300–1900 HV 1500–2500 HV
Roll forming tools, tools for very demanding conditions Inserts, dry forming tools
Source: Adapted from: Bolt, P.J. et al., Materialen — vormgeven van dunne metaalplaat, Vereniging FME–CWM, Zoetemeer, The Netherlands, 1996.
28.2.5 Lubrication in Sheet Metal Forming Forming lubricants are widely used in metal forming processes in order to control friction, enhance cooling, and prevent wear. The type of lubricant varies and includes: • • • •
Greases Forming oils Oil in water emulsions Protective tapes
The lubricant type and the required functionality of the lubricant is selected as a function of the forming process, see Table 28.10. The use of greases or pastes in metal forming is generally limited to low-volume applications such as the forming of panels with wooden tools or to applications at elevated temperature such as warm forming of aluminum. Typically, grease or paste is applied manually to the tooling with a brush after a series of, for example, ten products, see also Table 28.11. General multipurpose greases or pastes based on synthetic esters can be selected as forming lubricant. In many cases, it is difficult to remove the grease layers from products. An intermediate stage between greases/pastes and liquid forming oils are the so-called dry films or dry waxes. These films can be sprayed electrostatically to sheets and are usually formulated to be soluble or mixable with water so that the film can be rinsed off with water after it forms. Forming oils generally consist of a mixture of lubricant-based fluids of mineral, synthetic, or vegetable origin and additives for specific demands of metal forming applications. Additives can include, for example, oxidation inhibitors, antirust additives, foam inhibitors, viscosity modifiers, and biocides to control bacterial growth. A variety of chemical compounds is added to forming oils to meet the demanding requirements with respect to friction and wear control. These additives will create thin protective boundary layers on the sheet and tool material. The mechanism of boundary layer formation [14] can be used to classify the resulting layers, see Table 28.12. The viscosity of forming oils can be altered by viscosity
© 2006 by Taylor & Francis Group, LLC
Friction and Wear in Lubricated SMF Processes TABLE 28.10
28-15
Lubrication in Sheet Metal Forming Required functionality
SMF process Deep drawing Punch Die/blank holder Stretch forming Punch Die/blank holder Air bending Die Folding Punching Ironing
Control friction
Cooling
Prevent wear
Type of lubrication
−/ +
− −/
−/ +
Emulsion, oil, grease; in case of aluminum and stainless steel also tape
+ +
− −/
−/ +
Emulsion, oil; in case of aluminum and stainless steel also tape
−/ −/ −/ −/
− − + +
+ +
Thin oils; in case of aluminum and stainless steel also tape Emulsion, oil Emulsion
Source: Adapted from Bolt, P.J. et al., Materialen — vormgeven van dunne metaalplaat, Vereniging FME–CWM, Zoetemeer, The Netherlands, 1996. +: important, : less important, −: not important.
TABLE 28.11 Application Methods for Forming Lubricants Water mixable
Not mixable with water
Thin emulsion
Viscous emulsion
Soap solution
Oil 70 cSt
Oil 70–350 cSt
Grease
Paste
Tape
Manual Brush/rolling Spraying Immersing Sticking
+ + + −
+ −
+ + −
+ + −
+ + − −
+ − − −
+ − − −
− − − +
Mechanical Rolling Dripping Spraying Immersing Adhesive
+ + + + −
+ −
+ −
+ + + −
+ − + − −
− − − −
− − − −
− − − − +
Lubricant application method
Source: Adapted from Bolt, P.J. et al., Materialen — vormgeven van dunne metaalplaat, Vereniging FME–CWM, Zoetemeer, The Netherlands, 1996.
modifiers such as polymers and copolymers of olefins, methacrylates, dienes, or alkylated styrenes. The later polymers expand with increasing temperature to counteract oil thinning [15]. Heavy duty forming oils typically have a viscosity of about 300 cSt at 40◦ C. Mild forming operations of, for example, cold rolled steel can be done with forming oils of about 50 cSt or less. Some forming oils contain emulsifiers, sometimes in the form of self-emulsifying esters. Such oils can either be used directly on the sheet for severe forming applications or can be applied in the form of a oil-in-water emulsion in case of mild operations or in case cooling is required during the forming action. Finally, forming oils can be designed as vanishing oils. The base oil/solvent evaporates to the environment at a rate that depends upon the temperature, air circulation, and humidity, leaving a thin lubricant film that mainly consists of boundary layer forming additives. Vanishing oils have the advantage of reduced cleaning costs after forming. Yet, extra attention should be paid to ventilation and the risk of open flames or other sources of ignition. The boundary lubrication action of emulsions and vanishing oils is based on the same additive classes as listed in Table 28.12. Protective tapes can be applied to metal sheets or coils in order to prevent scratching during transport and storage. This kind of packaging is frequently used for aluminum and stainless steel sheet materials.
© 2006 by Taylor & Francis Group, LLC
28-16 TABLE 28.12
Handbook of Lubrication and Tribology Classification of Lubricant Boundary Layers
Boundary layer formation mechanism
Typical chemical compounds
Comments
Physical adsorption
Long chained alcohols and fatty acids
Chemical adsorption
Fatty acids, long chained fatty amides, and esters
Chemical reaction
Complex compounds based on P, S, B, and Cl
Blank feeder
Cleaning
Boundary layers consist of clustered, long chained hydrocarbons with a polar “head.” The polar group adheres to the contacting surface, by physical adsorption, forming high-viscosity hydrocarbon layers. This group of layers is meant to reduce friction and wear under mild loading conditions Boundary layers are formed by a combination of physical adsorption with chemical reaction with the surfaces, to form a metal soap. This group of layers is more resistant to increased contact temperatures, and therefore used for wear and friction reduction at moderate loading conditions Additives react with the surface to metal salts, with high temperature stability. These layers are suited for wear protection at severe loading conditions where extreme pressure (EP — additives), causes high contact temperatures. Some complex compounds are suspected to harm the environment and — without protective measures — human health
Forming lubricant
To press
FIGURE 28.10 Schematic drawing of blank cleaning and lubrication setup. Adapted from Bolt, P.J. et al., Materialen — vormgeven van dunne metaalplaat, Vereniging FME–CWM, Zoetemeer, The Netherlands, 1996.
In some cases, the protective tape is also used to prevent scratching during production and assembly of parts. The operation to remove the plastic from the sheet surface is usually performed manually, although mechanical solutions are available for specific applications. Tapes are available in a thickness range of 0.03–0.13 mm, and consist of (coextruded) polyethylene or polyvinyl chloride with an acrylic adhesive. Lubricant selection for sheet metal forming processes is a complex matter, since it involves many contradictory demands. Lubricants should, for example, adhere firmly to the sheet surface not only to maintain constant friction and constant wear prevention during the forming action but they should also be easily removable from the surface later on to avoid compatibility problems with subsequent welding or coating steps. Compatibility problems can also rise from the use of mill-applied preservation agents to protect the steel from corrosion during transport of the coils and during storage at the press shop. If such a compatibility problem exists it is necessary to use a setup as depicted in Figure 28.10. The forming lubricant is now applied after removing the preservation agent and before the actual forming operation. Such a cleaning step can be avoided in case no compatibility problems exist with the preservation agent. In that case, spot lubrication or cleaning by using washing oil is an option. Still, cleaning of the formed product is
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Friction and Wear in Lubricated SMF Processes
28-17
necessary, before a decorative coating can be applied. Cleaning becomes more and more important since many extreme pressure (EP) doped forming oils require the use of environmentally unfriendly solvents to remove boundary layers from the formed product. As such, much effort is put in dry forming or water rinseable forming lubricants.
28.3 Friction 28.3.1 Lubrication Modes Friction in lubricated systems is traditionally explained by referring to the Stribeck curve [16] and the lubrication modes related to this curve. The different lubrication modes are, respectively: • (Elasto)hydrodynamic lubrication ([E]HL). No physical contact between the asperities of the interacting surfaces occurs; the velocity difference between the surfaces is accommodated by shear in the lubricant film, which results in relative low-friction forces and, consequently, a low coefficient of friction. Typical value for f is of the order 0.01. • Boundary lubrication (BL). In this mode there is a physical contact between the asperities of the interacting surfaces. The entire load is carried by the interacting asperities. Shearing of boundary layers accommodates the velocity difference. Typically, f is in the range of 0.1 < f < 0.3. • Mixed lubrication (ML). This mode represents the intermediate regime between BL and (E)HL where a part of the load is carried by hydrodynamic action and the remaining part of the load is carried by the interacting asperities. The coefficient of friction ranges from (E)HL to BL values.
Coefficient of friction, f [-]
0.14
Boundary lubrication
Mixed lubrication
Full film lubrication
10
0.12
8
f 0.10
6
0.08 0.06
4
0.04 0.02 0.00 0.001
2
h/Ra
0 0.01
0.1 Velocity [m/sec]
FIGURE 28.11 Generalized Stribeck curve (friction) and separation.
© 2006 by Taylor & Francis Group, LLC
1
10
Film thickness/surface roughness = h/Ra [-]
It is shown that the Stribeck approach is also applicable to lubricated SMF-contacts [7]. As such, the coefficient of friction f , defined as the ratio of the friction force and the applied normal force, can be plotted in a diagram as a function of the velocity or, for instance, as a function of the dimensionless lubrication number L = ηv/(pm Ra ), introduced by Schipper [16], see Figure 28.11. The lubrication number consists of the main operational variables of study in lubricated contacts: the lubricant inlet viscosity (η), velocity (v), mean contact pressure (pm ), and the center line average roughness (Ra ). For SMF-systems it can be shown by using the operational conditions present in SMF that BL and ML will be the operational lubrication modes [7].
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Handbook of Lubrication and Tribology
28.3.2 Modeling Mixed Lubrication in Sheet Metal Forming A friction model for the ML mode can be constructed assuming that the total normal load FT acting on a contact is shared by the hydrodynamic action and the interacting asperities of the surfaces: FT = FC + FH
(28.13)
with FC being the load carried by the asperities and FH the load carried by the hydrodynamic component. 28.3.2.1 Load Carried by the Asperities The model of Greenwood and Williamson [17] can be used to estimate the load carried by the asperities: pa (x) =
2 nβσs 3
σs E F3/2 β
h(x) σs
(28.14)
with h being the separation between two surfaces, n the density of the asperities, β the average radius of the asperities, and σs the standard deviation of the height distribution of the asperities. F (h) is defined as: ∞ (s − h)3/2 φ(s)d s
F3/2 (h) =
(28.15)
h
where φ(s) is the height distribution of the asperity summits. Roughness is treated as a stochastic parameter with a known probability density. In this case, a Gaussian height distribution of the asperities has been assumed: 1 2 φ(s) = √ e−(1/2)s 2π
(28.16)
The elasticity modulus used in Equation (28.14) is defined as: 1 − υ12 1 − υ22 2 = + E E1 E2
(28.17)
with Ei being the elasticity modulus of the surfaces and υi the Poisson’s ratio. The pressure distribution in rough concentrated contacts can be calculated on the basis of Equation (28.18), which is based on the work on deformation of rough line contacts presented in References 18 and 19: pC = [1 + (a1 n a2 σsa3 W a2 −a3 )a4 ]1/a4 pmax
(28.18)
The dimensionless numbers n , σs , and W , are given in Equations (28.19) and (28.20), pmax is the maximum Hertzian pressure. 32 ∗ ∗ nR βR π
(28.19)
σs =
π σs 8 R∗
(28.20)
W =
FT BE R ∗
(28.21)
n =
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Friction and Wear in Lubricated SMF Processes
28-19
The values of the fit parameters ai in Equation (28.18) are: a1 = 0.953, a2 = 0.0337, a3 = −0.442, and a4 = −1.70, respectively. 28.3.2.2 Elasto Hydrodynamic Lubrication Based on calculations, Moes [20] formulated a function fit for the central film thickness in a lubricated line contact: Hc =
7/3
7/3 (3/7)s
HRI + HEI
−1 −7/2 −7/2 (−2/7)s s + HRP + HEP
(28.22)
with s=
1 (7 + 8e−2(HEI /HRI ) ) 5
(28.23)
HRI , HEI , HRP , and HEP are the dimensionless asymptotes as described in the Moes-diagram (see Equation [28.24]): HRI = 3M −1 HRP = 1.287L 2/3 (28.24)
HEI = 2.621M −1/5 HEP = 1.311M −1/8 L −3/4 in which E R ∗ 1/2 η0 v + ∗ 1/2 FT ER M= ∗ BE R η0 v + ∗ −1/4 ER L = αE η0 v + h H= ∗ R
(28.25)
The total pressure in an ML contact can be split into an EHL component and a BL component [21]. Figure 28.12 schematically shows the pressure distribution pT in an ML contact. In order to be able to use the results of the EHL calculations and the results of the dry contact or BL calculations, Equations (28.18) and (28.22) have to be adapted. For that purpose, two coefficients γ1 and γ2 are defined: γ1 =
FT FH
γ2 =
FT FC
(28.26)
It has been shown in References 21 and 22 that the results of the EHL and BL calculations can be combined in the ML regime, if E is replaced by E /γi , FT by FT /γi , and n by n · γi . For the EHL component this results in: −7/2 7/3 −14/15 7/3 (3/7)s −7/2 (−2/7)s 1/S Hc = γ1s HRI + γ1 HEI + HRP + HEP
(28.27)
with −2/5
s = 15 (7 + 8e−2γ1
© 2006 by Taylor & Francis Group, LLC
(HEI /HRI )
)
(28.28)
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Handbook of Lubrication and Tribology
pT
pH
pC
–b
0
b x
FIGURE 28.12
Pressure distribution in a ML contact.
hs hs
dd Mean height of summits
FIGURE 28.13
Mean height of surface
The contact between a rough and a flat smooth surface is drawn schematically.
and for the BL component in: pC 1 a2 a3 = [1 + (a1 n σs W a2 −a3 γ2a2 )a4 ]1/a4 pmax γ2
(28.29)
As the load increases or the surface becomes smoother, pC approaches the maximum Hertzian pressure corresponding to the fraction of load carried by the asperities. From all the equations given above so far, the fractions of load of the BL component and the EHL component can be calculated. 28.3.2.3 Extension for Rough Surfaces The film thickness that is calculated with Equation (28.22) is the central film thickness of a smooth contact and is therefore, for highly loaded contacts, a measure of the volume of fluid that passes through the contact. For rough surfaces, like in SMF, a comparable measure could be found. In Figure 28.13, the contact between a rough and a flat smooth surface is drawn. In this figure, two distances between planes have been defined, that is, hs being the distance between the smooth surface and the mean plane of the summits and hs the distance between the smooth surface and the mean plane (center line) of the rough surface. The distance between these two mean planes is given by dd : hs = hs − dd
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(28.30)
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For the extension, use is made of the definition of the film thickness given by Johnson et al. [21] as the average fluid volume between the two rough surfaces divided by the area:
hs h=
(hs − z)φ(z)dz
(28.31)
−∞
where φ(z) is the distribution of surface heights. For situations with a large film thickness, the average film thickness is the same in both models. This interpretation of h improves the prediction of the mixed lubrication model. Basically, for a given load partition, γ1 and γ2 , the hydrodynamic load yields the central film thickness, Equation (28.22), from which the value of h = hs giving the asperity pressure to the central pressure given directly by the asperity pressure, Equation (28.14), can be found. Equating this pressure to the central pressure given directly by the asperity load, Equation (28.29) is the criterion used to provide the self-consistent load partitioning of the problem [23].
28.3.3 Calculation of the Stribeck Curve The total friction force Ff is the sum of the friction force between the interacting asperities and the shear forces of the hydrodynamic component:
Ff =
N
τCi dACi +
i=1 A Ci
τH dAH
(28.32)
AH
with N the number of asperities in contact ACi the area of contact of a single asperity, i; τCi the shear stress at the asperity contact i; AH the contact area of the hydrodynamic component; and τH the shear stress of the hydrodynamic component. The coefficient of friction fCi of a single asperity can be written as: fCi =
τCi pCi
(28.33)
with pCi the normal pressure of a single asperity. Briscoe et al. [24] showed that the ratio of the shear strength and the local contact pressure is nearly constant. Since the coefficient of friction is by approximation constant for all asperity contacts, the first term in Equation (28.32) can also be written as: N
τCi dACi = fC FC
(28.34)
i=1 A Ci
where FC is the total load carried by all asperities. The value of fC is determined from experiments. For the shear force in the film different models can be used, for example, the isothermal Eyring model:
ηγ˙ τH (γ˙ ) = τ0 · arc sinh τ0
(28.35)
where η is calculated according to the Roelands equation, the pressure used to calculate the viscosity is the average pressure of the hydrodynamic component, that is, the total load carried by the hydrodynamic component divided by the total hydrodynamic contact area. γ˙ is the shear rate (v dif /h) and τ0 is the Eyring shear stress. v dif is the velocity difference between the two surfaces.
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Input (FT, v, h, ...)
Assume FC Eq. 3.1 / Eq. 3.14
Eq. 3.14
FH g1
g2
Eq. 3.15
Eq. 3.17
HC
PC Eq. 3.19
hs’ Eq. 3.18 hs Eq. 3.2 Pa(0) Equal? No
Yes, Eq. 3.25 f
FIGURE 28.14
Schematic representation of Stribeck curve calculation.
The coefficient of friction can thus be written as: fC FC + Ff = f = FT
AH τH (γ˙ )dAH
FT
(28.36)
Combining this result with Equation (28.35) results in the coefficient of friction for Stribeck curves: f =
fC FC + τ0 AH arc sinh(ηv dif /hτ0 ) FT
(28.37)
The Stribeck curve for SMF-contacts can now be calculated as demonstrated in Figure 28.14 by varying the sliding velocity. In Table 28.13 the experimental conditions of measurements performed by ter Haar [7] are given. In this table the value for Ra and dd are taken from actual measurements. The experiments were performed to simulate friction in sheet metal forming and to establish the Stribeck curve for sheet metal forming. The results of the measurements of ter Haar are plotted in Figure 28.15, together with calculations with the present model. Two different lubricants are used for the experiments. The presented calculations are done with the thicker oil of the two, calculations with the thinner oil is only slightly different. The friction in the (E)HL regime is slightly lower for Lub1.
28.4 Wear Tool wear and galling are the two main wear types that limit tool life in sheet metal forming applications. Based on the classification scheme presented in [25] both types can be defined as special cases of the
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TABLE 28.13 Experimental Conditions for UCS1-Material Property
Value 7.7 109 m−2 4.21 µm 2.19 µm 30 mm 231 GPa 50 mm 1.125 µm 1.85 µm 0.6 Pa sec (Lub1) 1.2 Pa sec (Lub2) 3.3 10−8 Pa−1 0.68 196.2 MPa 2.5 MPa 0.13 0.25% 350 N 72.7 MPa
n β σs B E R dd Ra η0 η0 α z pr τ0 fC Sep FT pav
Source: Adapted from ter Haar, R., Friction in sheet metal forming: the influence of (local) contact conditions and deformation, PhD Thesis, University of Twente, 1996.
0.16
Coefficient of friction, f [-]
0.14 0.12 0.10 0.08 0.06 0.04 0.02 0.00 10–5
Lub 1 Lub 2 Present model
10–4
10–3
10–2
Lubrication number, L [-]
FIGURE 28.15 Measured and calculated Stribeck curve. Adapted from Gelinck, E.R.M., Mixed lubrication of line contacts, PhD. Thesis, University of Twente, Enschede, The Netherlands, 1999.
sliding wear process (see also Reference 9). The velocity difference between the sheet and the tool during a forming action should preferably be accommodated by shearing of an interfacial layer, for example, a lubricant film. However, in some cases either the tool surface or the sheet surface is damaged due to the sliding action. The first case causes volumetric wear of the tool. The dominant wear mechanism in the sense of Reference 25 is abrasive wear. Reshaping of the tool surface is necessary in order to avoid products with unacceptable dimensions. The second case is characterized by severe scratching of the sheet
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28-24
Handbook of Lubrication and Tribology 15000 Lifetime [amount of products]
R = 0.7 mm
12500
Lifetime criterion: radius increase of 1 mm
10000 7500
GI
EZ
GA
5000 2500 0 Epoxy resin
Laminated wood
ZnAl4Cu3
GG-30
FIGURE 28.16 The lifetime of a forming tool as a function of the tool material for different zinc coated steel sheets. The triangular shaped area indicated by the arrow is sensitive to wear. The amount of products needed for a 1 mm increase of the top radius is used as lifetime criterion. Adapted from van der Heide, E. et al., Journal of Materials Processing Technology, 141/2, 197–201, 2003.
surface and is actually a form of volumetric wear of the sheet surface. This wear type is generally referred to as galling, and is related to pick-up of sheet material at the tool’s surface, a combination of abrasive and adhesive wear. Polishing of the tool surface is necessary in order to avoid an unacceptable surface condition of the product and extreme friction forces during forming.
28.4.1 Volumetric Wear of Forming Tools Volumetric wear of a forming tool is first visible at spots with high local pressure. Take for example, the tool given in Figure 28.16, which produces a box shaped product of 100 × 100 mm with a depth of 20 mm. The part sensitive to volumetric wear is indicated by the arrow. Results presented in Reference 26 show large differences in lifetime for this tool geometry as a function of the applied (soft) tool material and as a function of the processed sheet material. Clearly this points the system’s dependence of wear. A common approach in wear assessment is based on establishing a relation between volumetric wear and the operational conditions. It is shown, for example, by Reference 27 that an Archard type of wear equation [28], here presented in the nondimensional form Kw [29] by Equation (28.38) and in terms of the specific wear rate k by Equation (28.39), could be used for this purpose. Kw =
k=
V s · FN
V · Hsoft s · FN
mm3 (Nm)−1
(28.38)
(28.39)
in which Kw represents the coefficient of wear, V the wear volume, s the sliding distance, FN the applied normal force, and Hsoft the hardness of the softest contact partner. Achard’s wear equation has proven its value in comparative wear testing. But even more important, it can be used in the design environment based on the finite element method that is used extensively in sheet metal forming applications. One could, for example, build a routine that adjusts the geometry of the contacting tool surface based on the operational conditions that were applied on the surface during the forming step prior to the one of interest that is: the length of the sheet surface in contact with the element and the average normal force applied on the element during the forming action. The remaining task for a tribologist is to produce reliable data for the specific wear rate [26,30]. The simplest first guess is related to the hardness of the tool material: wear resistance increases with increasing hardness of the tool material. Equation (28.38) suggests the same: volume loss of the forming tool is inversely proportional to the hardness of the softest material. Abrasive wear tests like the rubber
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TABLE 28.14 Specific Wear Rate of Conventional Tool Materials Measured with Rubber Wheel Abrasion Test ASTM G65 at Fn = 100 N Using Quartz Sand as Abrasive (900 ± 50 HV, Grain Size 0.2–0.5 mm, Rounded Shape). WN 1.2379 is Hardened and Tempered to 60 HRc GGG 40 k [10−6 mm3 /Nm]
GGG 60
GGG 70L
WN 1.2379
1030
859
46
1357
Source: Adapted from van der Heide, E. et al., in Proc. NORDTRIB 04 Conference, Tønder, K. et al., Eds, Tromsø, Norway, 2004, pp. 355–364.
Volumetric wear [mm3]
0.50 0.40 0.30 0.20 0.10 0.00 0
20
40
60
80
100
120
140
Distance [m]
FIGURE 28.17 Rubber wheel abrasion test results, for hard chromium plated GGG 60. The specific wear rate increases significantly after the 50 µm Cr layer has worn off. TABLE 28.15
Slider-On-Sheet Test Results. FN = 100 N
Tool material Aluminum bronze Aluminum bronze GG-25 Aluminum bronze ZnAl4 Cu3 ZnAl4 Cu3 Epoxy resin
Sheet material
Lubricant
k-value 10−6 mm3 /Nm
AISI 430 BA AISI 430 BA DC 04 GA AISI 304 2B DC 04 GI DC 04 GA DC 04 GA
42 K 140-EL N 6130 140-EL N 6130 N 6130 N 6130
0.020 0.051 0.72 0.95 2.41 89.5 1450
wheel abrasion test or the Taber abrader test, confirm this, see for example Table 28.14, in which the specific wear rate is given for three commonly used cast irons and hardened tool steel WN 1.2379 (60 HRc). A further increase in hardness can be reached by the application of thin hard coatings like hard chromium plating or physical vapor deposition of CrN. The same rubber wheel test now shows a substantial decrease in worn volumes, until the layer has worn off, see Figure 28.17 for hard chromium plated GGG 60. Although the specific wear rates measured with the rubber wheel test or similar conventional wear tests can be used as a comparative measure of performance, it is clear that the calculated k-values cannot be used as input to FEM sheet metal forming applications. As such, dedicated test methods have been designed that take into account the tribological system that is used in lubricated sheet metal forming processes. Results with, for example, a slider-on-sheet configuration indicate the importance of measuring the wear response of materials in relation to the tribological system in which it is used. The specific wear rate can easily vary by a factor of 100 from one system to the other. Especially, the application of rough stainless steel can increase the measured wear rate substantially. For lubricated smooth stainless steel, specific wear rates in the order of 10−8 mm3 Nm−1 are found. For rough qualities this could easily be 10−6 mm3 Nm−1
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Handbook of Lubrication and Tribology TABLE 28.16 Process Characteristics Description
Value
Dimensions of the blank Dimensions of the die Die inner diameter Draw radius Punch force Blankholder force
Ø 440 × 2.1 mm Ø 440 × 80 mm Ø 257 mm 10 mm 800 kN 300 kN
Scratch depth [mm]
25 20 15 10 5 0 0
50
100
150
200
250
Scratch width [mm]
FIGURE 28.18
Depth and width of unacceptable scratches on an axisymmetric deep draw product.
depending on the lubricant used [30]. The same holds for the introduction of hard zinc layers like the galvanealled (GA) quality, see Table 28.15.
28.4.2 Galling The importance of control of galling can be illustrated by the production of 18 l expansion tanks for central heating systems. The upper and lower halves of the tank are made from 2.1 mm thick, cold rolled steel DC 03, by axisymetric deep drawing (see Table 28.16). A forming lubricant is applied on both sides of the blanks. The side in contact with the drawing die uses 24 kg of lubricant per 10,000 products. Cast steel 42CrMo4 , uncoated and not hardened, is used as tool material for the drawing die. During production, scratches appear on the products surface. The severity of scratching increases with the amount of products and reach an unacceptable level after 3,400 products, see Figure 28.18. Galling, in the context of metal forming, is associated with the tendency for lubricant film breakdown, resulting in pick-up of sheet material by the tool surface and subsequent scoring (severe scratching) of the work piece surface [31]. Scoring or severe scratching may, by definition, be due to local solid-phase welding or to abrasion. Galling mechanisms in SMF operations can be divided into three phases [32,33]: 1. Initiation 2. Lump growth 3. Severe scratching or seizure The initiation of material transfer occurs at tool surface defects like grinding marks or carbides. This can be understood by taking into account the contact situation given in Figure 28.19. As stated in Section 28.2.2, a hard and smooth tool surface interacts with a relatively soft and rough sheet material. The sheet material will plastically deform by applying normal force to the tool–work piece interface. Consequently, the sheet’s surface roughness will change and form roughness plateaus. Tool summits or surface defects, see Figure 28.20, now become important because they will plough through the plastically deformed
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28-27
Tool surface
hs Mean plane surface heights S Mean plane summit heights
FIGURE 28.19 Multi-asperity contact between a rigid (tool) surface and a plastically deforming (sheet) surface. Adapted from van der Heide, E. and Schipper, D.J., ASME Journal of Tribology, 126, 2, 275–280, 2004.
FIGURE 28.20 SEM image of roughness summits/surface defects on a tool steel surface. Adapted from Heikillä, I. et al., Proc. Innovations in Metal Foming, Brescia, Italy, 2004.
roughness plateaus at the sheet surface as a result of the sliding action, see Figure 28.19. The sliding action will, depending on the interfacial shear strength, the size of the summit, and the summits attack angle result in ploughing, cutting, or wedge formation [34]. Tool surface defects that operate in the wedge formation mode will initiate galling in unlubricated sheet metal forming contacts [35]. The interaction of tool summits and the work piece surface generates heat, which in turn results in a local surface temperature rise. The relation between frictional heating and surface temperature rise at the asperity level can be described by an extension of the flash temperature model, developed for ideally smooth surfaces by Bos [36]. From calculations [37,38] it follows that the flash temperature Tf at the summit level, strongly depends on the hardness of the sheet material, the thermal properties of the sheet material (conductivity and diffusivity), the operational conditions (load and velocity), and the thermal conductivity of the tool material. Galling initiation in lubricated sheet metal forming processes will occur in case the protective lubricant boundary layer, present at the interface of the tool summit–work piece contact, fails. The latter occurs, depending on the concentration of boundary layer forming compounds and depending on the boundary layer formation mechanisms [39,40], at a certain critical temperature, Tcr [41]. The model for frictional heating at the asperity level, can be applied to SMF tool surfaces, by calculating the relevant input parameters of the model for each individual tool summit. Introduction of
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the constancy of scoring temperature and the wedge formation mode as conditions for material transfer yields a tool for prediction of galling initiation. Based on model simulations, one can obtain two general strategies for avoiding the occurrence of galling [37]: 1. Exclude the occurrence of wedge formation, the necessary condition for material transfer, by: • The application of surfaces, for example, coatings, that introduce a low interfacial shear strength • Removal of possibly wedge forming summits, meaning not only the highest summits that operate in the cutting regime, but also the summits of intermediate height (mirror polish) 2. Exclude high flash temperatures, the necessary condition for lubricant failure, by: • The application of lubricants that resist high temperature • The application of tool materials/tool coatings with high thermal conductivity, which will decrease the maximum surface temperature rise to values below Tcr • The application of soft sheet materials, since a decrease in hardness causes a significant decrease in Tf , possibly below Tcr The first and fifth options are not as straight forward as the other options, because they require to change the sheet material, which in turn could influence the diffusivity of the material or the available area for heat conduction. Further, changes of the sheet material are generally impossible to implement, due to the specific demands of the application. The combination of a smooth tool surface with high thermal conductivity, however, is shown to be successful: galling initiation is avoided for a typical sliding contact with stainless steel sheet [42,9]. The second phase in galling processes, stresses the accumulative nature of galling: lumps on the tool surface grow as a function of the amount of products formed. Lump growth is controlled by the probability of wedge formation in the contact between a tool surface asperity and the flattened sheet asperities. A SEM image of a lump on tool steel after sliding contact with stainless steel sheet material is given in Figure 28.21. The mechanisms that control the growth rate can be simulated numerically for unlubricated sheet metal forming processes [35]. Lump growth continues to a certain point where the lumps reach a critical size and shape, which results in severe damage to the sheet surface. This third and final phase, more stochastic in nature than the previous phase, is characterized by severe scratching of the sheet material and possibly ends with seizure. The effect of lump growth on the severity of scratching of the sheet surface and its relation with friction is
WN 1.2379
AISI 304
FIGURE 28.21 Lump growth on tool steel WN 1.2379 after sliding contact with stainless steel sheet material AISI 304 2B. Adapted from Lovato, G. et al., High volume forming of stainless steel with easy to clean lubricants, ECSC — steel report, contract number 7210-PR-307, 2003.
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Scratch depth (mm)
Coefficient of friction [-]
Scuffing
lump growth
Initiation
Sliding length [m]
FIGURE 28.22 The effect of lump growth on the severity of scratching and friction. Adapted from van der Heide, E., Huis in ‘t Veld, A.J., and Schipper, D.J., Wear, 251–12, 973–979, 2001. TABLE 28.17 Results of Hat Drawing Tests with Coated Tools and Different Lubricants. The Maximum Test Run was 20,000 Hats. Sheet Material AISI 304 (EDT-) Skin Passed
WN 1.2379 WN 1.2379 WN 1.2379
Treatment
Lub. A 71 cSt
Lub. B 190 cSt
Lub. C 160 cSt
Lub. D 80 cSt
Lub. E 36 cSt
Hardened CVD TiN PVD TiCN
20000 20000 20000
2400 4300 20000
2000 20000 20000
500 19000 2000
500 2000 200
Lub. A, Cl containing reference lubricant. Lub. B–D, alternative lubricants (no Cl). Lub. E, preservation oil for DC type of sheet. Source: Adapted from Jordan, F. and Heidbuchel, P., IDDRG Working Group Meetings, IDDRG Birmingham, UK, June 1999.
given in Figure 28.22, which is based on slider-on-sheet test results [43]. It can be seen from Figure 28.22, that lump growth on the die results in increased depth of scratches on the sheet. A similar relation between scratch depth and lump growth is found, for example, by Murakawa [44] for a deep drawing application of aluminum. The selected forming lubricant largely influences the growth rate of lumps on the tool. It has been shown in forming trials, see for example, Table 28.17, with sheet materials sensitive to galling, that effective boundary lubrication delays or even prevents severe scratching.
28.5 Simulation of Tribological Contact Situations in Sheet Metal Forming 28.5.1 General Framework of Demands for Tribo Tests The effect of changes in SMF-contacts, like the application of surface treatment, coating technology, or enhanced lubricant chemistry, on tool life and product quality, should preferably be measured by actually performing tests at an industrial scale. Practical reasons like the unavailability of a trail press or lack of time and money to perform extensive trails, has created the need for tribological screening tests. In general, tribological testing is performed, assuming that industrial applications can be reduced to contact situations with input and output as illustrated by the system approach in Section 28.2. The extracted system is then simulated at a laboratory scale using the same system components and input. Since structure and input
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are similar, it is assumed that the output of the experiment, that is, friction and wear, correlates with the tribological performance in industrial practice [9]. Wear and friction tests related to SMF differ from conventional wear tests because of the specific system components involved and the required type of motion. Actually, the following requirements should be met [43]: 1. Since sheet-tool interaction has to be simulated, it is essential that a well-defined, reproducible contact between sheet and tool material is maintained during the test. 2. It is essential that the sheet material in contact is always fresh. Naturally, the general requirements for a tribotest should be obeyed, that is 3. The operational variables should be applied as in industrial practice. 4. Wear and friction should be measured directly. 5. Friction should be measured independently of the normal force. Because the method should also be economically feasible, final requirements are: 6. The test pieces used should be “easy” to make. 7. The test method should allow for sliding distances in the kilometer range, in case wear is the object of study.
28.5.2 Conventional Tribo Tests and Sheet Metal Forming The first two of the requirements of Section 28.5.1 imply that modified conventional tribo testers, like rotational and reciprocating testing devices, could only generate data linked to SMF for one turn or stroke. This type of testing is used often to measure the effect of sheet roughness and lubricant viscosity on friction as a function of sliding velocity and normal pressure, see for example, Reference 45. This kind of equipment is useful for generating frictional data for input of FEM sheet metal forming simulations, if the test pieces and the operational variables resemble the industrial application of interest (requirements 1 and 3). Wear measurements cannot be done because of the limited sliding distance. Yet, much tribological work related to lubricants is still done at conventional tribo tester like the Shell four ball machine, the Falex pin & vee block test machine or the Timken block-on-ring tester. Typically, one changes the standard test piece material to the materials used for the practical application, such as an aluminum ring vs. a tool steel block, and simply measures the extreme pressure characteristics of the lubricant of interest [2]. This kind of testing is able to rank two lubricants, one with and one without extreme pressure additives like chlorinated paraffin, correctly. Rankings found for a set of commercially available forming oils however, will have very limited value for sheet metal forming applications.
28.5.3 Simulation of Friction and Wear in Sheet Metal Forming Frequently used tribo tests for SMF-conditions make use of strip material drawn between two clamped cylinders, flat dies (flat die test), or over a simple radius (strip stretching test), see Figure 28.23. A strip of sheet material is pulled over a radius, during the strip stretching test, applying a constant pulling force and pulling velocity at point 1. The other end of the strip, point 2, is fixed. This test is also carried out allowing for linear displacement of both ends of the strip (radial strip drawing test). The back-pull force can be used to increase the normal force acting on the radius [46]. The flat die test geometry resembles the contact situation that exists in blank holder–work piece interfaces. The flat die test is typically used to generate friction data in the mixed or boundary lubrication regime, at different contact pressures. A modified test sequence is sometimes used to assess the galling characteristics of zinc coated steel sheet. A specific sequence referred to as the multi frottement test, is used by major European steel sheet manufacturers. The steel strip is oiled once for this test and pulled several times between two flat dies until galling appears. A maximum of 10 strokes of 150 mm is made in this sequence but, to increase the severity of the test, one can opt to use longer strokes or to pull a larger number of strokes.
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FBP Fp 2
Fn
Fn
1 Fp
FIGURE 28.23 Schematic drawing of the Strip Stretching Test setup (left) and the Flat Die Test setup (right). Adapted from Sniekers, R.J.J.M., Friction in deep drawing, PhD Thesis, Eindhoven Technical University, 1996.
Fp 1
2
FIGURE 28.24 Schematic drawing of the Draw Bead Test setup. Adapted from Sniekers, R.J.J.M., Friction in deep drawing, PhD Thesis, Eindhoven Technical University, 1996.
Sheet specimen Spring blades Elastic joint
Sliding tool Bellows Support friction force transducer
Rotating tool Support
Normal force transducer
Main support
FIGURE 28.25 Friction device for the cylinder on strip tribometer. Adapted from de Rooij, M.B., Tribological aspects of unlubricated deepdrawing processes, PhD Thesis, University of Twente, 1998.
Draw bead tests, can be regarded as a variant of this type of testing. Strip material is drawn through a set of three cylinders, as shown in Figure 28.24, by imposing a horizontal displacement at point 2, while suppressing vertical displacement at point 1 [46]. Tests are performed with cylinders that can rotate freely or with fixed cylinders. Again, this test is typically done to generate friction data for simulation of
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Plane surface test
Cylinder-plane test
Blank holder
Sheet metal
Die
Draw bead test
Cylindrical blank holder
Sheet metal
Die
Blank holder
Sheet metal
Draw bead
Die
Tool Working grab
Cam discs Blank holder lever
Drive train
FIGURE 28.26 PtU strip drawing test with exchangeable tool unit. Adapted from Groche, P., Filzek, J., and Nitzsche, G., in Wissenschaftliche Gesellschaft für Produktionstechnik (WGP): Annals of the German Academic Society for Production Engineering, Braunschweig, XI/1, Vol. 1, 2004, pp. 55–60.
the material flow near draw beads. In order to asses galling tendencies, one can for example apply the so-called multi-strip galling test sequence, which was developed for low carbon steel. This test method consists of pulling one oiled strip through the draw bead setup followed by the pulling of fresh nonoiled strips through the beads until galling occurs. A first indication of the galling tendency can be deduced from the number of strips that can be applied without the appearance of visual scratches on both sides of the strip. Friction can be measured direct and independent of the normal force, using a rotating tool in the cylinder on strip method shown in Figure 28.25 [7]. The sliding tool is used for friction measurements, the rotating tool for support during the test. By using a tensile tester for clamping the strip material, experiments can be conducted under conditions of controlled (plastic) deformation of the strip. A well-accepted intermediate stage between practice and laboratory testing uses a rather complex test piece geometry involving bending over a radius [47,48] or drawing a strip between two stationary test pieces, respectively simulating the sheet–tool contact at the die radius and the blank holder contact during
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y FN
v x
FIGURE 28.27
I
Schematic representation of the TNO slider-on-sheet test.
deep drawing. Recently a new PtU-strip drawing tester was designed in which a slitted coil is fed to an exchangeable tool unit [49]. Figure 28.26 shows the test set up and three possible tool units. The upper and lower tool units are equipped with piezoelectric force transducers in order to measure the normal load and friction force separately on each tool. The normal force is applied by means of a lever and a package of leaf springs. The sliding velocity can be adjusted by the rotational speed of the motor. One stroke of the intermitting test is typically 100 mm, a maximum of 15,000 strokes is performed for a standard wear test. Second, long sliding length experiments can be performed with the TNO slider-on-sheet tribo meter [43], see Figure 28.27. This tribo meter consists of a sliding contact between a ring, made of the tool material of interest and sheet material used in the application. Each track is made next to the previous track, in the same direction with sliding velocity v and under a normal load of Fn , thus assuring virgin sheet material in the contact. If the tracks are made 1 mm apart from each other, it is possible to realize 1 km sliding distance on 1 m2 sheet material. The friction force is measured with the help of a force transducer with strain gauges and a cantilever beam construction. The normal force is applied by means of pressurized air. The sliding velocity (0.001–1 m/sec), the normal force (50–1000 N) and the track length (100–2000 mm) can be adjusted within their ranges.
Acknowledgments The authors wish to thank the following people for the contributions to the work: Dr. ir. E.R.M. Gelinck at TNO for his input for Section 28.3 Friction and Dipl.-Ing. G. Nitzsche at PtU Darmstadt for his kind cooperation and supply of Figure 28.5 and Figure 28.26.
References [1] ASM Handbook, Vol. 14, Forming and Forging; ASM: Materials Park, OH, USA. [2] Schey, J.A., 1970, Metal Deformation Processes Friction and Lubrication, Marcel Dekker Inc., New York. [3] Czichos, H., 1978, Tribology: A Systems Approach to the Science and Technology of Friction, Wear and Lubrication, Tribology series, Vol. 1, Elsevier Scientific Publishing Company, Amsterdam. [4] Salomon, G., 1974, Application of systems thinking to tribology, ASLE Transactions, 17, 295–299. [5] Halling, J., 1986, The tribology of surface coatings, particularly ceramics Proceedings of the Institution of Mechanical Engineers, 200, C1, 31–40. [6] Greenwood, J.A. and Williamson, J.B.P., 1966, Contact of nominally flat surfaces, Proceedings of the Royal Society of London A 295, 300–319. [7] ter Haar, R., 1996, Friction in sheet metal forming: the influence of (local) contact conditions and deformation, PhD Thesis, University of Twente. [8] de Vin, L.J. et al., 1996, A process model for air bending, Journal Materials Processing Technology, 57/1–2, 48–54.
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[9] van der Heide, E. and Schipper, D.J., 2004, Tribology of Metal Forming, in Mechanical Tribology: Materials, Characterization, and Applications, Totten, G.E. and Liang, H., eds, Marcel Dekker Inc., New York, pp. 347–374. [10] Steels for Cold Work Tooling, 2001, Uddeholm Tooling, Hagfors, Sweden. [11] Doege, E. and Dröder, K., 1997, Einsatz von Keramik als Werkzeugwerkstoff in der Blechumformung, Bänder Bleche Rohre, 12, 16–21. [12] Kataoka, S. et al., 2004, Tribology of dry deep-drawing of various metal sheets with use of ceramics tool, Surface and Coatings Technology, 177–178, 582–590. [13] Bushan, B. and Gupta, B.K., 1991, Handbook of Tribology — Materials, Coatings and Surface Treatments, McGraw-Hill, New York. [14] Bowden, F.P. and Tabor, D., 1950, The Friction and Lubrication of Solids, Oxford, Clarendon Press. [15] www.lubrizol.com. [16] Schipper, D.J., 1988, Transitions in the lubrication of concentrated contacts, PhD. Thesis, University of Twente, Enschede, The Netherlands. [17] Greenwood, J.A. and Williamson, J.B.P., 1996, Contact of nominally flat surfaces, Philosophical Transactions of the Royal Society of London A, 295, 300–319, [18] Gelinck, E.R.M., 1999, Mixed lubrication of line contacts, PhD. Thesis, University of Twente, Enschede, The Netherlands. [19] Gelinck, E.R.M. and Schipper, D.J., 1999, Deformation of rough line contacts, ASME Journal of Tribology, 121, 3, 449–454. [20] Moes, H., 2000, Lubrication and Beyond, Lecture Notes, University of Twente, Enschede, The Netherlands. [21] Johnson, K.L., Greenwood, J.A., and Poon, S.Y., 1972, A simple theory of asperity contact in elastohydrodynamic lubrication, Wear, 19, 91–108. [22] Gelinck, E.R.M. and Schipper, D.J., 2000, Calculation of Stribeck curves for line contacts, Tribology International, 33, 175–181. [23] Gelinck, E.R.M. and Schipper, D.J., 2001, Stribeck and traction curves for highly loaded contacts, Report, University of Twente, TR-012263. [24] Briscoe, B.J., Scruton, B. and Willis, F.R., 1973, The shear strength of thin lubricant films, Philosophical Transactions of the Royal Society of London A, 333, 99–114. [25] DIN 50320, 1979, Verschleiß — Begriffe, Systemanalyse von Verschleißvorgängen, Gliederung des Verschleißgebietes, Beuth Verlag, Berlin. [26] van der Heide, E. et al., 2003, Wear of soft tool materials in sliding contact with zinc coated steel sheet, Journal of Materials Processing Technology, 141/2, 197–201. [27] Eriksen, M. and Wanheim, T., 1997, Wear optimisation in deep drawing, in Proceedings of the 1st International Conference on Tribology in Manufacturing Processes ’97, Dohda, K., Nakamura, T., and Wilson, W.R.D., eds, Gifu, Japan, pp. 128–133. [28] Archard, J.F., 1953, Contact and rubbing of flat surfaces, Journal of Applied Physics, 24, 981–988. [29] Shaw, M.C., 1977, Dimensional analysis for wear systems, Wear, 43, 263–266. [30] van der Heide, E. et al., 2004, Wear of aluminium bronze in sliding contact with lubricated stainless steel sheet material, in Proceedings NORDTRIB 04 Conference, Tønder, K. et al., eds, Tromsø, Norway, pp. 355–364. [31] Andreasen, J.L., Eriksen, M., and Bay, N., 1997, Major process parameters affecting limits of lubrication in deep drawing of stainless steel, in Proceedings of the 1st International Conference on Tribology in Manufacturing Processes ’97, Dohda, K., Nakamura, T., and Wilson, W.R.D., eds, Gifu, Japan, pp. 122–127. [32] Schedin, E. and Lehtinen, B., 1993, Galling mechanisms in lubricated systems: a study of sheet metal forming, Wear, 170, 119–130. [33] Schedin, E., 1994, Galling mechanisms in sheet forming operations, Wear, 179, 123–128. [34] Hokkirigawa, K. and Kato, K., 1988, An experimental and theoretical investigation of ploughing, cutting and wedge formation during abrasive wear, Tribology International, 21, 1, 51–57.
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[35] de Rooij, M.B., 1998, Tribological aspects of unlubricated deepdrawing processes, PhD Thesis, University of Twente. [36] Bos, J. and Moes, H., 1995, Frictional heating of tribological contacts, Journal of Tribology, 117, 171–177. [37] van der Heide, E., 2002, Lubricant failure in sheet metal forming processes, PhD Thesis, University of Twente, Enschede, The Netherlands. [38] van der Heide, E. and Schipper, D.J., 2004, On frictional heating in single summit contacts: towards failure at asperity level in lubricated systems, ASME Journal of Tribology, 126, 2, 275–280. [39] Frewing, J.J., 1943, The heat of adsorption of long-chain compounds and their effect on boundary lubrication, Proceedings of the Royal Society of London, A, 182, 270–285. [40] Spikes, H.A. and Cameron, A., 1973, Scuffing as a desorption process — an explanation of the Borsoff effect, ASLE Transactions, 17, 2, 92–96. [41] Blok, H., 1969, The postulate about the constancy of scoring temperature, in Interdisciplinary Approach to Friction and Wear, Ku, P.M., ed., Symposium Troy, New York, NASA SP-237, pp. 153– 248. [42] van der Heide, E. and Schipper, D.J., 2003, Galling initiation due to frictional heating, Wear, 254/11, 1127–1133. [43] van der Heide, E., Huis in ‘t Veld, A.J., and Schipper, D.J., 2001, The effect of lubricant selection on galling in a model wear test, Wear, 251–12, 973–979. [44] Murakawa, M., Koga, N., and Takeuchi, S., 1997, Utility of diamondlike carbon-coated dies as applied to deep drawing of aluminum sheets, in Proceeding of the 1st International Conference on Tribology in Manufacturing Processes ’97, Dohda, K., Nakamura, T., and Wilson, W.R.D., eds, Gifu, Japan, pp. 322–327. [45] Emmens, W.C., 1997, Tribology of flat contacts and it application in deep drawing, PhD Thesis, University of Twente. [46] Sniekers, R.J.J.M., 1996, Friction in deep drawing, PhD Thesis, Eindhoven Technical University. [47] Woska, R., 1982, Einfluß ausgewählter Oberflächenschichten auf das Reib- und Verschleißverhalten beim Tiefziehen. PhD Thesis, TU Darmstadt. [48] Schulz, A. et al., 1997, Deposition of TiN PVD coatings on cast steel forming tools, Surface and Coatings Technology, 94–95, 446–450. [49] Groche, P., Filzek, J., and Nitzsche, G., 2004, Local contact conditions in sheet metal forming and their simulation in laboratory test methods, in Wissenschaftliche Gesellschaft für Produktionstechnik (WGP): Annals of the German Academic Society for Production Engineering, Braunschweig, XI/1, Vol. 1, pp. 55–60.
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Maintenance
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III
29 The Degradation of Lubricants in Service Use 29.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
29-3
Controlled Deterioration of Lubricants • The Effects of Deterioration • Physical Causes of Deterioration • The Effects of Lubricant Chemical Deterioration • The “Bath-Tub Curve”
29.2 Field Tests for Lubricant Deterioration . . . . . . . . . . . . . .
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Direct Observation of Lubricant Condition • Field Kits for Lubricant Condition
29.3 Laboratory Tests for Lubricant Deterioration . . . . . . .
29-10
Viscosity and Viscosity Index • Trace Metals • Particulates and Ash in Lubricants • Acidity and Base Reserve • Water Content
29.4 Minor Methods of Investigating Lubricant Degradation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
29-33
Density, “Gravity,” or “Specific Gravity” • Flash Point of Degraded Lubricant • Foaming of Lubricants • System Corrosion (“Rusting”) with Degraded Lubricants • Demulsibility and Interfacial Tension of Degraded Lubricants • Instrumental Analytical Techniques
29.5 Case Studies of Degraded Lubricants . . . . . . . . . . . . . . . .
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A Degraded Lubricant Sample from a Heavy Duty Diesel Engine • A Degraded Grease Sample • A Degraded Lubricant Sample from a Gas-Fueled Engine • A Degraded Hydraulic Fluid • Overview of Degraded Lubricant Analyses
Malcolm F. Fox De Montfort University
Bibliography . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Reference . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
29-40 29-40
29.1 Introduction The very nature of lubricant service means that lubricants deteriorate during their service use. It is normal that lubricants degrade by partial evaporation, oxidation, and contamination. The purpose of lubricant formulation for a defined application is to control the deterioration of that lubricant in a planned manner over an established period of time, work, distance, or operation. The deterioration of a lubricant can either 29-3
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be planned and controlled by various means or be uncontrolled. Modern practice is strongly directed to the former.
29.1.1 Controlled Deterioration of Lubricants The way that a lubricant is changed in service use addresses the two extremes of one of the following: • A time- or distance-defined period of lubricant replacement, such as 500 h operation, annually or 10,000 km, without regard to the actual state of the lubricant. But custom and practice show that the service interval set is sufficient to ensure that excessive wear does not occur — a precautionary principle. This approach does not require sampling and analyses or “on-board” sensors and therefore of low-cost. The issue is that the lubricant is replaced with a substantial amount of remaining “life” in it, therefore tending to be wasteful of resources. • At the other extreme, a quantitative appreciation of the state of the lubricant is done by sampling at regular intervals and monitoring various parameters to give a collective assessment of the condition of the lubricant, “condition monitoring.” The time interval of sampling should be, at most, half of the anticipated service interval. The database built up over time has value for long term and is concerned with long-term trends in lubricant parameters such as wear metal concentrations, viscosity, and particulate levels. For a full condition monitoring program, the lubricant is replaced when its condition reaches a lower bound of aggregated parameters and it is judged to be, or close to being, unsuitable for its purpose of lubricating and protecting the mechanical system. • An interim position is to sum the overall performance of the system, be it engine or machine, from its last service interval by integrating power levels used in time intervals/distances traveled/time elapsed. The underlying assumption is that the level of performance and its time of operation are related to the degradation of the lubricant. Thus, 100 km of unrestricted daytime high-speed driving on an autobahn in summer is assumed to degrade a lubricant more than 100 km of urban driving in autumn or spring. Thus, the aggregates of high power level operation over time are weighted more than the same period of low power operation. Integration of the high and low power level operation is already used in some vehicles to indicate to the operator when the system’s service is due and the lubricant must be replaced. The objective at the end of the service period must be that the lubricant still be “in grade,” therefore specification, and that the engine or machine not to have suffered “excessive wear” or component damage. This “state of grace” is readily achieved by the vast majority of lubricants in operational service through the development and testing of formulations. The major current development is for service intervals to increase in terms of hours operated or distance traveled. Thus, for light vehicles, service intervals are progressively increasing to 20,000, 30,000, and 50,000 km for light vehicles. A target of 400,000 km is envisaged for heavy duty diesel engines or their “off-road” equivalent.
29.1.2 The Effects of Deterioration Lubricants are formulated from a base oil mixture and an additive pack, as described elsewhere in this volume. The base oil is usually a mixture of base oil types and viscosities chosen for their physical and chemical properties and their costs. The additives form part of an additive pack to protect oxidation, wear, acidity and corrosion, to remove and disperse deposits, maintain a specified operating viscosity range, and minimize foaming. A filter in the lubricant circulation system should remove suspended particulates above a certain diameter. Lubricant degradation occurs throughout its service life and the baseline for change is reached when its further deterioration would lead to a level such that it cannot protect the system from further excessive wear. This occurs because the lubricant has become physically unsuitable for further service use for several
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separate or joint causes: • It has become too laden with particulate dirt. • It’s viscosity has increased/decreased beyond its specification limit. • It’s additive pack has become depleted in one or multiple components. Often the additive component actions are interdependent, thus oxidants may protect other additive actions. • Abrasive and corrosive materials can cause bearing damage, or bore polish by removing the crosshatched honing marks, which maintain the lubricant film, or in extreme cases, “scuffing” of piston and bore. These effects are often interdependent and will cause further changes either directly related or through catalytic effects. When these lubricant deterioration effects occur in such complex systems as lubricant formulations, then a structured approach is needed to understand and solve the problem.
29.1.3 Physical Causes of Deterioration A lubricant formulation becomes physically unsuitable for further continued service use through a range of the following causes: 1. Internal sources: internal contributing sources are those which are either introduced into a system by the production or repair process, as: (a) Textile materials such as (production line) cleaning cloths, contributing “lint,” which compacts into obstructions of oilways. (b) Metallic materials such as metallurgical cutting residues and welding repair particulates or production grinding processes, or by the operational process, of either fuel or oxidative use, as follows: • Harder/softer particulates from the partial oxidation of lubricants as harder particulates from longer, C30 hydrocarbons, as in lubricant hydrocarbons, and softer particulates from shorter, C15 , hydrocarbons, as in diesel fuels hydrocarbons. • Through defective sealing systems, which allow ingress of silicaceous abrasive sources. • Fuel condensing into the lubricant and reducing its viscosity, or together with condensed water, forming an emulsion of low lubricity value. Cooling water ingress into the lubricant system through defective seals is another source of water contamination. 2. External sources: external contributing sources, predominantly grit and dust, are those either introduced into a mechanical system by: (a) Infiltration through exhausted and inefficient oil filters (b) Filling through unclean filler pipes/tubes (c) Lubricant reservoirs open to the (unclean) atmosphere (d) Through overwhelmed air filters, as in desert area operations The debris of system wear, abrasive wear products from combustion processes, and defective sealing materials are physical causes of lubricant deterioration. Another obvious physical cause of degradation is to add an incompatible lubricant to an existing formulation in an existing system — while the base fluids may be miscible, their additive packs may be incompatible and precipitate (“drop out”), leaving the circulating fluid as a simple base oil system with little mechanical/tribological protection. In most cases, the physical causes of lubricant deterioration are simply related to good maintenance, or the lack of its meaningful application, simply put as “good housekeeping.”
29.1.4 The Effects of Lubricant Chemical Deterioration Of all the chemical causes of lubricant deterioration, oxidation is the most important. It has extensive onward connections to the formation of organic acids, usually carboxylic acids, sludges that lead to resins/varnishes, which in turn bond carbonaceous deposits onto system components. Oxidation forms
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hard carbon from heavy hydrocarbons such as lubricant base oils, engines become very dirty and if the oxidation is sufficiently severe, then essential small orifices such as filters, minor oilways, and crucial orifices such as undercrown cooling jets become blocked and rapidly cause severe wear problems. Oxidation is temperature dependent and as a chemical reaction, is subject to the Arrhenius effect of reaction rates doubling/trebling for every 10◦ C increase in temperature. Thus, a reaction rate of unity at 300◦ C will increase to between 2 and 3 at 310◦ C, to between 4 and 9 at 320◦ C, and between 8 and 27 at 330◦ C, and so on — a compound increase. This has important implications for trends in increasing engine power densities, smaller lubricant volumes, and reduced cooling effects due to vehicle aerodynamics, which lead to increased engine operating temperatures, including its lubricant system. Future lubricants must withstand higher operating temperatures using smaller volumes for longer service intervals. Advanced lubricant formulations must be developed, which can operate at consistently higher temperatures to prevent their deterioration below levels that protect power train systems for extended, longer, service changes. The reserve concentration of unused, effective antioxidant in the lubricant during its service life is a crucial factor. Exhaustion of the antioxidant in the continuous use of a lubricant rapidly leads to the mechanical deterioration of the system. It is not sufficiently appreciated that heavier hydrocarbons, as used in lubricant base oils, have up to 10% of air dissolved or entrapped within it, the difference is semantic. The mechanical movement of the lubricant, as flow, agitation, or foaming, will maintain the air/oxygen concentration in the oil and increase the rate of oxidation. High temperatures will also affect the base oil molecules and additives directly. Thermal degradation is selectively used in refineries to reform hydrocarbons at temperatures similar to those by lubricants experienced within engines. Under the relatively uncontrolled thermal breakdown conditions within an engine, base oil molecules can break down into smaller molecules, “cracking,” or become functionalized with carbonyl groups, particularly, and undergo polycondensation to form varnishes and gums, which trap and sequester carbonaceous particles. The thermal stability of base oils is an important parameter in their selection. Additives are destabilized by high engine operating temperatures, dependent upon the extent and duration of their exposure to these high temperatures within the engine system, such as the ring zone and valve guides. The term “additives” covers a wide range of compounds, which can contain sulfur, phosphorus, and chlorine. Complex additives can break down to form a range of smaller compounds; thus, Zinc Di-alkyl Di-thio Phosphates (ZDDPs), antioxidant and antiwear agents break down in the ring zone of diesel engines to form organic sulfides and phosphate esters [1]. But reaction between additives — additive interaction — caused by exposure to high temperatures, not only depletes those additives but can also generate sludge deposits. The intermediates may also be corrosive to the system. There are several overall tests for the antioxidant reserve/antioxidancy of an oil, new or used, as either the ASTM 943, 2272, and 4310 tests, also the IP 280 tests. Of these are the following: • The Rotating Bomb Oxidation Test (RBOT), ASTM 2272 where a rotating bomb is loaded with a lubricant oil charge, pressurized with oxygen in the presence of a copper catalyst and water within a glass vial. The time recorded for the oxygen to deplete, by reaction, and its pressure to fall by a specified increment of 25 psi (1.74 bar). This method is operator-intensive and has a range of random errors greater than the other. • Pressurized differential scanning calorimetry (PDSC) method, CEC-L85T-99-5 is a relatively lowcost test with much improved reproducibility, where a small (8 mg) sample within a very small cup is held under 35 atmospheres pressure of air in a differential scanning calorimeter at 190◦ C. The time for the overall additive function to be exhausted by the combination of high temperature and the diffusion controlled oxidizing atmosphere and the residual hydrocarbon combustion to give an exotherm, as in Figure 29.1, is the “induction time.” New lubricant formulations will have longer induction times, which will gradually reduce for used samples of that formulation as its service life proceeds. A “zero” value for an antioxidant “induction time” indicates that the lubricant sample is substantially degraded and unprotected against further, and substantial, oxidative attack.
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Enthalpy
Exotherm for new lubricant Exotherm for used lubricant Induction times
tused
tnew
Time
FIGURE 29.1 PDSC induction time plots for new and used lubricant samples.
Wear or system failure
Terminal wearout/ failure state Initial rapid decrease in wear or system failure – “break-in”/“running-in”
Normal wear/system life — low level of steady state wear/failures Onset of terminal wear/failures
Operating time
FIGURE 29.2 The “bath-tub” curve for wear/system failure.
29.1.5 The “Bath-Tub Curve” All systems wear but at different rates in their serviceable life. The pattern of wear is well described by the “bath-tub” curve, which is a plot of “wear” against time, Figure 29.2. It can also be regarded as a plot of system failure against time. A “bath-tub” curve does not describe “wear” (or “failure”) for individual systems but is a statistical description of the relative wear/failure rates of a product population with time. Individual unit can fail relatively early but with modern production methods, these should be minimal, others might last until wear-out, and some will fail during the relatively long period, typically called normal life. Failures during the initial period are always caused by material defects, design errors, or assembly problems. Normal life failures are normally considered to be random cases where “stress exceeds strength.” Terminal “wear-out” is a fact of life due either to fatigue or material depletion by wear, from this it is self-evident that the useful operating life of a product is limited by the component with the shortest life. The “bath-tub” curve is used as an illustration of the three main periods of system wear/failure, and only occasionally is wear and failure information brought together into a database and the initial, normal, and terminal phases of
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system wear failure measured and calibrated. The timescales for these phases usually vary between one system and another. However, when condition monitoring is used to monitor the wear of a system, then a gradually increasing level of iron in each sample taken at service interval lubricant changes indicates that an engine has entered the final phase of its service life and its replacement and overhaul is becoming due. The necessary replacement arrangements can be made without failure or unexpected interruption of service. This saves costs because the engine is worn, but not damaged, readily and economically overhauled, the operation is planned and service interruption is minimized. Informed replacement of worn systems or components is usually estimated to have a direct benefit/cost ratio of 10:1, rising to 20:1 when indirect costs of unexpected interruptions of service are included.
29.2 Field Tests for Lubricant Deterioration Laboratory analyses of lubricants are necessarily done in laboratories; they are accurate but delayed unless, unusually, an operating site has its own laboratory. There is a good case for simple field tests, which may be less accurate but gives an immediate indication. Often the operation is physically separated from a laboratory, as in a merchant or naval ship, and needs reliable, simple tests.
29.2.1 Direct Observation of Lubricant Condition An experienced observer of lubricant condition will give considerable attention to the color of a lubricant sample — it is helpful to compare with an unused sample. Oxidative and thermal breakdown of a lubricant, often beyond exhausting its antioxidant reserve, gives a darker, more brown, color. The deepening in color is also associated with a very characteristic “burnt” odor, which is recognizable when experienced. The viscosity of the sample will also increase.
29.2.2 Field Kits for Lubricant Condition Various “field kits” are available to measure the essentials of lubricant condition, such as viscosity, water content, particulates, and degree of oxidation. These were called “spot tests” in the past but have improved in reliability to be acceptable for continuing analyses where access to laboratory tests is limited, such as on ships or isolated sites. Viscosity is readily measured by using a simple “falling ball” tube viscometer in the field on site. Comparison with an identical apparatus, often in a “twin arrangement” containing a new sample of lubricant gives a direct comparison of whether the used lubricant viscosity has relatively increased or decreased by the respective times taken for the balls to descend in their tubes. The simplest method to determine particulate levels in a sample of a degraded lubricant is the blotter test, where a small volume of oil sample is pipetted onto a filter paper or some other absorbent material. This is generally known as the “Blotter Test,” which can take various forms, either using a standard filter paper or a thin layer chromatographic (TLC), plate. The measurement concerned is the optical density (OD) of the central black spot. The higher the level of particulate, the denser (darker) the spot. The assumption is that the spread of the lubricant sample disperses carbon particulate within an expanding circle and that the optical density of the carbonaceous deposit is a direct measurement of the mass of particulate present in that sample. The system can be quantified by use of a simple photometer, for fieldbased simple systems, or a spectroreflectometer for laboratory measurements. Methods of automating these types of systems have included the following: • Automated, accurate, constant volume pipetting of the oil samples • Video measurement of the oil sample blot on the filter paper, thus its “OD” • Data recording of these results Despite many attempts and applications, these advanced methods have not achieved universal acceptance, possibly because of the increased complications built onto an initially simple test. Another,
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Black soot ring White oil ring
Development over time
FIGURE 29.3 TLC plate soot spot/oil ring dispersancy test.
and major, problem is the heterogeneous nature of the samples presented for analysis, which give different responses, arising from: • Different base stocks, such as the differences introduced by the mineral, semisynthetic, and synthetic base stocks used in modern lubricant and hydraulic formulations. • Different formulations, such as the differences between hydraulic, automotive, aerospace, and marine fluid formulations, a high dispersancy oil spreading its carbonaceous matter over a greater area than a low dispersancy oil. Marine lubricant formulations are an interesting case to consider. The lubricant volumes used per engine/vessel are very large, of the order of 103 l. The fuel used is high in sulfur, not being controlled to the same extent as land-based automotive diesel fuel, causing extensive additive and base oil degradation. The general case is for vessels to pick up the available top-up lubricants whenever they dock in various ports, leading to heterogeneity of base stock and additive formulation. These factors lead to scatter in the particulate signal/concentration plot. A further development of the“Blotter Test”is to use TLC plates, which are more uniform than paper. The intensity of the black spot from a 50 µl aliquot can be measured and, if its image is captured electronically, may be integrated across its area. But the black carbonaceous spot will also have a base oil ring extending beyond it, seen either as a change in white shade or a fluorescent area under UV illumination, Figure 29.3. The diameter of the white oil ring measures the movement of the lubricant and the black soot ring the movement of the soot particulate. This can be developed into a measurement of dispersancy for the oil sample. Dispersancy is a difficult property to measure, analysis of the dispersant concentration may indicate the amount of free dispersant in the sample together with a variable amount of dispersant desorbed from the particulate, an unsatisfactory measurement. The most effective way to measure dispersancy is to measure the dispersancy ability of a sample, not the concentration of dispersant. The dispersancy of a sample can be measured by the ratio of the black soot ring to the white oil ring. While this is not absolute, the change in dispersancy over the course of an engine test or the service life of a lubricant can be followed by the change in the spot/ring ratio, as the CEC97-EL07 development method. The method is very reproducible, provided that all of the following are considered: • Multiple samples are taken, which is much easier than the previous methods. • The sample images are captured using high resolution optical electronic methods and the area of each spot integrated, as the edges of the spots are often uneven in detail. • Each micropipetted sample is accurately and reproducibly dispensed. The ratio of the “spot” diameters for the white oil ring and the particulate measures the ability of the lubricant sample to disperse carbon particulates, a high ratio indicating a high level of dispersancy remaining in the lubricant. Equally, a low ratio of carbonaceous black spot to the radius of the oil blot indicates a low level of dispersancy. Dilution of a used lubricant sample with a light hydrocarbon such as “Petroleum Ether 60/80” and subsequent filtration through a standard filter paper will indicate the nature of the larger particulate debris, emphasizing metal particulate debris. Microscopic examination of the metal debris can show the nature of the larger metallic debris, which indicates the pattern of wear.
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Water content can be measured “in the field” by mixing a lubricant sample with a carbide tablet in a sealed stainless steel bomb. The measurement of water content is through the reaction of the carbide tablet (or calcium hydride in an alternative model) to generate gas pressure within the bomb. The pressure level generated is a measure of the water content of the sample. An alternative quick test for water content is the “crackle test,” where a small lubricant sample is suddenly heated. This can either be done by suddenly inserting a hot soldering iron bit into the sample — if water is present, a “crackling” noise is heard, which is absent for dry samples (the noise comes from steam generation in the sample) or small drop of sample can be dropped from a syringe onto a “hot” laboratory hot plate, when again a “crackle” will be heard if the sample is “wet.” From experience, the limit of detection is taken to be 0.1% water. The degree of oxidation can be measured by a simple colorimeter using a standard sample to measure color, ASTM D1500. The trend compared to previous values is the important observation. If the change occurs early in the service of the lubricant, then the antioxidancy reserve of the lubricant is being rapidly depleted or the lubricant is being contaminated. It is important to consider the change in color in combination with values and changes determined for Acid Number and viscosity for the same samples. Other simple tests are available in addition to those described above, as a suite packaged into a portable package for measurement of lubricant degradation in isolated situations such as remote mines and onboard ships.
29.3 Laboratory Tests for Lubricant Deterioration Some introductory general remarks are useful: • Results from laboratory tests for lubricant deterioration are of much greater value if the original, virgin, unused, lubricant is used as a benchmark. • Similar tests apply to most forms of lubricants as the deterioration challenges they face are chemically and physically similar. • However, the results from similar tests for different forms of lubricants must be considered in the context of each lubricant’s application. The advantage of laboratory tests is that they should have a background of both quality assurance and control. From this, they have serious weight in solving problems, assessing oil change intervals, what preventive maintenance is required from condition monitoring to conserve the system, and as well lubricant resources. The primary objective of a laboratory analysis program for lubricant samples is to ensure that they are fit for further service. If the lubricant is unfit, or becoming unfit, for further service, then it must be replaced. The benchmark for a laboratory program of sample analyses to assess a lubricant’s deterioration is to offer a rapid turnaround for analytical results, their assessment against limit values, and reporting back to the client. Isolated heavy plant mining operations can have lubricant analytical sample reporting times of weeks due to transport and communication issues; intensive transport systems in developed countries can expect less than 24 h reporting, such that a sample taken 1 day will be analyzed and reported upon before the next day’s operation commences and the appropriate action taken. An equally important benchmark is for laboratory to meet the various national or international standards, such as the ISO 9000 series. The use of certified analytical standards and accredited solutions is part of a complete package, which best involves a collaborative program of regular analyses of samples sent from and collated by central standards body. All apparatus and substances used should have an audit trail for standards and calibrations that are maintained. This is not only a good practice but necessary to respond to any implied liabilities, which may arise later. Of the many tests available, the major issues of lubricant deterioration are addressed by analyses of viscosity and viscosity index (VI), trace metals, particulates, ash, acidity/base reserve, and water contents. Other minor issues are color, demulsibility, foaming, rust testing, infrared spectroscopy, and to a certain
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extent, x-ray fluorescence (XRF). Gas and liquid chromatographies, x-ray diffraction, interfacial tension, and density are peripheral techniques that might be used to investigate unusual occurrences.
29.3.1 Viscosity and Viscosity Index Viscosity is the foremost quality of a lubricant to be measured. A lubricant must maintain its viscosity to effectively protect a system against seizure. Variations in viscosity are usually associated with effects, which show up in other analyses; therefore, a multidimensional approach is needed to consider the root cause of the change. Viscosities of lubricant samples are now measured by automated systems, taking samples from multiple sample trays, either circular or linear, and injecting them into either kinematic or absolute viscometers thermostated at either 40 or 100◦ C. If separate viscometers at these temperatures are used, then the VI of the sample can be calculated. Standards are inserted into the flow of samples through the system for quality control. Individual measurement using manual stopwatches and U-tube suspended viscometers are now rarely used in laboratories. A small increase in lubricant viscosity may be due to evaporation of the lighter ends of the base oil after prolonged high level operation. Beyond that, significant increases in viscosity, up to 10/20% being regarded as severe, result from the inadvertent replenishment with a higher viscosity lubricant, extensive particulate contamination, and extensive base oil oxidation. The particulate contamination as well as extensive oil oxidation will be readily seen, the latter on its own as an increasing dark brown coloration. The black particulate contamination will obscure the brown oxidation color. Oxidation effects will also appear in the Fourier Transform Infrared (FTIR) spectra and decrease in the PDSC antioxidant reserve time. A decrease in the viscosity of operating engines is usually due to fuel dilution, a characteristic occurrence when an engine idles for a prolonged period. A locomotive used for weekend track maintenance train duties will run its diesel engine at idle for periods of several hours and its lubricant will show a significantly decreased viscosity afterward due to fuel dilution. If subsequently used for normal, higher power duties, the increased lubricant temperature will evaporate the condensed fuel and the viscosity returns to its previous value. A more serious occurrence is when fuel and water are extensively condensed in the crankcase of a very cold engine at start-up. During short journeys, when the engine lubricant rarely becomes warm enough to evaporate the condensed fuel and water, the two contaminants can combine to cause the additive package to precipitate out from the lubricant formulation. The engine may then have “oil” but is then left with considerably reduced protection wear. Measuring fuel dilution in diesel lubricants is difficult and is discussed later in the subsection on “flash point.” Fuel condensed into a lubricant has the role of a solvent and the same effect of decreased viscosity is found when a solvent becomes entrained, such as a refrigerant fluid. Chlorofluorocarbons (CFCs) are well on their way to removal and nonreplacement from refrigeration systems but their replacements, the hydrochlorofluorocarbons (HCFCs) and hydrofluorocarbons (HFCs) have the same effect of reducing viscosity if allowed to leak through seals or rings and dissolve in a lubricant. Viscosity index improvers (VIIs), are long chain polymers of various basic units. Their different structures resist high rates of mechanical shear, as in bearings or in the ring pack/bore wall interface, to different extents. While there is a separate effect of temporary viscosity shear loss, lubricants with VIIs can suffer permanent viscosity shear loss due to breaking of the polymer chains. The initial lubricant selection process should have considered how robust the formulation was to permanent shear thinning. Tests for this include high temperature and high shear procedures such as ASTM D4683 and D4741. If the viscosity of a lubricant changes during its service use then its VI, will change necessarily. The major cause of a reduced VI is caused by breaking of some of the polymeric VII polymer molecules to give smaller chains of less effect. There are two effects — reduction in the molecular weight of the VII additive will reduce the viscosity of the lubricant formulation at both 40 and 100◦ C and also reduce the temperature related VII effect. The latter effect normally has the greater weight so that the permanent
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shear breakdown of polymeric VII additives reduces the lubricant’s VI. It is not unknown but rare for the VI of a used lubricant to increase in service use, often associated with extensive oxidation.
29.3.2 Trace Metals The term “trace metals” in a lubricant sample not only covers metals generated by wear in the system but also the elements from the additive pack. While the determination of trace metals for a “one-off ” sample gives some insight into the condition of a lubricant, the major value of trace metal determination lies with long-term condition monitoring. The “bath tub” curve of Figure 29.2 is recalled here — following the level of iron fine particulate in a series of regularly sampled lubricant from a system is an essential part of condition monitoring. The “break-in” or “running-in” phase, normal wear, and the gradual increase in wear element determination can be followed running over many hours and lubricant service changes. The onset of terminal wear can be detected and followed, with arrangements put in place to remove and replace the engine system. Levels of wear elements measured are usually iron (from bores and crankshafts), lead and copper from bearings, aluminum from pistons, and chromium from plating on piston rings. Others may be added to follow specific effects, for example, sodium levels indicate the ingress of cooling water and its additives, silicon levels indicate the ingress of sand and rock dust. It is important to recognize that the level of wear elements in a system’s lubricant is individual both to system design and to individual systems. Thus, levels of iron in the normal wear phase of engines will be different from one design to another; in addition, there will be some variation between the normal wear phase iron levels of engines of the same design. The quality of the lubricant used will also affect the level of wear metal, the higher the quality of lubricant, the lower the level of wear elements. The emphasis for assessing the condition of lubricated systems is placed upon the trend in wear element levels. While the iron level in the lubricant of one engine may be higher than another, it is the trend for successive samples over time in the measured levels, which is important. Wear processes in lubricated systems rarely occur for one metal. Increases in the levels of several wear metals can indicate the occurrence a particular wear process or contamination. Table 29.1 describes wear elements found in lubricants in service life. Wear metal analyses have additional effectiveness when combinations of enhanced element rates are considered, such as for a diesel engine. Combinations of enhanced wear elements are unique to each operating system design and its pattern of use. “Expert systems” applied to an extensive data system can be used to develop “rules,” which indicate which main assemblies or subassemblies are developing enhanced rates of wear and require attention for certain engine designs. The examples given in Table 29.2 are typical for certain applications — other systems may have different combinations for wear patterns, it is for the expert system to recognize them. More extensive combinations of elements indicating particular wear patterns by system components can be developed, such as using the “principal indicator” and associated “secondary indicator” elements. Cost-benefit analyses of spectroscopic oil analysis programs, with the acronym “SOAP,” have been demonstrated in many applications to be very significant. Continuously and heavily used plant, such as diesel express trains, where daily oil sampling and analysis gives an immediate cost-benefit ratio of 10 : 1 in direct costs and 20 : 1 for indirect costs when service reliability benefits are included. The analytical methods for wear metals have generally moved to inductively coupled plasma (ICP) atomic emission systems. A small sample is automatically extracted from a sampling bottle, diluted with kerosene and sprayed into the ICP analyzer plasma torch at 6000–8000◦ C. The very high temperature of the plasma excites the metal particulates to high energies, which emit light of a characteristic atomic wavelength. Duplicates (or more) are readily programmed. The emission from each metal present is detected and reported, the cost of additional wear element detection is marginal once the ICP system is set up. The ready availability of duplicate sample determinations and insertion of calibration standards gives a high level of quality control as precision, accuracy, and reproducibility to the final results. The analytical data generated by the ICP system is readily handled, quantified, and then placed into a file for that engine
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TABLE 29.1 Wear Elements in Lubricants and Their Sources Source Major elements aluminum Chromium Copper
Iron
Lead Molybdenum Nickel Silver Tin Titanium Zinc Minor elements Antimony Boron Magnesium Manganese Sodium Silicon
Primary component of piston alloys, also bearings, washers/shims and casings of accessories. From corrosion of engine blocks, fittings, and attachments Used as a hard(er) coating to reduce wear, indicates wear of chromium plating on engine bores, shafts, piston ring faces, some bearings and seals With zinc in brass alloys and tin in bronze alloy wearing components, copper present in journal, thrust, and turbocharger bearings, also cam, rocker, gear, valve, and small-end bushings. Also, fabricate oil cooler cores Still a major, massive component of engines, gearboxes, and hydraulic systems. Lubricant contact through cast bores, cylinder liners, piston ring packs, valve guides, rolling element bearings, chains, and gears. Difficult decision given by wearing component increased trace levels of iron In bearings, solder joints as “lead/tin alloy” and also seals A wear reduction coating on first piston ring faces for some diesel engines From valves, turbine blades, turbocharger cam plates, and bearings Alloys in bearings, bearing cages, and bushings for diesel engine small ends, turbochargers and rolling element bearing applications in gas turbines Common alloy in bearings with aluminum, bronze and brass fittings, seals, and also in cooler matrix solder Top end of market, gas turbine bearing hubs, turbine blades, and compressors With lead and tin in common alloys such as brass and also some seals May be used in bearing alloys Borates used as cooling system anticorrosion agents, presence in lubricant and hydraulic fluids shows leak in cooling system matrix Increasingly used as an alloy with aluminum for accessories and casings From corrosion of manganese steel alloys, occasionally in valves Usually sodium borate as cooling system anticorrosion agent. Increasing trace presence in fluids shows leak in cooling system matrix, marine applications indicate ingress of coolant sea water Piston wear. As silica, indicates road dust ingress, particularly damaging as hard particulate, which causes high levels of wear, shows air filter and breather system failure, particularly mining and deserts
TABLE 29.2 Elements
Some Indicative Combinations of Wear
Elements Sodium and Boron Lead and Copper Copper, Silver, and Iron Chromium and Iron Silver, Copper, and Lead Iron and Copper
Indicative cause Coolant leakage into lubricant, as through head gasket failure Main or big-end bearings Turbocharger bearings Piston rings Small-end bush Oil pump wear
system, which can then be compared with previous results. This is concentration level “trending” in its simplest form. The overall effect is to give a high throughput of high quality analyses at low cost. While the automated sampling ICP multiple element system has a high capital cost, of £150–200 k ($300–400 k) each, the high sample throughput can cut the unit cost per sample down to 50 p ($1). An atomic absorption (AA) apparatus can be used instead of the ICP system but suffers from the disadvantage of only determining one element per analysis from the nature of this method. The older emission system of an electric discharge between either still or rotating (“Rotrode”) carbon electrodes is still used but the advantages of the ICP system for high throughput of samples are gradually displacing it. The ICP spectroscopic technique and oil samples are brought together as a condition monitoring system.
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Handbook of Lubrication and Tribology TABLE 29.3
Corrective Levels for Lubricant Deterioration
Deterioration level Normal Alert Urgent Hazardous Dangerous
Action Within average, no action Within average ±2σ , action → increase sampling frequency Within average ±4σ , action → maintenance needed, can be deferred Beyond average ±4σ , action → immediate maintenance, no deferral. Or trend in analysis >60% average Trend in analyses up to 90% of alert level, action → shutdown/recall immediately/immediate urgent maintenance
It is meaningful to analyze trends in the wear element test data, which monitors the deterioration of the oil condition. Absolute and rate of change data concentration values can be used to assess the deterioration of a lubricant or hydraulic fluid — the ideal scheme, with regular sampling, servicing, and replenishment at preprogrammed intervals. It is rare for this regularity to hold; the reality is that sampling/servicing and replenishment of fluids occur irregularly and this must be adjusted numerically in the trend data. From these “trending analyses,” element concentration indicators can be developed by various statistical methods using system failure modes to set individual wear metal levels at which corrective or remedial measures must be taken for the deterioration of the lubricant, such as in Table 29.3.
29.3.3 Particulates and Ash in Lubricants The accurate measurement of particulates and ash in a lubricant sample is very important in assessing its deterioration for the excessive build-up of soot, dirt, or particulates in general can prevent the normal protective function of that lubricant. The term “particulates” covers a wide range, including insoluble matter, sediments, and trace metals as very fine diameter particulate. Larger metal particles such as metal flakes and spalled debris are not covered, these being covered by separate analyses and filtration. 29.3.3.1 Dirt and Particulates in Lubricants Controlling the cleanliness of any lubricant or hydraulic system as it deteriorated with use was very important in the past and will be even more important in the future, because of the following reasons: • System reliability is increasingly important and a major contributor to equipment failure is particulate contamination in the system operating fluid. • Systems perform at higher energy levels for longer periods and maintained to be “cleaner” so as to deliver that performance. • Equipment tolerances are decreasing for high precision components (∼5 µm clearance or less) and in automotive and hydraulic components they are increasingly common. Smaller particulates, for example, 2 µm dependent upon its nature, can agglomerate and clog sensitive components such as control and servo valves. • For automotive applications, two trends lead to increased particulate levels: Exhaust gas recirculation, for environmental exhaust emission reduction, primarily for NOx, having the additional beneficial effect for emissions of depositing particulate into the lubricant rather than being emitted. However, this creates a problem of enhanced particulate levels for the lubricant. Strong consumer pressure for increased service intervals, already up to and beyond 50,000 mi (80k km) for trucks and 30k mi (∼50k km, or every 2 yr) for some new 2005 light vehicles. Lubricant must last longer and yet meet enhanced performance standards. Enhanced levels of particulate are now envisaged, well above 1%, up to 2 or 3%, a steep challenge for the lubricant to remain effective under these conditions.
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29.3.3.2 Useful Definitions “Particulates” and “dirt” are descriptions, which require a more precise description and definition, as follows: • “Particulates” are small, up to 15 µm maximum, either carbonaceous, inorganic compounds or fine metal particles, where the metal particulates result from “rubbing wear.” • “Dirt” is road dirt ingested by faulty induction air filters, poor seals or defective/absent air breather components; the parts that survive are usually hard particles such as silicates (from sand, etc.). • “Metal debris” is comprised of larger metal flakes or spalled particulates resulting from catastrophic micro-failures or incipient major failures such as parts of gear teeth being separated. Hydraulic fluids develop haze or very light deposits over a considerable time of their service life, petrol/gasoline engines develop black particulates slowly over their service life, while diesel engines rapidly develop black particulates. The operating limit of circulating lubricant filters is in the range 10 to 15 µm whereas it has been shown that removal of the “larger,” >10 m particles from a circulating lubricant system can reduce catastrophic bearing failures by 25%. Further, for hydraulic systems it is claimed that 80% of failures can be avoided if particulates >5 µm are removed by filtration. While not going to these levels of filtration, higher levels of filter efficiency are now incorporated into new designs. This must happen to meet the enhanced levels of filtration required over the enhanced periods of service operation. However, one problem is that the enhanced levels of filtration can remove the small, fine, metal particulate, which is used for wear data and trend analysis. 29.3.3.3 Particulate Analyses There are a number of measurements available for the measurement of soot in lubricants. These measurement methods can be grouped into three categories, as where the particulate is: 1. Removed from the liquid, then oxidized while measuring the mass loss. 2. Separated by addition of solvents to the lubricant sample and the precipitated mass measured. 3. Measured within the neat, or diluted, lubricant sample for absorbance, scatter, or obscuration at a given wavelength. 29.3.3.4 The Enhanced Thermogravimetric Analysis, ASTM D5967 Appendix 4 (Colloquially Known as the “Detroit Diesel Soot Test”) Total particulate in a degraded oil sample is determined by thermogravimetric analysis (TGA), where 20 mg of oil in a pan on one arm of an electronic balance is heated under programmed temperature furnace environment in a nitrogen atmosphere. Differentiation is made between carbonaceous and incombustible ash by increasing the temperature and changing to an oxygen atmosphere. A 20 mg sample is larger than normal but is necessary because the final objective, the soot content, will be less than 1 mg. The temperature environment is held at 50◦ C for 1 min, raised to 550◦ C at a rate of 100◦ C/min, maintained isothermally for 1 min, and then raised to 650◦ C at 20◦ C/min. The method considers the residual sample at this stage to be composed of soot and incombustible material with liquid hydrocarbons removed. The atmosphere is then switched to oxygen and the furnace temperature raised to 750◦ C at 20◦ C/min and maintained for a stable weight for at least 5 min. The changes in weights at different temperatures and atmospheres are due to soot being the difference in weight between 650◦ C in nitrogen and 750◦ C in oxygen. The residual material is incombustible ash and metallic residues, assuming that all of the remaining lubricant base stock is driven off and oxidized at the higher temperatures under oxidizing conditions. 29.3.3.5 Optical Particulate Measurements A very desirable feature in particulate measurement is a linear relationship between particulate signal, by light absorption or scattering, and particulate concentration. This relationship generally holds as a linear relationship of a certain slope up to ∼1.5% particulate concentration, followed by a linear relationship with a higher slope at higher particulate levels. While an overall linear relationship is very desirable,
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Absorbance at 500 nm
1.0
CEC L-82-A-497 Optical particulate measurement sample dissolved in toluene
0.5
Particulate concentration
FIGURE 29.4 toluene.
CEC L-82-A-497 calibration plot for particulate determination in degraded lubricant by dilution in
the major problem is the change in relationship between signal, however derived, and particulate concentration in the region of 1.5% concentration. Two methods measure sample particulate concentrations, one infrared by direct sample absorption and one in the visible by dilution in toluene. The visible method, capillary electrophoresis chromatography (CEC) L-82-A-497, “Optical Particulate Measurement” dilutes the degraded oil sample in toluene, a solvent, which disperses all of the particulate, and then measures the absorbance of the diluted solution at 500 nm in a spectrophotometer. Standardization uses a lubricant or hydraulic fluid sample of known pentane insolubles content to construct a calibration curve, Figure 29.4. The method is quick, repeatable, and accurate, provided that the sample disperses well and does not cause light scattering, which will add to the apparent OD. This method was adopted by the CEC to measure soot developed in lubricant samples from the Peugeot XUD11BTE engine test and uses 0.1 g of oil sample in known aliquots of toluene. The solvent aliquot volume is increased to bring the OD within an acceptable range. The OD plot for lubricant samples dispersed in toluene and measured at 500 nm should be linear with a high correlation coefficient. The only drawback is that some additives or degradation products may cause light scattering and an incorrect result. 29.3.3.6 Infrared Measurements at 2000 cm−1 Soot does not absorb in any specific region of the infrared region but as small particulate scatters the incident radiation in a nonphotometric manner. Theoretically, light scattering of a spherical, uniform diameter, particulate is proportional to the fourth power of the wave number. From this, the background scatter in the infrared spectrum of a used lubricant containing particulate should decrease across the infrared region from 4000 down to 400 cm−1 . Background scatter does decrease but not as much as predicted by theory, probably because the particulate is not monodisperse and certainly not spherical. 2000 cm−1 is the chosen measurement point because there are no absorbing groups present in lubricants. Increase in lubricant absorption at 2000 cm−1 with engine run time are mainly dependent upon the mass of soot particulate present, with second order effects due to the effective particulate size and shape, therefore somewhat dependent upon engine type. High levels of soot particulate give high absorbance levels and inaccuracies in spectrophotometry, which can be overcome by using thinner path length cells. The results are in absorbance and need calibration for percentage soot. The advantage of the method is that it is a direct measurement on the sample, without the effects of adding solvents, the like, and that it arises from infrared measurements, which could be undertaken for another set of measurements in any case. The disadvantage of the method is that the sample spectra need to be the difference spectra, that is, the difference between the engine test run samples and the original, fresh oil, which may not always be available.
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29.3.3.7 Particle Size Distribution A more fundamental view of the nature of particulates in degraded lubricant and hydraulic fluids is the distribution of particulate sizes. This can be either done continuously, by light scattering, or discontinuously by a range of physical filters. The latter is self-explanatory and the first needs explanation. Particles suspended in a medium scatter incident light at an angle dependent upon particle size and also upon the wavelength of the incident light. The second is simplified by using a monochromatic source such as a laser. The particles are assumed to be spherical, a very broad-brush approximation. A variable correction factor is needed for the nonspherical nature of the particulates, such as a rodlike nature with a defined length/width ratio. Particulate light scattering optics uses a collimated laser light source, usually a He/Ne (red) laser, expanded by a lens into a broad beam, which diffuses the sample cell. The light is dispersed/scattered by the suspended particulate in the sample cell and then collected by a similar second lens and focused onto a detection plate. The detection plate samples the intensity of the scattered light at a large number of points and transformed into a particle size distribution by suitable software. The resolution of the method depends upon the spatial discrimination of the detector plate. Particle size distributions for a range of samples from engine runs using a range of related lubricant formulations show that these particulate distributions are interdependent, the smallest particulate size distribution leading to the successive growth of the larger particle size distributions. The interdependence of these particulate distributions measures the effectiveness of dispersants, for the particulate can successively agglomerate from the initial size of around 0.1 µm diameter to 1 to 7 to 35 µm and then larger diameters. If the dispersant within the lubricant is not degraded then the agglomeration process will be stopped or reduced. 29.3.3.8 Particulates in Hydraulic Fluids Hydraulic fluid cleanliness is crucial to the continued operation of hydraulic systems, avoiding component damage and failure. The level of cleanliness is many orders of magnitude down (better) from that accepted for lubricants. Instead of values of mass particulate, the emphasis for hydraulic fluids is on the number of particulates in the range of 2 to 15 µm, a range correlated to the probability of component problems. With this stimulus, several methods of electronic particle number counting have been developed, based upon the following: • Light absorption • Flow decay • Mesh obscuration These methods are continuous and easy to use; their main problem is the large amount of data that they generate for the size and number of particulates but without reference to the composition of those particulates. Wear metal or chemical analytical data is required to properly understand the complete picture of particulate composition in hydraulic fluids. A fundamental problem is the lack of suitable, repeatable reference standards. When used for equipment monitoring, it is very important that the response of the counter has a high particle size correlation with the size of particles, which cause damage to the fine tolerance components of the system. 5 µm diameter was regarded as the lower limit of damaging particles until recently, but this is now reduced to 2 µm as an indicator of potential damaging conditions, approaching the limit of discrimination between two such particles. One type of mesh obscuration particle counter uses three successive micro-screens of 15, 5, and 2 µm pore size, Figure 29.5. Laminar fluid flow through this array of screens generates pressure drops, caused by oversized particles partially blocking the respective pore size filter, recorded by differential pressure transducers. Count data from hydraulic samples is statistically derived through correlation with data from a calibration standard. This counter is effective for most oils of different levels of obscuration (lightblack) and is relatively insensitive to other counter-indicators such as entrained water and air in degraded lubricant samples. Another method of electronic particle size counting uses the blocking behavior of a particle size distribution in a degraded lubricant sample passing through a single, monosized micro sieve of either
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Handbook of Lubrication and Tribology Laminar flow
15 mm screen
5 mm screen
2 mm screen
FIGURE 29.5 The principle of the mesh obscuration particle counter. From Machinery Oil Analysis — Methods, Automation and Benefits, 2nd ed., Larry A. Toms, Coastal Skills Training, Virginia Beach, Virginia, USA (1998). With permission.
Laminar flow
Micro-screen
Plunger
FIGURE 29.6 The principle of the flow decay particle counter. From Machinery Oil Analysis — Methods, Automation and Benefits, 2nd ed., Larry A. Toms, Coastal Skills Training, Virginia Beach, Virginia, USA (1998). With permission.
15, 10, or 5 µm pore size, Figure 29.6. A correlation is assumed between the particle size distribution of an unknown sample and that of a standard. The measured parameter is the differential flow across the micro-screen, which converts flow decay measurements to an ISO cleanliness code. An optical particle counting method uses a path of collimated light passed through a hydraulic oil sample and then detected by an electrical sensor. When an translucent sample passes through the sample then a change in electrical signal occurs. This is analyzed against a calibration standard to generate a particle size and count database, linked to an ISO cleanliness value. Light absorption particle counter’s output values are badly affected by the following factors: • The opacity of the fluid raising the background value to the level that the instrument no longer works, overcome by sample dilution with a clear fluid. • Entrained air bubbles within the sample are counted as particles, which confuse the system, and are removed by ultrasonics and vacuum treatment. • Water contamination is more difficult to deal with, causing increased light scattering. But significant levels or water, such as >0.1 or >0.2% levels, will fail the oil anyway. The continued monitoring of particle cleanliness in hydraulic fluids within systems is a very important process to maintain the integrity and performance of complex hydraulic systems. 29.3.3.9 Ash Content The “Sulfated Ash” content of a lubricant is an important property and can be included under particulates in degraded lubricants. It gives a meaningful indication of the detergent additive content and is useful as a control test in the oil blending process. While it is a property only normally used for new
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formulations, results for degraded lubricants have considerable interference from both wear metals and other contaminants. The problem with sulfated ash arises from inorganic compound deposits in the ring zone and on the piston crown. The problem becomes very important when extensive deposits build up on the piston crown from low/medium power level operation, such as for a taxi engine in town. However, when such an engine is used at extended higher energy power levels, such as extended motorway journeys, the deposits on the piston crown become very hot, retaining heat and glow. They can become so hot that they melt part of the piston crown to the extent of penetration, that is, a hole, causing catastrophic deterioration of the engine, which is the downside of sulfated ash content. The upside of metallic detergent inclusion into lubricant formulations is their ability to reduce the deposition of carbonaceous substances and sludges in the ring zone and piston crown. The essence of the problem is to balance the level of metallic soap sulfonate in the original formulation and the amount of sulfated ash that results. Sulfated ash is a major contribution to the overall formation of ash, contributing to crown land deposits above the piston rings, valve seat deposits (and thus leakage through seat burning), and combustion chamber deposits. These deposits cause preignition of the gasoline/air mixture, leading to a decreased fuel octane rating for the same engine called octane rating decrease (ORD). It is beneficial to reduce the impact of this effect by minimizing ash deposits. Ash formed from lubricants can also contribute to diesel engine particulate emissions. Recalling that the sulfated ash content is important for new lubricants, the simplest test is the ASTM D842 Ash Test where the ash content of a lubricant is determined as a weighed sample, to constant weight, of oil burned for 10 min at 800◦ C. The mass measured is that of the incombustible solids, be they wear metals or other incombustibles such as fine metallic particles or silicaceous dust. The ASTM D874 Ash Test is an improved ASTM D842 method in that the oil sample is combusted until the carbon residue and metallic ash is left. Sulfuric acid is added, the sample is reheated and weighed to constant values. The last stage converts any zinc sulfate to zinc oxide. The sulfated ash tests indicate the concentration of the metal-based additives in fresh lubricant blends. Problems arise from (i) any phosphorus present forming pyrophospates of variable composition, giving higher and more variable results and (ii) magnesium sulfate being variably converted to its oxide. Carefully conducted, the sulfated test gives a reasonable measure of additive metals present in a lubricant formulation. The weight of metal present can be converted to the expected sulfated ash content by the conversion factors given below: To Estimate Sulfated Ash Content from Metal Content: Metal Conversion Factor — Metal % to Sulfated Ash Zinc Sodium Magnesium Calcium Barium
1.25 3.1 4.5 3.4 1.7
If the lubricant has been formulated with magnesium-based detergents or boron-based dispersants, then these methods of sulfated ash are unreliable. The sulfated ash test is also unreliable for used lubricants, due to the following reasons: • The presence of incombustible contaminants. • Additives will be degraded during service life and are thus changed chemically but the constituents will continue to appear in the ash residue at the same concentrations as for the new oil. • A trend toward ashless detergents, which undermines the relevance of the sulfated ash test as a measure of detergent in a formulation. It is important to check the sulfated ash method against reference blends wherever possible.
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29.3.4 Acidity and Base Reserve Determining the alkaline reserve or acid content of a degraded lubricant fluid should be straightforward by analogy to acid/base titrations in water. But this is the simplistic point that causes so many problems with determining “Base” and “Acid” numbers in degraded lubricant and hydraulic fluids. To thoroughly understand “Base Number,” an appreciation is needed to determine the following: • • • •
How it arises How it has been, and is currently, measured The problems of those analyses What this means for lubricant use/extended use and condition monitoring
While the idea of a “number” is simplistic and therefore appealing, the reality is complex and we need to look at the points made above, in order. 29.3.4.1 The Need for Base Number Measurement The need to measure the “Base Number” in some form as a property of a lubricant/degraded lubricant arises from the acidic products formed during the service life of that lubricant. The acid formation process can be rapid or slow, according to the stress that the lubricant is exposed to. The emphasis must be on the effect that the “service life” of the lubricant involves, in terms of either high temperature and pressure or over a short and intense, or a very long-term and less severe, service interval. The starting position is that most lubricant base fluids have some, maybe greater or lesser, basic properties that neutralize acidic components introduced into them. As the performance requirements of lubricants developed, it became evident that the naturally occurring antiacidic properties of unmodified base stocks were not sufficient to prevent lubricant and hydraulic oils becoming acidic and corroding the components of the system. The development of detergent additives had two effects: • The organic nature of the additives themselves had an additional, but marginal, antiacid contribution. • However, more importantly, the detergent additives had the ability to solubilize as inverse micelles alkaline, inorganic material such as calcium oxide/carbonate or the corresponding magnesium salts (much less used). These compounds react with acidic products formed in the lubricant to produce neutral salts, which bind the acidity as an innocuous product. Barium compounds are not used now because of toxicity problems. 29.3.4.2 Sources of Acidity-Induced Degradation Acidity in lubricants arises from two sources: • The (declining) sulfur content of fuels, forming sulfur oxides, primarily sulfur dioxide, SO2 . • The reaction (“fixation”) of atmospheric nitrogen by reaction with atmospheric oxygen in the high temperatures, 2000 to 3000◦ C, of the combustion flame front, forming nitrogen oxides such as NO, nitric oxide, and nitrogen dioxide, NO2 , primarily the former, which then slowly oxidizes to the dioxide. Sulfur and nitrogen dioxides, SO2 and NO2 , dissolve in any water present to give the mineral acids of sulfurous/sulfuric and nitrous/nitric acids. The two forms of each acid are given because the dioxides initially dissolve in water to give the first, weaker, acid and then oxidize to the stronger, second acid. Organic acids are formed by the partial oxidation of hydrocarbons. Normally, hydrocarbon oxidation is considered as going through to complete combustion with water and carbon dioxide as the final products. But combustion/thermal degradation can be partial, with hydrocarbon end groups forming carbonyl
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groups to make aldehydes, ketones, and carboxylic acids, the last as: R—C=O | OH Organic acids are not normally regarded as strong acids; acetic acid has a dissociation constant in water of 1.8×10−5 at 298 K and is regarded as a weak acid, the prime constituent of cooking vinegar. But various R-group substituents can increase the dissociation to make the acid stronger, such as for trichloroacetic acid. Two points to particularly consider for the strength of organic acids: 1. Acid dissociation constants increase with temperature, the higher the temperature, the stronger the acid. 2. The value given is for acetic acid in water. Acid : base interactions and equilibria are considerably different in other solvents, often making organic acids stronger. Applying these to organic acids in degraded lubricants, the lubricant is a drastically different solvent to water, which also operates at high temperatures. As an example of the strength of organic acids, the railways originally lubricated their steam engine cylinders with animal fats before hydrocarbon oils were available. The high steam temperatures within the cylinders degraded the fats into their constituent organic acids, which corroded the metals present, particularly the nonferrous metals such as copper, lead, zinc, and so on. The acidity generated within a degraded lubricant during its service life is a mixture of inorganic strong acids and weaker organic acids. This mixture is one of the causes of the analytical problems in determining the acidity of both the acids in, and the remaining alkaline reserve added to neutralize that acidity, in a lubricant formulation. This is the need to determine the Base Number in a lubricant, both new and used. It is a standard analytical measurement for degraded lubricants. 29.3.4.3 Measurement of Base Number An acid is normally associated with the bitter, corrosive, sometimes fuming in their concentrated form, properties of the mineral acids, classically sulfuric, nitric, hydrochloric, and phosphorous acids. There are others but these are the common mineral acids. Their common property is the ability to donate/give − a proton, H+ , to a base. Sulfuric acid then becomes an anion, such as sulfate, SO2− 4 , nitrate, NO3 , 3− − chloride, Cl , or phosphate, PO4 . The common bases as alkalis, such as sodium hydroxide, caustic soda, potassium, and ammonium hydroxides are strong bases with sodium carbonate as a mild alkali or weak base. Again, as for the acids, there are many others but these are the commonly used alkalis. The common feature of alkalis is the hydroxide group, OH− , which accepts the proton from the acid to form water, H2 O. Aqueous acids and bases in equal amounts neutralize each other to form a neutral salt and water, as in the standard neutralization of hydrochloric acid by sodium hydroxide: HCl + NaOH → NaCl + H2 O Whichever way this is done, by adding acid to alkali or the reverse, for equal amounts of acid and alkali, the end result is a neutral solution of pH 7. If the strength of one of the solutions is accurately known, then the concentration of the other solution can be calculated — basic chemical laboratory work. Neutralization is shown by an indicator with different colors in acid or alkaline solution, neutralization being shown by a color balance between the two forms. Litmus is one example of a neutralization indicator, being blue in alkaline and red in acid solution. Progress of acid/base titrations can equally be followed by other methods, such as: • The pH electrode combined with the standard calomel electrode to follow either the solution pH or the potential difference in millivolts, mV, between the electrodes.
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• The electrical conductivity of the solution between two platinum plate electrodes, because both the proton, H+ , and the hydroxide ion, OH− , have high conductivities relative to other ions and both H+ and OH− are at a minimum at the end point, pH 7. From these fundamental considerations, if the alkaline reserve (Base Number) of a degraded lubricant is a base, then it should be possible to titrate it against a standard acid solution to determine how much base is present. That is, the basis of Base Number determination, transferred over from water-based acid/alkali neutralizations to the analysis of new and degraded lubricants in a variety of organic solvent mixtures. Many acids have been used to titrate the alkaline reserve in a lubricant but they give different values, particularly for heavily used samples. 29.3.4.4 IP 177/ASTM D664 — Base Number By Hydrochloric Acid Titration This is a joint method developed by the Institute of Petroleum in the United Kingdom and ASTM in the United States and was the earliest methods for measuring the base content of a new or degraded lubricant or hydraulic fluids. It is still preferred by some operators and has essentially been reintroduced by the IP 400 method; see Section 29.3.3.7 later, with the same solvent and acid titration system but with a different detection system. The solvent for the titration of the lubricant/hydraulic sample must dissolve the sample and be compatible with the titrating acid. In this case, it is a mixture of toluene, isopropyl alcohol and a very small amount of water. The acid is dissolved in alcohol and the two solvents are completely miscible. The progress of the neutralization reaction is followed using a combination of a glass electrode and the standard calomel electrode, a standard nonaqueous solvent analytical procedure. The signal used is the potential difference between the electrodes expressed as mV. The neutralization works well for new and slightly used lubricants. The mV difference signal varies as a sharp sigmoidal form when mV is plotted against acid titration volume, Figure 29.7. The neutralization endpoint is at the mid-point of the sharp rise, as indicated. There is no problem with the analysis for new and lightly degraded samples, the neutralization curve is sharp, and the endpoint is clear. Problems arise as more extensively degraded lubricants are analyzed. The clear form of the neutralization curve slowly degrades with increased degradation of the lubricant sample until its form is lost and there is no clear endpoint, Figure 29.8. A procedure is suggested where an endpoint value to work to is used instead, but this is an unsatisfactory solution.
0.700
0.200
0.600
0.160
0.120 Endpoint
dE/ V
dE/ V
0.500
0.400
0.080
0.300
0.040
0.200 0.00
2.00
4.00
6.00
8.00
0.000 10.00
V/ml
FIGURE 29.7 method.
mV vs. volume plot for the titration of new/slightly degraded lubricants by the IP 177/ASTM D664
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0.400
0.060
0.340
0.048
0.280
0.036 dE/ V
dE/ V
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Endpoint
0.220
0.024
0.160
0.100 0.00
0.012
2.00
4.00
6.00
8.00
0.000 10.00
V/ml
FIGURE 29.8
mV vs. volume plot for the titration of heavily degraded lubricants by the IP 177/ASTM D664 method.
There are several strong arguments against the use of the IP 177/ASTM D664 method for Base Number: • The hydrochloric acid has an acid strength in the solvents used in this method, which only reacts with, and therefore determines “strong alkalinity,” >pH 11, in the lubricant sample. It does not determine “mild alkalinity,” up to pH 11, although it is not clear whether this is a crucial difference. • The method has poor reproducibility, although this is improved by using the replacement ASTM D4739 method, which uses a very slow potentiometric titration, 15 min/1 ml acid reagent added — an extremely slow method. • The sensitivity and fragility of the electrodes is important, the glass electrode is particularly fragile. Replacement glass electrodes must always be available, “conditioned” in the reaction solvent and ready for use. Another problem is that the electrode surfaces are gradually fouled by carbonaceous particulate in degraded lubricant samples and the electrode must be replaced. • This method is not unique as against the others, but all Base Number methods use chemicals with various forms of hazards, which are expensive to dispose of. The formal method uses a large test sample, 20 g, in 120 cm3 of solvent, the volume of which is increased by the ensuing titration. The test results are presented as milligrams of potassium hydroxide per gram sample equivalent. When applied to analyze successively degraded lubricant samples from engine bench or field tests, the IP 177/ASTM D664 Base Number method results tend to decline quickly in the initial stages of the test and then to decline more slowly, Figure 29.9, in contrast to results from other methods. It is generally held that a lubricant with a Base Number approaching a value of 2 should be replaced. Therefore, a Base Number of 2 for a degraded sample shows that its alkaline reserve equates to 2 mg potassium hydroxide per gram of sample. While the titration uses hydrochloric acid, this is related to its equivalent as potassium hydroxide. To sum up, the IP 177/ASTM D664 method suffers from the following: • • • •
Poor reproducibility, particularly for heavily degraded samples Lack of clarity in what it means Fragile apparatus Requiring large sample masses and solvent volumes
29.3.4.5 IP 276/ASTM D2896 Base Number By Perchloric Acid Titration This method is really a modification of the previous IP 177/ASTM D664 method, arising from the perception that changing the titrating acid from hydrochloric to the stronger perchloric, HClO4 , will react
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Base number
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IP276/D2896
IP 177/D664
Degraded lubricant, service life, h FIGURE 29.9 Base Number degradation values for same successive lubricant samples by IP 177/D664 and IP 276/D2896.
with both strong and mild alkalinity in lubricant samples. It is argued that the results using perchloric acid reflect the total additive content of the formulation. To accommodate the change in acid, the solvent must be modified as well and is a mixture of chlorobenzene and glacial acetic acid. The detection method is the same as for IP 177/ASTM D664, a combination of the glass electrode and the standard calomel electrode. The titration is the same and the plot of mV against acid volume has the same sigmoidal shape as given in Figure 29.7. Unfortunately, the method suffers from the same problems for heavily degraded samples, the plot then becoming indistinct with no clear endpoint, as in Figure 29.8. In this case, reproducibility is as poor as for the IP 177/ASTM D664 method. In this case, the method suggests a “back titration” with a much poorer range of reproducibility and repeatability. When this method is used to analyze degraded lubricant samples from engine bench or field tests, the IP 276/ASTM D2896 method Base Number results decline slowly throughout the test, in contrast to the results for the same samples analyzed using the IP 177/ASTM D664, as set out in Figure 29.6. There is no sharp decline in the initial stages of the test. There is a clear difference in results from the same samples between the IP 276/ASTM D2896 and the IP 177/ASTM D664 methods. As before, it is generally held that a lubricant with a Base Number approaching a value of 2 should be replaced. The test results have the same values as for the IP 177/ASTM D664 method. The solvents and chemicals used in IP 276/ASTM D2896 are even more hazardous and difficult/expensive to dispose of, than those used in the preceding IP 177/ASTM D664 method. The following summarizes, the IP 276/ASTM D2896 method: • It is a modification of the previous IP 177/ASTM D664 • It gives generally higher Base Number values, said to reflect the total, strong, and mild together, alkalinity present in a lubricant formulation • It has the same problem of an indistinct endpoint for heavily used samples • The solvents and chemicals used are hazardous and difficult/expensive to dispose of
29.3.4.6 ASTM D974 — Base Number by Color Indicator This method is worth noting but is now relatively little used. The method is very similar to IP 177/ASTM D664 method but uses an naphtholbenzein indicator color change to determine the neutralization endpoint. The results are expressed in the same way as IP 177/D664.
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Conductivity meter
Solution
Platinum electrodes
FIGURE 29.10
Diagram and picture of conductimetric cell.
29.3.4.7 IP 400 — Base Number by Conductimetric Titration IP 400 is relatively recent (there is no equivalent ASTM method) and directly addresses the problems of the previous methods. Chemically, it is identical to IP 177/ASTM D664 but the crucial difference is that it uses a conductimetric detection method to follow the progress of the neutralization reaction. It measures the resistance, or its inverse, the conductivity, of a solution between two platinum plates rigidly held in a glass tube, shown both as a diagram and picture in Figure 29.10. The plates are typically 10 mm square, welded to platinum wires, which exit through the wall of the glass tube to external connections. The conductimetric probes are very robust and work as well when bright metal or when coated with carbonaceous particulate. The only problem with electrode contamination occurs when the carbonaceous particulate coats the wall of the glass probe containing the electrodes sufficient enough to cause an electrical short circuit. The conductimetric cell does not need to be a special model, excellent results can be obtained using standard cells as used in initial physical chemistry laboratory experiments. Special cells are only constructed for automated systems, which use small volumes of sample, solvent, and titrating acid solution. The conductivity of the solvent plus sample is low, of the order of 2 µS (microsiemens) and increases linearly as the titration proceeds, Figure 29.7. At the endpoint, the gradient of the linear plot changes sharply. The endpoint is determined by the intersection of two linear plots, as shown in Figure 29.11.
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Handbook of Lubrication and Tribology 16 Conductance, mS
14
Conductance, mS
12 10 8 6 4 2 Endpoint 1.97 0
FIGURE 29.11
0
0.5
1
1.5 2 2.5 Volume HCI added, ml
3
3.5
4
Conductivity vs. titration volume plot for IP 400.
The crucial point about IP 400 is that the quality of the endpoint does not change with the sample condition, either new, lightly used, or heavily used, as shown in Figure 29.11. The intersection of the linear sections moves to lower values as the alkaline reserves of the samples reduce. Reproducibility is good, within the limits set for IP 276/ASTM D2896, with no deterioration with sample use, as shown. The test results are very close to those obtained using the IP 177/ASTM D664 method under its best conditions, which is not surprising as the chemistry is the same. Smaller volumes are used, the sample weight specified by the IP 400 procedure is 5 g but very good reproducibility and repeatability have been achieved down to 0.1 g for small volume and unique samples. The titration is relatively quick compared to the previous potentiometric methods, for during the titration the solution conductivity stabilizes as soon as the added aliquot of acid is thoroughly mixed. The IP 400 procedure is simple and straightforward. This demonstrates that the problems of the previous Base Number methods are associated with the following: • The potentiometric electrodes and their physical reactions. • The potential difference titrations of a number of substances in the used lubricant samples against their total conductivity. 29.3.4.8 Precision of Base Number Determinations The precision of these determinations has the following two forms, from cooperative test programs carried out between participating laboratories: 1. Repeatability, by the same operator, same laboratory 2. Reproducibility, by different operators and laboratories The format of precision is interestingly different from that normally encountered. It is set as a requirement that the results on the same sample should not vary by more than the stated limit values more than 19 cases out of 20, an interesting approach to a 95% confidence limit.
Repeatability For IP 177/ASTM D664, Base Number, By Manual Methods 7 mg Automatic Methods 6 mg
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Reproducibility 20 mg 28 mg
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Repeatability For IP 276/ASTM D2896, Base Number, New Lubricants 3% Used Lubricants 24% For IP 400 Base Number New and Used Lubricants 0.17x1/2 where x is the average of the results
29-27
Reproducibility 7% 32% with back titration 0.31x1/2
Note that there is no distinction in precision between new and used lubricant samples for the IP 400 conductimetric method for Base Number determination. 29.3.4.9 Fourier Transform Infrared Spectroscopy Methods Fourier transform infrared spectroscopy (FTIR), has been applied to the analysis of degraded lubricant and hydraulic fluids. The method is not direct in the sense of reading a value from a scale. The analysis is conducted indirectly by first obtaining various parameters derived from the difference spectra between the time sample and the original lubricant. Multivariate analysis and principal component regression (PCR), are then applied to these parameters to determine the Base Number. The overall process is now well established as a technique for measuring used lubricant properties. It has considerable potential as a nonwet, relatively “dry” method, which does not need to use hazardous laboratory chemicals. 29.3.4.10 Summary for Base Number Measurements For the reasons developed above, the “Base Number” value for a degraded lubricant is not a straightforward measurement. Any quoted values are not absolute and must be related to the method used to determine that value. The problems of the IP 177/ASTM D2896 and IP 276/ASTM D2896 methods lie with: • Repeatability/reproducibility difficulties introduced by the potentiometric electrode reactions. • Also, the interpretative differences seen between “strong” and “weak” alkalinity. The ASTM D974 colorimetric method is rarely used. The IP400 method is much better for Base Method measurement because of its clarity of endpoint, which is sustained for new, somewhat degraded and heavily degraded lubricant samples. The FTIR difference analyses combined with chemometric statistical analyses can predict Base Number and show considerable promise with very substantial reductions in the use of solvents and reactants. 29.3.4.11 Sources and Effects of Acidity In addition to acidity caused by combustion of inorganic compounds to give mineral acids, as described in Section 29.3.4.2, and also the oxidative degradation of hydrocarbon fuels and lubricants, hydraulic fluids will also degrade through localized high temperatures. Localized high thermal stress on a hydraulic fluid will, in due course but over a considerably longer period than for lubricants, cause thermal degradation and oxidation. These conditions will cause the physical properties of the hydraulic fluid to go outside its specification and it must be replaced. Acidity in degraded lubricant and hydraulic fluids corrodes system components. Corrosion combined with erosion gives enhanced wear rates, particularly in systems with mixed metals in contact by electrochemical effects. Corrosion can also generate solid debris within the system leading to clogging of tubes, filters, and obstruction of system operation. In collaboration with water, corrosion leads to rust formation.
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The areas that are prone to acidity attack are (i) bearing corrosion and (ii) cam and tappet corrosion and rusting. It may be surprising that acidity and alkalinity can exist together in a used lubricant. It is again helpful to go back to the explanation given in Section 29.3.4.2. The effect of hydrocarbons as solvents with their low permittivities (dielectric constants) has the effect of giving a greater range of acidity and also alkalinity to the components present in the system. Acid-base interactions can range from (i) complex formation, AHB- of acid, AH, and base, B-, together, with the acid proton shared between the acid and base, and (ii) to the full transfer of the proton from the acid to the alkali as normally understood by “neutralization.” In the former case, the substance can be both acid and basic (alkaline). Therefore, Base and Acid Numbers can coexist in the same system. Generally, in a new automotive lubricant sample, the Base Number will be high, of the order of 6–10 KOH units and the Acid Number can be of the order of 0.5–1.0 units. Acid Number is the corollary of Base Number but has not been subject to the same level of controversy as described for Base Number. Only recently, has there been difficulty with the Acid Numbers of synthetic ester lubricants used in gas turbine engines, addressed by an sampling and analytical error (SAE) method. Developments in applying conductimetric methods to the determination of Acid Number are also discussed as part of a method to determine Base and Acid Numbers sequentially for the same sample in the same apparatus.
29.3.4.12 Acid Number Determination by IP 177/ASTM D664 This method is directly analogous to the Base Number determination described previously in Section 29.3.4.4. The solvents for the sample are the same, a mixture of toluene, isopropyl alcohol, and water with the titrant being potassium hydroxide in alcohol. The method follows the neutralization of the sample solution by alcoholic alkali by using the glass and standard calomel electrode pair, giving a millivolt potential difference, between the electrodes against titration volume, V . The form of the titration is again a sigmoidal curve, with the endpoint at the change of gradient, the point of inflection, at the center of the sigmoidal plot. The endpoint is more clearly shown by the first derivative, d(mV )/dV plotted against V , Figure 29.12, which is an appropriate repeat of Figure 29.7.
Endpoint by first derivative
Electrode potential, mV
E(V )
E(mV )
E(V ) V(ml) Titration volume, ml
FIGURE 29.12 Acid Number determination by the IP 177/ASTM D664 method, mV vs. volume plot and first derivative plot.
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The test results are presented as milligrams of potassium hydroxide per gram sample equivalent, the same as for Base Number. The limits of repeatability and reproducibility are the same as for the determination of Base Number, as can be seen in the following table, as a requirement that the results on the same sample should not vary by more than the stated limit values more than 19 cases out of 20.
Repeatability By Manual Methods Automatic Methods
7 mg 6 mg
Reproducibility 20 mg 28 mg
29.3.4.13 D974 — Acid Number by Color Indicator Again, it is worth noting this method but it is relatively rarely used. It is very similar to IP 177/ASTM D664 but instead of a potentiometric method, it uses the color change of an indicator, naphtholbenzein, to determine the neutralization endpoint. The results are expressed in the same way, in milligram KOH per gram of sample. 29.3.4.14 Sampling and Analytical Error Determination of Acid Number in (Gas Turbine) Synthetic Ester-Based Lubricant Gas turbine lubricants are subjected to high temperatures in the center of the engine. But these lubricants are not exposed directly to combustion gases, as in a reciprocating engine. The high temperature within the central engine bearings causes breakdown of the esters to give acids and polyhydric alcohols and their degradation products. The organic acids give acidity to the ester lubricants, which cause corrosion to the engine unless controlled. It is important that this acidity is controlled for current gas turbines as future engines will operate at even higher internal temperatures and cause even more acidity. The results for Acid Number of the SAE method can be addressed by scrupulous attention to experimental detail in the IP 177/ASTM D664. 29.3.4.15 Simultaneous Conductimetric Determination of Base and Acid Numbers The conductimetric Base Number determination of IP 400 involves dissolving the lubricant sample in a toluene/isopropyl alcohol/water solvent and then titrating that solution with an alcoholic solution of hydrochloric acid, to give the well-known plot of Figure 29.13, Conductivity vs. Titration Volume. The sector A-B-C in Figure 29.13 is the Base Number titration, exactly the same as an IP 177/ASTM D664 titration for Base Number. This plot can be reversed by the addition of alcoholic alkali, which gives an almost exact symmetrically reversed plot, sector C-D-E. Further addition of alkali then titrates the original acidic content of the lubricant sample, sector E-F-G as the Acid Number titration. This method gives results within the IP and ASTM limits for repeatability. 29.3.4.16 Relationship Between Acid and Base Numbers of Degraded Lubricants The relationship between the Acid and Base Numbers for a degraded lubricant sample were developed for higher sulfur fuels and previous additive packages. Thus, the general rules, which developed were that if Acid Number rose to be greater than the declining Base Number, “crossing over,” then this was a condemning limit for the lubricant charge. Further, if Base Number declined below a value of 2, then this was a separate condemning limit for the lubricant charge. However, the gradual move to “low” and “lower” sulfur fuels for diesel fuel and, separately, modern additive packages can extend a system’s lubricant charge life. The condemning limits for degraded lubricants have changed considerably.
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Handbook of Lubrication and Tribology Br-tan oil Wayne kerr measurement of IP 177 BN/AN titration conductance term using isopropanol 50%/toluene 50% solvent
Mass of sample = 1.03 g BN
Ω–1 × 10–6
2.5
endpoint = 1.125 ml BN = 5.8
2.0 AN 1.5
endpoint = 2.65–2.12 = 0.53 ml AN = 2.8
1.0 0.5
1 V/ml HCI
2
1
2 V/ml KOH
3
4
Explanation of the BN/AN back titration curve KOH
Conductance
HCL
A
B
C Volume
D
E
F
The titration curve is seen to pass through a maximum of six different regions. The region up to point B represents the addition of HCI and from point B to point F represents the addition of KOH.
FIGURE 29.13
Sequential conductimetric acid and base titration of lubricant sample.
29.3.5 Water Content Water commonly contaminates machinery lubricant and hydraulic systems; its presence reduces the load carrying ability of a lubricant and increases wear. In addition, it promotes oxidation and corrosion. For synthetic polyol esters, water degrades the base stock back to its component acid and polyol. Maximum “safe” levels of water are usually taken to be 0.1–0.2%, higher for engines, lower levels for machinery and hydraulic systems. Water contamination of engine lubricant and hydraulic systems commonly arises from the following: • Combustion water, recalling that hydrocarbon combustion gives carbon dioxide and water as products. Some of the water passes into the crankcase as “blowby” down the side of the piston and condenses at the lower temperatures of that region of the engine. • Condensation of water in engines or hydraulic systems on standing or condensation into fuel tanks/hydraulic fluid reservoirs when operating at low/very low ambient temperatures. • Leakage into the fluids from cooling systems, such as circulating cooling water in engines by gasket failure, or leakage within the matrix of a heat exchanger. Almost all heat exchangers leak to some
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extent, acceptable ones leak very, very little but the leakage rate eventually increases with corrosion to become significant. • Water in lubricants degrades their formulation by absorbing acid gases to form strong acids. The presence of water in formulations can cause the additive package to precipitate out (“drop out”) as a severe form of degradation, which leaves the base oil only to lubricate the system. There are various methods to determine water in hydrocarbons and also lubricant and hydraulic fluids for the following reason: • The different nature of relatively pure hydrocarbons, as fuels, and lubricant and hydraulic fluids as complex formulations. • The different nature of the physical methods used to determine water in these fluids. • The varying nature of water at different concentrations. Water is a very complex physical substance for which complete models have yet to be accepted. Many models have been proposed for the physical properties of water but it is clear that “bulk” water, as a large polymeric but transient structure, has different physical properties from smaller groups of water molecules or, indeed, individual water molecules. Indeed, the infrared spectrum of very dilute water in organic solvents is a relatively narrow band centered on one frequency, whereas higher concentrations have appreciably wider bands at a shifted frequency. From these considerations, the various viewpoints of the methods used to measure the water content of lubricant and hydraulic fluids can be appreciated. 29.3.5.1 The IP 74/ASTM D95 — Water in Petroleum Products and Bituminous Materials by Distillation (the “Dean and Stark” method) The IP 74/ASTM D95 (“Dean and Stark Method”) for the determination of water in hydrocarbon fluids is a “total” method, rather gross and sensitive up to the 12% level. The method selectively distils water from petroleum products to separate and measure it using an organic solvent. It is an applied steam reflux distillation, which separates and concentrates the condensed water into a separate, calibrated, test tube, Figure 29.14. One problem in measuring the volume of water is complete separation of the water and hydrocarbon in the calibration test tube, which can be clear (complete) or hazy (incomplete), dependent upon the nature of the fluids and additive components present. The glassware apparatus for the Dean and Stark distillation is shown in Figure 29.11, note that the calibrated test tube in the system is positioned such that the water evaporated from the hydrocarbon fluid sample is collected and measured. The Dean and Stark method can be seen as a “total” water determination method as it collects all of the water from the sample that can be volatilized. Its limitation is that it uses an equilibrium water distribution between the (sample + organic solvent) and the (water + organic solvent), thus almost all of the water is removed to the measurement calibrated test tube. 100 g of oil sample is continuously distilled/refluxed with ∼100 cm3 of xylene, an aromatic solvent immiscible with water. The procedure is continued for 1.5 to 2 h to ensure that all of the water has been transferred. The percentage of water present in the sample is expressed as the volume of water in the graduated test tube multiplied by 100% and divided by the mass of oil sample. The method is direct with an unequivocal measurement of water but has the following disadvantages: • Lack of sensitivity • Occasional problems of measuring the water content because of incomplete separation of water/xylene in the measuring tube • The time of measurement, upward of 1.5 to 2 h per sample • Personnel intensive
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Condenser
Receiver
Distillation vessel
FIGURE 29.14
The Dean and Stark method apparatus for the determination of water in hydrocarbon fluids.
29.3.5.2 IP 356/ASTM D1744 — Determination by Karl Fischer Titration The Karl Fischer method of water determination is frequently discussed and results from it often quoted. It uses the reaction of water with iodine and sulfur dioxide in a pyridine/methanol solution, which is unpleasant to use. Iodine in a methanol/chloroform solution is an alternative reagent. The reagent reacts with hydroxyl groups, –OH, mainly in water but also in other hydroxylic compounds such as glycol, CH2 OH–CH2 OH, and depolarizes an electrode. The resulting potentiometric change is used to determine the endpoint of the titration and thus calculate the concentration of water in the oil sample. While the Karl Fischer method might be used to determine the water content of a formulated lubricant or hydraulic sample, it has never been approved for this purpose. The method was originally developed to determine the concentration of water in crude oil and can be used to determine water in fuels. When used to determine water in new and degraded formulated lubricants and hydraulic fluids, the method overdetermines a “water response” because the reagent not only reacts with water but also some of the additives present. This is a problem, because the Karl Fischer response for a new oil can give a blank value of 2%, mainly from the additive pack. But a failure limit for water in internal combustion engines is typically set at 0.2% or lower, thus the failure limit is an order of magnitude less than the blank value. Worse, however, is the problem of the oil additives degrading during service life, which may form unknown compounds which may or may not react with the Karl Fischer reagent. The “blank value” is now in doubt for used samples, it can be estimated but this leaves a possibly large margin of error. Therefore, the Karl Fischer titration method for the determination of water in new and degraded formulated lubricant and hydraulic fluids is fraught with difficulty. Some variations have been tried, such as gently sparging the oil sample with dry nitrogen and thus blowing the water content as vapor over into a Karl Fischer titration. This takes a long time to “complete” and it is uncertain at what point, if complete at all, when “all” of the water has been transferred for measurement. The Karl Fischer determination of water in formulated new/degraded lubricant and hydraulic fluids is unsuitable because of reactions with additives. When additives are absent, then the Karl Fischer method is a sound method to determine the water content of “pure” hydrocarbon fluids, such as base oils.
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29-33 0.5%
Bio-Rad Win-IR 0.4% 0.3 Absorbance
0.3% 0.2
0.2% 0.1%
0.1
0.05% 0.0%
0 3800
3600
3400
3200
Wavenumber (cm–1)
FIGURE 29.15 FTIR spectra of hydrocarbon fluids degraded with water. From Machinery Oil Analysis — Methods, Automation and Benefits, 2nd ed., Larry A. Toms, Coastal Skills Training, Virginia Beach, Virginia, USA (1998). Courtesy Bio-Rad Laboratories.
29.3.5.3 Water Content by FTIR Spectrophotometry The O–H group in water has a strong, broad, and distinctive infrared absorption from 3150 to 3500 cm−1 , centered on 3400 cm−1 . The absorption band is broad because the O–H group is hydrogen bonded for groups of water molecules. As the concentration of water decreases it becomes less hydrogen bonded or exists as more as smaller groups of bonded molecules or even individual molecules, and its molar absorption increases. Therefore, the calibration curve tends to be nonlinear at lower concentrations. A representative set of FTIR spectra for different levels of water contamination of hydrocarbon oils is given in Figure 29.15 and inspection demonstrates the nonlinearity of the water absorbance in the 0.0 to 0.2% concentration range. The method works well except for formulations using polyol ester base oils or with high dispersant/detergent additive levels. In the first case, the problem arises from the polyol ester infrared absorption in the previously used 3150–3500 cm−1 region, centered on 3400 cm−1 . Subject to detailed baseline corrections, the 3595 to 3700 cm−1 region is used instead to determine water contamination in these oil samples, this corresponds to a singly bonded O–H group. For high detergent/dispersant lubricant samples, the hydroxyl absorption band is not seen but a background increase in absorption occurs between 3000 and 4000 cm−1 . This effect is nonlinear and must be calibrated with standard solutions. It is separate from baseline shifts due to soot and particulates, which are unlikely to be present in this type of lubricant formulation. These two effects point to the main limitation of the FTIR method, which essentially reduces to the need to know the nature of the fresh, unused lubricant. This means that the FTIR method cannot be applied universally and will give errors occasionally, when samples of oils based on polyol esters or formulations containing high levels of dispersants/detergents. Other than this limitation, the FTIR method is very useful and shows great potential for the rapid and accurate determination of water degradation in many new and used lubricant formulations.
29.4 Minor Methods of Investigating Lubricant Degradation Description of the following methods as“minor”means that they are only used in particular and individual circumstances to investigate the degradation of lubricants and hydraulic fluids. They do not form part of a routine investigation of degraded lubricant samples.
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29.4.1 Density, “Gravity,” or “Specific Gravity” Density of a lubricant sample is also referred to as its “Specific, or API, Gravity” and has little value as a measure of the degraded lubricant’s fitness for purpose. The determination of lubricant sample density is now be readily measured to three significant figures using vibrating tube detectors, a very much shorter and accurate procedure than the density bottle method or glass hydrometer, ASTM D1298. But the density information gained has little importance, the density of a degraded lubricant should be close to that of the original material. Changes in density show contamination by a solvent, such as fuel dilution, a different product inadvertently added or a build-up of foreign material. The differences are nevertheless small and, as an example, fuel dilution needs to be extensive to see a significant change in density.
29.4.2 Flash Point of Degraded Lubricant Flash point determination of lubricant samples can now be considered more readily due to automated instruments now being readily available. Both manual and automated methods are based upon the PenskyMartins method, as in ASTM D93 for diesel lubricants. The method brings together considerations of volatility, combustion limits, and ignition temperatures to give a useful measure of great utility. Flash point values of degraded samples rarely increase, if they do, a higher viscosity fluid has been inadvertently added. Much more likely is a decrease in flash point for a degraded lubricant sample caused primarily by fuel dilution resulting from cold/low temperature engine operation. Thermal decomposition of the base oil under extended power operation may also generate lighter fractions which reduce the sample flashpoint. Reduced flashpoints of degraded diesel lubricants due to fuel dilution would normally be associated with a decreased viscosity value and a crosscheck should be done for this. The quantitative extent of fuel dilution is usually non-linear with respect to flashpoint and should be measured by either gas chromatogrphy or FTIR methods. Various method procedures exist of increasing accuracy, the Cleveland open cup, ASTM D92, the Pensky-Martins closed cup, ASTM D93 and the Setaflash small scale closed cup, D3828. Flash points for degraded petrol/gasoline fuel dilution in degraded lubricants are measured by ASTM D322.
29.4.3 Foaming of Lubricants Lubricant foam has a low load carrying ability. Excessive foam build-up in a reservoir or sump will rapidly lead to excessive wear and catastrophic failure of the system. Too high a level of lubricant in an engine sump, by overfilling or miscalibration of the level indicator (dipstick) causes the crankshaft and connecting rod big-end caps to whip up the lubricant into an all-pervading foam and rapid damage ensues. Air leaks into the oil flow or an open drop from a supply pipe into a hydraulic fluid reservoir can generate foam. Operationally, engines should not be overfilled, the level indicator correct, leaks stopped, and supply pipes extended to deliver return lubricant below the normal liquid surface level in a reservoir. While base oils have little foaming tendency, modern lubricant formulations contain many additives substances such as detergents, which can enhance their tendency to foam. Surface active additives will also increase the foaming tendency of a formulation. ASTM D892 measures the foaming tendency of a hydrocarbon fluid but is much more relevant to the fresh, unused, material under laboratory conditions than degraded samples in operating systems. Resolving a used lubricant foaming problem should be treated with great care, fortunately it is relatively rare. Foaming of the new formulation is controlled by the addition of liquid silicone polymers, which reduces the surface tension at the contact points of the foam cells. This allows the lubricant to drain away and the foam to subside. However, formulating the optimum silicone concentration requires extensive work as too little or too much silicone additive increases the foaming tendency of the formulation. Adding an antifoam silicone liquid in situ to a foaming, degraded lubricant, or hydraulic fluid should be approached very carefully and incrementally.
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29.4.4 System Corrosion (“Rusting”) with Degraded Lubricants Lubricant formulations contain rust inhibitors and a system which is maintained well, with a maximum water content of 0.1% for engines and 0.01% for other systems, should not have rust problems. Corrosion inhibitors are needed for systems and vehicles that are used very intermittently, such as military vehicles in prolonged storage or vehicles delivered from one side to the world to another. Repeated sequences of cold engine starts and very short drive distances up to dealer delivery causes condensation of water in the engine and ensuing corrosion of ferrous parts. The contaminants building up in a degraded lubricant system can adversely affect the action of the rust inhibitor present by competitive adsorption at metal surfaces. The ASTM D665 rust method applies to lubricant formulations, if rust corrosion is either suspected or present by discoloration, then degraded samples should be sent for test.
29.4.5 Demulsibility and Interfacial Tension of Degraded Lubricants The demulsibility characteristic of a lubricant is its ability to separate from water when emulsions are formed in a system. While the test is performed for new lubricants, the build-up of trace contaminants may reduce the separation from water in emulsions for degraded samples, hence the term “demulsibility.” Testing degraded lubricants for demulsibility in the laboratory may not be indicative of that sample’s performance in operating systems. ASTM D1401 is the demulsibility test for turbine lubricants and ASTM D2711 is for medium- and high viscosity lubricants. A slightly different procedure of ASTM D2711 is used for extreme pressure (EP) lubricants. Interfacial tension (IFT), is a measure of the surface energy of a fluid against a solid surface or an immiscible standard fluid. Additives contribute to that surface energy and a decrease in interfacial tension indicates that these additives are being deactivated, removed in some way, or depleted by oxidation. A decrease in interfacial tension is an early indication of oxidation before changes are noticed in Acid Number or viscosity. Alternatively, the circulating lubricant is collecting certain compounds in the system added as rust inhibitors, which have polar structures. Interfacial tension measurements of degraded lubricants are useful for rust- and oxidation-inhibited turbine and transformer oils by ASTM D971. The results should not be interpreted on their own but related and compared to changes in other measurements of the system, particularly viscosity and Acid Number.
29.4.6 Instrumental Analytical Techniques The spectroscopic, chromatographic, and x-ray analytical techniques represented by FTIR, gas and liquid chromatography (GC/LC), and x-ray diffraction (XRD/XRF) are increasingly used to investigate degraded lubricants. The long-term trend is for the cost of the instruments to decrease and their resolution to increase with enhanced information technologies. Fourier transform infrared technique is increasingly used to analyze degraded samples, particularly for sequential samples compared to new, unused, samples of the same lubricant. Selected regions of the infrared region are used to follow particular aspects of sample degradation. The method is given additional power through the use of multivariate data analysis. Toms describes the application of FTIR to the analysis of degraded samples. Chromatography, particularly liquid chromatography, may be used to analyse additives in lubricant formulations. CEC has very high resolution of additives and can follow their depletion in successive degraded samples. Gel permeation chromatography (GPC), can follow the degradation or scission, of polymer chain lengths and therefore mean molecular weights, of additives such as VIIs, dispersants, and other polymeric additives. X-ray diffraction is mainly used for quality control and to identify unknown deposits; of more importance is XRF and x-ray absorption fine structure (XAFS), used to identify the elements in compounds, liquids, and solids, found in operating systems.
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These instrumental techniques will have increasing importance in the analytical investigation of degraded lubricant samples. An important issue, already begun, is to bring these instrumental analytical techniques and their specific application developments into the set of standard analytical method for the lubricant and hydraulic fluid manufacture and service use industries.
29.5 Case Studies of Degraded Lubricants 29.5.1 A Degraded Lubricant Sample from a Heavy Duty Diesel Engine
New
Used
Standard
Odor
mid-brown, transparent mild
black, ← opaque diesel, slightly burnt ←
Viscosity at 40◦ C, mm2 /sec at 100◦ C Viscosity index Acid Number, mg/g/KOH Base Number, mg/g/KOH Water Percentage of Soot Percentage of Fuel Dilution
71.31 11.71 160 2.8 9.6 nil 0 0
61.82 10.58 162 4.5 4.5 nil 1.2% 4%
ICP Elements, mg/kg or ppm P Zn Ca Ba B Mg Na Fe Al Cr Cu Pb Sn Ni Mo Si
350 400 1100 <1 129 10 <1 <1 <1 1 <1 <1 <1 <1 <1 6
437 602 1267 <1 90 15 4 9 <1 <1 2 8 4 <1 <1 25
Appearance
D445
D664 D4739 crackle test TGA ← GC ←
Change
−13.3% −9.7% +1.25% +1.7 units −5.1 units
D5185 +24.9% +50.5% +15.2% −43.3%
The arrows conjoin the results for appearance, odor, soot, and fuel dilution to show how qualitative direct observation is supported by instrumental analyses. Overall Comment and Judgement: This degraded lubricant is still fit for service despite the identical Acid and Base Numbers. Once used as a “condemning limit,” the crossing of Acid and Base Numbers no longer means the end of a lubricant charge. Low sulfur fuels and modem additives can greatly extend the life of a lubricant charge in a system.
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The viscosity decreases at both temperatures are just within condemning limits and need to be followed carefully in future sampling. The additive elements, P, Zn, and C have increased, probably due to volatilization of base oil and the associated decrease in volume. The increase in silicon is considerable and is queried for a maintenance check — is the air filter allowing dust into the engine system? Overall, this oil can remain in the engine system as still being “fit for service” and able to protect it but will probably require replacement at the next sampling interval.
29.5.2 A Degraded Grease Sample
New
Used
Standard
Drop Point, ◦ C
181
143
Penetration Test — unworked worked Consistency, NGLI Appearance Texture Color Odor
310 319 1 bright smooth mid-brown very mild
450 D217 459 000 D217 dull coarse ← black ← unpleasant, oxidized/burnt ← 0.18 D95 4.2 2.6 D128
Water (% by weight) Ash (% by weight) Insolubles (% by weight) Changes in Wear Elements by ICP Pb
Soap Elements, mg/kg or ppm Ca Na K Li Pb FTIR Analysis
0.0. 2.7 0.06
D2265
Comment Substantial decrease ←
Softened ← Oil Separation?
Wear metal debris and ingress of dirt
Iron — major increase Al — minor increase Pb — minor decrease Si — traces detected
3000 100 <10 13,000 16,500 normal soap
5900 +97% 400 +300% <10 6100 — 53% 7300 — 56% (by microscope FTIR) separated regions ← of partially oxidized hydrocarbon oil, ← lithium, and calcium soap strand regions
Overall Comment and Judgement: The analyses of this degraded grease sample indicates that it is no longer fit for service and ought to be replaced immediately. Further thought might consider if it is the right grease for this purpose, see below. The first seven observations, from “Drop Point” down to “odor” indicate to an experienced operator that the grease is failing in its purpose. The grease is beginning to separate into a heterogenous gel, from a smooth to a coarse consistency, shown by its decreased NGLI classification. The microscope FTIR of the
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oil areas shows oxidation, possibly from overheating. This is borne out by the unpleasant/burnt smell and change to black in color from mid-brown, indicating overheating. The ash and insolubles values indicate that the grease is both collecting wear debris and picking up dirt. It is failing to protect against wear, shown by the increased Fe levels, and dirt is being ingested, from the increased Na and trace Si, possibly through faulty seals — a maintenance issue. The soap elements show contrary trends, calcium and sodium have increaseu possibly concentrated by loss of oil content ftom overheating. In contrast, the lithium soap thickener and lead EP compounds are decreasing, reflected in the decreased lead found in the spectroscopic analyses. Overall, there are conflicting indicators in this set of analyses. Clearly, this grease formulation is not doing well in this application. An experienced operator examining this analysis, and reviewing previous analyses, might well consider if this grease is the correct grade for the job. Further technical advice is needed and should be sought.
29.5.3 A Degraded Lubricant Sample from a Gas-Fueled Engine
Appearance Odor Viscosity at 40◦ C, mm2 /sec at 100◦ C Viscosity Index FlashPoint, ◦ C Acid Number, mg/g/KOH Base Number, mg/g/KOH UIUts Water Insolubles FTIR Oxidation Nitration ICP Elements, mg/kg or ppm P Zn Ca Ba Mg Na Fe Al Cr Cu Pb Sn Ni Mo Si
© 2006 by Taylor & Francis Group, LLC
New
Used
clear mild 125.6 13.3 101 >200 0.92 5.8
Comment ← reasonably clear ← slightly acrid/burnt 163 15.6 97 >200 1.78 2.6
nil 0 nil nil
nil 0.1% slight ← slight ←
500 560 2500 <1 370 <1 <1 <1 <1 <1 <1 <1 <1 <1 5
470 510 2100 <1 365 123 55 25 <1 33 5 2 6 3 25
Standard
D445
D92 D664 D2896
+29.8% +17.35 −4.0% +0.86 unit −3.2 units
Hotplate and crackle test D893
D5185 no major changes for any
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Comment: This degraded lubricant is still acceptable for service. Although thickened, its Viscosity Index has declined slightly. Base Number has depleted and should be monitored carefully but has not decreased below the Acid number value. It is interesting that the elemental data results have no major changes, the additive element concentration levels are stable, in contrast to the heavy duty diesel sample analyses presented in case study 29.5.1 above. This may be due to a combination of a concentration effect, from volatility of the lubricant formulation, and that ICP elemental analysis does not discriminate between active and inactive (“spent”) species. Note that the slightly different analytical parameters used for the overall analyses of degraded heavy duty diesel lubricants samples and those for a gas engine degraded sample.
29.5.4 A Degraded Hydraulic Fluid
New Appearance Color Viscosity at 40◦ C, mm2 /sec at 100◦ C Acid Number, mg/gKOH Water Percentage of insolubles, gravimetric, (as mg/l) increase n-pentane toluene resins, by difference ICP Elements, mg/kg or ppm Additive elements P Zn Wear Metal Changes Fe Cu Na Pb Cr Sn
Used
Standard
bright clear 2.9 58 7.3 7.3 1.0 nil 15
dark brown ← not clear 7.5 by dilution 70 8.8 2.1 nil 270
0 0
2.5 0.5 2.0
extensive degradation← D1500 D445 D664 crackle test substantial ← increase D893A substantial increase ← D5185
550 560
390 402
<1 1 <1 <1 <1 <1
52 ← 17 21 8 3 2
→ −29.9% → −28.2%
Overall Comment and Judgement: Hydraulic fluids normally remain much cleaner and less degraded than automotive lubricants. This degraded sample is, however, the exception and has several independent condemning features. The service age of this sample is unknown. The color has changed from bright/clear to dark brown/not clear. Color measurement on the degraded sample can only be done by dilution and the “not clear” report indicates extensive suspended particulates. The Acid Number is concerning, to have increased by 1.1 units from the (normally) low value of 1.0 when usual practice sees a decrease, indicates that oxidation is occurring. The insolubles by pentane/toluene comparison show a substantial production of resins, again related to oxidation, paralleled by the increase in particulates determined by gravimetry.
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The viscosities at both 40 and 100◦ C have increased by over 20%. The additive elements are declining in a uniform manner and the wear elements show an increase, mainly in iron, supported by increases in other minor wear elements. This hydraulic fluid has come to the end of its effective life, possibly beyond that, and needs rapid replacement for the continued mechanical health of the machine it operates. Overall, the sampling interval for fluid analyses should be reviewed to consider a shorter time. Or a regular sampling and analytical program should be set up for this machine.
29.5.5 Overview of Degraded Lubricant Analyses Review of the four examples given shows that each suite of analyses has been specifically targeted for the particular lubricant’s service application. A complete suite of analyses is neither needed nor cost-effective, only those that give an insight into the extent of degradation of the lubricant, that is, its “condition.” The use of conjoining arrows in each set of analyses shows how various effects are connected together by causal factors. Therefore, decisions on the extent of a lubricant’s degradation, and subsequent action should be broadly based to cover as many independent variables as possible. Degradation of lubricants will always occur but this can be controlled and contained by a suitably frequent program of sampling and appropriate analyses. With increasing reliability of machines occurring and expected, the monitored degradation of the lubricants for these machines is now essential.
Bibliography ASTM (www.ASTM.com) and IP (energyinst.org.uk) method numbers have been given in text and are readily accessed from the appropriate websites. Note that although the IP website is derived from the former Institute of Petroleum, hence “IP,” this Institute merged to form the Energy Institute in 2004, to explain the apparent discrepancy. There are many publications concerning aspects of lubricant degradation among other associated topics and a list would be extremely long. Very good introductions to lubricants and their degradation are contained in: 1. Machinery Oil Analysis — Methods, Automation and Benefits, 2nd ed., Larry A. Toms, Coastal Skills Training, Virginia Beach, Virginia, USA (1998). 2. Automotive Lubricants Reference Book, A. Caines and R. Haycock, 1st ed. published by SAE, USA (1996) and 2nd ed. by Mechanical Engineering Publications Ltd, London and Bury St Edmunds, UK (1996). ISBN 1 86058 049 1. These books, the ASTM and IP websites give an excellent introduction to studying the degradation of lubricants in their various formulations.
Reference [1] S.B. Saville, F.D. Gainey, S.D. Cupples, M.F. Fox, and D.J. Picken, A Study of Lubricant Condition in the Piston Ring Zone of Single Cylinder Diesel Engines Under Typical Operating Conditions, SAE 881586 (1988).
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30 Lubricant Properties and Test Methods 30.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.2 Lubricant Performance Tests . . . . . . . . . . . . . . . . . . . . . . . . . .
30-1 30-2
Lubricating Oil Tests • Semisolid Grease Lubricant Tests • Solid Grease Lubricant Tests
30.3 Oil Condition Tests . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Larry A. Toms Technical Services
Allison M. Toms GasTOPS Inc.
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Atomic Emission Spectroscopy — ASTM D5185, ASTM D6595 • Infrared (IR) Spectroscopy — ASTM (E2412) • Particle Counting — ISO 4406, ISO 11171 • X-Ray Fluorescence Spectroscopy • Water Determination by Karl Fischer Titration — ASTM D6304
References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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30.1 Introduction Proper lubrication of mechanical machinery is essential for the high reliability and maximum usable life of the machine. Thus, associated lubricating systems and fluids must be properly selected and managed. Lubricating oils and greases are formulated for a wide range of industrial, marine, and commercial applications and consequently require to fulfil a wide range of specific properties and functions. It follows that the range of test methods utilized to quantify these characteristics is also broad. Standard test methods for lubricants have evolved from a variety of sources including: the National Lubricating Grease Institute (NLGI) [1]; American Petroleum Institute (API) [2]; Society of Automotive Engineers (SAE) [3]; Coordinating Research Council, Inc. (CRC) [4]; Federal agencies (Army, Navy, Air Force); instrument manufacturers; and original equipment manufacturers (OEM) [5–8]. Most of the test methods utilized across industrial fields are described in the International ASTM [9,10] “Standards on Petroleum Products and Lubricants” Volume 05. A number of these methods are also joint standards with the API, Institute of Petroleum (IP) [11], Deutsches Institut fur Normung e.V. (DIN) [12], American National Standards Institute (ANSI) [13], and International Standards Organization (ISO) [14]. In this chapter, the test methods presented are primarily those most frequently used from the array documented by the International ASTM Standards on Petroleum Products and Lubricants. To facilitate coverage of the wide variety of test types, two primary testing categories [15] are presented: • Performance tests that define and quantify critical fluid properties, performance parameters, and specifications. These methods are generally utilized for new-oil specification, manufacturing quality control, and user acceptance for specific machinery applications. 30-1
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• Condition tests that measure operational parameters such as the accrual of fluid contamination and degradation characteristics. These test methods are generally utilized for in-service oil monitoring. The performance tests are subdivided into oil and grease applicability and then into the test methods specific to the respective lubricant properties or characteristics associated with the category. It is not the intent to cover all test methodology in detail, but rather provide a brief description of the most popular methods and practices with focus on the (1) property or characteristic parameter being measured, (2) the test methods and apparatus employed, and (3) the significance of the test and its results. If operational details are desired, reference should be made to the original source of the tests, that is, ASTM, etc. For ease in locating specific tests, they are presented alphabetically in each category.
30.2 Lubricant Performance Tests A modern lubricant is expected to operate efficiently over a wide range of temperatures and stresses while in contact with machinery metals, many different chemical compounds, and other debris resulting from combustion or other processes. Oil performance tests have been designed by the petroleum industry and OEM to define the performance or properties of oil products required for proper operation in each class of machinery application and for product manufacturing quality assurance and control (QA/QC). In addition, machinery owners often utilize selected performance parameters to determine the suitability of a specific new-oil product for use in a specific machine and to verify bulk product specifications upon delivery.
30.2.1 Lubricating Oil Tests 30.2.1.1 Additive Metals in Unused Oils — ASTM D4628, ASTM D4927, ASTM D4951, ASTM D6443, ASTM D6481 Most lubricating oils contain additive compounds that contain metals. These include detergents, dispersants, corrosion inhibitors, antioxidants, antiwear, and extreme-pressure agents. In unused fluids, the concentration of metals associated with additive compounds is a rough measure of the additive concentration. The ASTM D6443 determines the level of calcium, chlorine, copper, magnesium, phosphorus, sulfur, and zinc in unused lubricating oils by wavelength dispersive x-ray fluorescence spectrometry. Matrix effects (interelement interferences) are reduced by regression software. Results are reported in mass percent. High-concentration additives may be diluted to produce values within the calibration range of the instrument. Alternatively, the following methods may be utilized for determining additive metals in unused lubricating oils. Barium, calcium, phosphorus, sulfur, and zinc may be determined by ASTM D4927 wavelength dispersive x-ray fluorescence spectrometry. Phosphorus, sulfur, calcium, and zinc may be determined by the ASTM D6481 test utilizing an energy dispersive x-ray fluorescence spectrometer. Barium, calcium, magnesium, and zinc may be determined by ASTM D4628 test utilizing atomic absorption spectrometry. Barium, boron, calcium, copper, magnesium, phosphorus, sulfur, and zinc may be determined by utilizing the ASTM D4951 atomic emission spectrometry by an inductively coupled plasma spectrometer. Test methods for determining the level of chlorine and sulfur in lubricating oils are given below. 30.2.1.1.1 Comments The tests for determining additive metals are applicable only to unused lubricants and are not appropriate for evaluating the remaining additive concentration in in-service oil, as the metal component of the expended additive may remain present in the oil. In addition to these instrument methods, there are a number of single element methods such as the ASTM D1091, which determines the value of phosphorus.
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30.2.1.2 Aniline Point — ASTM D611 Hydrocarbon oils have a tendency to cause swelling of rubber parts such as elastomer hoses, seals, or gaskets. The aniline solubility temperature is a useful method to determine the compatibility of a lubricant with rubber parts. In the ASTM D611 test, specified volumes of the chemical aniline (a benzene derivative) and the oil sample are placed in a tube and mechanically mixed. The mixture is heated until the two phases become miscible. The mixture is then cooled at a controlled rate. The test endpoint is reached when the mixture separates into two phases. The temperature at the moment of phase separation is recorded as the aniline point for the oil sample. The measurement exhibits an inverse relationship in that the lower the aniline point, the greater the tendency for rubber swelling. Oils that are high in aromatics will have the highest solvency to aniline, and thus the lowest aniline point. Paraffinic oils have the least solvency and the highest aniline point. Napthenic, cycloparaffin, and olefin based oils exhibit an intermediate solvency to aniline and thus have intermediate aniline points. 30.2.1.2.1 Comments The main significance of the aniline point test is the determination of the swelling tendency of a given organic elastomer from exposure to the oil being tested. In this regard, the test method is only applicable for temperatures that fall between the freezing and boiling points of the sample mixture. Applicability: hydraulic fluids; steam, hydro, and industrial gas turbine oils; and electrical insulation oils. International equivalent standards include DIN 51775, IP 2, and ISO 2977. 30.2.1.3 Ash Tendency Test — ASTM D482, ASTM D874 Modern internal combustion engines require lubricant formulations that reduce the formation of carbonaceous deposits and sludge. The most common agents for deposit reduction are metallic detergent and dispersant additives containing compounds of calcium, magnesium, potassium, zinc, sodium, and barium. A rough measure of the level of these compounds in new-oil blends is provided by the ash content or sulfated ash content test. The simplest measurement is the ASTM D482 test where the ash content is the quantity or mass of incombustible solids left over after burning a specified quantity of oil for 10 min at a temperature of 800◦ C. The ash will include metallic additives and any contaminants containing metals. The ASTM D874 test is an improved method for determining the ash content of lubricating oils. In this method, the sample is burned until only the carbon residue and metallic ash remain. The residue is treated with sulfuric acid, reheated, and weighed. The sulfated ash test indicates the concentration of the metal-based additives in new-oil blends. 30.2.1.3.1 Comments For oils containing a magnesium-based detergent and a boron-based dispersant, the sulfated ash measurement will be unreliable. The ash test is also unreliable for used-oil measurements. In this case, the ash test results will also include contaminant compounds and wear metals in addition to the desired additive compounds. The test does not distinguish between consumed and remaining usable additives — the measured additive compounds will include both reserve (not-yet-used) and consumed additive constituents generating a concentration similar to new-oil. The trend toward ashless additive compounds renders this test a dubious measurement of detergent level. Applicability: The ASTM D482 is utilized for petroleum fuels, crudes, and lubricating oils. The ASTM D874 is generally utilized for unused lubricating oils. International equivalent standards include (D482) IP 4 and ISO 6245 and (D874) IP 163 and ISO 3987. 30.2.1.4 ASTM Color — ASTM D1500 The color of a lubricant provides no direct indication of its lubricating properties; however, a change in color does signify a change in chemistry or the presence of a contaminant. Lubricant color assessment is normally based on transmitted light, and the common color scales used for color interpretation are based on this principle. The basic assumption of the color test is that oil color can be related to deterioration or degree of refining and that any color change is due to deterioration or contamination. New-oil color is determined by the ASTM D1500 colorimeter method. The test compares the transmitted light of an oil with a set of standard glass color slides under controlled conditions. Note, new-oils complying with
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specific legislation (e.g., special tax exemptions) may have added dyes to indicate the special classification or status. 30.2.1.4.1 Comments The ASTM color test is only applicable to mineral oils that do not have dyes or other materials that change the oil’s natural color. The test is a rough measure of water and oxidation by-products in inhibited steam and hydro turbine lubricating oils. International equivalent standards include IP 196, ISO 2049, and DIN 51578. 30.2.1.5 Carbon Residue Characteristics — ASTM Dl89, ASTM D524 When oil is subjected to evaporation or is exposed to atmospheric air for long periods, oxidation of the nonvolatile constituents of the oil results in a carbonaceous residue. Because other oil ingredients present in lubricating oils often produce non-carbonaceous deposits, it is desirable to determine the ash content as well in order to establish the true value of the carbon content. The Ramsbottom (D524) and Conradson (D189) tests are used to determine carbon residue characteristics. In the more common Conradson test, the oil sample is heated in a porcelain crucible placed within a special steel crucible for a prescribed period of time. The increase in weight of the porcelain crucible indicates the amount of carbon residue. In the Ramsbottom test, the sample is heated at a specific temperature in a glass bottle with a small opening until all the volatile ingredients have evaporated and “coking” and “cracking” have further decomposed the oil. The weight of the remaining material is the carbon residue. 30.2.1.5.1 Comments The degree of carbon residue is of secondary importance for well-refined, quality-controlled lubricants and so the carbon residue test is not frequently performed. In addition, the test is of little value for synthetic lubricants with their exceptional thermal stability. In automotive applications, engine design, fuel selection, and operating conditions are as important in carbon deposition as the intrinsic carbon residue content of the oil. Consequently, the test is generally performed during the selection of hydrocarbon oils for such applications as heat-treating, air compressors, and high-temperature bearings. International equivalent standards include (Dl89) IP 113, DIN 51551, and ISO 6615 and (D524) IP 14 and ISO 4262. 30.2.1.6 Chlorine Content — ASTM D808 Chlorine compounds are deleterious to the environment and significant efforts have been undertaken to reduce the amount of these compounds in machinery fluids. Thus, in those machinery applications where it may be desirable to determine the presence of a chlorinated compound or contaminant the following test may be utilized. The D808 method utilizes a pressure vessel that provides a gravimetric determination of chlorine content. The method is suitable for either new or used oils as well as greases. The presence of other halogen compounds will obscure the results. 30.2.1.6.1 Comments These test methods only determine the amount of chlorine present and do not indicate fluid condition or effectiveness in a particular application. Note that chlorine is an oxidizer. For most machinery fluid applications, chlorine compounds are contaminants and their presence in any concentration will be considered harmful to machinery operations. For any machinery fluid still utilizing a chlorine-based additive, the optimum concentration must be monitored: too much may produce excessive corrosion, while an insufficient quantity may limit the fluid’s effectiveness. 30.2.1.7 Copper Corrosion Resistance — ASTM D130 Many lubricated machines are comprised of parts containing copper alloys. It is therefore essential that copper components be properly lubricated and that the oil be noncorrosive to metallic copper. Most petroleum crudes contain sulfur compounds, some of which are corrosive to copper. During the refining of high quality lubricants, most of the corrosive compounds of sulfur are removed. However, some
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lubricant types are blended with sulfur based emulsifying or extreme-pressure additives, which can also corrode copper surfaces. The degree of protection from copper corrosion provided by a lubricant is determined from the ASTM D130 test. In this test, a specified volume of oil is placed in a covered beaker at a temperature of 125◦ C. A polished copper-strip specimen is immersed into the oil for a specified period (usually 2 h) after which time it is checked for the degree of tarnishing or corrosion. The degree of corrosion is determined by comparison of the copper specimen with a set of ASTM (tarnish) standards. 30.2.1.7.1 Comments The copper corrosion resistance test method is suitable for hydrocarbon products having a Reid vapor pressure (D323-99a) no greater than 18 psi. The test will provide a rough measure of the corrosion prevention characteristics for oil utilized in machinery to lubricate native bronze and brass parts. The test is also suitable for determining the best candidate from a series of potential oil products. The test is not suitable for determining the overall corrosion prevention characteristics of lubricating oils. International equivalent standards include ISO 2160, IP 154, and DIN 51811. 30.2.1.8 Density and Specific Gravity — ASTM D1217, ASTM D1298 Density, a numerical expression of the mass-to-volume relationship of a fluid, is sometimes used as a measure of lubricant composition consistency or production uniformity. Density is also used as a rough indicator of hydrocarbon type or volatility. In the petroleum industry, density is usually expressed as kilograms per cubic meter (kg/m3 ) at 15◦ C. Kilograms per liter (kg/l) is also used. The ASTM D1217 test determines density and specific gravity of pure hydrocarbons or petroleum distillates that boil between 90 and 110◦ C. The test utilizes the Bingham pycnometer at the specified test temperatures of 20 and 25◦ C. The ASTM D1298 test determines the density, specific gravity, and API gravity values of petroleum and non-petroleum mixtures. The test is performed in a controlled environment and results are read off a glass hydrometer at convenient temperatures, and corrected to the standard temperature, 15.56◦ C (60◦ F), by means of the “petroleum measurement tables” issued by ASTM International and IP. The API gravity reading is determined from the hydrometer reading by the formula: API Gravity =
141.5 − 131.5 sp. gr. 60/60◦ F
30.2.1.8.1 Comments Density and specific gravity are generally utilized for establishing weight and volume factors for shipment, storage, delivery, and quality assurance. The American Petroleum Institute and ASTM provide API gravity and temperature corrections conversion tables that include density and specific gravity. Note paraffinic oils exhibit lower specific gravity than base oils that contain naphthenic or asphaltic components. Specific gravity measurements used in combination with other measurements (ASTM D2501) provide some indication of the hydrocarbon composition of oils. Such determinations, however, should be limited to oils that are predominately hydrocarbons. International equivalent standards include (D1298) IP 160, API 2547, DIN 51757, and ISO 650. 30.2.1.9 Emulsibility, Demulsibility Characteristics — ASTM D1401, ASTM D2711 A highly refined petroleum lubricant resists the tendency to form emulsion mixtures with water and will generally phase-separate upon standing. However, in pressurized circulating lubrication systems, the mechanical action of the pump and other components can cause oil and contaminating water to form an emulsion. Moreover, the system flow rate may be high enough to prevent sufficient standing time in the reservoir to allow phase separation to occur. The water separability or demulsibility property of a lubricant is determined from the ASTM D1401 test. In this test, 40 ml of oil sample and 40 ml of distilled water are placed in a graduated cylinder and stirred for 5 min at 54◦ C to create an emulsion mixture. The degree of separability is indicated by the time required for the oil/water phase separation to take place.
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The progress of the oil/water phase separation is measured at 5-min intervals. The D2711 test is utilized for medium- and high-viscosity lubricating oils. 30.2.1.9.1 Comments Oil emulsions make poor lubricants. Persistent emulsions will increase component wear, oil oxidation, and sludge and varnish deposits that can in turn clog filters, reduce cooling efficiency, and promote corrosion. Bearing, gear, and hydraulic systems that are prone to water contamination require good lubricant demulsibility to prevent severe wear damage to moving parts. Oils utilized for these components should be verified for oil–water separability before recommendation for service. International equivalent standards include (D1401) IP 19, DIN 51599, and ISO 6624 and (D2711) DIN 51353. 30.2.1.10 Extreme-Pressure Properties of Lubricants — ASTM D2782, ASTM D2783, ASTM D3233, ASTM D6121, ASTM D6425 One of the most important attributes of lubricating oil is the ultimate load that can be sustained without seizure or scoring of the lubricated sliding surfaces. The “seizure” condition relates to welding or fusion of metal asperities on the rubbing test pieces while the “scoring” characterizes the nature (furrowed scar) of the seizure. ASTM D3233 utilizes a Falex pin and vee block apparatus to determine the load carrying properties of fluid lubricants. The Falex device utilizes a pin rotating against a pair of v-shaped blocks immersed in the lubricant sample. The tester rotates the steel pin at 290 rpm. The load is continuously increased by a ratchet mechanism in 250 lbf (1112 N) steps until the end of the test. In both cases, the result is the load at which pin failure occurs. ASTM D2782 utilizes the Timken apparatus to determine the extreme pressure of lubricating fluids. In this method, the test machine utilizes a steel test cup rotating at 800 rpm against a steel block lubricated with the test sample. Addition or removal of weights varies the load between the cup and the block. The test results are indicated by the maximum weight at which no scoring occurs. ASTM D2783 utilizes the Falex four-ball test machine to determine the wear properties of extreme pressure lubricants. The four-ball tester employs a steel ball rotating at 1760 rpm against three stationary steel balls lubricated with the test sample. Addition or removal of weights is used to vary the load between the rotating and stationary balls. The test results are reported as the dimensions of scar marks made on the ball surfaces. Interpretation of the scarring up to seizure determines the load/wear index and the weld point. The ASTM D6121 test, commonly referred to as the L-37, evaluates the load-carrying ability of extremepressure (EP) lubricants under low-speed and high-torque conditions common in final hypoid drive axle applications. The final axle assembly is run in for a specified period at prescribed load, speed, and temperature after which the test phase is operated for 24 h at a prescribed load, speed, and temperature. The bearings and gears are evaluated for wear, corrosion, and deposits in accordance with CRC manual 17. The ASTM D6425 standard test for measuring friction and wear properties of EP lubricating oils utilizes the SRV test machine. This test determines the oil’s coefficient of friction and its ability to resist wear in machinery applications subjected to high-frequency, linear oscillation motion. The SRV machine utilizes a test ball oscillated at constant frequency, stroke amplitude, and load against a test disk. The result is determined from the mean wear scar diameter. 30.2.1.10.1 Comments Friction machine tests are a proven means for assessing the load carrying capacity of lubricants and the influence of various antiwear, antiscuffing, and extreme-pressure additives used in bearing and gear lubricants. Since these tests are intended to evaluate relatively low-viscosity oils for circulating systems, they are generally not suitable for evaluating compounded fatty lubricants used for open or hypoid gears. In addition, due to the subjective nature of the results, precise conclusions are not possible. However, these tests are useful screening tools and when supplemented with other test methods that provide antiwear
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and corrosion characteristics, type of agent, etc., a better interpretation of the relative overall lubricant performance may be made. For precise data, antiwear and extreme-pressure characteristics measurements should be performed on the specific bearings and gears and these should be operated under the same load, speed, temperatures, etc. that are representative of the intended machinery application. International equivalent standards include (D3233) IP 241, (D2782) IP 240, (D2783) IP 239, and (D6425) DIN 51834. 30.2.1.11 Flash and Fire Points — ASTM D56, ASTM D92, ASTM D93, ASTM D1310, ASTM D3828, ASTM D6450 The flash and fire points of a lubricant are measures of the fluid’s volatility and flammability. The flash point refers to the minimum temperature at which there is sufficient vapor to cause a flash of the vapor/air mixture in the presence of an open flame. The fire point is the minimum temperature at which the production of vapor is sufficient to maintain combustion. When selecting a lubricant for a given application, the main concern for the lubrication engineer is the assessment of its potential for explosion or fire under the anticipated conditions of operation or storage. Consideration must also be given to the potential for contamination by more volatile fluids, such as the dilution of an engine lubricant by fuel, which greatly increases the potential for damage due to explosion or fire. The flash and fire points of a lubricant are determined from several tests including the following. ASTM D56 flash point by “Tag closed cup” measures the flash point of liquids with a viscosity below 5.5 cSt (mm2 /sec) at 40◦ C (104◦ F) and a flash point below 93◦ C (200◦ F). The specimen is placed in a closed cup and heated at a slow constant rate. At specified intervals an ignition source is lowered into the cup. The lowest temperature, corrected to standard barometric pressure, at which the vapor above the sample ignites, is taken as the flash point. ASTM D92 flash point by “Cleveland open cup” measures the flash and fire point temperatures of all oil products except those with a flash point below 79◦ C. The test cup is filled with a specified amount of sample and the temperature is raised. At specified intervals an open flame is passed over the cup. The lowest temperature, corrected to standard barometric pressure, at which the vapor above the sample ignites, is taken as the flash point. ASTM D93 flash point by “Pensky-Martens closed cup” measures the flash point temperature of oil products with improved accuracy. The test cup is filled with a specified amount of sample and closed. The sample temperature is raised and at specified intervals an open flame is lowered through a shutter into the cup. The lowest temperature, corrected to standard barometric pressure, at which the vapor above the sample ignites, is taken as the flash point. ASTM D1310 flash point and fire point by “Tag open cup” measures the flash point and fire point temperatures of liquids having flash points between 0 and 325◦ F (−18 and 165◦ C) and fire points up to 325◦ F. ASTM D3828 flash point by “Small Scale Closed Cup” measures the flash point temperature of oil products with improved accuracy. It can measure the flash point or the occurrence of a flash at a specific temperature. The test cup is filled with a specified amount of sample and closed. The sample temperature is raised and at specified intervals an open flame is lowered through a shutter into the cup. The lowest temperature, corrected to standard barometric pressure, at which the vapor above the sample ignites, is taken as the flash point. ASTM D6450 flash point by “continuously closed cup” (CCCFP) test utilizes a 1-ml test sample placed into a closed but unsealed cup. Air is injected into the cup during the test. This test determines the flash point of fluids between 10 and 250◦ C. The flash point is the temperature reading at the moment the flash induced pressure increase is sensed. 30.2.1.11.1 Comments The flash point of an oil product is a rough measure of the fluid’s volatility or lower explosive limit corresponding to a vapor pressure of about 2 to 5 torr. The measure is an important property when selecting fluids for high-temperature machinery applications. In addition, the flash point temperature
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is useful as a quality control measure where a change in it can be related to a commensurate change in fluid volatility due to product changes or the presence of ingress contamination by more or less volatile fluids. International equivalent standards include (D56) IP 304, (D92) IP 36 and ISO 2592, (D93) DIN 51758, ISO 2719, and IP 34, and (D3828) IP 303 and ISO 3679. 30.2.1.12 Foaming Characteristics — ASTM D892, ASTM D3427, ASTM D6082 A reliable lubricant should release entrained air or other gas and resist foaming. Excessive foaming is detrimental to the operation of most machinery fluid systems. Foam can fill the internal spaces such as a separator or reservoir resulting in poor system efficiency or failure; cause a vapor block in filters resulting in oil starvation; and cause excessive wear of lubricated parts due to the poor load-carrying ability of entrained gas (air). Excessive foaming can result in oil loss due to the overflow of the reservoir through vents and create maintenance problems. The ASTM D892 test for foaming characteristics provides an empirical rating of the foaming characteristics of a lubricant sample under specified temperature conditions. A metered volume of dry air is blown through a diffuser immersed in the sample for a period of 5 and 10 min. The foaming tendency is determined as the foam volume (milliliters) at the end of the blowing period. The foaming stability is determined from the settling period for three or more sequences. Alternately, for oil temperatures above 93.5 to 150◦ C, the similar ASTM D6082 procedure is used. The ASTM D3427 air release properties test determines the ability of a lubricant to release entrained air at a controlled temperature. In this procedure, compressed air is blown through the sample for 7 min and the time taken to release the entrained air is recorded. The volume of entrained air remaining is determined from sample density measurements with a density balance. 30.2.1.12.1 Comments Foaming is attributed to air entrainment due to mechanical working of the oil during machine operation. In addition, the presence of water and surface active materials in the oil such as rust preventatives, detergents, etc. can cause foaming. Foaming may be controlled to some extent with the use of additives. Since these materials can increase the tendency of the oil to entrain air, the “optimum” amount of additive for the oil application must be determined. International equivalent standards include (D892) ISO 6247, DIN 51566, and IP 146, (D3427) DIN 51381, IP 313, and ISO 9120. 30.2.1.13 Hydrolytic Stability Characteristics — ASTM D2619 A reliable lubricant should resist the tendency to hydrolyze at machinery operating temperatures when in the presence of water and copper components. Poor hydrolytic stability gives rise to the formation of acidic by-products and insolubles, which in turn results in deposits of varnish and sludge and chemical leaching of copper and other machinery metals. This property is of particular importance for equipment utilizing ester-based lubricating or hydraulic fluids. The ASTM D2619 hydrolytic stability characteristics test determines the ability of a lubricant to resist the tendency to hydolyze (acidify) in the presence of water and machinery metals. The test employs a vessel containing 75 g of oil, 25 g of water, and a copper specimen. The vessel is closed and rotated for a period of 48 h at a temperature of 95◦ C (200◦ F). The hydrolytic stability value is determined from measurements of the resulting insolubles level, acid number, viscosity increase, and copper loss. 30.2.1.13.1 Comments The ASTM D2619 test is useful in the evaluation of a variety of potential hydraulic fluids. In particular, the effectiveness of hydrolytic stability inhibitors in synthetic fluids (i.e., phosphate, silicate esters) can be evaluated and compared. As with other similar tests, the most reliable information is interpreted from test results using actual application experience.
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30.2.1.14 Interfacial Tension — ASTM D971, ASTM D2285 Lubricating oils, as a result of oxidation or modification by additives, display differences in the degree of wetting, spreading, and boundary lubrication on various surfaces. Such differences are indicated by interfacial tension (IFT) values. The measurement of the strength of the boundary film (interfacial tension) employs a special device (tensiometer) which presents a planar ring of platinum wire, supported by a precision spring, to the interface between the test oil and water. The force required to detach the ring from the surface of the liquid with the higher surface tension (usually the water) is a measure of the IFT. The test is extremely sensitive and requires practice and control to obtain reproducible results. 30.2.1.14.1 Comments Strong mutual attraction between oil surface molecules is responsible for surface tension. When materials such as oxidation by-products are present, the film strength (surface tension) of the fluid is modified. The measurement of change in interfacial tension can compliment other performance tests such as neutralization number as a means of detecting excessive fluid oxidation, and life-limiting values may be applied. Note that a rigorous chemical analysis method such as Fourier transform infrared (FT-IR) is preferable for analyzing the degree and type of oxidation by-products present in used-oils. IFT measurement on new oils, with the exception of electrical insulating oils, is generally not appropriate. International equivalent standards include (D971) ISO 304 and DIN 53914 and (D2285) ISO 9101. 30.2.1.15 Low-Pressure Volatility of Lubricants — ASTM D2715 Machinery operating in reduced atmospheric pressure or vacuum conditions requires lubricant products with low volatility and decomposition rates. The ASTM D2715 encompasses the necessary apparatus and procedure to determine evaporation and decomposition rates of lubricating materials. In the D2715 procedure, a prescribed quantity of the lubricant is exposed in a vacuum thermal balance device. As the evaporated material collects on a condensing surface, the decreasing weight of the original sample is recorded as a function of time. With additional instrumentation, it is possible to obtain quantitative evaporation data along with the identity of the volatile and decomposition by-products. The procedure is repeated at increasing temperature levels and the evaporation rate(s) determined for each temperature. The vapor pressure may also be determined when the molecular weight of the test lubricant is known. 30.2.1.15.1 Comments Determination of evaporation rates under prescribed conditions may provide a basis for the estimation of an approximate useful life if attrition of the lubricant can be attributed to evaporative loss. Data interpretation will be more reliable when the other factors that influence low-pressure performance of lubricants are also evaluated. This test is limited to speciality low-vapor pressure liquid and solid lubricants. 30.2.1.16 Neutralization Number (Acid/Base Number) — ASTM D664, ASTM D974, ASTM D2896, ASTM D3339 New lubricant products will normally exhibit acidic or alkaline characteristics, depending on their intended function, additive mix, manufacturing process, and the presence of contaminants or degradation by-products formed during service. For example, internal combustion engine lubricants are formulated to exhibit an alkaline reserve, while turbine lubricants tend to be somewhat acidic. The degree of acidity or alkalinity of a lubricant is derived from its neutralization number. The ASTM D664 acid number by potentiometric titration may be used to determine the change in relative acidity of a lubricant regardless of color or other properties. In this test, the sample is dissolved into a mixture of toluene, isopropyl alcohol, and water and titrated with alcoholic potassium hydroxide. The meter readings are plotted against the respective volumes of the titrating reagent utilized. The acid number is indicated as the quantity of base, expressed in milligrams of potassium hydroxide per gram of sample, required to titrate the sample from its initial value to that of a calibration standard. The ASTM D974 acid or base number by color indicator titration is a simple qualitative method for determining acid or base number. In this test, the sample is dissolved into a mixture of toluene, isopropyl alcohol, and water and titrated with an alcoholic base or acid solution until the endpoint is indicated
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by a color change of the added naphtholbenzein solution. The acid number (AN) is determined as the quantity in milligrams of potassium hydroxide (KOH) required to neutralize all the acidic by-products in 1 g of lubricant sample. The base number (BN) “alkalinity reserve” is determined as the quantity of acid, expressed in equivalent milligrams of KOH, required to neutralize all basic constituents in the lubricant sample. The ASTM D2896 base number by potentiometric titration test may be used to determine the change in relative base number of a lubricant regardless of color or other properties. In this test, 20 g of oil sample is dissolved into 120 ml of titration reagent glacial acetic acid and chlorobenzene (or mixed xylenes) solution which is titrated with perchloric acid. The potentiometric readings are plotted against the respective volumes of titrating solution. The base number is calculated from the quantity of acid needed to titrate the solution, expressed in milligrams of potassium hydroxide per gram equivalent. For testing smaller samples, utilize the ASTM D3339 method. This test is especially intended where the sample volume is too small to allow accurate analysis by D974 or D664. 30.2.1.16.1 Comments The D664 and D2896 potentiometric methods are more versatile and provide more reproducible data than the D974 colorimetric method. These methods may also be utilized to analyze dark oils and oils with weak-acid, strong-acid, weak-base, or strong-base constituents even though the dissociation constant of a strong acid or base may be up to 103 greater than the next strongest component. The constituents that contribute to overall oil acidity include phenolic compounds, resins, esters, inorganic and organic acids, heavy metal salts, and salts of polybasic acids. The constituents that contribute to alkalinity include organic and inorganic bases, salts of polyacidic bases, amino compounds, etc. Some amphoteric additives such as inhibitors, detergents, salts of heavy metal, soaps, and other fillers can produce either an acid or alkaline response. An increase in used-oil acidity is generally due to the presence of oxidation by-products. Neutralization number is useful for production quality control, specification, and the purchasing of lubricants. It is also used in combination with other tests, for example, interfacial tension, to ascertain changes that occur in service under oxidizing conditions. When used to monitor lubricant condition in systems such as steam turbines, hydraulic systems, and transformer and heat transfer units that are normally free of interfering contaminants, neutralization number may be used to indicate when the oil must be changed, restored, or reclaimed. Note that rigorous chemical analysis methods such as FT-IR will provide better data granularity on the individual constituents contributing to changes in an oil’s acid or base number. International equivalent standards include (D664) IP 177, DIN 51558, and ISO 6618 and (D974) IP 139 and ISO 6618. 30.2.1.17 Oxidation Stability — ASTM D943, ASTM D2272, ASTM D2893, ASTM D5763, ASTM D5846, ASTM D6514 All oil products will begin to oxidize and degrade from the moment of use. Exposure to atmospheric oxygen promotes the formation of acidic by-products that degrade into sludge and varnish deposits. Good oxidation stability is one of the key requirements of a lubricant and an important factor in the estimation of remaining useful service life. Oxidative by-products are generally characterized as insoluble resins, varnishes, and sludges, or soluble organic acids and peroxides. The usable service life of in-service oil tends to be inversely proportional to the level of these contaminants. These tests are generally utilized to determine the oil’s oxidation stability and provide an indication of the remaining usable life. The ASTM D943 lubricant oxidation characteristics test was developed to determine the oxidation stability of petroleum lubricants in the presence of oxygen, water, copper, and iron. A volume of oil is reacted with oxygen in the presence of water and a catalyst of iron and copper at a temperature of 95◦ C for a period of 500 to 1000 h. Samples are removed periodically during the test period for analysis of acidic content. The ASTM D664 acid number test is typically performed to determine sample acidity. The oxidation characteristics test is discontinued when the AN reaches 2.0 mg KOH/g or above. Note, some new oils may exhibit AN of up to 1.5 mg KOH/g. The D943 test is of limited applicability for these oils.
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ASTM D2272 — rotating pressure vessel oxidation test (RPVOT) is a relatively short duration test procedure and generally preferred over other methods. This test utilizes a rotating pressure vessel in which the oil sample, water, oxygen, and a copper catalyst coil is placed. The vessel is charged with oxygen to 90 psi and is rotated at 100 rpm in a temperature-controlled bath. The test is timed until a specific drop in oxygen pressure signals the endpoint. The remaining oil life is reported as minutes of bomb life. The D2893 test was developed to determine the oxidation stability of extreme pressure lubricants. In this test, dry air is bubbled at a rate of 10 l per hour through 300 ml of oil at a temperature of 95◦ C for a period of 312 h. Samples are removed periodically during the test period for analysis of viscosity and precipitation number. ASTM D5763 test determines the oxidation and thermal stability characteristics of gear oils using the universal glassware apparatus. In this test, 100 g of oil is exposed to a continuous flow of dry air at 120◦ C for 312 h. The result is determined from an evaluation of the total sludge, viscosity change, and oil loss. ASTM D5846 evaluates the oxidation stability of petroleum-based hydraulic and turbine lubricating oils using the universal oxidation test apparatus. The oil is exposed to air at 135◦ C until the acid number increases by 0.5 mg KOH/g over the new, untested oil or until insoluble deposits are observed. The ASTM D6514 evaluates the high-temperature oxidation stability and deposit-forming tendency of inhibited steam and gas turbine lubricating oils in the presence of air, copper, and iron metals. The test is performed for 96 h at 155◦ C after which the acid number, viscosity, and sludge level are reported. 30.2.1.17.1 Comments The soluble by-products of oxidation tend to increase oil acidity and viscosity and promote corrosion of metal parts. The rate of oxidation is basically a function of time. However, high operating temperature and the presence of contaminants such as water and catalysts such as copper, accelerate the oxidation process. In general, the oxidation rate of petroleum oil will double for every 10◦ C (18◦ F) increase above the nominal operating temperature of the machine’s fluid system. The main problem with these test methods is that the associated test procedure does not duplicate equivalent in-service conditions. The presence of oxidation by-products and some additives (acting as catalysts) will reduce the effectiveness of these tests for in-service oils, generally causing the oxidation stability results to read low. Applying an oil purifier to remove oxidation by-products before sampling will generally improve oxidation stability test results. Oxidation resistance is critical in steam turbine oils because of the seriousness of turbine bearing failures. Other applications such as gear, transformer, hydraulic, heat transfer, and gas turbine systems also require lubricants with excellent oxidation stability. Although direct data comparisons are not always reliable, oxidation stability test results form the basis for comparison of the relative life of competitive lubricant products as well as new and in-service oils. International equivalent standards include (D943) IP 157, DIN 51587, and ISO 4263. 30.2.1.18 Pentane Insolubles of Lubricating Oils — ASTM D893, ASTM D4055 When the oil is diluted by a prescribed volume of pentane, insolubles such as resins, dirt, soot, and metals tend to drop out of suspension. These precipitates are collectively called pentane insolubles. A high level of pentane insolubles indicates a high level of oxidation or contamination of the lubricant. Further treatment with toluene will release additional precipitates collectively referred to as toluene insolubles. A high level of toluene insolubles indicates a high level of external contamination such as soot or dirt. The difference between pentane insolubles and toluene insolubles indicates the level of oxidation by-products in the oil. The ASTM D893 centrifuge or D4055 micro-filtration methods may be utilized to determine the degree of pentane and toluene insolubles in lubricating oils. These methods also determine the pentane insolubles of high-detergency engine oils, where a coagulant is used to release insolubles held in suspension by the detergent/dispersant additive. In the ASTM D893 method, pentane insolubles are determined by placing 10 g of sample into a centrifuge tube, filling the tube to the 100-ml level with pentane, and centrifuging. The precipitate is washed twice with pentane, dried, and weighed to the nearest 0.1 mg. The percent pentane insolubles is
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calculated from the before and after weight of the centrifuge tubes. In addition, the pentane insolubles residue can be further treated with toluene to dissolve and separate out the resins formed during the oil oxidation process. In this procedure, 10 g of sample is placed into a centrifuge tube, filled to the 100-ml level with pentane, and centrifuged. The precipitate is washed twice with pentane, once with a toluene alcohol solution, once with toluene, dried, and weighed to the nearest 0.1 mg. The residue left over after this test is referred to as toluene insolubles. In the ASTM D4055 method, pentane insolubles in lubricating oils are determined by membrane filter analysis. This procedure utilizes a submicron filter membrane as a means to extract the insolubles precipitate. The filter membrane is cleaned and weighed to the nearest 0.1 mg. 1 g of sample is placed into a volumetric flask and filled to the 100-ml level with pentane. The sample solution is filtered and the membrane is re-weighed to the nearest 0.1 mg. The percent pentane insolubles is calculated from the before and after weight of the membrane filter. 30.2.1.18.1 Comments Pentane and toluene insolubles are generally utilized to indicate the amount of soot and unburned fuel by-products in used diesel engine oils. It is usually preferable to utilize a more rigorous chemical analysis method such as FT-IR, which will provide individual results for soot and acidic by-products in used diesel engine oils. 30.2.1.19 Pour and Cloud Point — ASTM D97, ASTM D2500, ASTM D5771, ASTM D5772, ASTM D5773, ASTM D5949, ASTM D5950, ASTM D5985, ASTM D6749 Many types of machines must be started while cold and the ability of a lubricant to flow and lubricate properly during (and just after) start-up is an important factor in the selection of the lubricant. A general classification of oil response to low temperature conditions may be determined by two temperature measurements: pour point and cloud point. When an oil product is cooled sufficiently, a point is reached when it will no longer flow under the influence of gravity. This is referred to as the pour point. At this temperature, the lubricant cannot perform its primary functions. However, agitation by a pump will break down the wax structure and allow paraffinic type oils to be pumped at temperatures below their pour point. Naphthenic oils contain little or no wax and their pour point is reached by virtue of the increase in viscosity due to the low temperature. Naphthenic oils cannot be pumped readily near their pour point. Just prior to reaching the pour point, a paraffinic lubricant becomes cloudy as a result of the crystallization of waxy constituents. These materials crystallize in a honeycomb structure at low temperatures. The temperature at which a cloudy appearance is first observed is referred to as the cloud point. At this temperature, most oils will not perform effectively enough for reliable machine operation. Under these conditions, very viscous oil or waxy constituents may accumulate and plug oil passages or filters. Note that the cloud point should not be confused with cloudiness or color changes due to high lubricant stress or contamination. In the ASTM D97 pour point test, the oil is first heated to ensure solution of all ingredients and elimination of any influence of past thermal treatment. The oil is then cooled at a specified rate, and checked at temperature increments of 3◦ C. The test apparatus removes the sample vessel from its cooling jacket and tilts it at a 90◦ angle as prescribed by the D97 test method until no flow is observed for a 5-sec interval. The pour point result is then reported as 3◦ C higher than the temperature at which the sample ceased to flow. In the ASTM 2500 cloud point test, the oil is first heated to ensure solution of all ingredients. In this test, the oil’s temperature is measured in 1◦ C temperature changes. The instrument typically utilizes a coaxial fiber optic sensor positioned above the test sample to determine and record temperature changes. The temperature indicated at the initial appearance of crystallization signifies the cloud point. ASTM D5771 (stepped cooling method), D5772 (linear cooling rate method), and D5773 (constant cooling rate method) tests determine cloud point of petroleum products by utilizing automatic systems for sample cooling and optical sensors to detect the cloud point. During the controlled cooling process,
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the optical sensor monitors the transparency of the sample for a change in crystalline structure. The temperature at which wax crystals are sensed is taken as the cloud point. The test apparatus may be operated over a temperature range of −40 to +49◦ C with a resolution of 0.1◦ C. ASTM D5949 standard test for determining pour point of petroleum products utilizes the automatic pressure pulsing apparatus. This test applies a controlled burst of nitrogen gas onto the sample surface while the sample is being cooled. The instrument detects sample movement over a temperature range of −57 to +51◦ C with an optical sensor. ASTM D5950 standard test for determining pour point of petroleum products utilizes the automatic tilt apparatus. This instrument automatically tilts the test lubricant fixture during cooling and detects movement of the surface of the test specimen with an optical sensor. The test may be operated over a temperature range of −57 to +51◦ C. This test method is not intended for use with crude oils. ASTM D5985 standard test for determining pour point of petroleum products utilizes the rotational method. This test determines pour point via an automatic instrument that continuously rotates the test specimen against a suspended detection device during the cooling process. The no-flow point is detected by a change in the fluid viscosity or crystal structure that is sufficient to impede flow. The test may be operated over a temperature range of −57 to +51◦ C. This test method is not intended for use with crude oils. ASTM D6749 standard test for determining pour point of petroleum products utilizes the automatic air pressure tester. This test determines pour point by an automatic application of positive air pressure onto the sample surface during the cooling process. The lowest temperature at which surface deformation of the sample is observed is the pour point. The test may be operated over a temperature range of −57 to +51◦ C. This test method is not intended for use with crude oils. 30.2.1.19.1 Comments The pour point depends to a great extent on the type of crude from which the oil is made. Naphthenic oils exhibit naturally lower pour points. These materials, however, also tend to thicken more rapidly as the temperature lowers. Paraffinic base oils, on the other hand, contain waxy materials that solidify and become insoluble at temperatures near or slightly below the pour point. In these cases, flow may occur if the oil is subjected to mechanical shear or agitation. However, paraffinic oils must have sufficient flow characteristics to be able to be drawn into the pump. In contrast, naphthenic oil, although free of waxy structures, will not respond to shear forces when the pour point is reached. Again, pumpability will suffer unless the pour point can be depressed. The pour point of some paraffinic base oils may be improved by using refining processes or formulation with pour depressant additives. The cloud point of an oil is the temperature at which a distinct cloudiness or haze is discerned in the bottom of the test jar as paraffin wax and other materials, normally soluble, begin to separate. Since the test is a visual observation, the measurement is limited to oils that are transparent. Operating capillary or wick-fed oil systems at the cloud point will restrict lubrication effectiveness. In this case, wax particles or any fine dispersion contained in the oil will retard or may even prevent the flow of lubricant. International equivalent standards include (D97) IP 15 and ISO 3016. 30.2.1.20 Precipitation Number — ASTM D91 Semi-refined petroleum oils or black oils often contain naphtha insolubles, which are referred to as asphaltic materials. When determining the suitability of an oil product for use as a basestock lubricant, the asphaltic contamination level must be determined. In addition, in-service lubricants accumulate soluble and insoluble contaminants and the by-products of oil oxidation. These materials are detrimental to oil performance and long service life. Contaminants that are insoluble in oil may be separated and quantified by centrifuging the sample. However, soluble contaminants must be precipitated out of an oil sample by treatment with a solvent. The ASTM D91 precipitation number test method determines the asphaltic residues remaining in black oils and steam cylinder stock after refining. These residues are largely insoluble in paraffinic naphtha and, therefore, they may be readily extracted. Asphaltic residues are, generally, deleterious to in-service
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lubricants as they exhibit lower thermal and oxidative resistance, thus acting as a catalyst to increase oil oxidation. The ASTM D91 method defines the precipitation number of a lubricant as the quantity in milliliters of precipitate or sediment formed when 10% by volume of oil and 90% by volume of naphtha solvent are centrifuged under prescribed conditions. 30.2.1.20.1 Comments The D91 precipitation test may be applied to any oil but is primarily intended for steam cylinder and black oils. The test may be used to show the presence of foreign matter in the oil. Since the D91 test measures total solids that are insoluble in naphtha, separate analysis methods are required to quantify individual insolubles and any naphtha soluble materials that may be present. International equivalent standards include IP 75. 30.2.1.21 Refractive Index — ASTM D1218, ASTM D1747 Refractive index is a consistent fluid property and therefore a useful technique for checking fluid uniformity. The property is also useful in characterizing basestocks for fluids that have equivalent molecular weights. The refractive index of oil increases with increase in molecular weight, from paraffins to naphthenes to aromatics. Refractive index is defined as the ratio of the velocity (of a specified wavelength) of light in air to that in the test fluid. In the case of process oils and other petroleum products, the D-line of sodium (5893 Å/589 nm the most frequently used wavelength. To obtain the refractive index, a drop of sample is placed on the measuring prism face and the light source is adjusted in line with the refractometer telescope while the sample is uniformly lighted. After alignment, the scale is read and the value converted to refractive index at the appropriate temperature. The test temperature and light source are also reported with the refractive index reading. The ASTM D1218 test method determines the refractive index of transparent and light-colored hydrocarbons that have a refractive index ranging from 1.33 to 1.50, and at temperatures from 20 to 30◦ C. The test utilizes a high-resolution refractometer. The ASTM D1747 test determines the refractive index of viscous materials and low melting point solids that have a refractive index ranging from 1.33 and 1.60. These materials are tested at temperatures between 80 and 100◦ C. 30.2.1.21.1 Comments In conjunction with viscosity and specific gravity, refractive index may be used for hydrocarbon type identification. 30.2.1.22 Rust Prevention Characteristics — ASTM D665, ASTM D1748, ASTM D3603, ASTM D6557 Iron and steel alloys are the major metals used for fabricating machinery. Any contact between the native metal with air and moisture will cause rusting. It is therefore necessary to protect internal machinery surfaces from contact with dissolved or free moisture that may be carried along by the lubricant. In addition to surface damage, rust particulates will promote oil oxidation, cause abrasion of wear surfaces, clog oil passages and filters, and damage sensitive components such as servo-valves, etc. The ASTM D665 test utilizes a vertical steel rod immersed in a mixture of 300 ml of test oil and 30 ml of water (distilled or sea). The mixture is heated to 60◦ C (140◦ F) and stirred continuously during the test. After a period of 4 h the rod is examined for rusting. At the end of the time period, the test is reported as pass or fail depending on the observance of any rust. The ASTM D1748 test measures the ability of metal preservatives to prevent rusting of steel under conditions of high humidity. Polished steel panels are dipped in the test oil, allowed to drain, and suspended in a humidity cabinet at 48.9◦ C (120◦ F) for a specified test period. The size and number of rust spots on the test panel is interpreted as the result. A pass is taken to be no more than three spots of rust, where none is greater than 1 mm in diameter.
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In the ASTM D3603 method, a polished steel cylinder is immersed into a prepared sample, consisting of 275 ml of the test lubricant and 25 ml of distilled water for a 6-h period. At the end of the time period, the test is reported as pass or fail depending on the appearance of any rust. In addition to the above methods, the ASTM D6557 test may be used to evaluate the rust preventive characteristics of automotive engine oils. This test utilizes a series of test tubes containing the sample and a steel ball. As the tubes are mechanically shaken, air and an acidic solution (to provide a corrosive environment) are continuously applied to each tube for an 18-h period. The steel balls are imaged by an optical system to quantify the rust prevention capability of each sample oil. This standard method was designed to replace the ASTM D5844. 30.2.1.22.1 Comments The main significance of the rust prevention test is the determination of the ability of a given lubricant to maintain a reliable lubricant film and deny exposure of the metal surface to entrained and atmospheric oxygen. Applicability: hydraulic fluids; steam, hydro, and industrial gas turbine oils; and electrical insulation oils. International equivalent standards include (D665) ISO 7120, DIN 51566, and IP 146 and (D3603) NACE TM-01-72. 30.2.1.23 Saponification (Sap) Number ASTM D94 Many lubricants have fatty compounds added to increase their fluid film strength and improve water emulsibility characteristics. The fatty material displays a strong affinity for metal surfaces and enables oil to combine physically with the water instead of being displaced by it. The saponification (sap) number indicates the amount of these compounds (i.e., the degree of compounding). The sap number of a lubricant can be determined from two basic titration methods, one colorimetric and the other potentiometric. In the ASTM D94 sap number test, a specific quality of KOH is added to a specific quantity of sample and heated. In the reaction that follows, the fatty compounds are converted to a soap or saponified. Excess KOH is then neutralized by titrating with hydrochloric acid. The sap number is the quantity (milligrams) of KOH that was consumed in the reaction with the oil. 30.2.1.23.1 Comments In practice, the sap number is usually considered together with the neutralization number to determine the relative levels of acidic and fatty compounds in the lubricant. Note that the test results of new oils will vary depending on the level of acidic additives or other constituents in the formulation. In used oils, the sap value tends to increase in conjunction with the neutralization number. In addition, the nature and degree of contaminants will further distort used-oil test results. Considerable experience is necessary to properly interpret the results and meaning of sap number data. International equivalent standards include IP 136. 30.2.1.24 Sulfur Content — ASTM D129, ASTM D1266, ASTM D1552, ASTM D2622, ASTM D4294, ASTM D5453 All crude oils contain sulfur in elemental or various compound forms. Some sulfur compounds are acidic and corrosive, others have functional uses, for example, as naturally occurring antioxidants. During the lubricant refining process, the nonharmful sulfur compounds are retained to increase oxidation resistance. Additional sulfur compounds may be added to improve oxidation resistance or extremepressure performance for a particular application. The concentration of active, combined, and total sulfur content of lubricating oils may be determined from the ASTM DI29 sulfur content by pressure vessel oxidation method, the ASTM D1266 sulfur by titration method, the ASTM D1552 high-temperature method, the ASTM D2622 or D4295 x-ray fluorescence (XRF) spectroscopy methods, and the ASTM D5453 total sulfur by ultraviolet fluorescence method. The ASTM D129 sulfur by pressure vessel oxidation technique is generally applicable to all lubricating oils for measuring the total sulfur content. The procedure involves the ignition and combustion of a
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small sample of oil under pressurized oxygen. The products of combustion are collected and the sulfur is precipitated and weighed as barium sulfate. For measurement of sulfur levels from 0.01 to 0.4 mass%, the ASTM D1266 method is used. In this test, the sample is burned in a furnace and the combustion gases are titrated. The ASTM D1552 high-temperature method determines the concentration of sulfur in lubricating oils to 0.005% from the reaction of a known amount of potassium iodide with the sulfur dioxide given off by burning the oil sample in a high-temperature oxygen stream. The ASTM D2622 sulfur by XRF spectroscopy method determines the concentration of sulfur in lubricating oils by x-ray fluorescence spectroscopy. Modern high-power wavelength and energy dispersive XRF instruments can measure sulfur concentrations of 0.005% and are not affected by the presence of dissolved metals. The test utilizes an automated spectrometer system and requires much less sample preparation and analysis time than other sulfur determining methods. Alternatively, the ASTM D4294 sulfur by XRF spectroscopy may be used to determine the total sulfur content from 0.05 to 5 mass %. The ASTM D5453 test method determines total sulfur in liquid hydrocarbons that boil from approximately 25 to 400◦ C, with viscosity between 0.2 and 20 cSt, at room temperature. The sulfur is oxidized to sulfur dioxide, which absorbs energy from ultraviolet (UV) light. The amount of sulfur is derived from the resulting UV fluorescence.
30.2.1.24.1 Comments The undesirable consequences of sulfur compounds in oil are corrosion of metal surfaces, particularly copper. However, in machinery applications utilizing hypoid gears or metal cutting tools, the greater extreme pressure and increased load-carrying ability of the lubricant formulation effectively offsets the corrosion problem. With the exception of “active” sulfur containing lubricants, sulfur content of general-purpose lubricants is usually limited by machinery operating specifications. International equivalent standards include (D129) IP 61, (D4294) ISO 8754, and (D5453) ISO 20846.
30.2.1.25 Thermal Stability — ASTM D2070, ASTM D5579, ASTM D5704 Thermal stability refers to the resistance to oil degradation or property change due to thermal stress and is characteristically important for machinery oils. The thermal stability of hydraulic fluids is determined by the ASTM D2070 test. The test fluid is heated to 135◦ C for 168 h in a beaker containing copper and iron rods. At the end of the test, the rods are rated visually for discoloration. Additionally, the change in viscosity of the oil, the amount of sludge formed in the oil and the weight loss of the copper rod can be determined. The ASTM D5579 test determines the thermal stability of heavy-duty manual transmission lubricants utilizing a cyclic durability test stand. The thermal stability performance of the lubricant is determined from a specified number of shifting cycles that can be performed without failure of synchronization when the transmission is operated while continuously cycling between high and low range. The results are determined from a visual inspection of the transmission parts. The ASTM D5704 test determines thermal and oxidative stability of lubricating oils used in manual transmissions and final drive axles. The test, commonly referred to as the L-60-1 test, utilizes a heated gearcase containing two spur gears, a test bearing, and a copper catalyst. The lubricant is heated to a specified temperature and operated for 50 h at prescribed loads and speeds during which air is bubbled through it. Thermal stability is determined from viscosity, insolubles, and gearcase cleanliness.
30.2.1.25.1 Comments These tests are useful for determining oil deterioration characteristics under operating conditions and are suitable for determining additives and base oils for thermal degradation tendencies. International equivalent standards include (D5704) STP512A L-60-1.
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30.2.1.26 Viscosity and Viscosity Index — ASTM D445, ASTM D2270, ASTM D2983, ASTM D4683, ASTM D5481, ASTM D6080, ASTM D6616, ASTM D6821 Generally speaking, viscosity is the measure of a lubricant’s internal friction or resistance to flow or the ratio of shear stress to shear rate. In a given machine, the optimum thin film thickness separating the moving surfaces can be maintained only if the operational viscosity range is correct. Consequently, viscosity is probably the most important property of a lubricant specification. Changes in the viscosity of in-service oil can indicate ingress contamination or fluid degradation. Contamination of oil by fuel, light oils, or solvents will lower its viscosity level. Oxidation and combustion by-products tend to thicken the oil, a condition indicated by an increase in viscosity level. Dynamic (absolute) viscosity is the measure of the tangential force required to shear one parallel plane of a fluid over another. The force is proportional to the fluid viscosity, the planar area being sheared, and the rate at which the adjacent parallel planes are being forced to slide over each other, divided by the thickness, of the film. At any given temperature, the dynamic viscosity is related to force, fluid density, film thickness, and area by the following formulae: Dynamic Viscosity = (Force/Area) × (Film Thickness/Velocity) = Kinematic Viscosity × Density The force/area parameter is reduced to units of pressure — Pascal (Pa). The film thickness/velocity parameter is reduced to units of time — seconds (sec). Thus, the dynamic viscosity is expressed as Pascalseconds or Poise (P). For most lubricant applications, Pascal-seconds and Poise are inconveniently large units and dynamic viscosity is generally reported in milliPascal-seconds (mP) or centipoise (cP). The absolute viscosity is generally measured by a rotating spindle or ball-type viscometer such as the Brookfield or tapered bearing simulator (TBS). Note that consistent measurement results will depend on reproducibility of the oil type and measurement conditions. Multigrade oil will exhibit a decrease in viscosity as the shear rate increases; thus, absolute viscosity values will change with different instrument types unless exact shear rate and temperature conditions can be duplicated. The Brookfield viscometer measures dynamic viscosity under low shear rate conditions. To ensure effective lubrication of high shearrate machines, lubricating oils should also be tested with a high shear-rate viscometer such as the TBS. The TBS viscometer measures absolute viscosity under high and very high shear-rate conditions and allows the evaluation of monograde and multigrade oils at or near the shear-rates of modern high-performance reciprocating machines. The ASTM D445 kinematic viscosity test is the most popular measurement used to determine the nominal viscosity of machinery oils. The unit of measurement is the Stoke and is equal to one square centimeter per second. For most lubricant applications, the Stoke is an inconveniently large unit. A smaller unit, the centistoke (cSt) — equal to one square millimeter per second — is preferred. The D445 test is usually performed at a lubricant temperature of 40 or 100◦ C to standardize the results obtained and allow comparison among different users. Note, always report the oil temperature when performing kinematic viscosity. The ASTM D445 test determines the kinematic viscosity of liquid lubricants by measuring the time taken for a specific volume of the liquid to flow through a calibrated glass capillary viscometer under specified driving head (gravity) and temperature conditions. The method is applicable both to transparent and opaque fluids. The kinematic viscosity is the product of the time of flow and the calibration factor of the instrument. The ASTM D2983 utilizes the Brookfield rotary viscometer. The Brookfield instrument measures dynamic viscosity as a function of the resistance to the rotational force of a metal spindle immersed into the oil. This test quantifies dynamic viscosity under low-temperature, low shear rate conditions. Note, always report lubricant temperature when performing viscosity tests. The test reports viscosity readings in centipoise (cP). The D2983 test procedure was developed to determine the viscosity performance of crankcase, gear, industrial, and hydraulic oils over the temperature range of +5 to −40◦ C. Note that a variation of this method is widely used to determine the viscosity of new and used lubricating and
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hydraulic oils. In this case, the test is generally performed at ambient temperature. Performing the test at ambient temperature produces ambiguous results. Utilizing an immersion bath to control the sample temperature during the test will improve results. Alternatively, the ASTM D6080 may be used to determine the viscosity of synthetic and petroleum hydraulic fluids. The ASTM D5481 determines the apparent viscosity at high-temperature and high-shear rate (HTHS) utilizing a multi-cell capillary viscometer. This viscometer is instrumented for pressure, temperature, and timing. The test is conducted at a temperature of 150◦ C and a shear rate corresponding to an apparent shear rate at the viscometer tube wall of 1.4 million reciprocal seconds (1.4 × 106 sec−1 ). The viscosity is determined from calibrations previously established for Newtonian oils over a range of 2 to 5 mPaS at 150◦ C. Note that the results determined by this method may not correspond with viscosity measurements generated by other instruments. The ASTM D6616 viscosity at high shear rate by TBS viscometer utilizes a tapered rotor that rotates at 3600 rpm in a close fitting stator containing the oil under test. The TBS measures viscosity as a function of the resistance to the rotational force of tapered journal and plain bearing set operating under a prescribed load. The shear rates are applicable to high-speed reciprocating machinery. High shear-rate viscosity measurements may also be performed by the ASTM D4683 using the TBS. ASTM D6821 determines the viscosity for driveline lubricants (gear oils and automatic transmission fluids) utilizing a constant shear stress viscometer at temperatures from −40 to 10◦ C. The viscosity is determined by applying a prescribed torque and measuring the resulting rotational speed after a controlled temperature regime. The result is reported as milliPascal seconds (mPa/sec). Viscosity index (VI) is defined as the extent to which oil resists viscosity change due to temperature change. The higher the viscosity index, the less an oil’s viscosity will change as a result of temperature change. Viscosity index data is generally used to assess the viscosity performance of oil over the normal operating temperature range of a machine including its start-up and shutdown conditions. The viscosity of most machinery fluids will tend to decrease as the fluid temperature increases. The viscosity index will change with oil chemistry. However, the viscosity of all lubricants does not respond to temperature changes in a consistent manner. Monograde oil (Newtonian) will generally exhibit a constant viscosity/temperature at all (except very high) shear rates. For monograde oils, the viscosity will decrease as the temperature rises, and vice versa. However, multigrade oil is a blend of light basestock oil and a thick viscosity index improver (VII) agent. Multigrade oil is a non-Newtonian fluid and its viscosity will not be as affected by temperature change as will be similar viscosity monograde oil. The ASTM D2270 procedure determines viscosity index from kinematic viscosity (ASTMD445) readings taken at 40 and 100◦ C. The D2270 calculation is rather awkward and complex and Table 30.1 in the practice will greatly speed the process of determining the viscosity index for a particular lubricant. 30.2.1.26.1 Comments Each machinery application will require a specific viscosity grade lubricant depending on operating temperature, load, and speed conditions. In general, low speed, high-load, or high-temperature machines will utilize higher viscosity oil than machines that are lightly loaded or operate at higher speeds or temperatures. Thus care is essential in selecting the correct lubricant for each specific application. First, the viscosity results determined by each ASTM approved method will generally not correspond to the viscosity reading of another. Since oil manufacturers may utilize an instrument of choice, care is required when comparing the viscosity of different products. Second, not all oil manufacturers’ brands will exhibit the same viscosity for a given oil type or grade when the same viscosity measurement method is utilized. Table 30.1 is a compilation of three manufacturers’ oils by grade. The table compares the range of viscosity for ISO, SAE, and AGMA oil products. The table also indicates the acceptable minimum and maximum viscosity range for each ISO oil grade. Note that the viscosity for SAE types such as SAE 30 or SAE 15W40 can vary by more than the 10% limit imposed on ISO types. Therefore, determining the actual viscosity of SAE and AGMA oils before use is an important consideration if maximum lubrication reliability is to be obtained. International equivalent standards include (D445) IP 71, DIN 51566, and ISO 3104, and (D2270) IP 226 and ISO 2909.
© 2006 by Taylor & Francis Group, LLC
Lubricant Properties and Test Methods TABLE 30.1
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Range of Viscosity for ISO, SAE, and AGMA Oil Products SAE multigrade
Viscosity at 40°C
Low limit
High limit
ISO R&O, AW, EP
2.2
1.98
2.42
ISO 2
3.2
2.88
3.52
ISO 3
4.8
4.14
5.06
ISO 5
6.8
6.12
7.48
ISO 7
10
9.0
11.0
ISO 10
15
13.5
16.5
ISO 15
22
19.8
24.2
ISO 22
32
28.8
35.2
ISO 32
SAE 5
46
41.4
50.6
ISO 46
SAE 10
50
45.0
55.0
68
61.2
74.8
68
61.2
74.8
68
61.2
74.8
85
76.5
93.5
100
90
110
100
90
110
SAE 10W40
100
90
110
SAE 15W40
100
90
110
115
103
126
SAE 30
SAE 5W50
115
103
126
SAE 40
SAE 10W40
115
103
126
130
117
143
SAE 30
SAE 20W40
130
117
143
SAE 40
SAE 75W90
150
135
165
SAE 40
SAE 80W90
150
135
165
SAE 50
SAE 20W50
180
162
198
220
198
242
220
198
242
320
288
352
ISO 320
SAE 60
460
414
506
ISO 460
SAE 140
680
612
748
ISO 680
1000
900
1100
ISO 1000
1500
1350
1650
ISO 1500
2200
1980
2420
ISO 2200
3200
2880
3520
SAE monograde
AGMA R&O
AGMA EP
#1 SAE 0W30
ISO 68
SAE 20
SAE 5W30
#2
2EP
#3
3EP
#4
4EP
SAE 10W30 SAE 20W 20 SAE 30 ISO 100
SAE 30
SAE 5W40 SAE 5W50
SAE 75W90
SAE 15W40
ISO 150
ISO 220
SAE 40
SAE 20W50
SAE 50
SAE 80W140
#5
5EP
SAE 85W140
#6
6EP
#7
7EP
SAE 90
#8
8EP
SAE 140
#8A
8AEP
SAE 250
#9
9EP
#10
10EP
30.2.1.27 Wear Prevention/Load-Carrying Properties — ASTM D2670, ASTM D2882, ASTM D4172, ASTM D5182 Load-carrying ability is generally a function of the oil’s basestock lubricity and the antifriction and antistick characteristics provided by additives. The complex relationships between the oil and the various additives require different test methods and apparatus depending on the oil blend and intended machinery application. ASTM D2670 wear properties of fluid lubricants by Falex pin and vee block determines the wear properties for fluid lubricants by means of the Falex pin-on-vee-block machine operating under prescribed conditions. This test determines the load carrying ability and wear properties of fluid lubricants by the degree of wear of a steel pin (journal) rotating at 290 rpm against a pair of v-shaped blocks immersed in the lubricant sample for a period of 15 min. The load on the pin is ratcheted to a prescribed load for the
© 2006 by Taylor & Francis Group, LLC
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duration of the test. The wear scars on the pin and vee blocks are interpreted to determine the test fluid’s wear properties. ASTM D2882 wear properties of hydraulic fluids by constant volume pump determines the degree of wear on pump components by high-pressure hydraulic fluids. The test utilizes a rotary vane pump that pumps the test fluid through a standardized circuit at 2000 psi for 100 h. The results are determined from the weight loss of the pump cam ring and vanes. ASTM D4172 wear preventive characteristics of lubricating fluids determines the antiwear characteristics of fluid lubricants utilizing the four-ball wear test machine. The tester utilizes a steel ball rotating at 1200 rpm against three stationary steel balls lubricated with the sample. The sample is temperature controlled at 75◦ C and the load between the upper and lower balls is set at 147 or 392 N. The test is run for 60 min and the results are determined from the scar diameters worn in the three lower balls. ASTM D5182 standard test method determines the scuffing load capacity of oils used to lubricate hardened steel gears. The test utilizes the Forschungstelle fur Zahnrader und Getriebebau (research site for gears and transmissions) visual method, commonly referred to as the FZG visual method. The test is primarily used to assess the resistance to scuffing of low-additive oils such as industrial gear oils and transmission and hydraulic fluids. In this test, the FZG gear machine is operated at constant speed for a prescribed period at increasing loads until failure criteria is reached. The FZG method is not appropriate for evaluating high EP oils as the load capacity of these oils generally exceed the capacity of the test rig. 30.2.1.27.1 Comments Gear tests are a proven means for assessing the load-carrying capacity of lubricants and the influence of various antiwear, antiscuffing, and extreme-pressure additives. Since these tests are intended to evaluate relatively low-viscosity oils for circulating systems, they are not generally suitable for evaluating compounded fatty lubricants used for open or hypoid gears. In addition to the above methods, the WAM high-speed load capacity test method has been submitted to SAE E-34C for approval in AIR4978 “Temporary Method for Assessing the Load Carrying Capacity of Aircraft Propulsion System Lubricating Oils.” The test utilizes the WAM1, WAM3, or WAM4 machines to provide Ryder-like load capacity data of gas turbine and gearbox oils. International equivalent standards include (D2882) IP 281 and DIN 51389, (D5182) DIN 51354, IP 334, and IP 166.
30.2.2 Semisolid Grease Lubricant Tests 30.2.2.1 Apparent Viscosity of Greases — ASTM D1092 The apparent viscosity of grease is defined as the ratio of shearing stress to rate of shear calculated from Poiseuilles’ equation for viscosity using capillary tubes. The ratio varies with shear rate that is proportional to the linear velocity of flow divided by the capillary radius. Rheological properties of grease such as apparent viscosity can be determined with a rather complex apparatus the principal elements of which are a pressure cylinder containing a floating piston and attached capillary flow tubes. A calibrated hydraulic pumping system and temperature-controlled test chamber complete the test setup. In ASTM D1092, the chamber temperature is controlled over the range of −54 to 38◦ C. The procedure employs eight capillary tubes and two pumping speeds to determine the apparent viscosity at 16 different shear rates. Because of grease thickness, measurements are limited to ranges of 25 to 100,000 P at 0.1 sec, and 1 to 100 P at 15,000 sec. At very low temperatures, the shear rate range may require reduction because of the force needed to move the grease through the smaller capillaries. 30.2.2.1.1 Comments Apparent viscosity of grease is of general use in considering flow problems occurring in the distribution and dispensation of grease from a central source through a network. Volume rate of flow, pressure drop in the system, etc. can be predicted and adjustments made to accommodate the lubrication needs of specific systems and various operating conditions. Apparent viscosity depends on the type of oil and the type and amount of thickener used in the grease.
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30.2.2.2 Cone Penetration Test — ASTM D217, ASTM D1403 Grease consistency describes the relative softness or hardness of grease. Grease consistency depends on the type and viscosity of the base oil and the type and proportion of the thickener used in the formulation of the grease. The NLGI classifies lubricating grease according to the ASTM D217 cone penetration test. The D217 test instrument measures grease penetration depth at a temperature of 25◦ C, for four categories of grease preparation: 1. Unworked sample: The grease sample is subjected to minimum disturbance in its transfer to the grease test vessel. 2. Worked sample: The grease sample is subjected to the shearing action of 60 double strokes in a standard grease worker. 3. Prolonged worked sample: The grease sample is subjected to the shearing action of 60 double strokes at a temperature of 15 to 30◦ C, followed by an additional 60 double strokes at a temperature of 25◦ C (standard grease worker). 4. Block penetration sample: This test indicates the penetration at 25◦ C on the freshly prepared face of a cube of grease that is sufficiently hard to maintain its shape. The results of the ASTM D217 cone penetration test are indicated by the depth (in tens of millimeters) up to which the cone descends into the grease sample. Alternatively, the cone penetration number of grease products may be determined with one-quarter and one-half scale instruments using the ASTM D1403 standard method. 30.2.2.2.1 Comments The higher the consistency number, the harder the grease product, as indicated by the lower depth penetration distance. Harder greases have less mobility and will tend to stay in place longer; however, flow and pumpability characteristics may suffer. Greases that are too soft can increase rotational friction due to churning. Consequently, it is important that grease consistency is sufficient to maintain a reliable lubricant film for the machinery application. Most general-purpose greases tend to have a NLGI consistency of one or two. Table 30.2 indicates the standard NLGI consistency grades as determined by the ASTM D217 cone penetration test. TABLE 30.2 NLGI Consistency Grades determined by the ASTM D217 Cone NLGI consistency Number
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Worked penetration range (tens of millimeters)
000
445 to 475
00
400 to 430
0
355 to 385
1
310 to 340
2
265 to 295
3
220 to 250
4
175 to 205
5
130 to 160
6
85 to 115
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When the cone penetration test is performed at a temperature of 0◦ C on the same grease type, the sample will generally be one to two NLGI consistency numbers higher (harder). Conversely, grease will generally indicate one NLGI consistency number lower (softer) when the test is performed on the same grease at 43◦ C. International equivalent standards include (D217) IP 50 and ISO 2137. 30.2.2.3 Copper Corrosion Resistance — ASTM D4048 Copper alloys are utilized in many greased components and it is essential that these parts are properly lubricated by the grease and that the grease is noncorrosive to metallic copper. Since grease is essentially thickened lubricating oil that may contain compounds that are corrosive to copper, it is desirable to test grease products for their corrosion tendency. In the ASTM D4048 test, a polished copper strip specimen is immersed into a specified volume of grease at 100◦ C for a period of 24 h after which it is compared with the copper strip corrosion standard to determine the degree of tarnishing or corrosion. 30.2.2.3.1 Comments The copper corrosive tendency results have some value in providing a product-to-product or batch-tobatch assessment of the compositional quality, consistency, and uniformity of the grease. International equivalent standards include (D4048) IP 112, ISO 2160, and DIN 51811. 30.2.2.4 Dropping Point of Grease — ASTM D566, ASTM D2265 The dropping point is defined as the temperature where grease passes from a semisolid to a liquid state. The dropping point temperature determines the maximum service temperature range for grease. The ASTM D566 test procedure determines the dropping point of grease under controlled conditions. The test apparatus includes a test cell that is immersed in a 400-ml Pyrex™ bath for heating at the prescribed rate. An integral heater, temperature controller and stirrer maintain uniform control of the temperature and its rate of rise up to 288◦ C. The test sample is heated incrementally until the dropping point is observed. The temperature at which a liquid drop first falls from the sample cup is reported as the dropping point of the sample. The ASTM D2265 standard test method may be utilized to determine grease dropping point over a wider temperature range up to 316◦ C. This test is usually performed on high-temperature grease products. 30.2.2.4.1 Comments Grease must not soften and flow during normal operating temperatures. Note, grease consistency and firmness will change when the machinery warms up from ambient to operating temperature. The dropping-point property is useful for checking the quality and uniformity of a specific manufacturer’s product and for comparing various grease brands for a specific application. Do not estimate consistency or dropping point by look-and-feel observations. Grease will be much softer in summer than in winter and also when the operating temperature is above ambient. Pay particular attention to automatic grease dispensers that must operate over wide temperature ranges. Since most grease products are semi-solid mixtures of oil and one or more kinds of soap or thickening agents, they do not usually exhibit well-defined melting points. Therefore, the dropping point only approximates this characteristic and has limited significance in terms of service performance. Many other factors (i.e., design, speeds, loads, evaporation losses, thermal cycles, etc.) greatly influence useful operating temperature of the grease and, therefore, must be recognized in lubricant selection for a particular application. International equivalent standards include (D566) IP 132, ISO 2176, and DIN 51806, and (D2265) IP 132, DIN 51806 and ISO 2176. 30.2.2.5 Extreme-Pressure Properties of Grease — ASTM D2509, ASTM D2596, ASTM D5706 One of the most important attributes of grease is the maximum load it can sustain without seizure or scoring of the lubricated parts.
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The ASTM D2509 utilizes the Timken wear test machine for rating the load-carrying and EP properties of grease lubricants. The Timken apparatus employs a hardened steel ring rotating against a flat steel block that is lubricated by the test lubricant. The maximum load applied without scoring is the Timken OK load (pass). The minimum load that causes scoring is reported as the score load. The ASTM D2596 test utilizes the Falex four-ball test apparatus where four half inch diameter steel balls (AISI E-52100) are arranged with three balls clamped together so as to cradle and load the top ball. This top ball is then rotated on a vertical axis. The EP property of a grease product is determined from the load-wear index. This property is derived from ten wear tests of varying loads up to and including the point at which seizure occurs (weld point). The ASTM D5706 test determines the EP properties of lubricating greases using a high-frequency, linear-oscillation SRV machine. The SRV machine utilizes a steel ball oscillating against a lubricated steel disk. Load is increased in 100 N increments until seizure occurs. 30.2.2.5.1 Comments These tests compare the relative EP characteristics of grease products and are useful in grease product quality control and to compare different grease products for a particular application. Note that there is no general correlation between EP test results and anticipated in-service performance. International equivalent standards include (D2509) IP 326 and (D2596) DIN 31350. 30.2.2.6 Leakage Tendencies of Wheel Bearing Grease — ASTM D1263, ASTM D4290 Greases used to lubricate automotive wheel bearings must operate under conditions of high speed, high load, and excessive braking forces. Greases subject to oil leakage from mechanical or thermal softening effects can lead to bearing failure. The ASTM D1263 test method determines the leakage tendencies of wheel bearing greases when tested under prescribed laboratory conditions. The grease is applied to a modified front wheel hub bearing assembly. The hub is rotated at a speed of 660 rpm for 6 h at a temperature of 105◦ C. The result is determined from the leakage of the grease or oil, or both in grams and the condition of the bearing surfaces. The ASTM D4290 accelerated leakage tendency test evaluates the leaking tendency of automotive wheel bearing greases in a modified automotive front wheel hub-spindle-bearings assembly. The D4290 test employs severe conditions including 111 N thrust load, 1000 rpm, and a spindle temperature of 160◦ C to induce grease deterioration and failure. The test continues for a 20-h period, after which leakage of grease and oil is measured and the bearings are washed and examined for deposits of gum and varnish. The amount of leakage is the test result. 30.2.2.6.1 Comments The test provides a basis for comparing the relative tendencies of different greases to leak from a simulated wheel bearing operation. Since the evaluation is performed under accelerated conditions, it does not equate to actual longtime service nor reproduce actual highway stresses involving load, vibration, and temperature. It is possible, however, to distinguish products having decidedly different leakage characteristics and to judge qualitative changes (i.e., slumping, softening, etc.) in a grease. The value of results is enhanced when used in combination with other stability tests, for example, oxygen bomb stability, roll stability, etc. 30.2.2.7 Life Performance of Grease Products — ASTM D1741, ASTM D3336, ASTM D3337, ASTM D3527 Usable life is an essential consideration when purchasing lubricating grease products. The usable life of grease is related to its ability to resist oxidation and mechanical stresses in the presence of machinery metals and atmospheric oxygen and water. The ASTM D1741 test describes the measurement of life performance and leakage characteristics under prescribed operating conditions at temperatures up to 125◦ C. The test apparatus utilizes a cleaned 306 ABEC Class 3 bearing packed with the candidate grease product. The bearing is “run in” at 3500 rpm for about 2 h to generate an equilibrium state and then run to failure. The time to failure is the reported
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result. In the leakage evaluation test, the test bearing is operated on a 20/4-h on/off cycle until failure is observed. The ASTM D3336 determines usable life performance employing a high temperature test (up to 371◦ C). The test utilizes a clean 204 size ball bearing packed with a weighed amount of grease equivalent to 3.2 cm3 . After “working in” the grease by hand rotating the bearing, the system is driven at 10,000 rpm for about 1.5 h to establish equilibrium test conditions and then operated until failure occurs. The time to failure is the reported result. The ASTM D3337 test employs a small R-4 size, AFBMA Class 7 ball bearing. The method specifies that the bearing is packed 13 full with the grease-under test and immediately installed on the test spindle. The bearing is “run in” for about 100 revolutions at a speed of less than 200 rpm. The test is conducted at 12,000 rpm until failure. During the test, total test hours, torque-meter tare, torque meter reading, net torque, torque in gram-centimeter, and outer-race bearing temperature are recorded. The ASTM D3527 life performance test evaluates the high-temperature stability of automotive wheel bearing greases in a modified automotive front wheel hub-spindle-bearing assembly. The D3527 test employs severe conditions (25 lbf [111 N] thrust load, 1000 rpm, 160◦ C spindle temperature) to induce grease deterioration and failure. The test continues in a 20/4 h on/off cycle until grease breakdown causes the measured drive motor torque to increase past an established endpoint. The number of hours to failure is taken as the test result. 30.2.2.7.1 Comments The above test methods are screening techniques that may be used for the evaluation and selection of candidate greases for particular applications. Since these tests do not replicate in-service conditions, improved operation in the test rig may not correlate to improved in-service performance. The tests cannot distinguish between minor differences in performance life. 30.2.2.8 Low-Temperature Torque Characteristics — ASTM D1478, ASTM D4693 Greases become more solid-like at very low temperatures and consequently require a greater force to overcome resistance to movement. The ASTM D1478 test measures the torque (gram centimeters) required to withhold the movement of the outer ring of a 204 size open ball bearing packed with the test grease as the inner ring is rotated at one rpm at a prescribed test temperature ranging from −75 to −20◦ C. The running torque is determined after 10-min rotation of the bearing. The ASTM D4693 test determines the low-temperature torque of grease-lubricated wheel bearings. The test grease is packed into a specially manufactured, spring-loaded, automotive type wheel bearing assembly. The assembly is heated and cold soaked to −40◦ C. The spindle is rotated at one rpm and the torque required to prevent rotation is measured after a period of 60 sec. 30.2.2.8.1 Comments As with apparent viscosity, the low-temperature torque displayed by a grease depends on the type and amount of thickener as well as the viscosity characteristics (including temperature coefficient) of the lubricating fluid. Low-temperature torque of greases is of special interest for military, space and certain applications involving small bearings and relatively limited available torque, for example, mechanical timing devices. 30.2.2.9 Oil Evaporation and Oil Bleed — ASTM D972, ASTM D2595 The proper operation of grease-lubricated components requires the oil base, additives and thickener to remain consistent during machinery operation. When the ratio of oil to thickener decreases, the grease hardens and lubrication is impaired. Oil separation is the result of two different processes, evaporation and bleed:
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1. Oil evaporation refers to the loss of volatile light-ends of the base-oil or additive constituents at high operating temperatures. Oil evaporation causes grease to thicken into a dry paste-like consistency unsuitable for lubrication. 2. Oil bleed refers to the process of oil separation or seeping from grease during operation. This condition is easily recognized by the oily appearance of the greased area and often the formation of little pools of oil. Oil bleed will harden grease, rendering it unsuitable for use. The ASTM D972 and ASTM D2595 tests determine the evaporation loss tendency of lubricating grease over a wide temperature range. The test apparatus consists of an evaporation cell placed into a temperature-controlled chamber. A helical heating tube wrapped around the cylinder circumference provides the source of a controlled flow of evaporating air as it passes over the surface of the sample and then out through a central stack. 20 g of the grease sample is placed in the cell and heated to the desired temperature for a period of 22 h. The D972 test is limited to a temperature up to 149◦ C. The D2595 augments D972 by utilizing a similar test cell placed into a thermostatically controlled aluminum block oven to determine evaporation losses up to a temperature of 316◦ C. The sample is weighed before and after the test and the difference is taken as the weight loss due to evaporation. 30.2.2.9.1 Comments Evaporation losses may include by-products from thermal cracking in addition to egress of volatile ingredients in the grease blend. As these materials evaporate, the grease will also undergo an increase in viscosity. Ultimately, the lubricant will be completely degraded leaving a hard or tarry residue. Note the evaporation test is static and the results cannot be directly interpreted in terms of dynamic in-service conditions. 30.2.2.10 Oil Separation from Lubricating Grease — ASTM D1743, ASTM D6184 Oil loss due to separation causes lubricating grease to thicken and harden leading to a serious lubrication problem. The ASTM D1743 standard test method for oil separation from lubricating grease during storage determines the tendency of lubricating grease to separate during storage in a 35-lb pail. The sample is placed on a sieve inside a special test cell and subjected to 0.25 psi (1.72 kPa) air pressure at constant temperature. Any oil that bleeds from the grease during a 24-h period is collected in the cell and weighed. The ASTM D6184 determines the tendency of oil to separate from lubricating grease at an elevated temperature. The test apparatus consists of a 60-mesh nickel gauze cone with wire handle, mounted inside a 200-ml covered beaker. The grease sample is placed in the wire gauze cone and heated to a test temperature for a period of time. The sample is weighed before and after the test to determine the weight loss from oil separation. The test method is conducted at 100◦ C for 30 h unless other conditions are required by the grease specification. The test is not suitable for greases having an ASTM D217 penetration greater than 340 (softer than NLGI No. 1 grade). 30.2.2.10.1 Comments Grease products that are stored for extended periods and products that are intended for use in hightemperature applications should be verified for oil separability before use. Note the ability of grease to be reliably delivered by a pump is governed by the properties of pumpability and slumpability. Grease pumpability is the ability of a grease product to be pushed through plumbing and nozzles of a powered grease applicator. Slumpability is the ability of a pump to draw grease from a reservoir. Oil separation and subsequent grease hardening will affect the pumpability and slumpability properties of the grease. The pumpability and slumpability of grease products are related to the following factors: 1. Grease products made from highly viscous base oils generally exhibit poor pumpability at low operating temperatures. 2. Grease products with a smooth buttery consistency will generally exhibit good pumpability but poor slumpability.
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3. Grease products with a fibrous consistency will generally exhibit good slumpability but poor pumpability. Choosing the correct grease for use in a powered applicator is essential to effective lubrication of downstream components. International equivalent standards include (D1743) DIN 51817 and IP 121. 30.2.2.11 Oxidation Stability — ASTM D942 As with oil lubricants, grease products will also oxidize and degrade from the moment of use. Atmospheric oxygen will react with grease constituents and form acidic by-products, which will eventually degrade into sludge and varnish deposits. Since grease lubrication systems generally contain a small quantity of grease, good oxidation stability is a key requirement to a reasonable service life. The ASTM D942 oxidation stability test determines the oxidative characteristics of grease by means of a pressure drop in oxygen after a prescribed time period has elapsed. The test apparatus includes five tiered shelves that support a number of grease sample dishes. The shelves are placed into a steel pressure vessel and initially pressurized with oxygen to 758 kPa and heated to a temperature of 99◦ C. At the end of the test period, the pressure drop indicates the amount of oxygen that was consumed by the grease. 30.2.2.11.1 Comments Changes in the observed oxygen gas pressure are the combined result of gas absorption and reaction processes. A plot of pressure vs. time will normally show a characteristic induction period (gradual decrease) after which the pressure drops sharply indicating the beginning of autocatalytic oxidation process and thus the end of usable life. Greases exhibiting lower pressure drops over the prescribed test period will generally have a lower tendency to deteriorate in storage. International equivalent standards include DIN 51808 and IP 142. 30.2.2.12 Rust Prevention Characteristics — ASTM D1743, ASTM D6138 The ability of grease products to provide a barrier to water and atmospheric oxygen and prevent the rusting of machinery components is an important property. The ASTM D1743 test determines the corrosion preventive properties of greases using grease-lubricated tapered roller bearings stored under wet conditions at 52◦ C for 48 h after which they are checked for the degree of tarnishing or corrosion. Prior to the test, the bearings are operated under a light thrust load for 60 sec to distribute the lubricant. This test method is based on CRC Technique L 41 that correlates laboratory results with in-service grease-lubricated aircraft wheel bearings. The ASTM D6138 corrosion preventive properties test for lubricating greases utilizes a clean lubricated double-row self-aligning ball bearing. The bearing is packed with grease, partially immersed in water and operated under a prescribed sequence with no load at a speed of 80 rpm for a period of one week after which it is checked for the degree of tarnishing or corrosion (Emcor test). 30.2.2.12.1 Comments These methods can distinguish between good and poor-performing greases, but are not suitable for ranking manufacturer brands within a specific product category. International equivalent standards include (D6138) DIN 51802, ISO 11007, and IP 220. 30.2.2.13 Water Wash out Characteristics — ASTM D1264, ASTM D4049 Many grease-lubricated components are normally exposed to environmental moisture. Water contamination generally affects grease reliability by reducing its load-carrying ability due to the formation of water–oil emulsion or by changing its texture, consistency, and/or tackiness. The ASTM D1264 water washout test evaluates the ability of lubricating grease to adhere to operating bearings when subjected to a water spray. A grease sample is packed into a ball bearing test specimen and subjected to a steady water stream under controlled test conditions. The test result is the percentage of grease (by weight) washed out in a 1-h period, tested at 38 and 79◦ C.
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The ASTM D4049 water washout test evaluates the ability of lubricating grease to adhere to a metal surface when subjected to a water spray. The grease sample is applied to a stainless steel test panel and subjected to a direct water spray at a specified pressure (40 psi), temperature, and time period. The test result is the percentage of grease (by weight) sprayed off after a 5-min period. 30.2.2.13.1 Comments Response of greases to free water can involve rather complex chemical and physical effects. Some grease products naturally resist water ingress or incorporate it as droplets; other greases absorb water while forming “emulsions,” performing normally and otherwise remaining relatively unchanged. The water washout test does not duplicate in-service conditions and results are subjective with relatively poor precision. However, the results are useful for assessing the water washout behavior of greases and comparing candidate greases for a particular application. International equivalent standards include (D1264) IP 215 and DIN 51807. 30.2.2.14 Wear Prevention/Load-Carrying Properties — ASTM D2266, ASTM D3704, ASTM D4170 Load-carrying ability is generally a function of the oil’s basestock lubricity and the antifriction and antistick characteristics provided by additives. The complex relationships between the oil and the various additives require different test methods depending on the oil blend and intended machinery application. Falex, Timken, or similar bearing and gear type wear machines are generally used to determine the wear and antiwear or extreme-pressure additive characteristics of lubricating oils and greases. The ASTM has standardized a number of methods utilizing these wear machines. The ASTM D2266 wear preventive characteristics of lubricating grease by Falex 4-ball determines the wear characteristics of grease lubricants in sliding steel-on-steel applications. The tester employs a steel ball rotating against three stationary steel balls lubricated with the test sample. The results are reported as dimensions of scar marks made on the ball surface. The ASTM D3704 wear preventive characteristics of lubricating grease by Falex block-on-ring test machine in oscillating motion determines the wear properties of grease lubricants in oscillating or sliding steel-on-steel applications. The test is used to differentiate between greases with different sliding wear properties. The tester employs a steel ring oscillating against a steel block lubricated with the test sample. The test speed, load, time, angle of oscillation, and specimen finish and hardness is varied to simulate real world machinery conditions. The results are reported as dimensions of scar mark made on the block by the ring oscillation. The ASTM D4170 fretting wear protection by lubricating greases determines the fretting wear potential and level of protection of greased components and is utilized to differentiate greases with superior antifretting characteristics. The test utilizes two antifriction thrust bearings lubricated with the test sample and operating under prescribed vibratory and load conditions. The test is run at room temperature for 22 h. The result is the weight loss to the upper and lower bearing races. 30.2.2.14.1 Comments These tests are useful screening tools for comparing greases for friction and wear characteristics. However, for precise antiwear results, tests using the specific bearings and/or gears and operated under the load, speed, temperatures, etc. that are representative of the intended machinery application are recommended. Note fretting is a destructive wear mode caused by low-level oscillatory motion between machinery parts. It frequently occurs in stored, standby, and emergency machinery that are shut down for long periods.
30.2.3 Solid Grease Lubricant Tests 30.2.3.1 Adhesion of Solid Film Lubricants — ASTM D2510 Effectiveness of a solid film lubricant system is dependent on its adhesion to the surfaces to be lubricated. It is important therefore to measure the “bonding” quality of the lubricant to these surfaces.
© 2006 by Taylor & Francis Group, LLC
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In ASTM D2510 test, specially prepared aluminum test panels are spray coated to produce a dry film thickness of between 0.005 and 0.013 mm. After measuring the actual dry film thickness deposited, the panels are immersed in a beaker of water (or other fluid) to half the depth of the coating for a 24-h period at room temperature. The panel and coated surface are then dried and the wetted and unwetted surfaces scratched with a stylus. Masking tape placed perpendicular to the parallel scratches is pressed on with a roller. The tape is then stripped abruptly from the surface. Any evidence of film damage or exposure of bare metal is reported as well as any other test conditions or observations that are important to interpretation of the results. 30.2.3.1.1 Comments The procedure provides a means of assessing the effectiveness and continuity of a solid-film lubricant coating process. Since several techniques may be used for applying a particular coating, testing the relative “adhesiveness” of one type of material vs. another should be conducted on a comparable basis for meaningful results. Test conditions may be varied to accommodate specific application requirements. 30.2.3.2 Corrosion of Solid Film Lubricants — ASTM D2649 Since solid film lubricants may contain chlorine, sulfur, and other elements, the corrosiveness of solid film lubricants is a necessary consideration. In ASTM D2649 test, a solid film lubricant is applied to precleaned aluminum test panels to produce a cured film thickness of between 0.005 and 0.013 mm. A coated panel is then located against an uncoated panel, positioned at a 90◦ angle in a channel fixture, and placed under a constant load by applying a torque of 2.8 Nm to the nut and bolt clamp. The test assemblies, including blanks, are preheated at 65.6◦ C for 2 h and then placed in a humidity (95◦ RH) chamber (49◦ C). After 500 h, the test units are disassembled and both coated and uncoated surfaces inspected for any pitting, etching or other evidence of corrosion. 30.2.3.2.1 Comments The method provides a technique for evaluation of corrosion tendencies of applied solid film lubricants. The test, being qualitative, is subject to interpretation. The results can be highly dependent on the procedure, materials and treatments used to obtain the film. Although the evaluation is performed using aluminum, consideration should be given to use of other test metals. It should also be recognized that solid film lubricants generally do not exhibit rust preventive characteristics. 30.2.3.3 Lubricating Qualities of Solid Lubricants — ASTM D1367 Graphite and similar types of solid materials have proven to be useful lubricants alone and when combined with other lubricants such as oils or greases. The ASTM D1367 is designed to evaluate the lubricating qualities of graphite and also may be used for evaluating other similar solids. To conduct the test, 36 g of the powdered solid are added to 204 g of paraffin base oil having a viscosity of 20.5 to 22.8 cSt at 37.8◦ C. A cleaned, weighed double-row test ball bearing is mounted on a test drive shaft and the assembly is lowered into a special beaker containing the test mixture. After a 2-h operation at 1750 rpm at room temperature, the test bearing is removed, cleaned, and weighed to determine its weight loss. 30.2.3.3.1 Comments The test method provides a means for quality control checking the lubricating quality of solid lubricants. Although not a precision method, results may be used for comparative evaluation of different types of solid lubricants used in combination with other lubricating fluids. When the carrier fluid is run separately to establish a “norm,” a means for identifying the contribution of the solid lubricant to the lubricating process can be determined. Addition of the “lubricating” solid can cause either an increase or decrease in the wear and load carrying capacity of the fluid lubricant. Possible interactions between additive components in fluids and solid lubricants also may be examined.
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30.2.3.4 Thermal Shock Sensitivity of Solid Film Lubricants — ASTM D2511 In applications where extreme temperature variations occur, measuring the capability of a solid film to withstand thermal shock is useful. To perform a test, the solid film lubricant is applied on test panels so as to obtain a finished film thickness of between 0.005 and 0.013 mm that is accurately determined by micrometer measurement. The test panels are first heat soaked at 260◦ C for 3 h and then immediately cooled to −54◦ C for 3 h. Subsequently, the test panels are returned to room temperature and examined visually for damage. If the surfaces are intact, additional testing to determine the adhesion of the film (ASTM D2510) may be needed. The same criteria of failure, that is, flakes, cracks, blisters, etc. are evidence of surface coating deficiencies and are noted in the report of the results. 30.2.3.4.1 Comments As with the adhesion test (ASTM D2510), this evaluation assesses the bond strength of solid film lubricants subjected to extreme thermal stress. The evaluation is more qualitative than quantitative and, hence, is subject to individual judgment. 30.2.3.5 Wear and Load-Carrying Capacity of Solid Lubricants — ASTM D2625, ASTM D2981 Wear characteristics and load-carrying ability of solid film lubricants are important for metal-to-metal sliding applications. The ASTM D2625 employs the Falex lubricant tester. To determine wear life, two lubricant coated stationary steel V-blocks are pressed at a specific load against a rotating lubricated steel pin until the solid film is depleted, evidenced by a torque increase of 1.13 Nm above the steady state value. For measuring load-carrying ability, the test is repeated with incremental increases in load until a sharp increase in torque (1.13 Nm) or breakage of the pin occurs. Load-carrying capacity is defined as the highest gage load sustained for a 1-min period. The test equipment is identical to that employed for measurement of the wear properties of fluid lubricants. The ASTM D2981 test evaluates the wear life of a bonded solid film lubricant under oscillating motion by means of a block-on-ring friction and wear test machine. The test machine comprises a coated steel test ring oscillating against a steel test block. The oscillating speed is 87.5 cpm at 90◦ arc. The test is run in under a load of 283 kg and then operated at a prescribed load until failure. 30.2.3.5.1 Comments The ability of the lubricating film to adhere to the applied surface and provide effective separation between moving parts is a perquisite. These methods offer an effective means of evaluating lubricity and life performance characteristics of various solid lubricants and are useful for comparing the consistency of production lots. Since the test involves the coating of solid film lubricant on test pieces, the results obtained are highly dependent upon the application technique.
30.3 Oil Condition Tests As all machinery and fluids will begin to deteriorate from the moment of use, fluids should be monitored on a regular basis to determined suitability for continued use. In addition to laboratory analysis, highvalue or critical machinery may utilize real-time sensing systems to provide a continuous data stream. The data from both laboratory and real-time sensing are interpreted in conjunction with known machine failure modes, machine usage, configuration and maintenance data to render an expert opinion on the condition of the machine/oil system [15]. Oil condition testing seeks to determine the level and nature of contaminants. The analysis will comprise methods to detect and quantify the concentration of metals, water, dirt, fuel, insolubles (soot), glycol (antifreeze), degradation by-products, and any other process contaminant that may be key to fluid condition. The primary test methods utilized for oil condition monitoring are presented.
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30.3.1 Atomic Emission Spectroscopy — ASTM D5185, ASTM D6595 Elemental spectroscopy determines the elemental constituents of lubricant fluids by raising their atoms to an excited energy state in a high-temperature source. The ASTM D5185 inductively coupled plasma (ICP) spectrometry utilizes a radio frequency induction heater to create a hot argon gas plasma torch. The oil sample is atomized and carried by argon gas into the torch. The very high temperature excites metal atoms, which radiate their characteristic emission lines. The emission lines are captured and measured by a digital array sensor or a series of photomultiplier tubes. The ICP can provide parts-per-billion (PPB) sensitivity for compounds and wear metal particulates less than 3 µm in size. The ASTM D6595 rotary disk emission (RDE) spectrometry is so called because the sample fluid is transported into a high-temperature arc by means of a rotary carbon disc. The disc is immersed into the sample vessel and picks up oil and wear metals as it turns. The arc raises the energy states of the metal atoms in the oil causing them to emit their characteristic emission lines. The lines are measured by the optical system. RDE spectrometers measure particles up to about 10 µm in size at parts-per-million (PPM) sensitivity.
30.3.1.1 Comments Contaminants and additives that contain metals can be quantified by AES methods. Note elemental spectrometers are not created equal. Because different instruments utilize different methods of excitation, different excitation energies and different detection systems, readings do not correlate. If good oil analysis performance is to be achieved, the same instrument type should be utilized to analyze all samples.
30.3.2 Infrared (IR) Spectroscopy — ASTM (E2412) Used-oil condition monitoring requires the evaluation of a number of the fluid condition and any contaminants including additives, oxidation by-products, and ingress contaminant materials. Molecular analysis by Fourier transform infrared spectroscopy provides a reproducible means of determining a wide range of fluid degradation (oxidation, nitration, and sulfation by-products) and contamination (water, fuel, antifreeze, and soot) faults. The FT-IR spectrometer analyzes the chemical structures in an oil sample by evaluating the spectral response of a mid-infrared beam from ∼4000 through ∼650 cm−1 wavelength range. The beam passes through the oil sample and is altered by the characteristic absorbencies of the various oil and contaminant molecules. A detector and electronic circuit pick up and convert the beam into an audio frequency, which is converted to spectral data by a fast Fourier transform (FFT) software program.
30.3.2.1 Comments The multiple oil degradation and contamination symptoms generated by in-service oil applications must be individually quantified to determine the cause and thus the remedy. FT-IR provides an excellent means of determining oil contamination and degradation progress as it quantifies the chemical structures (failure mode symptoms) associated with these materials. Note certain oil additives interfere with IR water detection methods such as demulsifier additives found in hydraulic and steam turbine oils. For these oil types, water concentration may be determined by Karl Fischer titration (KFT). In addition to benchtop FT-IR analyzers, there are also a number of real-time, IR sensors available. In most cases, IR sensors are miniaturized, ruggedized solid-state versions of traditional IR spectrometers that can be mounted on operating machinery fluid systems. In this regard, they can withstand machinery operating temperatures and vibrations, are autonomous, and eliminate the need for sampling and remote lab analysis.
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30.3.3 Particle Counting — ISO 4406, ISO 11171 Fluid cleanliness is also a key factor for machinery system reliability, especially for hydraulic and precision bearing systems. Insoluble particulates due to dirt ingress and oil degradation will increase with equipment use. These materials will cause silting, gumming and abrasion of machinery parts. In practice, there is a direct relationship between oil cleanliness and component usable life — the cleaner the system, the longer the life of the oil and oil wetted parts. With supplementary filtering and clean oil makeup, the particle count can be expected to maintain a nominal value or “dynamic equilibrium.” This “baseline” represents the normal condition of the system. Any increase in particulate counts above the baseline indicates an increase in contamination, regardless of whether the particles are wear metals, ingress dirt or oil degradation by-products. Electronic particle counting is the preferred method of monitoring system cleanliness. Since the number of particles per milliliter increases dramatically as particle size diminishes, both the size and count data must be interpreted to determine to potential effect on a given machinery/fluid system. The interpretation is further complicated by the fact that even small changes in the count bin sizes have dramatic effect on counts recorded. To overcome this problem, the ISO community established the ISO 4406 and ISO 11171 standards for fluid cleanliness monitoring. The 4406 standard characterizes fluid cleanliness in increments from 0.01 to 2,500,000 particles per milliliter of sample. The 4406 standard specifies three bin ranges and a series of numerical codes to indicate the count in each bin, where: • The first code indicates the particle count above 4 µm • The second code indicates the particle count above 6 µm • The third code indicates the particle count above 14 µm The system is open-ended and can expand in either direction, above or below the current codes. In practice, particle count data from a counter is compared to the ISO 4406 table (Table 30.3) for determination of the ISO cleanliness rating for each size range. For example, a sample containing 80–160 particles per milliliter greater than 4 µm; 20–40 particles per milliliter greater than 6 µm; and, 5–10 particles per milliliter greater than 14 µm, would generate a cleanliness code of 14/12/10. The ISO 11171 standard defines the requirements for particle counter measurement calibration and data reproducibility. The standard cover flow rates, coincidence error, resolution, and sensor and volume accuracy. It also ensures that all instruments are calibrated to the same National Institute of Standards and Technology (NIST) traceable standards. 30.3.3.1 Comments For simple hydraulic systems, the particle count will directly correlate with system reliability. However, for machinery systems containing bearings and gears, particle counting does not indicate the nature of the debris and other analysis techniques should be used when high particle counts are observed. There are a number of particle counters available in bench top and portable versions. In addition, there are a number of on-line particle-count sensors available for real-time monitoring of equipment fluid systems. The sensors are extensions of the benchtop light-extinction particle counter technology, packaged in a rugged casing to withstand the rigors and high pressures of on-line machinery fluid systems. The sensors evaluate count data with preset limits and generate automatic alarms for immediate operator attention such as green-light/red-light. It should be noted that while most instruments have good measurement repeatability, different manufacturers’ instruments and sensors will not give the same results for the same sample if the instrument does not meet ISO 11171. In addition, entrained water and air bubbles affect all particle counters. Some models handle air bubbles. However, water bubbles are counted as particles by all models. These problems can result in unreliable counts especially for the smaller particle sizes.
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30.3.4 X-Ray Fluorescence Spectroscopy XRF Spectroscopy is similar to AES except that x-ray energy rather than heat is used to stimulate the metal atoms in the sample. The x-ray source raises the energy level of the atoms in the sample, resulting in a corresponding release of x-ray energy from the excited atoms. The ASTM D4927, ASTM D6443, and ASTM D6481 are specifically for determining additive metals in unused oils. 30.3.4.1 Comments XRF can monitor solids, fluids, or dissolved metals in oil (whether naturally occurring, introduced as additives, or resulting from normal wear processes in machinery) and can measure particles of all sizes. In addition to benchtop XRF spectrometers, there are a number of XRF sensors for off- and on-line elemental analysis. These sensors utilize both flow-through and filter patch technologies.
30.3.5 Water Determination by Karl Fischer Titration — ASTM D6304 The ASTM D6304 test determines water level by titrating a measured amount of the sample and the Karl Fischer reagent. The reagent reacts with the OH molecules present in water and depolarizes an electrode. The corresponding current change is used to determine the titration endpoint and calculate the concentration value for water present. 30.3.5.1 Comments Note many other compounds including some oil additives contain OH molecules. These will be counted as water, skewing the results. Note the most common interfering materials are mercaptans and sulfides.
TABLE 30.3
Determination of the ISO Cleanliness Rating
ISO code
Min. count
Max. count
ISO code
1
0.01
0.02
15
160
320
2
0.02
0.04
16
320
640
3
0.04
0.08
17
640
1,300
4
0.08
0.16
18
1,300
2,500
5
0.16
0.32
19
2,500
5,000
6
0.32
0.64
20
5,000
10,000
7
0.64
1.3
21
10,000
20,000
8
1.3
2.5
22
20,000
40,000
9
2.5
5
23
40,000
80,000
10
5
10
24
80,000
160,000
11
10
20
25
160,000
320,000
12
20
40
26
320,000
640,000
13
40
80
27
640,000
1,300,000
14
80
160
28
1,300,000
2,500,000
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Min. count
Max. count
Lubricant Properties and Test Methods
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Also, high-additive package lubricants such as motor oils will generate erroneous results due to additive interference. Consequently, the Karl Fischer test should not be used to evaluate water in crankcase lubricants.
References [1] [2] [3] [4] [5] [6]
[7] [8] [9] [10] [11] [12] [13] [14] [15]
National Lubricating Grease Institute (NLGI), www.nlgi.org. American Petroleum Institute (API), www.api.org. Society of Automotive Engineers (SAE), www.sae.org. Coordinating Research Council, Inc. (CRC), www.crcao.com. Caterpillar Machine Fluids Recommendations, Document SEBU6250-11, Caterpillar Inc. October 1999. General Motors Maintenance Lubricant Standard LS2 for Industrial Equipment and Machine Tools, Edited by General Motors Corporation, LS2 committee on maintenance lubricant standards, revised 1997. Special Manual Lubricants Purchase Specifications Approved Products, Document 10-SP-95046 Cincinnati Milacron (Cincinnati Machine) Company, July 1995. Lubricating Oil, Fuel and Filters Engine Requirements, Detroit Diesel Corporation, 1999, www.detroitdiesel.com. International ASTM “Standards on Petroleum Products and Lubricants” Volumes 05.01, 0.02, 0.03, and 0.04. www.astm.org. Rand, Salvatore J., “Significance of Test for Petroleum Products,” 7th ed., International ASTM, 2003. Institute of Petroleum (IP). www.energyinst.org.uk. Deutsches Institut fur Normung e.V. (DIN), www.normung.din.de. American National Standards Institute (ANSI), www.ansi.org. International Standards Organization (ISO), www.iso.org. Toms, Larry A., Machinery Oil Analysis — Methods, Automation & Benefits, 2nd ed., Coastal Skills Training Inc., Virginia Beach, VA, 1998.
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31 Contamination Control and Failure Analysis Definitions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 31.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 31.2 Contamination Control Program . . . . . . . . . . . . . . . . . . . .
31-1 31-2 31-5
Failure and Criticality Analysis • Functions of an Engineering System • Types of Mechanical Failures • Phases of Failures • Causes of Failures • Tribological Analysis • Contamination • Mechanical Modes of Failure • Hydraulic and Lubrication Failures
31.3 Contamination Balance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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Oklahoma State University (OSU) Model • Dynamic Contamination Control (DCC)
31.4 Monitoring Procedures . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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Sampling • Detection • Diagnosis • Prognosis • Action • Postmortem
Jacek Stecki Department of Mechanical Engineering Monash University
31.5 Training . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 31.6 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Appendix . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Standards and Drafts of ISO/TC 131/SC 6 Contamination Control and Hydraulic Fluids . . . . . . . . . . . . . . . . . . . . . . . . . Further Reading . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
31-53 31-54 31-54 31-56 31-56 31-57
Definitions For the purposes of this chapter we use the following definitions. Cause. Cause is the means by which a particular element of the design or process results in a Failure Mode. Contaminant. Any foreign or unwanted solid, liquid, or gaseous substance present in the lubricant and working fluid that may have a detrimental effect on the condition of a machine. Criticality. The criticality rating is the mathematical product of the Severity (S) and Occurrence (O) ratings. Criticality = (S) × (O). This number is used to place priority on items that require additional quality planning.
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Handbook of Lubrication and Tribology
Current Controls. Current Controls (design and process) are the mechanisms that prevent the cause of the failure mode from occurring, or which detect the failure before it reaches the next operation, subsequent operations, or the end user. Detection. Detection is an assessment of the likelihood that the sensor will detect the cause of the failure mode or the failure mode itself, thus preventing it from reaching the next operation, subsequent operations, or the end user. Effect. Effect is an adverse consequence that the next operation, subsequent operations, or the end user. Failure Mode. Failure modes are sometimes described as categories of failure. A potential failure mode describes the way in which a product or process could fail to perform its desired function (design intent or performance requirements). Function. Function could be any intended purpose of a product or process. FMEA functions are best described in verb–noun format within engineering specifications. Lubricant. Any substance interposed between two surfaces in relative motion for the purpose of reducing friction and/or reducing mechanical wear. Lubricant/Working fluid based condition monitoring and diagnostics. Monitoring of the characteristics of the lubricant (liquid or solid) and/or working fluid to determine its current state (diagnosis) and to predict its future state (prognosis). Nonsolid contamination. Small portions of liquid or gaseous matter (e.g. air, water, chemical substances) present in a liquid lubricant and working fluid, which are the result of ingression into lubrication and hydraulic/pneumatic system. Occurrence. Occurrence is an assessment of the likelihood that a particular Cause will happen and result in the Failure Mode during the intended life and use of the product. Particulate based condition monitoring and diagnostics. The monitoring of the characteristics of solid particulate present in the lubricants and working fluid of a system to determine the current state of the system (diagnosis) and to predict its future state (prognosis). Risk Priority Number (RPN). Risk Priority Number is a mathematical product of the numerical Severity (S), Occurrence (O), and Detection (D) ratings. RPN= (S) × (O) × (D). This number is used to place priority on items that require additional quality planning. Severity. Severity is an assessment of how serious the effect of the potential failure mode is on the next operation, subsequent operations, or the end user. Solid contamination (particulate). Small portions of solid metallic and nonmetallic matter present in lubricants and working fluids, which are the result of mechanical wear of surfaces in a machine, electrochemical processes (e.g. electrical discharge), or ingression from the environment. Symptom (contaminant analysis). A measured or otherwise identified physical, chemical, or morphological characteristic of particulate present in a sample of lubricant/working fluid that provides pertinent information about a particular characteristic of the state of the machine. Syndrome. A set of symptoms that identifies a certain fault of the machine. Tribology. The science and technology of interacting surfaces in relative motion, and of related subjects and practices. Tribology based condition monitoring and diagnostics. The detection of changes to the tribological characteristics of a system to determine the current state of the system (diagnosis) and to predict its future state (prognosis). Working fluid. Fluid used as an energy transmission medium (produced from petroleum products or aqueous or organic materials) or for other purposes (e.g. coolant).
31.1 Introduction In systems employing fluids (produced from petroleum products or aqueous or organic materials) as lubricants or working media, wear particles (metallic wear particles, friction polymers, carbon deposits, etc.) are carried away from the wearing surfaces and circulated throughout the system. In addition, fluids
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can contain contaminants (undesirable solid, liquid, or gaseous material) that are the result of ingression from the system’s environment. Control of contamination is specially important in hydraulic power and control systems which, due to their advantages over other systems in providing flexible and accurate control of motion and forces, are widely used in automotive, aerospace, machine tools, earthworking machinery, and other areas of engineering. The presence of contamination, resulting from contaminant ingression or from wear in system components, if not controlled, will affect hydraulic system reliability and accuracy of its control actions. Experience shows that the majority of hydraulic systems fail due to damage or operational malfunction caused by the presence of contamination in the system. The level of fluid contamination in the system is the result of a multitude of factors that interact in a complex way. Some of the more important factors that affect the contamination level are: the rate of contaminant entry to the system from its environment (e.g. dust), rate of wear debris production by system components, filter characteristics (e.g. β-rating [1], dirt holding capacity), type of filtration system (e.g. high-pressure, low-pressure, by-pass), system duty cycle, design of reservoir, fluid loss, and contamination tolerance of system components. The function of a filtration system is usually defined as removal of contamination from the fluid no filtration system will remove all contamination from the hydraulic or lubrication system. The best we may expect is that the level of contamination is contained to the quantity below the components’ tolerance level. Thus, this function must be redefined as maintaining the contamination level of the fluid below the tolerance level of the system components. Although the above is recognized by most of the users of fluid power systems, in practice, many hydraulic systems have inadequate protection against contamination. The filtration system is usually the first casualty when the price of a hydraulic system must be competitive. High-quality filters are expensive, thus to lower the price of the system, the quality of contamination control is lowered by applying smaller size, lower filtration rating filters, or by offering a better contamination control system at an extra cost (this is usually rejected by the customer as in most cases the user has no understanding of filtration requirements). The effects of “cutting corners” when selecting a filtration system will only become apparent later on when system damage or malfunction occurs, forcing modifications to the system, which are usually expensive. A very common reason for inadequate contamination control is lack of understanding of contamination control technology. Usually the filters are selected by following the equipment manufacturer’s recommendations, which specify the required fluid cleanliness level (e.g. in terms of ISO 4406 [2] or NAS 1638 [3] standards), absolute maximum mesh size and required filter β-rating necessary to maintain correct operation of the various hydraulic components. Filtration system selected on the above basis, without consideration of expected rates of ingressed and system generated contamination with which the equipment will have to cope and no knowledge of components tolerance to contamination, may be totally inadequate in the case of a complicated multi-branch system, consisting of many hydraulic control elements and operating in a dirty environment. The concept of contamination balance, first proposed by Fitch [4], which considers the generation and removal of contamination within the system, provided a mathematical tool to describe interactions between various factors affecting level of contamination in a system and thus led to a better understanding of the contamination characteristics of hydraulic and lubricating systems. An important step forward was the recognition of the significant economic impact of contamination on the life cycle costs of the plant and that contamination control, both in hydraulic and lubricating systems, must be considered as an integral part of a quality program of the plant. This in turn provided a driving force behind further advances in contamination control theory [5], the development of on-line contamination sensors [6, 7], widespread application of on-line monitoring and root-cause and on-condition maintenance, availability of microprocessor based signal and knowledge processing (expert system, fuzzy logic, neural nets), and a marked progress in further development of the filtration technology [8–12]. The design of a contamination control system consists of three major tasks, Figure 31.1: • Setting of appropriate targets for cleanliness of the system in order to achieve desired life cycle of the system
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Contamination control program
Set target
Implement
How clean system has to be to meet desired life cycle
Eliminate and reduce ingression sources
System design Filter system
Detect failures
Predictive
Prevent failures
Proactive
Monitor
FIGURE 31.1
Contamination control program.
• Implementing contamination control • Applying condition monitoring strategies (predictive and/or preemptive) to prevent systems failures Contamination control is thus concerned with contamination and its effects on a system in design, manufacturing, commissioning, and operation/maintenance phases of system life: Design: formulating objectives and determining goals, setting specifications for fluid cleanliness in a system, identifying of possible root causes of contamination problems. Manufacturing: using materials that will minimize effects of contamination on wear of components (wear resistance), applying manufacturing techniques that will result in clean components and will inhibit production (wear) or ingression of particles (seals). Commissioning: using approved test procedures, following correct system flushing procedures, using correct fluids, seals etc. Operation and maintenance: following correct operating procedures, implementing monitoring procedures and corrective actions. Implementation of a contamination control program will fail unless management is fully supportive. Management must produce an environment that will support rethinking of current maintenance practices in order to introduce a “cleanliness ethic” in the plant. To be successful the program must aim to move decision making down to the workplace by giving operators/maintenance staff authority to act in a knowledgable manner. Thus, the personnel must be trained to have knowledge and understanding of the program, techniques, and tools. The program goals must be stated in terms of what is required and not how it should be done. The program must include proper reporting and assessment procedures. Some factors (constraints) that must be considered are: • Money: investment in equipment, cost of maintenance staff, insurance, cost of spares, equipment loss, production loss • Machines: type of equipment (mobile, stationary), complexity, failure rates, effectiveness of contamination control system • Methods: selection of techniques, applicable monitoring techniques, detection and prognostic methods, sampling procedures, data acquisition — permanent units, data collectors, data processing — knowledge based, computer based, manual • Materials: type of contaminants, types of wear, failure modes • Minutes: expertise acquisition time — learning curve, sampling turn-around time, time to achieve guidelines • Manpower: current experience of staff, technical levels and skills of staff, training required
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31.2 Contamination Control Program Setting up, implementation, and monitoring of a contamination control program requires thorough knowledge of the plant that will be included in the program. The reliability of diagnosis and prognosis of the monitoring program will be directly affected by the selection of detection methods, thus the selection of a proper mix of monitoring methods that will be able to identify changes in machine condition is important to the success of the program. Specifically, setting up of contamination control program comprises the following tasks: • Identification, by using functional and failure analysis techniques, of failure modes and effects • Analysis of tribological systems present in selected machines to determine root causes of wear and contamination conditions in each machine • Identification of root cause and tribological symptoms of machine state that should be monitored and selection of sampling points • Selection of techniques, guidelines, and procedures to carry out root cause and tribology based condition monitoring of a machine A flowchart of setting-up tasks is shown in Figure 31.2, only when the results of these tasks are known the decision can be made how the contamination control program should be executed. Experience shows that if setting up of contamination control does not follow the above general procedure the results of the program will not gain predicted benefits.
31.2.1 Failure and Criticality Analysis Failure analysis techniques are usually employed to detect potential safety problems in newly built, or existing systems and application of these techniques is now being extended to analysis of new, still in the design stage, systems. In engineering practice there is a large, approximately 60, number of available techniques for failure analysis, some of them are, (Figure 31.3): • Failure mode analysis (FMA) — identification of unwanted conditions of the system • Failure modes and effects analysis (FMEA) — identification of effects of component failure on the system operation and safety • Hazard analysis (HA) — identification of potential hazards during system operation • Failure modes, effects, and criticality analysis (FMECA) — identification of effects of component failure on the system operation and safety, probabilities of occurrence and their criticality The FMEA analysis is concerned with hardware failures and malfunctions, it does not include hazards due to errors caused by human operators, effects of environment, and other operating and hazardous conditions outside the scope of design limits of the components. A functional failure is defined as a component’s or system’s inability to perform the intended function (e.g. piston speed low). A structural failure is defined as the component’s failure due to changes in material (e.g. fluid viscosity), geometry of component parts (increased orifice size), or tribological action (e.g. increased friction), Figure 31.4. The basic objective of FMEA is to identify all possible failure modes, that is, how system/parts can fail (e.g. hydraulic line leakage, valve seized) and deduce the consequences of these faults, that is, what happens if failure occurs (e.g. piston speed too low, no flow through the valve). Other information sought during analysis is the nature of the failures and whether there are any redundancies, backup components, or sensors that can safeguard components against failures. The FMEA technique supplemented with analysis of criticality of failures becomes FMECA. The basic objective of this analysis is to identify possible failure modes of the components, deduct what are the consequences of these failures and determine criticality of each failure mode. A methodology of FMEA is shown in Figure 31.5, guidelines for preparing FMEA/FMECA are the subject of a number of national and international standards [13–16]. Although FMEA/FMECA appears to be a rather straightforward task, its execution requires a clear understanding of system operation, availability of details of system components and their interaction,
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Handbook of Lubrication and Tribology
System
Identify failure mechanisms Identify cause & effects Identify symptoms of failures
Identify system boundary Functional analysis Identify functions Failure analysis
Cost of spares, maintenance
Identify critical equipment
Identify tribological pairs
Wear analysis
Identify wear parameters
Contamination analysis
Identify wear modes Set target and tolerance levels Select set of symptoms
Select monitoring methods
FIGURE 31.2
Identify system's components
Select monitoring techniques
Prepare function diagram
Cost of production loss
Identify sources and types
Ingression levels
Contamination balance
Develop guidelines Set monitoring procedure
Setting up of contamination control system.
a knowledge of the functions that the system and its various parts must perform, and a knowledge of interfaces with external systems. As a system may operate in various modes (e.g. emergency mode, start-up mode, test mode) the analysis should be carried out for all modes of system operation. Failure analysis of complex modern mechanical engineering systems is very time consuming and thus attempts were, and are, made to automate it in application to fluid power systems [17–19]. The current approaches to automated, computer-aided FMEA of hydraulic power and control systems are based on investigation of suites of models (functional, behavioral, structural, and teleological) using qualitative reasoning, fuzzy logic, and other approaches [20–24].
31.2.2 Functions of an Engineering System The Oxford dictionary definition of a system is “a set or assemblage of things connected, associated, or interdependent so as to form a complex unity.” Implied in this definition is the purpose of such a set which, in engineering terms, is the transmission of generalized information that can be in the form of energy,
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Contamination Control and Failure Analysis
31-7
Failure mode and effects analysis Failure analysis
Failure mode,effect, and criticality analysis Fault tree analysis
Fuzzy logic
Methods
Artificial intelligence
Neural nets
Expert systems
Failure patterns Modeling and simulation Cause and effect
FIGURE 31.3
Failure analysis methods.
mass, or control information. The purpose of a mechanical system is to transform, transmit, and generate mechanical forces and motions, that is, mechanical systems are concerned with energy and work. The action of the system in meeting the system’s objective is described in terms of functions that the model (system) must perform to process input information into output information, Figure 31.6. Whenever an engineering problem is investigated, a designer can either investigate the system by direct experimentation on the real system or carry out the investigation on the basis of some type of a model. The system can be conceptually separated from its environment by defining a boundary that encloses the system and a number of points that are common with other systems or the environment. These points are information transfer terminals, and the information passing through these points will define the character of interaction with other systems and the environment. After defining the boundary of the system and points of information transfer we are able to investigate the system in isolation from other systems and its environment. Although the selection of a system’s boundary is arbitrary, the choice of boundary is influenced by the objective of the investigation. The real engineering system has a great number of these attributes and the investigation of the system would be difficult if all were considered, thus a skilful designer will select only these attributes that are relevant to the subject of investigation. To simplify the investigative task some attributes are totally ignored while others are idealized, on the basis of a certain set of criteria and assumptions accepted by the designer. The subset of selected variable attributes of the system that represent information transfer between the environment and the system can be separated into two sets. Those variable attributes that represent information transfer from the environment to the system and that can be controlled or manipulated are designated as inputs to the system. Variable attributes that represent information transfer from the system to the environment and can be measured or observed are designated as outputs. Other attributes that represent the physical or geometric attributes of a system, for example, kinematic viscosity of fluid, or that cannot be or were chosen not to be controlled and manipulated are designated as parameters. Parameters do not have to have constant value, for example, viscosity of fluid will vary with system temperature. The process of delineation of a system boundary and identification of its attributes is called modeling. The objective of modeling is to construct a system, a model, which is a subset of the real, physical system.
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31-8
Handbook of Lubrication and Tribology
Function
Function
Function
Function
Function
Function
Functional diagram
High range effect 1 downstream, (e.g., damaged support)
Noise (e.g., friction) Component function definition Control parameters (e.g., pressure)
Measured variable (e.g., force) Apply force
Upper limit Acceptable operating range
Failure modes and effects
Lower limit Effect 2 downstream (e.g., failed to lift) Low range
Physical component (e.g., actuator)
Dynamic model
Energy
Energy
Failure modes and effects
Energy processing
Component representation (e.g., drawing)
FIGURE 31.4
Load velocity etc.
Tribological model
Wear friction
Failure modes and effects
Tribology model
Functional and tribological models.
Investigation of the model will yield information about functional, behavioral, and structural properties of the real system under consideration. The correct choice of modeling assumptions is the one factor that will usually seriously affect the quality of the system investigation and the magnitude of modeling and simulation tasks. The model of a real system, developed on the basis of such a set of assumptions, should in all important aspects be equivalent to the original real system. Structural description of a system S defines inner working of the system by showing the interaction between elements (parts) of the system and relevant properties of these elements: S = {A, P, R}
(31.1)
where A is elements of the system, P is relevant properties of the system, R is interrelation between elements. Such a model describes, for example, a tribological system. Functional description of a system defines system behavior in terms of input–output relations: O = F (I )
(31.2)
where F is function relating inputs I to outputs O and I is inputs. Functional model is used to describe, for example, a mechanical systems where function F represents system transfer function. In most general case a system’s model is represented as a blackbox with technological implementation and structure (topography) of the system unknown. The blackbox power model of a system accepts
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Contamination Control and Failure Analysis
31-9 Components
Step 1
Identify system components
Step 2
Identify component functions
Step 3
Identify failure mechanisms
Step 4
Identify failure modes
Step 5
Identify cause and effect
Assembly
Failure Causes Failure Caused by
FIGURE 31.5
Step 6
List corrective measures
Step 7
Assess reliability
Failure
Failure analysis procedure.
How?
Why? Failure modes
Control parameters
Critical parameters
Function
Noise parameters
Function response
Nominal response Failure modes
Controled by designer Outside control of designer
FIGURE 31.6
Function definition.
the inputs, processes them, and transfers the output information to the other systems and environment, Figure 31.7. Each technical system consists of a number of subsystems and parts, each performing some function [25]. Thus we usually may identify a hierarchical structure of the system and identify functions at each hierarchical level, Figure 31.8. Identification of functions and their hierarchy, Figure 31.9, is a rather difficult task but the effort may be greatly reduced by using value analysis techniques like FAST (Function Analysis System Technique) and standards [26,27].
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31-10
Handbook of Lubrication and Tribology Output ys
Input us Real system pS
Input uM
Output yM Model pM
pM
uM
yM
FIGURE 31.7
Modeling process.
Energy Signal
Energy Input / output
Signal
Matter
Matter
Primary function
Function
Function
e.g., increase torque
FIGURE 31.8
Secondary function
Function
Function
Function
Function
Function
Function
Functions of an engineering system.
In general, most mechanical systems fall into two types: stationary, force transmitting systems (e.g. machine base) and moving parts, power transmitting systems (e.g. hydraulic power unit). Usually a mechanical system is composed of both these types of systems. When carrying out functional analysis on stationary systems we are concerned with flow paths of forces through the structure whereas in the case of moving part systems we are concerned with energy flow paths through the system. Identification of failures is greatly facilitated by the application of Fault Tree Analysis (FTA). The method is based on finding out what are the possible unwanted conditions or failures of the system and finding out what are the operating conditions, component faults and failures that caused these unwanted conditions. The application of FTA can be a backward (we know the condition, or top event, and trying to find what
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Contamination Control and Failure Analysis System
31-11 Functions provide locking force
connector How?
transfer force
dogs
Why? Piston/cam ring
exert force
FIGURE 31.9 System-function.
Failure
Failure
Forward chain
Backward chain
and
or
Failure
or
Failure
Failure
and
Failure
Failure
Failure
FIGURE 31.10 Fault tree analysis.
causes it) or forward chaining (we know the modes of components failures and trying to find out how these failures affect hierarchically higher components and the system), Figure 31.10. The resultant failures can be either caused by a serial combination of preceeding failures in different paths — represented on the diagram as an and gate, or by a parallel combination of any preceeding failures — represented by an or gate. FTA technique is well formalized and computer programs are available to automate development and analysis of fault tree diagrams. In the case of a fault tree that represents a system with a built-in redundant path, care must be taken to identify failure modes common across the redundant branches. The occurrence of such faults will reduce the inherent reliability of a redundant system. A typical fault that will have such an effect would be, for example, the failure of a contamination control system that will affect all the components of a hydraulic drive system. Improper maintenance procedure, failure of an indicator or sensor could have a similar effect. In addition, we should also consider the effect of some events, which although they are not directly associated with failure, may cause a higher level failure. An incorrectly set pressure relief valve may cause the failure of a pump and therefore a system. Disconnecting a pressure switch may result in the failure of a pump by not switching the system off when the pressure is at the limit.
31.2.3 Types of Mechanical Failures Failure of a machine or its components is defined as an inability to perform a required function or functions within the constraints imposed by system specification (i.e. why it is very important to have good equipment specification). We may identify the following types of failures during the life of equipment.
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Handbook of Lubrication and Tribology
31.2.3.1 Incipient Failure The machine is in incipient failure mode when its performance or condition is gradually degraded and it will eventually fail if no corrective action is taken. This type of failure can often be tolerated, however, its progression must be monitored to prevent occurrence of a catastrophic failure. The decision “to live” with this type of failure will depend on the mechanism of wear. For example, fatigue wear in gears usually progresses in a gradual fashion, thus we may tolerate this type of wear and consider the gears to be in an incipient failure mode. The gears should be monitored to see if there is a change in the rate of wear and if so necessary corrective action should be taken, for example, reduction of the load carried by the gears. On the other hand, detection of scuffing (sliding wear) indicates lubrication problems and as this type of wear progresses very rapidly it may lead very quickly to destruction of the gears and immediate corrective action should be taken. 31.2.3.2 Misuse Failure The equipment is used beyond the specified limits or is used to perform a function it was not designed to perform. Faults due to wrong maintenance action or operation also fall into this category. For example, replacement of hydraulic fluid with fluid of wrong (e.g. too low) viscosity may lead to excessive temperature in the system and destruction of the pump. An example of wrong operation is closing the inlet valve to a hydraulic pump causing pump failure due to cavitation. 31.2.3.3 Catastrophic Failure This is a sudden and total failure of the system that is not expected to occur. A typical example of such a failure is jamming of the valve due to contamination.
31.2.4 Phases of Failures We may identify the following phases of failures: Wear-in failures: After commissioning of equipment there is a period of time during which various tribological pairs are subjected to large tribological stresses. This produces often excessive wear, which in most cases is not posing a danger to the future operation of the machine. For example, gears during the wear-in (running-in) stage of operation may produce an abnormal amount of fatigue wear particles (pitting). This is caused by abnormal contact stresses between gears’ teeth caused by asperities and teeth surface irregularities, which are the results of machining process. However, if the gears were correctly designed, the production of wear particles will be significantly reduced after the “high spots” are removed from the gear teeth surfaces. This type of fatigue wear occurring during wear-in period is called “arrested pitting” (under some circumstances the fatigue wear process is not arrested and the gears enter into a “progressive pitting” mode, which inevitably leads to catastrophic failure of the gear set). The wear-in occurs very rapidly, tests at Monash University have shown that wear-in in gears will occur within a few thousands of gear revolutions. More dangerous failures during the infancy period are due to design or manufacturing errors or defects. Very often mistakes are made during assembly, installation, or commissioning stages. Random failures: Random failures occur in the mid-life of the equipment and are caused by overloads or under-designed components, deficient lubrication, contamination, corrosion, etc. These types of failure are difficult to predict. Wear-out: Wear-out (degradation) failures are caused by fatigue (both high and low cycle), corrosion, erosion, or by design (planned obsolescence). Failure of a hydraulic system manifested by pump leakage caused by pump wear-out due to a prolonged exposure to contamination is an example of this phase of wear. The failure rate for each phase of failure may be represented using “bath tub” curve, see Figure 31.11. In the figure indices A, B, and C refer respectively to wear-in, random, and wear-out failures.
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Failure rate
Contamination Control and Failure Analysis
31-13
C
A B
Time in service
FIGURE 31.11 Bath tub curve. Stress concentrations
Surface finish
Residual stresses Surface treatment
Overheating
Manufacturing deficiencies
Overloading
Service conditions
Contamination
Material hardness Material deficiencies
Material selection Combination of materials
Lubrication
Causes of failures
Improper procedures
Contamination
Tolerances
Maintenance procedures
Design deficiencies
Overstressing
Strength, rigidity
Assembly procedures
Wrong replacements
Damage to seals
Damage during transport
Damage to surfaces
Clearances, fits
Incorrect handling
Clearances, fits
FIGURE 31.12 Causes of machine failures.
31.2.5 Causes of Failures Mechanical failures are most commonly caused by (see Figure 31.12): Machine (design and manufacturing): lack of knowledge of loads, wrong manufacturing procedures, design deficiencies — tolerances, clearances, fits, strengths, rigidity, manufacturing deficiencies — residual stresses, surface finishes, stress concentrations. Assembly procedures — overstressing, clearances, fits, damage to seals, damage to surfaces, manufacturing defects.
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Handbook of Lubrication and Tribology
Material: inadequate material properties, wrong material selection criteria, mismatch of material properties with environment (e.g. corrosion), wrong combination of materials, surface treatment, hardness. Man: inadequate training, inexperience with a new type of system, reliance on computer analysis, incorrect operation of the system, erroneous or omitted design calculations (stress analysis). Methods: incorrect specification (wrongly specified constraints), lack of quality control, inadequate manufacturing methods, inadequate design methods, inadequate maintenance methods, incorrect operating procedures (service conditions — overloading, overheating, contamination, lubrication), maintenance procedures — contamination, wrong replacements, improper procedures. Incorrect handling, damage during transport (fretting). Money: emphasis on cost at the expense of quality of design and manufacturing. Minutes: wrong timing of maintenance activities, cutting corners to meet time schedules, deterioration of material properties over time. The above list is not exhaustive and other factors may also contribute to failures. In addition to the above factors failure can be caused by effects, that is, qualitative factors, like environmental conditions, radiation, noise. Investigation of failed components shows that most failures can be attributed to: • Excessive tribological stresses resulting in wear and corrosion. This type of failures are usually gradual and result in changes of geometry of the affected components • Excessive elastic deformation • Excessive plastic deformation • Fracture of component • Misuse or abuse • Assembly errors • Manufacturing defects • Improper maintenance • Fastener failure • Design errors • Improper material • Improper heat treatments • Unforeseen operating conditions • Inadequate quality assurance • Inadequate environmental protection/control To illustrate some causes of failures we may consider rolling element bearings. Bearings are the most common elements in mechanical equipment and should be very long-lasting components. The bearings should have a very low wear rate, and they rarely fail due to material faults. Why then do bearings fail earlier than their design life? Experience shows that most bearings fail by accident (96% of failures according to some literature sources) and approximately 10% of these failures are due to fatigue. A fair number of bearings fail due to lubrication problems, seal problems, and improper mounting. Some of the bearings are already damaged due to mishandling before they are assembled on the machine (e.g. fretting damage during transport). Typical causes of damage to rolling element bearings are shown in Figure 31.13.
31.2.6 Tribological Analysis The study of wear in mechanical systems is part of a scientific discipline called tribology (tribo — to rub in Greek). Wear in mechanical systems is the result of tribological action, and is defined as the progressive loss of substance from the surface of a solid body due to contact and relative motion with a solid, liquid, or gaseous body. Tribological analysis of a machine leads to identification of types of tribological actions in the machine and therefore identification of possible wear modes and their severity. This information combined with identification of internal and external contamination sources is used to develop sampling and detection procedures of monitoring and diagnostic purposes.
© 2006 by Taylor & Francis Group, LLC
Contamination Control and Failure Analysis
31-15
Bearing failures
Mounting
40.0%
Defects
26.0%
Lubrication
FIGURE 31.13
Dirt ingression
11.0%
3.0%
Fretting
4.0%
Corrosion
6.0%
Overload
10.0%
Causes of bearing failures.
TABLE 31.1
Examples of Tribological Systems Elements of tribological systems
System Guide bush Disk brake Gear box Hydraulic pump Hydraulic motor Directional control valve Rolling element bearing
Body
Counterbody
Interface
Environment
Guide Pad Pinion Slipper Piston Spool Roller
Rod Disc Gear Plate Cylinder block Valve block Race
Lubricant — Lubricant Hydraulic fluid Hydraulic fluid Hydraulic fluid Lubricant
Air Air Air — — — Air
Wear results in producing wear particles and in changes in the material and geometry of the tribologically stressed surface layers of the components forming a tribological pair, that is, two components that are in contact with each other. Normally wear is unwanted however, in certain circumstances for example, during running-in, wear may be beneficial. A typical example of a tribological pair is a pair of gears in which tribological action occurs on the meshing surfaces. Elements of a tribological pair may be in direct contact (e.g. not lubricated gears) or contact between these surfaces may be via interfacing medium (e.g. lubricant) that modifies interaction between the elements of a tribological pair. Examples of tribological systems are listed in Table 31.1. The body and counterbody are members of a wear couple where wear is of particular importance. The interface medium may have a wear reducing effect (e.g. lubricant) or wear increasing effect (e.g. dust). The environment may also play a major role in the wear process. The external operating variables that act on the elements of the tribological system form the operating variables of the tribological system. The wear characteristics describe the nature of material loss that occurs through the action of the operating variables. Tribological action requires both contact between the wearing couple and relative velocity between elements of the couple, examples of such actions are shown in Figure 31.14. External actions on the element for example, bending, shear are not considered to be tribological actions. Flow chart of the analysis of the wear process is shown in Figure 31.15. System operating parameters that must be considered in analysis of wear are [28]: • • • •
Normal load over tribological contact area (exception — fluid flow) Relative velocity between wear couple (there is an exception — erosion due to fluid flow) Thermal equilibrium state — operating temperature Duration of wear process — period of application of tribological stress
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31-16
Handbook of Lubrication and Tribology
Surface roughness Humidity Temperature Force Velocity
Velocity
Hardness
Fluid
Type of motion Load Heat transfer characteristics
Rolling
FIGURE 31.14
Sliding
Spinning
Impacting
Flow
Sliding and rolling
Tribological actions.
Identification of wear mechanisms includes investigation of material characteristics, wear processes present, and characteristics of wear/contamination particles, Figure 31.16. The characteristics of debris that are generated during tribological action and the appearance of surface differ for different wear processes. Sources of wear in mechanical systems are shown in Figure 31.17. Changes in the quantity of debris generated in a system with size of particles and severity of wear are shown in Figure 31.18. The regime of wear depends on a number of factors that are shown in Figure 31.19. Manufactured surfaces are far from smooth when viewed under a microscope. Typically the surface consists of a multitude of peaks (called asperities) and valleys randomly distributed over the surface. When two surfaces with similar hardness are brought into contact, the number of asperities touching and the contact areas are very small. However, as the normal load increases the contact stresses in the points of contact (asperities) rapidly increase and the plastic flow value of the material and the contact points deform plastically in such a manner that the total contact area is now finite (i.e. increases) as shown in Figure 31.20. Wear is a complex process involving the removal of material from sliding surfaces. The complexity of the process can be seen in the plethora of defined wear modes. Two modes, abrasive wear and adhesive wear, are probably the best known. Although these two modes are well defined, in practice one seldom finds pure adhesive or abrasive wear but rather both processes are evident on the same worn surface. Other well-recognized forms of wear include fretting, particle erosion, corrosive wear, and fatigue wear. Wear is a natural occurrence and occurs when any two surfaces rub together. The designer cannot eliminate wear but, by understanding the mechanisms of wear, he or she can try to minimize wear by proper selection of materials, lubrication, and design. There are a number of wear mechanisms that describe wear, however the terminology used may be confusing to the beginner, e.g. scoring is called scuffing by some people and abrasion by others. The most
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Contamination Control and Failure Analysis
Soft Hard
Identify machine functions
Type of material
Character Type
Materials Gas Liquid
31-17
Identify tribological pairs
Interface medium
Motion Magnitude Direction
Solid E.g., rolling
Tribological action
Tribological parameters
Evaluate wear characteristics
Wear-out Identify descriptors
Thermal equilibrium state Duration of wear process
Identify symptoms
FIGURE 31.15 Tribological analysis of a system.
Pitting Surface appearance
Scuffing Fretting etc. Shape
Investigation of wear
Particle morphology
Texture Color etc. Fatigue
Wear process
Adhesion Abrasion etc.
FIGURE 31.16 Investigation of wear.
© 2006 by Taylor & Francis Group, LLC
Contact area
Relative velocity
Running-in Normal operation
Normal load
Operating temperature
31-18
Handbook of Lubrication and Tribology Particle type
Source Boundary lubrication Normal
Machine start-up Rubbing wear
Break-in Abrasive contamination
Cutting wear
Sharp edges Fatigue wear spheres, laminar particles
Hard/soft materials Long operating time Overload
Severe sliding wear
Overheating
Abnormal
Excessive speed Oxide particles
Lubricant starvation Corrosive wear Corrosive contaminants Acidic lubricant
Spheres
Cavitation erosion High temperature (melting)
Se ve r
ity
of
we a
r
Sources of wear particles.
Quantity of particles
FIGURE 31.17
1
10 100
Spectrom
etry
Ferrograp hy Particle co untin
g
FIGURE 31.18
Particle S
ize mm
Magnetic
plugs
Progression of wear.
comprehensive classification of modes of wear is that employed by Godfrey [29]. He classifies wear into the following categories: • Adhesive wear • Mild adhesive wear
© 2006 by Taylor & Francis Group, LLC
Contamination Control and Failure Analysis
31-19
Surface roughness Humidity Temperature Force Velocity
Velocity
Hardness
Fluid
Type of motion Load Heat transfer characteristics
FIGURE 31.19 Tribological parameters. Bulk fluid
Force F Velocity v
Wearing surfaces
Fluid
Film Enlarged view of the surface
Boundary lubricant Asperities
A
Section A-A
FIGURE 31.20 Contact areas between surfaces (dry friction).
• • • • • •
Severe adhesive wear or scuffing Abrasion Erosion Fatigue Delamination Corrosive
© 2006 by Taylor & Francis Group, LLC
A
Height of asperities
31-20
• • • • •
Handbook of Lubrication and Tribology
Electro-corrosive Fretting corrosion Cavitation damage Electrical discharge Polishing
Each of these types of wear categories has its own mechanism and symptoms. In practice, wear usually results from a combination of a number of wear processes, which may be concurrent or follow each other. Before a wear problem can be diagnosed, contributing wear mechanisms must be identified. Previous history of failures (and results of postmortem) may provide useful information in this regard. 31.2.6.1 Adhesive Wear Adhesion wear is the result of formation and rupture of interfacial adhesive bonds (e.g. “cold welded junctions”, “scuffing”), Figure 31.21. The volume of material removed during adhesive wear is normally high during the initial “running-in” period in which the high asperities are plastically deformed and sheared and some material is removed by micro ploughing of the softer material. This initial high wear rate is usually followed by a marked decrease in the rate at which the material is removed, referred to as a stabilized (equilibrium) wear, see Figure 31.22. An important characteristic of adhesive wear is the transfer of metal from one surface to the other. Although transfer of material can also occur during abrasive wear, fretting and corrosive wear does notin itself identify any particular form of wear; simultaneous occurrence of adhesion and transfer
Force F
Fluid
Velocity v
Cold joints
FIGURE 31.21
Mechanism of adhesive wear.
Wear volume
Failure Unstabilized wear Wear-in Stabilized wear
Time
FIGURE 31.22
Running-in process.
© 2006 by Taylor & Francis Group, LLC
Contamination Control and Failure Analysis
31-21
identifies severe wear conditions resulting from dry sliding. The character of adhesive wear can be influenced by sliding velocity, load, presence of lubricant, material properties, and presence of a gaseous environment. If the load and speeds are high, the equilibrium condition may not be reached and the wear rate may remain high causing premature failure of the machine component. As the speed and load is varied, the wear regime may change from mild adhesive wear (also called continuous wear, rubbing, or normal wear by some authors) to severe adhesive wear in which case the machine will fail catastrophically. Mild adhesive wear has following characteristics: • • • •
Repetitive wear on metal surfaces Small flakes of metal Less than 10 µm in size Occurs throughout life of machine
Mild (normal) wear particles are shown in Figure 31.23. Severe adhesive wear or scuffing can be recognized by gross amounts of transferred material and is characterized by: • Metal to metal contact — no lubricant film • Visible parallel striations on co-surfaces • Particle sizes between 5 and 5000 µm (usually 15 to 50 µm) This type of wear is caused by lubricant loss, overloading, and wrong oil. It typically occurs in places were sliding occurs — tips of gear teeth, pistons, loose bearings — typical particles are shown in Figure 31.24 and Figure 31.25. Surface films have a marked influence on both friction and adhesive wear. Wear can be reduced by employing low shear strength surface films on a hard base material. This technique allows the load to be supported through the soft film by the base material, while any shear takes place within the surface film. Suitable types of film include oxides, chlorides, sulfides, other reaction products, soft materials such as lead, silver, copper, and many other nonmetallic materials. To minimize the bond strength between asperities, the sliding pair should be selected in such a way that the surface film is mutually insoluble in the base metals or form intermediate compounds. 31.2.6.2 Surface Fatigue Surface fatigue wear is caused by tribological fatigue stress cycles resulting in crack formation in surface regions of contacting surfaces and separation of material (“pitting”), see Figure 31.26.
FIGURE 31.23 Normal wear.
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Handbook of Lubrication and Tribology
FIGURE 31.24
Severe sliding wear — 1000× (max. particle sizes 35 µm).
FIGURE 31.25
Sliding wear particle — 450× (particle size approx. 60 µm).
Fatigue cracks
FIGURE 31.26
Fatigue wear — formation of spalls.
Fatigue is recognized by surface and subsurface cracks. Surface fatigue occurs when repeated sliding, rolling, or impactive loads are applied to a surface. These loads subject the surface to repeated cyclic stress that initiate cracks near the surface. With time these cracks spread, link up, and form discrete particles that will detach from the surface and can contribute to three-body-wear abrasion and rapid deterioration of the surface may occur due to the flaking off of further fragments. Edges of the cracks are usually sharp and angular and are distinguished from corrosion pits by the presence of other cracks at the edges of pits and surfaces with “beach” marks or ripples indicating progressive stepwise cracking. Fatigue wear is characterized by: • Subsurface stressing in gears, rolling element bearing etc. • Large particles between 5 µm and 5 mm in size (usually 15 to 50 µm)
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Contamination Control and Failure Analysis
31-23
Fatigue wear can be observed in rolling element bearings, the pitch line of gear teeth, and some other tribological pairs where high contact, cycling, stresses are present. Surface fatigue or rollingcontact fatigue is particularly important in the operation of ball and roller bearings, and is usually caused by excessive loads. A fatigue type particle generated during failure of a bearing is shown in Figure 31.27. The fatigue wear particles are often accompanied, specially in the case of rolling element bearings, by spheres (<5 µm diameter, Figure 31.28) and so-called laminar particles. The exact origin of spheres is still in dispute. It has been observed that spherical particles appear, and at constant load on the bearing the quantity of spheres rises exponentially, long before any indication of bearing trouble, characterized by spalling, is detected. Laminar particles are generated by the flattening of fatigue particles during their passage between rolling elements and cages, Figure 31.29. Gears are subjected to both sliding and rolling motion; thus both adhesive and fatigue wear is present depending on gears loading speed and position, Figure 31.30. Adhesion wear may occur at the root and tip of gear teeth as these regions of teeth are in sliding motion. Typical gear fatigue particles are shown in Figure 31.31 and severe wear (scuffing) particle in Figure 31.32.
FIGURE 31.27 Rolling fatigue wear particle — 1000× (particle size approx. 60 µm).
FIGURE 31.28 Spheres.
© 2006 by Taylor & Francis Group, LLC
31-24
FIGURE 31.29
Handbook of Lubrication and Tribology
Laminar particle. SEM 2500× (approx. 40 µm).
Overloading
Load
Catastrophic failure — toothfracture
Fatigue wear — pitting Severe sliding wear — scuffing
Speed
FIGURE 31.30 Wear in gear sets.
FIGURE 31.31
Gear fatigue particles. SEM 1000× (maximum particle size 25 µm).
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Contamination Control and Failure Analysis
31-25
FIGURE 31.32 Gear scuffing (adhesion) particle. SEM 2500× (max. particle size 25 µm). Two-body abrasion Soft surface
Hard surface Asperities Three-body abrasion Contaminant
Hard surface
Soft surface
FIGURE 31.33 Mechanisms of abrasive wear.
The fatigue wear may occur at the pitch line (pitting). We recognize two types of pitting: • Arrested pitting: Arrested pitting occurs in the initial stages of gear operation when there are large local contact stresses. Usually if the gears are operating within design limits this pitting stops after a while and gears will reach their design life. • Progressing pitting: In progressive pitting mode the damage becomes progressively more severe until the gear fails. 31.2.6.3 Abrasive Wear Abrasion (cutting) wear is caused by the removal of material by ploughing (micro-cutting process). Abrasion is recognized by the presence of clean furrows (grooves) in the direction of sliding. There are two basic mechanisms of abrasive wear — two-body and three-body, see Figure 31.33. During the two-body abrasion process, grooves are produced by hard asperities scraping the softer material. The wear can be controlled to some extent by careful selection of relative hardness and surface finish of the sliding pair during design, although other factors such as micro-structure and elastic modulus can have some influence. It is common practice to make one surface considerably softer than the other in order to confine any damage to the component that is easily replaceable. The harder surface is also given a good surface finish to minimize abrasion by protruding asperities. During the three-body abrasion, hard particles that have been generated due to severe adhesive wear or grit from an external source get embedded in a softer material and provide a cutting tool. The wear can be controlled by eliminating dust or grit from the sliding surfaces by seals and minimizing the abrasion from
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Handbook of Lubrication and Tribology
FIGURE 31.34
Cutting wear. Particle
Eroded particle
FIGURE 31.35
Erosion process.
other wear debris, that is, by reducing fatigue and adhesive wear in the system. A special case of three-body abrasive wear can arise when hard wear particles and external contaminants become partly embedded in the softer counter-face and act like sand paper cutting grooves on the harder counter-face and produce fine swarf referred to as wire-wooling. Typical cutting wear particles are shown in Figure 31.34. Factors that contribute to the appearance of abrasive wear are low-quality machining, contaminants, and misalignment of components of tribological pairs. Abrasive wear is characterized by: • • • • •
Abrasive contact between two surfaces Two-body wear (metal/metal) — softer metal is removed Three-body wear (metal/contaminant/metal) — harder metal is removed by contaminant Curls of metal (like lathe swarf) appearance of particles Particle sizes between 5 and 5000 µm (usually 10 to 85 µm)
31.2.6.4 Erosive Wear Erosive wear is a special case of abrasive wear and is produced by the impingement of sharp particles on a surface, see Figure 31.35. The particles are often transported in a moving liquid or gas and can also occur due to cavitation of pumps at inlet. The worn surface is characterized by short grooves due to the micro machining action of the abrasive particles. Important factors that affect selection of materials for components, which could be subject to erosion wear by abrasive particles, are: velocity, impact angle, hardness, quantity, density, and surface contour. At low impact angles harder materials offer the best wear resistance to erosion whereas at large impact angles the appropriate material depends on the particle velocity. At low impact angles, the abrasive particles remove the material by a cutting process (abrasive wear) hence hard materials offer the best solution. At
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Contamination Control and Failure Analysis
31-27
Platelet (flake)
FIGURE 31.36
Delamination process.
FIGURE 31.37
Delamination particle (approx. 60 µm).
large impact angles and low velocities rubbers and resins are used because these materials can absorb the energy of impact much more efficiently than harder materials. 31.2.6.5 Delamination Delamination wear is the loss of metal due to formation and propagation of subsurface cracks parallel to the surface, see Figure 31.36. Delamination occurs due to the plastic flow of the material in conjunction with sliding causing subsurface cracks to propagate parallel to the surface, see Figure 31.37. 31.2.6.6 Corrosive Wear Corrosive wear has been defined as a wear process in which the predominant factor affecting chemical or electrochemical reactions is the environment. As chemical reaction rates increase with increase in temperature — corrosive wear tends to accelerate at higher temperatures. The surface of metals subjected to corrosive wear appears as rough pits or depressions that are irregular and discolored. 31.2.6.7 Electro-Corrosive Wear This type of wear results from stray currents, galvanic action, and streaming potential. 31.2.6.8 Fretting Corrosion Fretting is a form of surface damage that occurs when loaded surfaces are subjected to small amplitude oscillations. To reduce fretting corrosion the obvious solution is to eliminate oscillations. However many systems are required to oscillate, in which case it may be possible to increase the magnitude of oscillations or to rotate one surface continually. The other method to reduce fretting corrosion is to reduce the amount of oxygen present at the contacting surfaces by using antioxidants and surface coatings.
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31.2.6.9 Cavitation Wear Cavitation wear produces very rough but clear surfaces that may sparkle under intense light. This type of wear occurs at pumps’ inlets. 31.2.6.10 Electrical Discharge Electrical discharge across a gas or oil film can cause micro craters or tracks in which there is evidence of molten metal. 31.2.6.11 Polishing Wear This type of wear produces a smooth mirror finish and can occur in engine bores, piston rings, etc. The basic mechanism is attributed to fine scale abrasion, corrosion, and electrolytic corrosion.
31.2.7 Contamination Contamination of a fluid is a general term encompassing debris in hydraulic fluid — generated in the system or ingressed from outside. Solid, hard, and sharp particles (metallic, rust, sand) may contribute greatly to component wear; whereas soft, nonmetallic particles (seal residue, fibrous material, paint) will usually have an operational effect due to blockage of clearances and orifices. Even a small quantity of large solid particles will invariably cause catastrophic failure of hydraulic components; on the other hand a large quantity of small solid particles may cause catastrophic failures due to silting. Thus the effect of both large and small particles on system integrity must be considered when selecting a filtration system. 31.2.7.1 In-Built There are basically two types of built-in contaminant, one which is introduced during filling-up with “new fluid” and the other resulting from system construction. “New fluid” is not necessarily with “clean fluid.” Tests at the Centre for Machine Condition Monitoring, Monash University carried out under the auspices of a Basic Research Program showed that “new fluid” samples may have contamination levels well in excess of acceptable levels. Thus filling a system with a “new fluid” through a suitable filter is of great importance to give a system a “clean” start! Hydraulic system components are manufactured with great care. However, inevitably after manufacturing, assembly of the system, and system flushing a certain amount of contaminants remain in the system. Typical built-in contaminants include burrs, sand, weld splatter, paints. The efficiency of the removal of built-in contaminant during flushing is dependent on the velocity of the flushing fluid; high velocities are required to dislodge built-in particles. Most manufacturers provide system flushing procedures that should be followed to the letter. 31.2.7.2 Ingested Environment contributes greatly to system contamination. The contamination enters the system via the fluid reservoir (air breathers and access covers), and any sealing pairs (cylinders seals, pump, and motor seals). It is estimated that approximately 50 to 60% of ingressed contaminants enter via the cylinder seals and it can be expected that the amount of ingression will increase with seals wear. Ingression rates from 0.03 to 0.2 g/h were reported in literature [30,31]. 31.2.7.3 Maintenance Generated Following improper maintenance procedures may be a major reason for widespread contamination of the system. A “new fluid” added to the system without appropriate precautions is a major source of contaminant. Other activities, like replacing a fitting may add 6,000 to 60,000 particles >5 µm to the system. Thus following proper maintenance procedures is extremely important in maintaining the clean condition of the system. 31.2.7.4 Internally Generated The internally generated contaminants are the result of wear processes (abrasion, adhesion, fatigue), corrosion, cavitation, oxidation, and fluid breakdown. From the operational point of view the rate of wear
© 2006 by Taylor & Francis Group, LLC
Contamination Control and Failure Analysis
31-29
in hydraulic components cannot be great — otherwise the performance of the components will be affected due to resultant leakage (pumps, valves) or friction. The generation of the contaminant will be accelerated by the contaminant already present in the system, thus the rate of generation of the contaminant is closely aligned with the magnitude and rate of the other sources of contamination (ingested, maintenanceinduced, or in-built). The actual amount of wear will depend on the size and the material of the solid contaminants, the ratio of the particle size to the working clearances, the shape of the particles, the lubrication regime (boundary lubrication, hydrodynamic), the pressure, the flow velocity, and the fluid type. Presence of wear and contamination in hydraulic and lubrication systems is probably the most important factor that affects their longevity and reliability [32–35].
31.2.8 Mechanical Modes of Failure In the literature there are many different classifications or categorizations of mechanical failure modes. Categorization of failure modes proposed by Collins [36,37] is presented in Table 31.2 and is a basis for describing the common mechanical modes of failure listed in Table 31.3.
31.2.9 Hydraulic and Lubrication Failures Proper design of components and system play a great part in reliability of hydraulic and lubrication systems. The incorrect design of a contamination control system will inevitably impose severe maintenance problems. In the case of a hydraulic system, looking at Table 31.4 we may notice that leakage, temperature, and specially contamination control play an important part in counteracting problems in hydraulic systems.
TABLE 31.2
Categories of Mechanical Failures Category
Manifestation of failure
Subcategory
Elastic deformation Plastic deformation Rupture or fracture Material change metallurgical Chemical Nuclear
Failure inducing agents
Failure locations
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Force
Steady Transient Cyclic Random
Time
Very short Short Long
Temperature
Low Room Elevated Steady/transient Cyclic Random
Reactive environment
Chemical Nuclear Human
Body type Surface type
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Handbook of Lubrication and Tribology TABLE 31.3
Mechanical Failure Modes
Category
Subcategory
Elastic deformation
Category
Force induced Temperature induced
Subcategory
Wear
Surface fatigue Deformation
Yielding
Impact
Brinnelling
Fretting
Ductile rupture
Impact
Fracture
Brittle fracture
Deformation
Fatigue
High-cycle Low-cycle Thermal Corrosion Fretting
Creep
Corrosion
Direct chemical attack Galvanic Pitting Intergranular Selective leaching Cavitation Hydrogen damage Biological Stress
Thermal relaxation Stress rupture Thermal shock Galling and seizure Small spalling Small radiation damage Buckling Creep buckling Stress corrosion
Wear
Adhesive Abrasive Corrosive
Corrosion wear Corrosion fatigue Creep and fatigue
Aeration
Flow rate Wear debris analysis Vibration and noise
FIGURE 31.38
Fretting
Visual inspection State of components
Leakage
Fluid condition
Temperature
Fluidborne
Pressure
Fluid analysis Monitoring of hydraulic system
Wear/ contamination Leakage
Wear Fatigue Wear Corrosion
Airborne
Temperature System steadystate behavior
Velocity, position
Temperature
Vibration and noise
Velocity, position
System dynamic behavior
Pressure variations
Flow rate variations
Force variations
Pressure shocks
Flow rate
Leakage
Monitoring of hydraulic system.
Usually, monitoring of hydraulic systems is limited to monitoring of pressure (and occasionally flowrate) in various parts of the circuit and to visual monitoring of the system for leakages, fluid aeration, water contamination, reservoir temperature, etc. [38], Pressure measurement allows identification of excessive pressure losses, leakages etc. Some, more important, techniques are shown in Figure 31.38.
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Contamination Control and Failure Analysis TABLE 31.4
31-31
Characteristics of Hydraulic Systems
Characteristics High power density (ratio of weight/transmitted power) High speed operation High contact stresses (valve plate/cylinder block, gears, etc.) High temperature operation Shock loading Leakage Predominance of sliding tribological pairs (pistons, spools, poppets, vanes, etc.) High precision of equipment (servovalves, pumps, etc.), small clearances Low cycle fatigue (high start-up loads) Fluid as transmission medium
Result in
Preventive measure
High operating stresses
Choice of material, design
Noise, vibrations Fatigue wear
Proper design of system Choice of material, design
Fluid oxidation, distortion, poor lubrication High operating stresses Entry of contaminant, increased temperature Damage due to scoring, sliding wear. Production of wear debris
Temperature control (heat exchange, reservoir design) Pressure control, design Seals, close tolerances, contamination control Selection of materials, contamination control
Very high sensitivity to contamination and wear, silting (valves)
Seals design, contamination control
High stress level
Design, pressure control
Sensitivity to temperature, contamination, aeration
Reservoir design, temperature control, aeration control, contamination control Design of components, choice of materials Noise/vibration control, design Seal design, contamination control Design, pressure control Contamination control, aeration control Education, training, quality control (design audits)
Flow throttling as means of control
Erosion, aeration
Pressures/flow pulsation Need to seal pressure volumes Unskilled operators Poor maintenance practices
Noise, vibrations Close tolerances, seals, leakage Overloading, incorrect operation Contamination, sealing problems
Improper system design procedures
Operational, maintenance, safety problems
For example, Figure 31.39 shows a diagram of a simple hydraulic system with marked locations of contamination entry points and tribological pairs, then using this information appropriate monitoring techniques can be selected and criteria for detection of abnormal conditions of the system can be established. Tribological pairs and wear mechanisms in common hydraulic components are shown in Table 31.5. Visual observation of the state and the behavior of a system (e.g. erratic action of the actuator, loss of motor speed, aerated fluid in the reservoir) were used to detect faults and to identify probable causes using troubleshooting guides widely published by the manufacturers of hydraulic equipment. This type of monitoring is still widely practised. Although monitoring of pressure and visual observation of the system condition are still valid techniques of condition monitoring, the development of a wide range of sensors for monitoring of vibration, noise, flow, contamination, water content, and other parameters of interest and better understanding of phenomena affecting operation of hydraulic systems (e.g. contamination, aeration, flow forces) led to the development of advanced monitoring techniques that are able to detect early stages of failures. Typical effects of aeration and temperature on dynamic behavior of hydraulic systems are shown in Figure 31.40. Contamination control in hydraulic and lubricating systems is fundamental to reliability and performance and no longer needs justification. Ideally, the selection and design of a proper contamination control system should be addressed at the system design stage. However, proper design of a filtration system requires a lot of effort on the part of the designer to find the necessary information about the environment in which the system will operate, the types of components and their tolerances to contamination, user’s maintenance procedures, etc. The amount of effort expended on designing an efficient
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Handbook of Lubrication and Tribology
Sliding wear
Leakage
Sliding wear Fatigue wear
Aeration contamination
Leakage Sliding wear erosion
Leakage
Erosion
Sliding wear erosion
Leakage Leakage Sliding wear Fatigue wear Aeration contamination water
FIGURE 31.39 Wear and contamination in a hydraulic system.
contamination control system usually depends on the complexity and importance of the hydraulic system — for most systems the design task is simply reduced to the selection of appropriate filters on the basis of manufacturer’s recommendation.
31.3 Contamination Balance The technology associated with contamination control is by no means mature. Using available engineering methods, it is neither possible to predict, the cleanliness level in a hydraulic or lubricating system during the design phase, nor to estimate the remaining life of filters in a system under operation. In contamination control, hydraulic components can be classified into four different groups: • Components that capture and retain contamination (filters, strainers) • Components that produce contamination due to wear (pumps, cylinders, valves) • Components that allow induction of contamination from the system’s working environment (breathers, rod seals, rotating shaft seals); manufacturing and assembly residues in the system can be added to this group • Components that mix the contamination level and allow the contamination level to fluctuate; the reservoir is the main component in this group Any scientifically based method of design for contamination control must be based on a fundamental relation of equilibrium between particle ingression and the removal of particles. The contamination characteristics of the system are dynamic in nature and dependent on a number of variables and factors: • Type of hydraulic system • Filtration ratios of filters
© 2006 by Taylor & Francis Group, LLC
Contamination Control and Failure Analysis TABLE 31.5
Hydraulic Systems — Wear Mechanisms
Hydraulic component
• • • • • •
31-33
Tribological pair
Mechanism of wear
Check valve (poppet)
Poppet/body Poppet/seat
Impact, sliding Flow erosion
Check valve (ball)
Ball/body Seat
Small impact Flow erosion
Vane pumps and motors
Bearing ring Vane/rotor Vane/casing
Rolling fatigue Sliding Sliding
Gear pumps and motors
Rolling element bearing Gear pair Gear/thrust plate
Rolling fatigue Rolling fatigue, sliding Sliding, abrasion
Piston pumps and motors
Rolling element bearings Piston/block Block/valve plate Piston/slipper Slipper/swash plate Block/shaft
Fatigue Abrasion Sliding, abrasion Sliding Sliding Fretting
Directional control valves
Spool/body Spool
Sliding Erosion
Proportional and servo valves
Spool/body Jet nozzle Spool
Sliding Erosion Erosion
Pressure control valves
Spool/body Seat/poppet Seat and poppet Orifices
Sliding Impact Flow erosion Flow erosion
Flow control valves
Spool/body Seat/poppet Seat and poppet Orifices
Sliding Impact Flow erosion Flow erosion
Cylinders
Cylinder rod/gland bush Seal/cylinder Swivel bush/pin
Sliding Sliding Sliding
Accumulators (piston)
Piston/cylinder
Sliding
Rotary actuator (vane)
Bearing Vane/body Vane/rotor
Rolling fatigue Sliding Sliding
Rotary actuator (rack/pinion)
Seal/cylinder Rolling element bearing Gear pair Cylinder rod/gland bush Seal/cylinder
Sliding Rolling fatigue Rolling fatigue Sliding Sliding
Environment Expected contaminant ingression rates Target cleanliness levels for system components Type of filtration Duty cycle, fluid loss Type of system (e.g. by-pass) and others
Basic relationship between particles entering and leaving a system is shown in Figure 31.41 and on this basis three different methods have been developed. The first step in developing such a scientific method
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Handbook of Lubrication and Tribology
Higher
Speed of response
Leakage
Temperature
Viscosity
Lower Damping Speed of response
Natural frequency
Laminar losses
Cavitation Fluidborne noise
Higher Bulk modulus
Lower Stiffness Flow forces Higher Temperature water contamination Lower
FIGURE 31.40
Hysteresis losses
Higher
Higher
Lower
Viscosity
Aeration
Bulk modulus
Lower
Lower
Higher
Shocks
Higher effect Lower effect
Effect of temperature and aeration on hydraulic system. Number of particles entering the system
Number of particles generated in the system
S
Increase of particles number in the system
Number of particles removed from system
FIGURE 31.41
Contamination balance.
was the Oklahoma State University (OSU) model of particles balance in a hydraulic system developed by E. Fitch [4]. D. Anderson presented a slightly different approach to determine contamination balance, however, both results are equivalent. The development of the concept of contamination balance, equating the generation and removal of contamination within the system, provided a mathematical tool to describe some of these interactions and thus led to a better understanding of the contamination characteristics of hydraulic control and lubrication systems.
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Contamination Control and Failure Analysis
31-35
R1
R2 Nr
b
Q Nu
FIGURE 31.42
Nd
OSU model.
The importance of considering contamination balance in the system, that is, matching ingression rates to the contamination tolerance levels of components in the system and filter characteristics, is now well recognized. The concept is simple — to maintain the contamination within the allowable contamination level in the system for any amount of contamination that is ingested or generated within a system, an equal amount of contamination must be removed. The concept is, however, difficult to apply in practice especially in the case of multi-branch systems consisting of many hydraulic components, as both ingression rates and contamination tolerance levels, as well as contamination levels in various points of the circuit, are difficult to quantify. The original work at OSU that led to the development of the concept of contamination balance is limited to simple systems.
31.3.1 Oklahoma State University (OSU) Model Assumption of a constant rate of wear and constant rate of contamination. The equation for material balance in the reservoir, see Figure 31.42 Nr V =
(R1 + R2 )dt +
Nd Qdt −
Nu Qdt + No V
(31.3)
where V is volume of fluid, No is initial contamination level (number of particles greater than size d per unit volume), Nr is contamination level in reservoir (number of particles greater than size d per unit volume), Nu is contamination level upstream of filter (number of particles greater than size d per unit volume), Nd is contamination level downstream of filter (number of particles greater than size d per unit volume), R1 is rate of particles ingression (number of particles greater than size d per unit time), R2 is rate of particles generation by system components (number of particles greater than size d per unit time), and Q is flowrate. Characteristic size of particle is denoted by d. The ratio of Nd to Nu (for particles size >µm) is defined as β-rating of the filter. The contamination level upstream of the filter is expressed by: Nu =
β (R1 + R2 ) (1 − e−t /τ ) β −1 Q
(31.4)
where time constant τ is equal to: τ=
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β V β −1Q
(31.5)
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Handbook of Lubrication and Tribology
Machine
x(1–ai)
N(ai,n)
Pump
b
FIGURE 31.43 Anderson’s model.
Steady-state contamination level upstream of the filter (t −→ ∞) is then equal to: Nu =
R1 + R 2 β Q β −1
(31.6)
Nd =
R1 + R 2 1 Q β −1
(31.7)
and downstream of the filter:
These equations assume that β-ratio, Q, V , and particle ingression/generation rates are constant. A more complete filtration model is provided by Anderson’s Model. Anderson developed his model to determine equilibrium of particles in a system under following assumption, see Figure 31.43: • During one cycle full volume of oil passes through the machine, that is, if a pump has 100 lpm flowrate and total volume of fluid in the system is 500 l it will take 5 min to complete one cycle (assuming that all oil enters the pump). • During one cycle machine produces x particles per unit volume. • The removal efficiency of size i particles, where i refers to characteristic length of the particles, is denoted as ai , 0 ≤ ai ≤ 1. • During each cycle ai x particles are destroyed during passage through system components, lost in the reservoir or removed by the filter. After one cycle a number of particles in a fluid volume (particle concentration) is equal to: N (ai , 1) = x(1 − ai )
(31.8)
where N (ai , 1) is concentration of particles of size i after one cycle. After two cycles concentration is equal to: N (ai , 2) = x + x · (1 − ai ) · N (ai , 2) = x + x · (1 − ai )
(31.9)
and ai [x + x · (1 − ai )] particles are removed. After n cycles particle concentration is: N (ai , n) = x + x · (1 − ai ) + x · (1 − ai )2 + · · · + x · (1 − ai )n−1
(31.10)
and substituting y = 1 − ai N (ai , n) = x · (1 + y + y 2 + · · · + y n−1 ) = x
r=n r=1
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y r−1
(31.11)
Contamination Control and Failure Analysis
31-37
then concentration is equal to:
N (ai , n) =
x · (1 − y n ) x · (1 − (1 − ai )n ) = 1−y ai
(31.12)
when n =⇒ ∞, the concentration approaches: N (ai , ∞) ⇒
x ai
(31.13)
We now define quantity α, α 1, such that after R cycles N (ai , R) ≤1−α N (ai , ∞)
(31.14)
and we can calculate a number of cycles R to reach the given percentage of equilibrium concentration 1 − α:
R≥
ln α ln(1 − ai )
(31.15)
βi − 1 βi
(31.16)
As the efficiency of the filter is equal to:
ai =
where βi is filtration ratio, we may now express Equations (31.12), (31.13), and (31.15) in terms of βi :
N (ai , n) = x
βin − 1 (1−n) β βi − 1 i
(31.17)
βi βi − 1
(31.18)
N (ai , ∞) = x
R≥
ln α ln(1 − (βi − 1)/βi ))
(31.19)
Thus as the efficiency of filtration for small particles is small, thus a large number of cycles is required before equilibrium is reached.
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Handbook of Lubrication and Tribology
Both OSU and Anderson models show that the following factors will affect length of time needed by the system to reach particle equilibrium: • Filtration — number of times a particle of a given size and material, on average, passes through the filter. The better the filter the shorter the time to equilibrium. Deterioration of filter (due to gradual clogging of the filter, by-passing etc.) will affect particle removal efficiency — specially small particles. • Fluid cycle rate — expressed in volume per unit time divided by the volume of lubricant in the system. Varies from 5 times per minute to once per hour or longer. It affects material loss by filtration. • Location of slow moving oil — the particles may settle or adhere to surfaces, for example, bottom of sump. • Dispersion quality of lubricant — these additives (detergents) prevent agglomeration of the particles and also discourage their adherence to the surface, thus increasing their life expectancy. • Breakdown of particles — during repeated passage through wear contacts the large particles will be reduced to smaller sizes. • Oil loss due to leakage will remove some particles from the system (but it is not the recommended way of controlling contamination!) • Oxidation and chemical attack will change the size of particles. The model used by Anderson, similar to the OSU model, has the following deficiencies: • Assumption of constant wear rate (constant x), this assumption is only applicable to normal wear. • Assumption of constant ai ignores removal of particles due to oil losses and deterioration in filter efficiency. • All wear debris producing components and contamination ingression points were lumped together and represented as a single source of particles. A real system has many points where wear is produced and contamination from the environment is ingressed into the system. The rate of wear generation and contamination ingression varies between components. • Only one fluid path is considered, real systems have usually more than one fluid path. The particle concentration in each flow path will depend on the flow rate and the particle generation/ingression rate by the components in the path. • Changes in contamination condition due to leakages, oil addition, changes in efficiency of filters etc. cannot be included in the model. • Each model gives concentration of particles in the reservoir, in practice we are interested to know the exposure of various components to wear and contamination particles. Nevertheless both OSU and Anderson’s models provide better insight into the behavior of contamination control system and provide a basis for sampling guidelines. The models show that large particles will reach equilibrium much sooner than small particles, thus care should be taken to assure that sufficient time elapsed to allow concentration of particles in the system to reach an equilibrium before the samples are taken. This should be taken into consideration when using monitoring methods that are not able to detect small particles for example, spectrometric oil analysis (SOAP), see Figure 31.44.
31.3.2 Dynamic Contamination Control (DCC) The concept of contamination balance provided sufficient insight to permit recognition and understanding of the controlling parameters of the system and to permit adequate description of the influence of each parameter on the system as a whole. To describe quantitative relationships between variables and parameters describing fluid contamination in hydraulic control systems a mathematical model in the form of a Dynamic Contamination Control (DCC) was developed at the Centre for Machine Condition Monitoring, Monash University in Australia [39,40]. A DCC allows analysis and optimization of contamination control in a fluid power and lubrication system. The model of the contamination control system is
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Contamination Control and Failure Analysis
31-39
Normal wear
Run-in
Abnormal wear
Concentration
Large Large
Small
Large — after oil change
Small Time Small — after oil change
FIGURE 31.44 Equilibrium of small and large particles in a system.
Q
N3
R2
Control valve
R3
N3
N2
B1
N4
N2 1 b1
B2 N1
R0
N5
R3/Q
R2/Q Filter
Filter R1/Q
N1
R1
Pump
Actuator
N0
Tank 1 (V/Q)s + 1
1 b2
R0/Q Ni
N5
N0 (a)
N0...N5 contamination level R0...R4 ingression level Q flowrate
(b)
FIGURE 31.45 DCC Model (Monash University).
solved using digital simulation, the results provide information about the contamination level at various locations in the circuit. The model also allows investigation of the contamination levels at various points in the circuit in response to variation in the size distribution and the level of ingressed or generated contamination. It also allows investigation of the effects of changes in the filtration ratios of filters and their locations in the system. This model has been proven to be accurate through experiments carried out at the Norwegian University of Science and Technology [41]. 31.3.2.1 Single Path Model A single path hydraulic system, shown in Figure 31.45(a), is used as an example to illustrate the DCC model. The system includes a return line and pressure line filters, a hydraulic pump, a reservoir, a directional control valve, and a cylinder. The contamination control system is represented by a block diagram in which each block represents the function of a single hydraulic component in the contamination control system. The variable of interest in contamination control is contamination level Ni (number of particles larger than a specified size per unit volume of fluid). Thus, the input and output of the block for an individual hydraulic component are the upstream and downstream contamination levels at the location of the component. The DCC model
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Handbook of Lubrication and Tribology
of a whole hydraulic control system can be obtained by linking the contamination block diagrams of the components of the system. Output of a component block diagram becomes the input of the following component block diagram according to the hydraulic circuit diagram. The DCC model of the system is shown in Figure 31.45(b). Some of the variables and parameters in a contamination control system such as flowrate Q, filtration ratios of filters, particle ingression rates R vary with time, they may also be affected by changing size distribution of contaminant particle, changing duty cycle, filters clogging, or changes in operating environment. Although these variations in values of parameters are not currently modeled, this technique can be readily extended to include these variations in the model when and if the information about interaction between various parameters becomes available. In our example, as the objective of the model is to use it in assessing final levels of contamination in the system when the system contamination levels reach the state of equilibrium, we assume mean flowrate Q, and constant filtration ratios β1 , β2 and ingression rates R0 , . . . , R3 . The block diagram of DCC model shows the location in the system where contamination ingress and/or wear debris generation occurs (summing junctions) and the location of filter(s). The topology of the block diagram is identical to the modeled real system. The contamination level at any point in the circuit can be directly calculated from the closed-loop DCC model. For the system shown in Figure 31.45, the following equations can be written: Ni =
N0 R1 R2 R3 R0 + + + + β1 β2 β1 β2 Q β2 Q β2 Q Q
(31.20)
where Ni , N0 are input and reservoir contamination levels (No. of particles >d µm per volume), β1 , β2 are filtration ratios, Q is mean flowrate in the system (volume per time), R0 , . . . , R3 are contaminant ingression rates (No. of particles >d µm per time). The dynamic change in contamination level in the reservoir is described by the following equation: N0 =
Ni (V /Q)s + 1
(31.21)
where V is reservoir volume, s is Laplace operator. Substituting Equation (31.20) into (31.21) we obtain: Ni N0 = = (V /Q)s + 1
N0 R1 R2 R3 R0 + + + + β1 β2 β1 β2 Q β2 Q β2 Q Q
1 (V /Q)s + 1
(31.22)
and N0 −
N0 1 = β1 β2 (V /Q)s + 1
R1 R2 R3 R0 + + + β1 β2 Q β2 Q β2 Q Q
1 (V /Q)s + 1
(31.23)
then N0
β1 β2 [(V /Q)s + 1] − 1 = β1 β2 [(V /Q)s + 1]
R1 R2 R3 R0 + + + β1 β2 Q β2 Q β2 Q Q
1 (V /Q)s + 1
(31.24)
1 (V /Q)s + 1
(31.25)
and rearranging: N0
[(V /Q)s + 1] − 1/β1 β2 = [(V /Q)s + 1]
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R1 R2 R3 R4 + + + β1 β2 Q β2 Q β2 Q Q
Contamination Control and Failure Analysis
31-41
1/ß2
1/ß1
1/ß2
from pump R2/Q
Input-2
from valve
Nm+No/ß1ß2
1
No
R1/Q Input-1
1/ß2
+ (V/Q)S+1
R3/Q Input-3 from actuator R0/Q
Input-4
from reservoir 1/ß1
1/ß2
Multi-Input
Feedback path input No/ß1ß2
Nm=R1/Q(ß1ß2)+R2/Qß2+R3/ß2+R0/Q
FIGURE 31.46 Reduced model.
finally, the contamination level at exit from the reservoir is described by: N0 =
{R1 /β1 β2 Q + R2 /β2 Q + R3 /β2 Q + R4 /Q} Nm = [(V /Q)s + 1] − 1/β1 β2 [(V /Q)s + 1] − 1/β1 β2
(31.26)
where Nm =
R1 R2 R3 R4 + + + β1 β2 Q β2 Q β2 Q Q
(31.27)
is the equivalent contamination input to the system, and the contamination level in the reservoir is equal to, see Figure 31.46: N0 =
Nm Nm = [(V /Q)s + 1] − 1/β1 β2 (V /Q)s + (1 − 1/β1 β2 )
(31.28)
Filtration ratio (β-ratio) of a filter is defined as the ratio of contamination input (measured as a number of particles above a certain size per fluid volume) to contamination output. The β-ratio is measured by subjecting the filter to a Multi-pass test ISO 4572-1981 and may vary from 2 to 70 (β-ratio = 1 indicates no filtering action). As good quality filters have β-ratio much greater than 1 then: 1−
1 >0 β1 β2
(31.29)
In time domain solution (31.28) yields: N0 (t ) =
Nm Nm − e−(Q(1−F )/V )t 1 − F (β) 1 − F (β)
(31.30)
where: F=
1 βi
i = 1, 2, . . . , n
Time constant of the system is defined by: T =
© 2006 by Taylor & Francis Group, LLC
Q(1 − F ) V
(31.31)
31-42
Handbook of Lubrication and Tribology Q3
Circuit Form
N31
N33
1/ß31 Rz31/Q3
1/ß21
Branch
Rz32/Q3
N32 Q2
N21
N23
1/ß22 Rz21/Q2
N24
N22
Rz22/Q2
Q1 1/ß11
N11
N12
N4
Rz12/Q1
Rz11/Q1
N2
N13
Q
1/ß1
R0/Q
R1/Q Pressure line
N1
N5
1 (V/Q)s +1
No
No
1 (V/Q)s +1
1/ß2
Ni
Ncom
Ni
Feedback Form
Return line
+
Fcom
FIGURE 31.47
Multi-branch contamination model.
and time to reach equilibrium is approximately equal to: T0 ≈ 4T = 4
Q(1 − F ) V
(31.32)
31.3.2.2 Multi-Path Systems Most hydraulic systems include more than one flow path — the above outline modeling approach can be easily applied to systems having multiple flow paths each having different flowrate and different ingression rates, a DCC model of a multi-branch system that has three flow branches is shown in Figure 31.47. Each branch flow can be treated as a fraction of total flow, thus: Qi = Ai Q
k = 1, 2, . . . , n
(31.33)
where Qi is flowrate in ith branch of the circuit, Ai is fraction of total flow in ith branch. It can be seen that:
Qi = Q A=1
(31.34)
For the contamination system shown in Figure 31.47 that has three parallel paths, we may write: Q1 + Q2 + Q3 = Q
(31.35)
To convert the original system to feedback form we express N4 as follows: N4 = A1 N13 + A1 N24 + A1 N33
© 2006 by Taylor & Francis Group, LLC
(31.36)
Contamination Control and Failure Analysis
31-43
where: N13 =
N0 Rz11 Rz12 R1 + + + β1 β11 A1 Q A1 Q β1 β11 Q
N24 = N33 =
N0 Rz21 + β1 β21 A1 Q N0 Rz31 + β1 A3 Q
Rz22 1 R1 + + β22 A2 Q β1 β21 β31 Q
(31.37)
Rz32 1 R1 + + β31 A3 Q β1 β31 Q
and A1 =
Q1 , Q
A2 =
Q2 , Q
A3 =
Q3 Q
Input contamination level Ni is equal to: Ni =
N4 R0 + β2 Q
(31.38)
and substituting above expressions we obtain the expression for Ni : Ni =
N0 Rz11 Rz12 R1 + + + β1 β11 A1 Q A1 Q β1 β11 Q
+ +
N0 Rz21 + β1 β21 A1 Q
N0 Rz31 + β1 A3 Q
A1 β2
Rz22 1 R1 A1 + + β22 A2 Q β1 β21 β31 Q β2
Rz32 1 R1 + + β31 A3 Q β1 β31 Q
R0 A1 + β2 Q
(31.39)
We can now express Ni in this form: Ni = Ncom + Fcom N0
(31.40)
where Ncom is combined contamination input to the reservoir, Fcom is combined filter effect. For our circuit: Ncom = A1 +
R1 R1 R1 Rz11 + A2 + A3 + β1 β2 β11 β1 β2 β21 β22 Q β1 β2 β131 Q β2 Q Rz12 Rz21 Rz22 Rz31 Rz32 R0 + + + + + β2 Q β2 β22 Q β2 Q β2 β31 Q β2 Q Q
(31.41)
In general, these equations for a multi-path circuit become:
Fcom =
i=m i=1
© 2006 by Taylor & Francis Group, LLC
Ai βi
(31.42)
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Handbook of Lubrication and Tribology
and
Ncom = Npr + Nbr + Nrt =
k Ri i=1
Q
Fcom +
l m i=1 j=1
Rrij R0 + n βmn Q
(31.43)
where Npr is contamination level in the pump delivery line, Nbr is contamination level in each branch flow path, Nrt is contamination level caused by contamination ingression in the reservoir, Rrij is the jth ( j = 1, 2, . . . , l) ingress rate in the ith (m = 1, 2, . . . , m) branch path, n is number of filters in the i-line between ingression source Rzij and entry to the reservoir, k is number of ingression sources in pressure line, βmn is filters’ β-ratios, m is number of flow branches, and l is number of the ingression sources in the branch. Time solution for linear multi-branch system can be obtained using Equation (31.30) and time to reach equilibrium is calculated using Equation (31.32) in which we replace: Ncom ⇒ Nm
(31.44)
Fcom ⇒ F
Modeling procedure outlined above is shown in Figure 31.48. Typical packages that can be used to model the contamination system are Matlab’s Simulink, ACSL, or Vissim.
R3
1 be R2
Q
Re
System b1
b2
Qs
1 (V/Q)s +1
N0
Ni
R1 bb
R0 V
Re N0
K (V/Q)s +1
Qb
C + N0 * F
F DCC N3
N2
1 b1
N1
FIGURE 31.48
R2
Q
R1
N0
be R3 R0
1 (V/Q)s +1
Ni
Qs
1 b2 Ni
DCC modeling procedure.
© 2006 by Taylor & Francis Group, LLC
Reduced DCC Model
Model
N0
Re V
Contamination Control and Failure Analysis
FIGURE 31.49
31-45
Block diagram of contamination system shown in Figure 31.46.
31.3.2.3 Simulation Using the above modeling approaches we may assess contamination levels analytically. However, an analytical approach to this task is rather tedious as in practice we are interested in likely effects of system parameter changes (e.g. β-rating of filters, locations of filters, size of reservoir volume) on contamination levels at particular locations in the system. Availability of simulation packages makes it easy to investigate the system and allows experimentation with different parameters affecting contamination of the system. We may also apply a contamination model in parallel with a dynamic model of the system to investigate the effects of changing contamination characteristics of the system on system operation. A Vissim simulation block diagram model of the system presented in Figure 31.45, is shown in Figure 31.49. An example of results obtained from digital simulation is shown in Figure 31.50. Modeling and simulation of contamination in hydraulic systems using the approach outlined in this section allows evaluation of contamination levels in the system and makes it possible to investigate the effects of design changes. This technique could easily be extended to include determination of other characteristics of contamination of a control system, for example, assessing the dirt capacity of the system that will provide information about how often a filter must be changed. Also, nonlinear effects, for example, filter rating change with exposure to contaminant could be easily included.
31.4 Monitoring Procedures Contamination control program is an important component of proactive maintenance and predictive maintenance strategies, see Figure 31.51: • Predictive maintenance strategy is based on monitoring the condition of the system and, when changes in system condition that could lead to failure are detected, corrective action is taken.
© 2006 by Taylor & Francis Group, LLC
31-46
Handbook of Lubrication and Tribology No. of particle / 100 ml - reservoir
200
No. of particles / 100 ml
175 150 125 100 75 50 25 0 0
250
500
750
1000 1250 1500 1750 2000 2250 Time (sec)
FIGURE 31.50
Result of simulation — level of contamination in the reservoir. Based on condition of the equipment The maintenance action when system is diagnosed to be in failure mode or modes
Generalized load
On-condition Maintenance
Large range of condition monitoring techniques Effective in cases when no past history of failures is available
Capacity
Duty
Used if there is a possibility of multiple modes of failures Used if there is a low probability of a particular failure T2
T1
T3
Time Corrective action Proactive Maintenance
Eliminates conditions which could lead to failure of the system Does not replace on-condition Maintenance Includes monitoring of root-causes of the failures
Generalized load
Based on monitoring of system duty and changes in system elements
Capacity
Duty T2
T1
T3
Time
FIGURE 31.51
Comparison of predictive and preemptive maintenance.
This strategy is now commonly used in many industries. The monitoring techniques based on the investigation of the wear debris and contaminant carried by a lubricating or hydraulic fluid were recognized long ago as valuable detection and diagnostic tools. Analysis of lubricant/working fluid characteristics provide additional information about the state of a machine, possible wear, and friction regimes and thus in combination with particulate based analysis enhance both monitoring and diagnostic characteristics of tribology based condition monitoring methodology. The onset of damage in machinery can be detected by examining the wear and contaminant particles present in the lubricant. The shape, size, size distribution, and number of wear particles present in a sample as well as the morphology and the condition of the particles give many clues as to the state of a machine and the possible incipient damage, and thus help the analyst to judge the likely future
© 2006 by Taylor & Francis Group, LLC
Contamination Control and Failure Analysis
Root cause
31-47
Water in fluid
Oxidation
Proactive
Lubricant deterioration
Effect
Increased wear Predictive System failure
FIGURE 31.52
Relation between proactive and predictive monitoring strategy.
performance of the machine. The reliability of diagnosis and prognosis of the condition monitoring program will be directly affected by the selection of detection methods. Selection of a proper mix of monitoring methods that will be able to identify changes in machine condition is thus important. • Proactive strategy is based on the monitoring of root causes of failures and taking corrective actions to eliminate or minimize them. Although monitoring procedures closely resemble those used in predictive maintenance the purpose of monitoring is to provide guidance for eliminating conditions that could lead to failures; Figure 31.52 shows the comparison between proactive and predictive monitoring approaches. Monitoring procedure is shown in Figure 31.53 and consists of a number of steps: 1. Sampling: (a) Obtaining particulate and lubricant/working fluid that are representative of tribological state of the machine (b) Obtaining information about presence of conditions that may cause failure, for example, presence of water, high temperature 2. Detection (a) Detecting symptoms of an abnormal condition of a machine that indicates an incipient failure of a machine by comparing the results of particulate and fluid analysis with previously determined guidelines (b) Detecting changes in properties of particulate and fluid analysis with previously determined guidelines to determine initiation of condition for failure 3. Diagnosis (a) Diagnosing, by comparing a set of symptoms with a known syndrome of machine condition, the mode, severity, location, and mechanism of the wear process that may lead to failure (b) Diagnosing what cause changes in system conditions, for example, leaking seal resulting in contamination 4. Prognosis, on the basis of historical data and diagnosis, of the future state of the machine and remaining service life before failure 5. Action required to correct and/or rectify identified problems 6. Postmortem (a) Investigation of failed components (b) Investigation of what caused changes in system condition
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
Sampling Particulate characteristics Detection
Symptoms Fluid characteristics
System OK
Yes
Condition change?
Compare
Guidelines - level - trend - comparison
No Syndrome= Set of Symptoms
Known failure modes
Compare
Diagnosis
Identify failure - mode - severity - location
Prognosis
Compare
Historical data
Estimate residual life
Action
Identify causes of failure
FIGURE 31.53
Rectify problems
Postmortem
Condition monitoring procedure.
Monitoring program must be designed to evaluate fluid properties, fluid contamination, and wear/contamination condition of the system. Proactive maintenance strategy is based on similar steps; however it is concerned with monitoring root causes, that is, system parameters and variables that indicate presence of conditions, which may lead to development failure symptoms.
31.4.1 Sampling Usually sampling is treated implicitly, that is, it is considered to be a part of the detection step. However, the importance of obtaining a sample representing the true condition of the fluid, fluid contamination, and machine wear/contamination warrants treating sampling as a separate step in the condition monitoring procedure. The whole process of condition monitoring is based on the premise that the quality of sampling is acceptable and that samples provide a representative indication of the condition of the machine. Methodology of sampling varies depending on the type of monitored equipment. There are three basic methods of taking samples of a lubricant or a working fluid, see Figure 31.54: • Off-line (sampling from reservoir or fluid lines, taking grease samples). In this method a sample fluid is removed from the machine for further manual processing. • On-line (sampling from reservoir of fluid lines). In this method a sample of fluid is drawn from a fluid line or reservoir by the test instrument and automatically returned to the reservoir or fluid line after the test. On-line Particle Counters use this method of sampling. • In-line (fluid lines). In this method a sensor or instrument is placed in a fluid line and provides continuous measurement of parameter/variable of interest. Magnetic Plugs use this method of sampling.
© 2006 by Taylor & Francis Group, LLC
Contamination Control and Failure Analysis
Sample container
Off-line sample
31-49 Off-line sample
sensor Flow
Sample container sensor
In-line sample
Flow
Reservoir
sensor
On-line sample Flow
On-line sample (a) Fluid line
(b) Reservoir
FIGURE 31.54 Fluid sampling methods.
Methods of taking a sample from systems that use liquid lubricants will depend on the method of delivery of the lubricant to components that require lubrication. Systems using splash lubrication may be sampled off-line or on-line. All three methods can be used for sampling of liquid lubricants in recirculation lubrication systems and for sampling of working fluids. A sample should be taken at the time when concentration of debris in a system reached equilibrium for a given set of operational parameters. When a hydraulic or a recirculation lubrication system operates under steady wear and contamination conditions, concentration of the wear and contaminant particles in the fluid (number of particles per unit volume) achieves an equilibrium level for each given set of operational parameters. To achieve such an equilibrium level, since wear debris is continually generated in any operational mechanical system and contamination also always enters the system, the rate of particles removal from the fluid must be at the same rate at which they are generated or ingressed. Each machine has a characteristic operating time necessary to return to its normal equilibrium level. The equilibrium level will be upset when wear or contamination conditions change due to an increased rate of wear or contamination. Some factors that influence the time required to reach equilibrium level are listed below: • Filtration efficiency: Filtration efficiency is related, in the case of hydraulic filters, to their β-ratings. In general the better the filters (high β-ratings) the shorter the time to particle equilibrium. • Particles cycle rate: Particles cycle rate is equal to the ratio of pump flowrate (volume per unit time) to the volume of lubricant or hydraulic fluid in the system. The higher the cycle rate the more often the particles are exposed to the filter. The particle cycle rate, that is, the number of times a particle of a given size and material may, on the average, pass through the filter is affected by the size of the reservoir. • Dispersive qualities of the fluid: In systems where the fluid contains detergent additives to prevent agglomeration of the particles and to discourage their adherence to surfaces, particle life expectancy is increased. • Locations of slow moving fluid: If there are locations where the fluid moves slowly, the particles may settle down or adhere to the surfaces. Bottom of sumps, oil tanks, pipe bends etc. are examples of such locations. • Particulate loss: Known and suspected particle losses are the result of the following factors: (a) Placement of filters in the system (b) Settling rate of particles in the reservoir, in sharp bends, silting (c) Adherence of particles to solid surfaces (d) Breakdown of particles during repeated passage through tribological pairs (e.g. gears, bearings) (e) Oxidation (f) Chemical attack (g) Loss of lubricant or working fluid from the system
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
• Fluid change intervals: Systems where the oil change time is of the same order as the time taken to reach equilibrium are difficult to diagnose and are particularly prone to misinterpretation. This is specially so when diagnostic techniques give a warning on achieving certain absolute particle concentration levels. In these cases the particles concentration level should be recorded through the fluid life cycle and compared with a predetermined norm for that system. • Fluid addition: Following the addition of a lubricant or a hydraulic fluid to the system to replenish fluid loss, the concentration of wear debris and contaminant in the fluid will be lowered. Thus, concentration of debris in samples taken after addition of fluid should be adjusted to account for such changes. • Location of filter: In-line filters can profoundly modify the particle size distribution in the machine’s fluid. The filter changes the particle population in following ways: (a) A filter lowers the concentration of particles in the fluid. (b) The average particle that remains in the fluid of a filtered system is generated more recently than in the case of a system with no filter. (c) A filter removes large particles more effectively than the small ones so that the concentration of larger particles is reduced. Thus, to detect the presence of large particles that are indicative of severe wear and/or contamination in the system, samples should be taken from locations in the system upstream of the filter.
31.4.2 Detection The objective of predictive monitoring is to determine by analysis of a sample whether the condition of the machine has changed sufficiently enough over a period of time to indicate incipient failure of the machine and thus warrant the diagnostic step. After sampling, detection of changes in the condition of a machine is the second, most important step in condition monitoring procedure. For example, presence of silica in a sample is a symptom of dust contamination, presence of solid particles larger than 10 µm is a symptom of abnormal wear. Each measured or observed change in a characteristic of particulate or lubricant/working fluid and of any tribological variable or parameter can be considered a symptom of machine condition, see Figure 31.55. Usually a set of symptoms is required to reliably identify the condition of the machine. Effectiveness of detection is dependent on the selection of a set of symptoms that is representative of machine distress, availability of detection techniques to detect these symptoms, and Energy Inputs
Energy outputs
Machine
Energy losses
Material losses
Temperature
Noise
Performance
FIGURE 31.55
Condition monitoring symptoms.
© 2006 by Taylor & Francis Group, LLC
Strain
CM Techniques
Vibration
Wear Analysis
Contamination Control and Failure Analysis
31-51
availability of guidelines with which the symptoms are compared. In proactive monitoring the objective is to identify changes in basic parameters of fluid, for example, moisture level, contamination level, temperature, which are root causes of subsequent changes in the wear regime of the plant. Each change in a measured or observed root cause characteristic of lubricant/working fluid, or a system (e.g. increased temperature) or in contamination level can be considered a symptom of as operational and/or maintenance problem. 31.4.2.1 Symptoms A set of symptoms reliably describing root causes (proactive) and the state of the machine (predictive) should be determined after a systematic evaluation of the monitored equipment using failure analysis techniques, followed by analysis of tribological systems present in the machine. The above procedure leads to identification of pertinent characteristics of a system that should be monitored in order to detect any changes in root causes and in the health condition of a machine or a system. 31.4.2.2 Guidelines Detection of root causes, damage, failure, or malfunction is carried out by comparing results of analysis of characteristics of the particulates or fluid with predetermined criteria (guidelines). The three approaches to detection are shown in Figure 31.56. Establishing allowable absolute level and trend guidelines requires keeping of historic data on the past history of the machine. The levels/trends provided in literature are conservative and may be used only as starting points for setting detection limits. • Absolute level: An allowable absolute level is used to detect changes in the state of the machine by comparing the results of analysis with the allowable absolute level of system variable or parameter that defines the symptom. For example, the caution level may be set at 10 parts per million of Fe (ferrous material) for spectrometric oil analysis. Using absolute levels for detection purposes is not very satisfactory as the same machines operating under a different load condition or in a different environment may show different absolute values of symptoms. Levels are used when the detection technique provides a single number that indicates a value of some condition monitoring variable. Spectrometric Oil Analysis (SOA), Particle Counting, Direct Reading Ferrograph techniques are examples of detection techniques that provide this type of information.
Detection
By level
By trend
Level of x
Level of x Dx
Time
FIGURE 31.56 Detection methods.
© 2006 by Taylor & Francis Group, LLC
By comparison
Time
~ = Dt
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Handbook of Lubrication and Tribology
• Trend: Trending of analysis data over time and comparing it with an allowable trend offers a better indication of changes in the health condition of the machine and should be used in preference to detection based on levels. For example, for particle counting the caution level may be set at five particles larger than 10 µm per 10 h of operation. • Comparison: When a sample yields qualitative data, the detection process is based on comparison of qualitative properties of the sample with a reference model or pattern. Thus assessment of wear/contaminant debris is carried out by comparing shape, morphology, color of the debris with reference to particles whose origins are known. For example, presence of parallel striations on the surface of a particle may indicate the sliding mode of wear. Effective detection of changes in the machine’s condition is based on a number of factors listed below: (a) Representative samples. The samples should be representative of the condition of the machine. (b) Processing of samples. The procedure for processing of samples should be clearly laid down and strictly adhered to. (c) Reporting and data recording procedures. The procedures for recording and reporting of data should include relevant information about the machine, samples, and results of analysis. • Detection techniques. A suitable mix of detection techniques should be used to detect a relevant set of symptoms. 31.4.2.3 Results of Detection The result of the detection step is a list of measured/observed symptoms and assessment, based on the results of comparison with guidelines, of their severity. The severity of symptoms is usually classified as low, normal, caution, high, or critical.
31.4.3 Diagnosis Diagnosis is the task of identifying the current state of the machine by comparing the set of symptoms of the machine state identified in the detection step with a known syndrome that defines a certain state of the machine. For example, a partial set of symptoms to diagnose the condition of a gear box is listed in Table 31.6. Syndromes of machine state used for diagnosis may be identified by carrying out FMEA, FMECA, or other type of failure analysis. In addition, historical data and other documentary evidence of past maintenance actions should be maintained in order to develop a good data base for future use. Effective diagnosis is dependent on: • Knowledge of syndromes of known conditions of the machine. • A reliable and complete set of symptoms (identified during the detection step), for example, increased trend of wear debris quantity in successive samples. If a measured set of symptoms is a subset of a required set of symptoms then the reliability of diagnosis is diminished and might lead to errors in diagnosis. Diagnosis relying on only one symptom might be unreliable, and often such diagnosis will be wrong. There are, however, cases when a single symptom is dominating and it is the only one that is monitored. TABLE 31.6
Gear Box — Partial Set of Symptoms
Symptom Abnormal wear morphology of particles Solid particles larger than 10 µm Water contamination Dust contamination presence of silica Lubricant deterioration
© 2006 by Taylor & Francis Group, LLC
Method Ferrography, Filtergram Particle counting IR spectrometer AE spectrometer Total Acid Number
Strategy Predictive Proactive/predictive Proactive/predictive Proactive Proactive
Contamination Control and Failure Analysis
31-53
Even a more disastrous situation can, and often does, occur when the wrong condition variable (a symptom) is monitored as then no diagnosis is possible. For example, a gas turbine that was monitored using vibration technique obtained a clean bill of health shortly before the turbine failed catastrophically due to a bearing failure (hydrodynamic bearings were not monitored). 31.4.3.1 Results of Diagnosis The result of diagnosis is the identification of the mode, severity, location, and mechanism of wear or contamination and changes in chemical/physical properties of the lubricant/working fluid that caused the change in the machine’s health condition.
31.4.4 Prognosis The objective of the condition monitoring program is to be able to predict when the machine will need to be serviced or replaced. The prediction of the residual life of a machine is very dependent on availability of historical data on the past condition of the machine, history of failures of the machine, or of similar machines. The most difficult point in prognosis is to decide how long it should be operated before taking corrective action. To be able to make this type of decision the results of diagnosis should be correlated with the following data on the monitored machine or similar machines: • Records of past failures — records on monitoring, diagnosis, and corrective action taken • Records of reasons why this failures occurred — postmortem reports • Records of times lapsed between detection of conditions indicating failures and times when failures actually occurred In practice it is difficult to develop a data base for prognosis as population of machines that are critical to a given production process may be small, thus no sound statistical data can be obtained. 31.4.4.1 Approaches to Prognosis The prediction methods are still in the research stage of development. The usual way to predict residual life is to apply some type of curve fitting into data obtained from detection/diagnosis steps on the assumption that the failure will follow a trend discovered by monitoring. Techniques like Grey Theory, polynomial fitting, and more advanced artificial intelligence techniques (neural net and fuzzy logic) are being applied to provide recognition of the pattern of failures and prediction of future trends in the progression of machine failure. 31.4.4.2 Results of Prognosis The prognosis step should ideally result in an estimate of how much time is left before the machine will fail in service.
31.4.5 Action Maintenance action (replacement, repair, etc.) required to bring the machine to an acceptable operational state is dictated by the results of diagnosis and prognosis.
31.4.6 Postmortem In the event of machine failure, postmortem examination should be carried out to determine the causes of the failure, appearance of failed components, etc. This should be related to the results of detection, diagnosis, and prognosis to provide a historical record for future use. It must be stressed that reliability of diagnosis and prognosis of machine condition monitoring will greatly increase if such examinations are carried out and are properly recorded.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
In most companies the postmortem on failed machines is seldom carried out routinely, and even when it is carried out the records are often not complete. When the problem is fixed there is generally another more urgent job than completing the postmortem report.
31.5 Training Implementation of contamination control requires thorough understanding of the aims of the program, knowledge of a system, and fluid monitoring technology. Contamination control should become an important component of quality improvement process in the majority of industrial, aviation, mining, metalworking companies that use fluids for power/control. Contamination control should be incorporated in a Company Quality Improvement program which aims to increase reliability, productivity, and profitability of the plant. A successful implementation of contamination control program is, however, conditional on the availability of clearly defined procedures, well-trained personnel, and monitoring equipment. Training of personnel should include three groups of personnel: • Management: To be successful the program requires involvement and acceptance by all personnel involved in design, manufacturing, and operation/maintenance of the system and this must be fully supported but not imposed by management. Management must fully understand that wellimplemented contamination control will be profitable if all essential components of the program are in place. Management must set life cycle improvement targets that must be achieved using contamination control and then must provide necessary resources — manpower, equipment, training, etc. The life improvement obtained via contamination control should be checked against targets and the program adjusted if necessary. Management must understand that a contamination control program is based on proactive rather then reactive approach, that is, it is aimed at avoiding failures rather than at catching failures. Furthermore management must provide a working environment in which a contamination control program can be successful. This may require structural, functional reorganization of the workplace, change of maintenance strategy and acceptance of new ideas, new technology and new attitudes. Management training should include the fundamentals of a contamination control program, impact on other quality improvement programs, profitability assessment, strategy, organization and resources needed. • Supervisory staff: Engineering staff must be trained to understand the contamination control approach. Implementation of contamination control should not be the responsibility of only the maintenance personnel as staff dealing with purchasing, design, manufacturing, commissioning, and operation can have pronounced influence on the success of the program (e.g. equipment purchased or designed without consideration how it will fit into the plant’s contamination control). Basic training of engineering staff associated with a contamination control program should include, in addition to basic contamination control training exposure to the following: (a) Tribology (lubrication, wear, hydraulic and lubricating fluids) (b) Hydraulic power and control systems (principle of operation, troubleshooting) (c) Functional and Failure analysis (FMEA, FMECA, mechanical failures analysis) (d) Tribology based monitoring and diagnostics (wear debris analysis, monitoring techniques, sampling, interpretation, etc.) (e) Condition monitoring (predictive and proactive) • Operators and maintenance staff: Operators and maintenance personnel are best placed to detect any changes in operating regimes of the machines — the most important first step in both predictive and proactive monitoring. Maintenance staff dealing directly with all aspects of contamination control for example, sampling, filter changing, servicing of equipment must be trained in the basics of contamination control so that they can perform their work knowing what they are trying to achieve. They should also be equipped with tools that will facilitate detection without waiting for lab results (e.g. filtergrams, portable contamination comparators) and thus reacting to changes in a very short time. Training should include instructions on use and capabilities of various monitoring
© 2006 by Taylor & Francis Group, LLC
Contamination Control and Failure Analysis
31-55
tools, sampling, sensory detection. Basic understanding of lubrication and tribological processes will be an advantage in the execution of their tasks.
31.6 Conclusions We must recognize contamination control as a vital factor in operation of high-performance systems. Contamination control is more than just ordering a filter; we must understand various factors that are involved in assuring proper contamination control. A contamination control program integrates our knowledge of contamination control and its effect on quality improvement programs. To be successful, contamination control program requires management and personnel commitment to quality improvement and understanding that its success is based on clearly defined goals, staff education, selection of right technology, and an efficient management system.
References [1] “ISO 16889:1999 Hydraulic Fluid Power Filters — Multi-Pass Method for Evaluating Filtration Performance of a Filter Element”, ISO, 1999. [2] “ISO 4406:1999 Hydraulic Fluid Power — Fluids — Method for Coding the Level of Contamination by Solid Particles”, ISO, 1999. [3] “NAS 1638 Cleanliness Requirements of Parts used in Hydraulic Systems. National Aerospace Standard”, NAS, 1964. [4] E.C. Fitch, An Encyclopaedia of Fluid Contamination Control for Hydraulic Systems, Hemisphere Publishing Corp., Stillwater, USA, 1979. [5] L. Chao, A System Approach to Monitoring and Controlling of Contamination/Wear in Hydraulic Control Systems, PhD thesis, Monash University, Australia, 1995. [6] L. Chao and B.T. Kuhnell, “The Comparison and Development of On-line Wear & Contamination Monitors”, The Bulletin of Centre for Machine Condition Monitoring, Vol. 4, No.1, pp. 20.1–20.8, 1992. [7] C.H. Pek and J.S. Stecki, “On-Line Contamination Monitoring Using Nephelometry”, Proceedings of the 2nd International Machinery Monitoring and Diagnostic Conference, pp. 121–125, Los Angeles, 1990. [8] T. Hong and J.C. Fitch, “Model of Fuzzy Logic Expert System for Real-Time Condition Control of a Hydraulic System”, Proceedings of the 3rd International Conference on Machine Condition Monitoring, Windsor, U.K., 1990. [9] L. Chao and J.S. Stecki, “Intelligent Expert System for Contamination Control in Hydraulic Control Systems”, Proceedings of 6th International Conference on Industrial and Engineering Applications of Artificial Intelligence and Expert Systems, pp. 538–543, Edinburgh, June 1993. [10] J.S. Stecki and L. Chao, “Computer-Aided Techniques in Contamination Control, Design and Performance”, Eighth Bath International Fluid Power Workshop, pp. 126–140, Bath, UK, Research Studies Press & John Wiley and Sons, ISBN 0 471 96128 0, 1996. [11] G.J. Schoenau, J.S. Stecki, and R.T. Burton, “Utilization of Artificial Neural Networks in the Control, Identification and Condition Monitoring of Hydraulic Systems — An Overview”, SAE 2000 Transactions, Vol. 109, Journal of Commercial Vehicles, Section 2, pp. 205–212, 2000. [12] J.S. Stecki and G. Schoenau, “Application of Simulation and Knowledge Processing in Contamination Control”, SAE 2000 Transactions, Vol. 109, Journal of Commercial Vehicles, Section 2, pp. 331–347, 2000. [13] “MIL-STD-1629A, The Procedures for Performing a Failure Mode, Effects and Criticality”, DOD, 1984. [14] “SAE J 1739: Potential Failure Mode and Effects Analysis (FMECA) Reference Manual”, SAE, 1994.
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[15] “ECSS-Q-30-02A Space Product Assurance — Failure Modes, Effects and Criticality Analysis (FMECA)”, European Cooperation for Space Standardization, 2001. [16] “BS5760:1991: Reliability of Systems, Equipment and Components, Part 5, Guide to Failure Modes, Effects and Criticality Analysis (FMEA and FMECA)”, BS, 1991. [17] D.R. Bull, W.J. Crowther, C.R. Burrows, K.A. Edge, et al., “Approaches to Automated FMEA of Hydraulic Systems”, Proceedings of I. Mech. Congress Aerotech95, Seminar C505/9/099, Birmingham, UK, 1995. [18] W.J. Crowther, C.R. Burrows, K.A. Edge, et al., “Automated FMEA for Hydraulic Systems”, Proceedings of 12th Fluid Power Technology Colloquium, IFAS, Aachen, Germany, 1996. [19] P.G. Hawkins, R.M. Atkinson, et al., “An Approach to Failure Modes and Effects Analysis Using Multiple Models”, IFMA Proceedings, V4, Athens, Greece, June 1996. [20] D.R.Bull, J.S. Stecki, K.A. Edge, and C.R. Burrows, “Failure Modes and Effects Analysis of a ValveControlled Hydrostatic Drive”, in Fluid Power Engineering: Challenges and Solutions, Tenth Bath International Fluid Power Workshop, Research Studies Press Ltd., pp. 131–143, 1997. [21] C.E. Peleaz and J.B. Bowles, “Using Fuzzy Cognitive Maps as a Model for Failure Modes and Effects Analysis”, Information Sciences, Vol. 88, pp. 177–199, 1996. [22] J.S. Stecki, F. Conrad, and B. Oh, “Software Tool for Automated Failure Modes and Effects Analysis (FMEA) of Hydraulic Systems”, Fifth JFPS International Symposium on Fluid Power, Nara, Japan, 2002. [23] J.S. Stecki, “Maintenance Aware Design (MAD) Methodology”, The Eighth Scandinavian International Conference on Fluid Power, SICFP’03, May 7–9, Tampere, Finland, 2003. [24] J.S. Stecki and B. Oh, “Failure Analysis Using Energy Disturbance and IA Techniques”, The Eighth Scandinavian International Conference on Fluid Power, SICFP’03, May 7–9, Tampere, Finland, 2003. [25] G. Pahl and W. Beitz, Engineering Design: A Systematic Approach, Springer-Verlag, Berlin, 1988. [26] L.D. Miles, Techniques of Value Analysis and Engineering, McGraw-Hill, New York, 1972. [27] “ECSS-E-10 Space Engineering — Functional Analysis”, European Cooperation for Space Standardization, 1999. [28] H. Czichos, Tribology. A System Approach to the Science and Technology of Friction, Lubrication and Wear, Tribology Series 1, Elsevier, New York, 1978. [29] D. Godfrey, Wear Control Handbook, ASME Publication, 1980. [30] R.K. Tessman and E.C. Fitch, “Field Contaminant Ingression Rates — How Much?” Proceedings of Eighth Annual Fluid Power Research Conference, paper P74-47, October 1974. [31] L.E. Bensch and W.T. Bonner,“Field System Contaminants — Where, What, How Much”, Proceedings of Seventh Annual Fluid Power Research Conference, paper P73-CC-1, pp. 187–193, October 1973. [32] W. Backé, “Wear Sensitivity of Hydraulic Displacement Units to Solid Contaminats”, Ölhydraulik und Pneumatik, Vol. 33, No. 6, pp. 29–38, 1989. [33] R.K. Tessman, “Speed is not a Factor in the Contaminant Wear of Hydraulic Pumps”, BFPR J., paper 8, pp. 1–11, 1979. [34] G. Silva, “Wear Generation in Hydraulic Pumps”, SAE Technical Paper, 9011679, 1990. [35] R.K. Tessman, “Component Contaminant Generation”, Proceedings of Eighth Annual Fluid Power Research Conference, paper P74-46, October 1974. [36] J. Collins and B. Hagan, The Failure-Experience Matrix: A Useful Design Tool, ASME Journal of Engineering for Industry, 1976. [37] J. Collins, Failure of Materials in Mechanical Design: Analysis, Prediction, Prevention. Wiley Interscience, 1993. [38] B.T. Kuhnell and J.S. Stecki, “Monitoring of Hydraulic System”, Proceedings of 2nd International Fluid Power Conference, pp. 729–734, Hanghzou, China, March 1989. [39] L. Chao, J.S. Stecki, and P. Dransfield, “Dynamic Contamination Control Model for Fluid Contamination”, Proceedings of International Conference on Fluid Power Control and Robotics, pp. 342–247, Chengdu, China, 1990.
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[40] J.S. Stecki and L. Chao, “Simulation of Contamination Control Systems”, The Bulletin of Centre for Machine Condition Monitoring, Vol. 2, No. 1, pp. 10.1–10.6, 1990. [41] J.S. Stecki and M. Grahl-Madsen, “Distribution of Particulate Contamination in Multi-branch Hydraulic System”, The Sixth Scandinavian International Conference on Fluid Power, SICFP’99, May 26–28, Tampere, Finland, 1999.
Appendix Standards and Drafts of ISO/TC 131/SC 6 Contamination Control and Hydraulic Fluids • ISO 2941:1974 Hydraulic fluid power — Filter elements — Verification of collapse/burst resistance. • ISO/CD 2941 Hydraulic fluid power — Filter elements — Verification of collapse/burst pressure rating. • ISO 2942:1994 Hydraulic fluid power — Filter elements — Verification of fabrication integrity and determination of the first bubble point. • ISO/CD 2942 Hydraulic fluid power — Filter elements — Verification of fabrication integrity and determination of the first bubble point. • ISO 2943:1998 Hydraulic fluid power — Filter elements — Verification of material compatibility with fluids. • ISO 3722:1976 Hydraulic fluid power — Fluid sample containers — Qualifying and controlling cleaning methods. • ISO 3723:1976 Hydraulic fluid power — Filter elements — Method for end load test. • ISO/CD 3724 Hydraulic fluid power — Filter elements — Verification of flow fatigue characteristics. • ISO 3724:1976 Hydraulic fluid power — Filter elements — Verification of flow fatigue characteristics. • ISO 3938:1986 Hydraulic fluid power — Contamination analysis — Method for reporting analysis data. • ISO 3968:1981 Hydraulic fluid power — Filters — Evaluation of pressure drop vs. flow characteristics. • ISO/FDIS 3968 Hydraulic fluid power — Filters — Evaluation of pressure drop vs. flow characteristics. • ISO 4021:1992 Hydraulic fluid power — Particulate contamination analysis — Extraction of fluid samples from lines of an operating system. • ISO/AWI 4405 Hydraulic fluid power — Fluid contamination — Determination of particulate contamination by the gravimetric method. • ISO 4405:1991 Hydraulic fluid power — Fluid contamination — Determination of particulate contamination by the gravimetric method. • ISO 4406:1999 Hydraulic fluid power — Fluids — Method for coding the level of contamination by solid particles. • ISO 4407:1991 Hydraulic fluid power — Fluid contamination — Determination of particulate contamination by the counting method using a microscope. • ISO/DIS 4407 Hydraulic fluid power — Fluid contamination — Determination of particulate contamination by the counting method using an optical microscope. • ISO/WD 7744 Hydraulic fluid power — Filters — Statement of requirements. • ISO 7744:1986 Hydraulic fluid power — Filters — Statement of requirements. • ISO/CD 10949 Hydraulic fluid power — Component cleanliness — Methods of achieving, evaluating, and controlling component cleanliness from manufacture through installation. • ISO/TR 10949:1996 Hydraulic fluid power — Methods for cleaning and for assessing the cleanliness level of components.
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• ISO/CD 11170 Hydraulic fluid power — Filter elements — Procedure for verifying performance characteristics. • ISO 11170:1995 Hydraulic fluid power — Filter elements — Procedure for verifying performance characteristics. • ISO 11171:1999 Hydraulic fluid power — Calibration of automatic particle counters for liquids. • ISO 11500:1997 Hydraulic fluid power — Determination of particulate contamination by automatic counting using the light extinction principle. • ISO/WD 11500 Hydraulic fluid power — Determination of particulate contamination by automatic counting using the light extinction principle. • ISO 11500:1997/Cor 1:1998 No title. • ISO 11943:1999 Hydraulic fluid power — On-line automatic particle-counting systems for liquids — Methods of calibration and validation. • ISO/CD 16009 Hydraulic fluid power — Filters — Method for determining filtration performance of an off-line filter element. • ISO/CD TR 16144 Hydraulic fluid power — Calibration of liquid automatic particle counters — Procedures used to certify the standard reference material. • ISO/TR 16386:1999 Impact of changes in ISO fluid power particle counting — Contamination control and filter test standards. • ISO/CD TS 16431 Hydraulic fluid power — Assembled systems — Method for verifying cleanliness. • ISO/CD 16860 Hydraulic fluid power — Filters — Method of test for differential pressure devices. • ISO/DIS 18413 Hydraulic fluid power — Cleanliness of parts and components — Inspection document and principles related to sample collection, sample analysis, and data reporting.
Further Reading 1. J.S. Stecki and L. Chao., “Selection of Filters for Hydraulic Control Systems”, Proceedings of International Conference on Maintenance Management, paper 27, Melbourne, September 1991. 2. L. Chao, P. Dransfield, and J.S. Stecki,“PCEX — Expert System for Evaluating Pump Contamination Sensitivity”, Proceedings of 2nd JHPS International Symposium on Fluid Power, pp. 527–532, Tokyo, September 1993. 3. L. Chao and J.S. Stecki, “Knowledge-Base for Evaluation of Pump Wear Tolerance”, The Bulletin of Centre for Machine Condition Monitoring, Vol. 5, pp. 130–135, November 1993. 4. L. Chao, P. Dransfield, and J.S. Stecki, “Expert System Aided Filtration System Design”, Supplement Proceedings of 3rd (ICFP) International Conference on Fluid Transmission and Control, pp. 6–11, Hangzhou, China, September 1993. 5. J.S. Stecki (ed), Workshop on Advances in Hydraulic Control Systems, Monash University, ISBN 7326 0987 5, p. 150, Melbourne, 1996. 6. J.S. Stecki (ed), Workshop on Total Contamination Control, Fluid Power Net Publications, ISBN 0-646-35306-3, p. 177, Melbourne, 1998.
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32 Environmental Implications and Sustainability Concepts for Lubricants 32.1 The Environmental Drivers . . . . . . . . . . . . . . . . . . . . . . . . . .
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Concepts of Sustainability and Sustainable Development • Sustainable Development at Global, Continental, and National Levels • Environmental Management Systems (EMS)
Responsible Care . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . The Valdez Principles . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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Life-Cycle Assessment, LCA
A Case Study of Life Cycle Analysis . . . . . . . . . . . . . . . . . . . . . . . . . . 32.2 Environmental Implications of Lubricant and Hydraulic Fluids Production . . . . . . . . . . . . . . . . . . . . . . . . . .
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The Size of the Problem • Life Cycle Assessment • Quantifying Lubricant Chemistry Production
32.3 Environmental Benefits of Lubricant and Hydraulic Formulations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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Overview of Lubricant Use
32.4 Lubricant and Hydraulic Fluids as Wastes . . . . . . . . . . .
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Definitions of Waste • Types of Waste • Development of a U.K. National Waste Strategy, “Making Waste Work” • Introduction to the “Waste Strategy 2000” • Waste Reduction • Waste Recovery • Used/Waste Lubricants as Fuels • Incineration • Waste Disposal to Landfill • Directions in the Disposal of Used Lubricants
32.5 Environmental Pollution by Used Lubricants . . . . . . .
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The Polluting Effects of Used Lubricant • The Biodegradability of Lubricants
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32.6 The Environmental Future and Lubricants . . . . . . . . . .
Malcolm F. Fox De Montfort University
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Annual Statistics and Used Arisings in the United Kingdom • Trends in Waste Lubricant Arisings • The Environmental Business Economics of Recycling Lubricants • Barriers to Acceptance of Recycled Base Oils • Relative Costs • Conclusions
References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
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32.1 The Environmental Drivers Lubricants are commercial consumer products and are now subject to analyses, as is the case for any other product, regarding their impact on the environment over a new concept, their “life cycle.” The effects of producing and using lubricants in society are wide-ranging, which commences from: • • • •
Their formulation, through to Their production, from main base oil and additive component parts Their beneficial effects during use Concluding with the effects of their eventual disposal
The environmental effects of lubricants can be beneficial or negative. But first, it is helpful to examine the concepts and management techniques that are used to assess the importance of these issues. The framework has changed and the issues of environmental impact and resource sustainability for lubricant production, use, and disposal have risen, and are rising further, in the public domain. These issues may not be popular, but cannot be ignored. Lubricants must be used more effectively with longer lifetimes for increased sustainability. Their production and use is controlled for environmental reasons, equally their subsequent disposal after use is now strictly controlled and should be recycled. It is no longer acceptable to tip used lubricants onto the ground, into watercourses, or to burn them without specialized furnaces. That is the nature of the environmental impact of the end use of lubricants, their sustainability, and sustainable development.
32.1.1 Concepts of Sustainability and Sustainable Development Ecological and economic systems are now widely recognized as being interlinked and that unrestrained development and use of resources leads to serious resource depletion, residual pollution, and survival problems. It follows that economic development should meet the test of “sustainability.” The concept of “sustainable development” is the key principle of environmental policy formulation and development and was introduced by the 1987 Report of the World Commission [1] on Environment and Development, “Our Common Future.” Chaired by the ex-Prime Minister of Norway, Gro Haarlem Brundtland, it is known as the “Brundtland Report” and defined sustainable development as: “development which meets the needs of the current generation without compromising the ability of future generations to meet their own needs.” which contains the concepts of: • Intergenerational equity — we should pass on to the next generation at least an equivalent resource endowment to allow them to meet their needs. • Intragenerational equity — that poverty should be eradicated so that differences in living standards are reduced to meet the needs of the majority of people in the world. • Carrying capacity — that ecosystems are not stretched beyond their capacity to either renew themselves or absorb wastes. The concept of sustainable development therefore includes issues of social progress, economic growth, and environmental protection. These concepts may appear to be abstruse and theoretical but are gradually
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entering legislation at continental and national levels in the form of detailed environmental law and regulations. These, in turn, become technical requirements and standards and the next section demonstrates how environmental policies are transmitted downwards into targets and actions, for lubricants as for other products. This explains the background for the changes in policies and standards, which increasingly guide the production, use, and disposal of lubricants.
32.1.2 Sustainable Development at Global, Continental, and National Levels 32.1.2.1 At the Global Level The concept of sustainable development has developed over the last 30 to 40 years and its significant markers along the way commence in 1972 at a United Nations conference in Stockholm, Sweden: 1972 — The United Nations Conference on the Human Environment (UNCHE) [2] in Stockholm, Sweden, considered the need for a common outlook and principles to inspire and guide the world’s peoples to preserve and enhance the human environment. 1987 — The World Commission on Environment and Development publishes The Brundtland Report [1] bringing concepts of sustainable development onto the international agenda. It also provides the most commonly used definition of sustainable development. 1992 — 180 countries met at the “Earth Summit” (the UN Conference on Environment and Development) in Rio de Janeiro, Brazil [3] to discuss how to achieve sustainable development. The Summit agreed the Rio Declaration on Environment and Development, setting out 27 principles supporting sustainable development. It agreed a plan of action, Agenda 21, and a recommendation that all countries should produce national sustainable development strategies. The 1992 Earth Summit also established the Framework Convention on Climate Change and the Convention on Biological Diversity. 1997 — A special UN conference [4] reviewed implementation of Agenda 21 (as Rio+5), repeating the call for all countries to have sustainable development strategies in place. 2002 — Ten years after the Earth Summit in Rio in 1992, the 2002 Johannesburg Summit [5] in South Africa, the World Summit on Sustainable Development, brought together many participants including heads of State and Government, national delegates and leaders from nongovernmental organizations (NGOs), businesses, and other major groups to focus the world’s attention and direct action toward meeting difficult challenges. These include improving people’s lives and conserving natural resources in a world that is growing in population with ever-increasing demands for food, water, shelter, sanitation, energy, health services, and economic security. These global policies set the pattern for policies at continental and national levels. 32.1.2.2 At the Continental Level Whilst the world’s continents are legally autonomous and the implementation of global commitments may appear to be different due to differences in legal systems, the overall thrusts of environmental legislation are broadly the same. This section examines the environmental actions of the European Union, then how that has devolved down to the national level, as in the United Kingdom. The European Union (EU) and its European Environment Agency, EEA, is recognized as one of the most important international innovators for environmental issues, standards, and enforcement, comparable to the United States and its Environmental Protection Agency. There are over 600 EU legislative measures that apply to the environment, although some are either revised or have amended previous legislation. The earliest EU legislation for the protection of the environment commenced in 1970 and applied to vehicle emissions. This was the start of a program closely analogous to vehicle emission control programs in North America and Japan and illustrates the force of international environmental action. The First EU Action Program on the Environment was adopted in 1973 by the six original member nations of the EU. Gradual expansion of the EU over subsequent years led to the current (2004) membership of 25 nations. The Single European Act of 1986 amended the original Treaty of Rome of 1957 to explicitly
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include measures to protect the environment. Since then, EU Environmental Action Programs have been increasingly aimed at pro-actively protecting the environment: 1993 — “Towards Sustainability,” the Fifth Environmental Action Programme of the EU [6] sought to integrate environmental concerns into other policy areas to achieve sustainable development. It recognized that legislation cannot be used to solve all environmental problems and that other mechanisms are needed to achieve sustainable development, including market-based instruments such as environmental management systems, eco-audits, and eco-labeling. 1999 — changes to the Treaty of Rome were made by the Treaty of Amsterdam [7] to give sustainable development a much greater prominence within the EU. 2001 — The EU’s Sixth Environmental Action Programme (2001–2010) entitled “Environment 2010: Our Future, Our Choice [8],” takes a wide-ranging approach to these challenges and gives a strategic direction to environmental policy over the next decade as the EU prepared to expand its number of member nation states. The new program identifies four priority areas as: 1. Climate Change 2. Nature and Biodiversity 3. Environment and Health 4. Natural Resources and Waste The new Program set out five approaches to achieve improvements in each of these areas, emphasizing that more effective implementation and innovative solutions were needed. The EU Commission recognized that a wider constituency must be addressed, including businesses, who can only gain from a successful environmental policy. The five key approaches are to: 1. 2. 3. 4. 5.
Ensure the implementation of existing environmental legislation Integrate environmental concerns into all relevant policy areas Work closely with consumers to identify solutions Ensure better, more accessible environmental information for citizens Develop a more environmentally conscious attitude toward land use
32.1.2.3 Sustainable Development at the National Level — the United Kingdom In its response to the EU Directives, the U.K. Government published its strategy for sustainable development in 1999, “A Better Quality of Life” [9], replacing a previous 1994 strategy document “Sustainable Development — the UK Strategy.” The 1999 Strategy highlights the need for integrated policies to achieve a better quality of life for current and future generations by ensuring social progress that recognizes: 1. 2. 3. 4.
The needs of every person Effective environmental protection The prudent use of natural resources Maintenance of high and stable levels of economic growth and employment
The U.K. Local Government Act 2000 [10] requires local authorities (government, such as councils at different levels) in England and Wales to prepare a community strategy for promoting, or improving, the economic, social, and environmental well-being of their community. This should also contribute to the achievement of sustainable development of the United Kingdom. The concept of ’sustainable development’ is set out at some length to emphasize that it is real, it is a reality now, and will become even more important as a decision parameter to include in business policies in the future. 32.1.2.4 Example — The Reduction of Diesel Fuel Sulfur Content in Europe An example of the impact of the changes driven by environmental policies at the continental level is the sulfur content of diesel fuel. This is now the major source of urban sulfur dioxide from diesel vehicle vehicles, not because its source has dramatically increased but because other sources of sulfur dioxide from coal and heavy oil heating fuels have sharply decreased over the last two decades by replacement with natural gas. The successive Clean Air Acts of the United Kingdom and the increased use of natural gas
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have ousted higher sulfur liquid and solid fuels from space heating. Diesel fuel itself has seen a dramatic reduction of sulfur content in the United Kingdom over the last 20 years from 0.3% (or 3000 ppm) to a current (2004) level of ∼40/45 ppm in the United Kingdom with projected reductions to a maximum of 10 ppm sulfur content by January 2009 [11]. The driver for these reductions in fuel sulfur content arises from international treaty obligations entered into by Western nations. Good as that is for the (urban) environment particularly, it has very serious implications for diesel fuel pump wear. The sulfur compounds at high concentrations in diesel fuel act as inherent lubricants for the injector pump. Their removal can be effectively countered by a program of fuel lubricity additives, which is being successfully implemented in Europe. But there are serious variations across the regions of the United States, with East Coast U.S. diesel fuel being comparable to EU quality whereas for the rest of the U.S. regions, the High Frequency Reciprocating Rig, the HFRR [12] test, shows much higher injector pump wear rates [13]. 32.1.2.5 Environmental Management Systems (EMS), and Life Cycle Assessment (LCA) There are many different analytical tools that businesses can use to evaluate and improve their environmental performance and to then demonstrate those improvements. It is worth examining two of them, Environmental Management Systems (EMS) and Life Cycle Assessment (LCA).
32.1.3 Environmental Management Systems (EMS) The very visible and tangible environmental impact of some industries in the last few decades and the increasing pool of sophisticated knowledge amongst both professionals and the public, led to increased pressure on organizations to become more “environmentally friendly.” These pressures include questions from shareholders at annual meetings and insurers who are increasingly concerned about contingent liabilities. The contingent liabilities from asbestos injury claims in the insurance industry are a stark reminder. The introduction of an EMS is one of the best approaches to building environmental protection into an organization. It can be effectively applied to a whole organization and used to ensure consistency of action. Just as quality systems were developed during the 1980s to aid manufacturing efficiency, production, and competitiveness, EMS guidelines and standards were developed to enable organizations to become less wasteful, more profitable, and less detrimental to the environment. The procedures for EMS analyses, Quality Control procedures, and Health and Safety programs are very similar and can be integrated into an overall program, leading to the aphorism that “— a quality production company is a safe company, is an environmentally responsible company which is also a profitable company” and experience shows this to be broadly true. There are several different standards to which an organization’s EMS can be certified, the two most important being: • EMAS, the European Union’s Eco-Management and Audit Scheme • ISO 14001, the standard produced by the International organization for Standardization Another industry-wide EMS standard that is important for the chemical, petrochemical, and polymer industries is “Responsible Care.” This system originated in Canada as a specific system for the chemical industry and was introduced to other countries, including the United Kingdom in 1989 by the Chemical Industries Association (http://www.cia.org.uk/industry/care.htm) [14]. “Responsible Care” incorporates health and safety as well as environmental management and its main principles are set out below. The issue is whether this is a real, continuing, commitment or a panacea:
Responsible Care “Responsible Care is a voluntary scheme under which organisations commit to provide a high level of protection for the health and safety of employees and associates, customers, and the public; and
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for the environment. organisations make a commitment to sustainable development and continual improvement by promising to adhere to the following principles: • Policy: We will have a Health, Safety and Environmental (HS&E) policy which will reflect our commitment and be an integral part of our overall business policy. • Employee involvement: We recognise that the involvement and commitment of our employees and associates will be essential to the achievement of our objectives. We will adopt communication and training programmes aimed at achieving that involvement and commitment. • Experience sharing: In addition to ensuring our activities meet the relevant statutory obligations, we will share experience with our industry colleagues and seek to learn from and incorporate best practice into our own activities. • Legislators and regulators: We will seek to work in co-operation with legislators and regulators. • Process safety: We will assess and manage the risks associated with our processes. • Product stewardship: We will assess the risks associated with our products, and seek to ensure these risks are properly managed throughout the supply chain through stewardship programmes involving our customers, suppliers and distributors. • Resource conservation: We will work to conserve resources and reduce waste in all our activities. • Stakeholder engagement: We will monitor our HS&E performance and report progress to stakeholders; we will listen to the appropriate communities and engage them in dialogue about our activities and our products. • Management systems: We will maintain documented management systems which are consistent with the principles of Responsible Care and which will be subject to a formal verification procedure. • Past, present and future: Our Responsible Care management systems will address the impact of both current and past activities.” Whilst EMSs are voluntary schemes they are also market-based tools designed to internalize environmental costs. They demonstrate to producers and consumers the need to use natural resources responsibly and minimize pollution and waste. EMSs introduce market forces in the environmental field by promoting competition between industrial activities on environmental grounds. The assumption is that the market rewards organizations that establish EMSs with greater market share and, consequently, market pressures encourage others to set up their schemes. The overall outcome should be that more organizations become active in environmental management and their environmental performance will improve. EMSs also tend to spread down through the “supply chain” — once an organization has introduced its own EMS, it starts to look at its suppliers and their environmental performance as well. The advantages of an EMS include: • • • • • • • •
Improved environmental performance Improved public relations Improved relations with regulators Control of environmental legislation requirements Formalizing/coordinating existing systems Raises environmental awareness within the organization Improves relationships with local communities Effective, nonbureaucratic documentation system
*competitive advantage *marketing tool *cost savings *cheaper insurance *pro-active approach
The disadvantages of an EMS can include: • Exposing environmental liabilities (advantageous if forestalling regulatory action), long term commitment, including resources • Implementation and certification costs • Increased workload initially, which then decreases as not being dependent on the fire-fighting approach for tackling environmental emergencies • Increased training requirements (which occur anyway in an enlightened company)
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Commitment.
Initial Review.
Policy. Organisation and personnel
Review
Audits
Record
The stages form part of the standard specification as assessable parts of the overall system
Evaluation and register of environmental aspects
Operational control
Register of Regulations
Objectives and targets Management manual
Management programme
FIGURE 32.1 Establishing an environmental management system (EMS).
• Uncontrolled environmental information provided to the public may not always be advantageous as it highlights areas that an organization might not want in the public domain. Once released, there is no control over what the public or media do with it. But there is a general move toward a requirement for transparency of information, including environmental information and companies will have to learn to live with this approach. EMS standards do not outline the expected environmental performance of an organization, although they do require the relevant legislation and regulations to be adhered to. Instead, compliance is centered on the ability of the organization to meet its own stated objectives. These objectives are expected to change and be modified over time as the organization strives for continual improvement. The basic stages of establishing an EMS are shown in Figure 32.1. An ISO 14001 analysis requires that an Environmental Statement is published following an initial environmental review and also after the completion of each subsequent audit cycle. It should be designed for the public and written in a concise, comprehensible form. Many companies, especially large companies with high public profiles, are keen to show their environmental credentials. Their insurers increasingly require an EMS to be in place so as to limit and define the company’s environmental contingent liabilities, both immediate and long term. There are also The Valdez Principles, produced after the oil spill from the Exxon Valdez tanker in Alaska in March 1989, a comprehensive and exacting list of principles that can provide a useful starting point for developing a company environmental policy.
The Valdez Principles 1. Protection of the Biosphere — minimize releases of any pollutant that cause environmental damage to air, water, or earth. We will safeguard habitats in rivers, lakes, wetlands, coastal zones and oceans and avoid contributing to greenhouse effects, depletion of the ozone layer, acid rain, or smog.
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2. Sustainable Use of Natural Resources — make sustainable use of renewable natural resources like water, soils, and forests. We will conserve non-renewable natural resources through efficient use and careful planning. We will protect wildlife, habitat, open spaces and wilderness, while preserving biodiversity. 3. Reduction and Disposal of Waste — minimize waste, especially hazardous waste, and whenever possible, recycle materials. We will dispose of all waste through safe and responsible methods. 4. The Wise Use of Energy — use environmentally safe and sustainable energy to meet our needs. We will invest in improved energy efficiency of products we use or sell. 5. Risk Reduction — minimize environmental, health, and safety risks to employees and communities where we operate by employing safe technologies and operating procedures and by being constantly prepared for emergencies. 6. Marketing of Safe Products and Services — to sell products that minimize environmental impacts and are safe as consumers commonly use them. We will inform consumers of the environmental impacts of our products or services. 7. Damage Compensation — take responsibility for any harm we cause to the environment by making every effort to restore the environment and to compensate those persons who are adversely affected. 8. Disclosure — disclosure to our employees and to the public, incidents relating to our operations that cause environmental harm or pose health or safety hazards. We will disclose potential environmental health or safety hazards posed by our operations. 9. Environmental Directors and Managers — At least one seat on our Board of Directors is designated for an environmental advocate. We will commit management resources to implement these principles, including the funding of an office of Vice President for Environmental Affairs or an equivalent executive position to monitor and report on our implementation efforts. 10. Assessment and Annual Audit — We will conduct and make public an annual self-evaluation of our progress in implementing these principles and in complying with all applicable laws and regulations. We will work towards the timely creation of independent environmental audit procedures to which we will adhere.
32.1.4 Life-Cycle Assessment, LCA Consumers are increasingly concerned about the environmental impact of products and services. But it is often not straightforward to determine which is the best environmental option because the problem becomes more complicated as it is realized that products, services, and processes have impacts throughout their life cycles, from the “cradle to the grave.” Producers are increasingly aware that they must take life cycle responsibility for their products and services. Both producers and consumers need a reliable, credible, and recognized method of assessing environmental impacts for different options and determining which is “best.” Life Cycle Assessment (LCA) is an objective process that is used to evaluate the environmental burdens or impacts associated with a product, process or activity. The procedure should (i) identify and quantify the inputs/outputs, for example, the energy, the materials used, and the wastes released to the environment, (ii) assess the environmental impacts of the energy and materials used, and the wastes released, and (iii) evaluate and implement environmental improvements. Life Cycle Assessment includes the entire life cycle of the product, process, or activity from “cradle to grave” including production, distribution, and consumption. It covers elements such as extraction and processing raw materials, manufacturing, transportation and distribution, use/reuse/maintenance, recycling, and final disposal. LCA is a brilliantly simple concept that is very complex to work through. Assessing impacts from a life cycle analysis allows the focus to be moved away from “end-of-pipe” remedies for environmental problems towards more effective and efficient remedies that can exist at other
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points in the life cycle of a product, for example, upstream in the design of the product or process. It also goes beyond assessing individual environmental issues toward an integrated approach that aims to measure overall environmental sustainability. Life Cycle Assessment has been used to define and compare the environmental burdens/impacts for finished products or services and to guide development work. LCA is used by an organization to guide a specific project or set overall strategies. It can provide environmental information to consumers and establish criteria for environmental labeling schemes — the EU Directive on eco-labeling states that criteria should be established using a life-cycle approach. Already, companies have shown that they do not wish to have “toxic to ecology” labels on their technical products and have sought alternative nonecotoxic formulations with an equivalent or better performance. LCA is most useful in assessing the environmental impacts of the production of goods and the provision of services, or to compare the impacts of different products, processes, or services. Examples would be, for example, polystyrene vs. china cups, aluminum vs. steel beverage cans, mineral base oils against vegetable base stocks. 32.1.4.1 The Functional Unit It is difficult to make comparisons between unlike products in LCA therefore the idea of the functional unit concentrates on the function served by a product, for example, it is unrealistic to directly compare a returnable milk bottle to a carton because the bottle can be reused a number of times. But the more useful functional unit for LCA comparative purposes would be the packaging of, say, 1000 l of milk, which could be by glass bottles, cartons, or other material. Reuse and recycling can then be brought into the overall LCA. Care must be taken in choosing the functional unit for the LCA because it may alter the outcome of the assessment. There may be thresholds where one product becomes more or less environmentally acceptable than another, for example, milk bottles may need to be re-used a certain number of times before they are more acceptable than cartons.
A Case Study of Life Cycle Analysis The Netherlands Ministry of Housing, Physical Planning and Environment studied the environmental performance of “crockery” made either from polystyrene, paper/cardboard, and porcelain. The functional unit was 3000 uses assuming that the polystyrene was not recycled and the product was used on a “use once/wash once” basis. The basic question was then: — how many times must porcelain crockery be used and washed so that its environmental impact is equal or less than that of disposable polystyrene or paper/cardboard crockery? The answer depended on whether water consumption and waste treatment, energy consumption, air pollution or landfill wastes are considered in terms of: — water consumption, where porcelain always does worst, because ‘disposable’ crockery is never washed, — energy consumption, where porcelain is used 640 times before it competes with polystyrene and 294 times before it competes with paper/cardboard crockery, — air pollution, where porcelain needs to be used 1800 times to compete with polystyrene and 48 times to compete with paper/cardboard, — waste, where porcelain needs to be used 126 times before it competes with polystyrene and 99 times to compete with paper/cardboard. Porcelain always overtakes the other products in terms of conserving energy, air, and waste well within the functional unit limit of 3000 uses. But it always has the worst performance in terms of water consumption. The overall problem is then the summation of whether the positive energy, air, and waste impacts for porcelain’s use as crockery outweighs its negative result impact for water. Lubricants do not fit readily into this category because they are not sold by decanted/fed volume, as from large containers, possibly the case in large service center garages, but normally from either sealed containers
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Output Waste heat
Fuel and energy
Raw materials
FIGURE 32.2
S Y S T E M
Air emissions Solid waste Water effluents
System inputs and outputs.
of 1 l or resealable 5 l. What is relevant is their service replacement interval and their contribution toward energy efficiency by reducing friction.
32.1.4.2 Inputs, Outputs, and System Boundaries Schematically, any extended industrial system can be represented by Figure 32.2 with the ensemble of operations enclosed within the conceptual box. The conventions are similar to system thermodynamics where the concept box denotes the system boundary separating the system from its surroundings and defines the system environment. This acts as the source of all inputs to the system and also as the sink for all outputs from the system. The physical description of the system, or inventory, is a quantitative description of all flows of materials and energy across the system boundary either into or out of the system itself, identical to those of conventional thermodynamics. Consequently, relatively few new procedures are needed to manipulate the data. Many arbitrary decisions introduced by some analysts in recent years arise from a lack of appreciation of the close relationships between life-cycle analysis and thermodynamics. The system is a physical entity that obeys all physical laws of matter, including laws of mass conservation and thermodynamics. These conditions exclude any nonsense or inconsistencies such as more mass coming out than is put in, or similarly, that more energy is given out than is put in, as in a perpetual motion machine. These laws provide a useful check on the validity of any proposed description, any violations of the laws invalidates the description. For extended industrial systems, the data describing the overall performance of the extended system comes from a number of different, separate operations, each taking the output from an upstream operation and processing it into a product for the next operation downstream. Each physical subsystem has the same characteristics as any other system, their function must be specified and obey the standard scientific laws. Figure 32.3 shows the different stages or subsystems throughout the life cycle of a product or operation, the use of materials and energy, emissions to the environment and the points at which reuse and recycling can occur between subsystems. An example of recycling is the reuse of glass bottles, whereas an example of product remanufacture is the recycling of glass. Material recycling to take the product back to the raw material stage is not applicable here for glass is not converted back to its raw materials of sand and soda. But material recycling can be carried out for ground glass (“cullet”), certain plastics converted into hydrocarbons as raw materials for new plastics, or used lubricants re-refined back into base oils; The further a recycling loop goes back up the flow chart, then in general, the more expensive the recycling process becomes. This occurs because each stage of the flowchart adds value to the materials passing through it. Materials recycled back to the beginning of the process will then have to compete with cheaper virgin materials
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Environmental Implications and Sustainability Concepts for Lubricants Raw materials and energy Emissions to air, water, and land Raw materials and energy Emissions to air, water, and land Raw materials and energy Emissions to air, water, and land
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Material acquisition Material re-cycling
Emissions to air, water, and lan
Material manufacture Product re-manufacture Product manufacture Product re-cycling
Raw materials and energy
Process materials energy
Process materials energy Emissions to air, water, and land Process materials energy Emissions to air, water, and land
Use Emissions to air, water, and land Raw materials and energy Disposal Emissions to air, water, and land
FIGURE 32.3 The life cycle of a product or operation.
coming forward. Recycled materials will always have to compete with virgin materials on cost, quality for purpose, and availability. LCAs require input data from a wide range of sources, unfortunately such data is often inaccurate and incomplete. To address these considerable shortcomings in available information, many industrial organizations have commenced projects that examine and analyze their working practices to give the necessary information for wider dissemination and use. The use of unreliable data cannot be justified and it is essential that any data is reliable and validated for proper analyses. In this way, opposing views of environmental impacts can be examined objectively.
32.2 Environmental Implications of Lubricant and Hydraulic Fluids Production 32.2.1 The Size of the Problem The total global production of lubricant oils, hydraulic fluids, transformer and condenser (capacitor) oils, heat transfer fluids and metal working coolants is approximately 39 Mt/pa. Of this total mass, the mass of lubricant additives contributes between 1.5 to 2 Mt/pa, and synthetic lubricant base fluids contribute approximately 0.7/0.8 Mt/pa. To put these masses into perspective, “chemicals” have an annual total global production of 400 Mt/pa, which includes additives and synthetic base fluids but not mineral oils. For if mineral oils are included and recognized as chemical products, for example, as in the Netherlands,
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then the lubricants industry is responsible for ∼8.5% of total global chemical production by weight. But the manufacture, at the front end, and eventual disposal at the end of its life cycle of this very large mass of material presents an environmental burden. Therefore, from their reaction activation energies it is not surprising that the environmental aspects of lubricants/hydraulics and their application methods have been of prime importance because of the complexity of local disposal. This continues to be an active area of “research,” in reality, mainly development.
32.2.2 Life Cycle Assessment Proper quantification of the overall environmental impact of lubricants and hydraulics, taken to include the other hydrocarbons described previously, requires a detailed LCA to cover manufacturing, use, and “end of life” fate. A systems approach identifies its boundaries as including petroleum, petrochemical, oleochemical, and engineering industry activities. This is a very complex process due to the broad scope required of the assessment and also because of the particular issues characteristic of the industry and its applications. An immediate complexity is that lubricants are not produced alone but usually as coproducts in an integrated product network based upon optimized petroleum and oleochemical refining or chemical processing. Consequently, the allocation of resource requirements and environmental impacts to the lubricant/hydraulic elements of these networks must be either exhaustively calculated or considered as necessarily arbitrary calculations. But also, since different lubricant types vary very much in their performance and in the amount required for a particular purpose, a detailed knowledge of the application performance is necessary to define appropriate functional comparisons. LCA comparisons are made on equivalent outputs, so that a simple comparison of different lubricant types based solely on their resource requirements per mass (kg) or per volume (liter), will give misleading results. Even more so, as lubricants are widely used in many different products and applications, tracing their fate at the end of life can be very difficult because the “end of life” treatment of industrial lubricants used at a single manufacturing site is readily controlled. But tracing the eventual fate of used car engine oils is a more challenging problem because that eventual fate of much lubricant production is not well known. Even when data is, in principle, publicly available it may be regarded as confidential to the product manufacture users and may not be openly published. Therefore, for LCAs to be publishable generally requires coordination by an independent body facilitates “pooling” of commercially sensitive information. An example of the extent of this task is given by studies from related chemical industry sectors including the 1995 European Life Cycle Inventory for Detergent Surfactant Production. This study considered energy and resource requirements used to produce seven major surfactant types. There were no assessments of impacts previously described. The overall study took 2 years to prepare and involved 17 technical professionals from 13 companies, a considerable resource input, and the “Eco Profiles” of the European Plastics Industry. A key conclusion from these and other studies was that the technical basis existed to support a general environmental superiority claim either for an individual product type or for various options for sourcing raw materials from petrochemical, agricultural, or oleochemical feedstocks. For these formidable reasons no comparable comprehensive Lubricant LCA has been attempted as yet for lubricant and related fluids. Lubricant and hydraulic fluids have been considered as contributing elements of LCA studies focused on particular applications such as forestry hydraulic equipment applications, municipal cleansing, and domestic refrigerators. But companies involved in the lubricants and hydraulics industry carry out detailed analyses of their product’s environmental impacts in their strategic and internal decision making. It could be argued that the environmental implications of their products have now had to be considered in depth. But the results of these studies are not generally available external to the companies but some limited studies have been revealed. Single company studies are limited by access to the company data involved. This restriction can also cause concerns for the independence of the conclusions reached from the commercial interests of that company.
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32.2.3 Quantifying Lubricant Chemistry Production The ideal situation would be to have full “cradle to grave” LCA of lubricant and hydraulic fluid production and use. But this lack of analysis does not imply that no information is available — “Green Chemistry” concepts reinforce and complement those of LCA. This section summarizes some relevant, available, information and reviews its implications in terms of Green Chemistry principles to arrive at general qualitative conclusions. 32.2.3.1 Production of Lubricant and Hydraulic Fluid Base Oils — Mass Efficiency and Process Energy Requirements Any product manufacturing process has distinct steps that can be analyzed and broken down, for lubricant and hydraulic fluids these broadly divide into manufacture of the base fluid, manufacture of the additive, and formulation of the base fluid and additive into a final product. Environmental impacts for each of these production process steps can be assessed. Because each of these process steps involve energy inputs, their efficiencies can be assessed on the basis of each process step efficiency in terms of reaction efficiencies, which is readily measured, and the energy requirements for each process, again readily measured. For the first point, the typical energy requirement for the petroleum processing of mineral base oil fluid production is estimated as 9 MJ/kg, ∼21% of the Energy of Material Resource (EMR), or the total energy of the product, ∼42 MJ/kg. Process energy requirements in a modern integrated refinery come from feedstock combustion in several ways. Refineries produce a wide range of products from LPG to asphalt so that all of the input raw material is mainly converted to products or is used for energy production to cause that transformation. The mass efficiency of mineral oil base fluid production is very close to 100% if that part of the feedstock used for energy production is not included in the overall mass and energy balance. Therefore, every part of the crude petroleum input is used in refinery operations, otherwise residual, unusable, and unpleasant materials would accumulate at a fast rate at each refinery. A considerably smaller proportion of lubricants are based on natural oils. Natural oils have very large and broad market applications in food and agriculture and lubricants are a minor application. To allocate the resource requirements of agriculture to the minor sector of lubricant base fluids is rather arbitrary and, as such, highly sensitive to the assumptions made for the input mix of raw materials, their geographical origin, thus implying transport costs, the related energy inputs of mechanical equipment (tractors, harvesters, and oil extraction machinery), and the energy input into fertilizer production. Taking all of these into account, a typical overall energy requirement of ∼3 MJ/kg for vegetable oil production is taken as being typical. It is possible to refine this value further for a particular natural oil but to little numerical effect. 32.2.3.2 The Production of Lubricant and Hydraulic Oil Additives The wide range of additives used in lubricant formulations, treat rates for individual additive types, and the chemical reactions involved, combine to make it difficult to quantify mass efficiencies and process energies for lubricant and hydraulic additive manufacturing. But from the general manufacture of specialized chemicals and polymers it is reasonable to assume that mass efficiencies and process energies will vary considerably between products and in this way affect their manufacturing costs. However, these values and costs are regarded as very commercially sensitive and equally difficult to obtain. But additives are not unique as chemicals and polymers and these is much published information on similar and analogous compound production of similar scale. This information shows that reaction mass efficiencies for specialist chemicals and polymers are generally 95–98%, the 2–5% losses arising from reactor washouts and filtration losses. These losses reduce if the same product is sequentially made in the same reactor. Therefore, it is readily assumed that lubricant and hydraulic additives are mostly made in a batch reactor. The reaction solvent is usually a mineral oil that is included in the final product, with little further processing of the reaction products. Again, there is very little available information on the process energies for lubricant and hydraulic additives because of the industry’s intensely competitive nature. But again by analogy to available values for materials that can be judged as being quite closely related, the estimate for the overall process energy requirements for the manufacture of lubricant and
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hydraulic additives are close to the EMR of the final products, 35/40 MJ/ kg. Variance of product types can double or halve that value. Stepping back and reviewing what these values mean, lubricant and hydraulic additives are not produced for their energy content but for the energy or resource that they can save. The insight of a LCA for a friction reducer takes into account the energy that its application will save over its service life. Similar analysis for an antiwear additive must take into account the increased life of a machine given by that additive. 32.2.3.3 Production of Synthetic Lubricant and Hydraulic Base Oils Very similar analyses to those used previously are used to assess synthetic base fluid production. The benchmark for process mass efficiencies of most synthetic base fluid production is >85%, with lower values for phosphate ester and silicone production by organic compound substitution of a chlorinated reactant such as POCl3 or SiCl4 , respectively. These reactions evolve HCl and the reaction efficiency overall depends whether it is recycled into another reaction or used to make hydrochloric acid. PAOs and PAGs have higher values, produced by addition polymerization that can be configured to give very high yields. Overall, the process energies for manufacturing synthetic fluids are at most comparable to lubricant additives. As these compounds may be produced in a specified state, for example, by varying the polymerization conditions for a PAO, then distillation/fractionation is unnecessary. On this point, the final distillation process energy for mineral base fluids is not needed for some synthetic fluids. As most synthetic fluids are hydrocarbons, their EMR is very similar to that for the mineral base fluids at ∼42 MJ/kg. Whilst it is useful to step back and review what these values are and what they mean, synthetic fluids are not produced for their energy content but for the energy or resource that they can save and their LCA takes that into account. 32.2.3.4 Production of Lubricant and Hydraulic Formulations Base oil and additives have to be brought together as a formulation for sale and application. This is necessarily a mixing process or liquid blending operation usually in a batch reactor. The mixture must be heated and mixed and if a careful progressive sequence of products is followed, then there should be 100% nominal mass efficiency. From analogous mixing operations, the input energy required for mixing and heating the components during blending is estimated as 3.6 MJ/kg. Longer mixing times may be required for incorporating some types of VII materials because they have the initial physical composition of toffee into the formulation. The length of mixing time must be balanced against the possibility of shear and thermal degradation of the VII. 32.2.3.5 Lubricant and Hydraulic Formulations — Conclusions Taking the overall composition of a lubricant formulation as being 15% additive and 85% base oil and using the values discussed in the preceding sections on a pro rata basis, then lubricant production has high mass efficiencies, minimum 85% and on average much higher in the 90 to 95%+ region. It has relatively low unit process energies, mainly because of the preponderance of the mineral base oils with their low energy requirements. They contribute to low process energies because of their natural origins and being (a small) part of the very high production volume petroleum industry.
32.3 Environmental Benefits of Lubricant and Hydraulic Formulations 32.3.1 Overview of Lubricant Use The primary purpose of lubricants is to reduce friction in machines, “machines” defined in the pure engineering sense. It follows that reducing friction in turn reduces the consumption of energy for the same amount of delivered work. Therefore reduced energy consumption is equal to increased energy efficiency. As almost all energy conversion processes cause some form of environmental pollution, then reduced
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energy consumption means reduced emissions. This can be a “win-win” situation with considerable returns for improved lubricant and hydraulic fluid performance. To extend the position of lubricants in the petroleum/petrochemical context, the total lubricant production worldwide as approximately 39 Mt/pa but constitutes only ∼1% of worldwide refinery crude petroleum throughput. In comparison, the total production of other nonfuel uses such as solvents, waxes, and bitumen is approximately 1% total petroleum refinery throughput. The main petrochemical feedstocks such as ethylene and propylene are approximately 3% of worldwide refinery crude petroleum throughput, the remaining 95% of crude petroleum refinery throughout is used as fuel in various forms as about 3.5 to 4.0 Bn/Tn/pa. These figures of volumes, masses, etc., are continually changing due to economic activity levels and should be carefully reviewed on the basis of contemporary consumption. For the developed world, consumption of lubricant and hydraulic fluids is decreasing. Elsewhere in the world, consumption of lubricants is higher per head but decreasing overall. But overall lubricant consumption outside of what is currently regarded as “the developed world” is tending to increase due to increased vehicle ownership and use. The further context is that crude oil only contributes approximately 40% of world energy consumption, varying in percentage terms from one field of energy to another. Thus, crude oil products contribute heavily to mobile transport but relatively little to fixed energy generation. The other energy sources are coal and gas as fossil fuels, then nuclear energy with renewable energy sources well down the scale at about 1 to 2% but increasing relatively rapidly. The total world energy “system” produces about 22 BillionTonnes/annum of carbon dioxide and this emission is now recognized as the major air pollutant contribution by far compared to other nitrogen and sulfur oxide emissions and also solid particulate matter. Carbon dioxide is said to be the main contributor to the “Greenhouse Effect” and its uncontrolled, increasing, emissions are associated with global warning. 32.3.1.1 Energy Conservation and Lubricants After separating out the direct use of fuel to heat premises, plant and processes, most fuel is still used as the energy driving force for machines as transport and power generation. These processes require the relative movement of surfaces in some form of bearing where the lubricants’ function is to reduce friction between those surfaces. The lubricant should, overall, reduce the loss of energy into frictional losses and to maximize the conversion of energy into useable work, thereby reducing the energy input. If the “correct,” that is, matched for purpose lubricant is used then the efficiency utilization of energy can be improved by up to 10%, more realistically 5%. But improved lubricants properly applied can save even more. To a certain extent this is semantics, for there “is always a better lubricant.” But the lubricants have a role to play in overall energy conservation. 32.3.1.2 The Wear and Conservation of Machines by Lubricants The Jost Report [15] on “Tribology” commissioned by the U.K. Ministry of Technology in the mid1960s showed that the main benefit of properly applying tribological principles was improved machine life and system reliability. These benefits feed through into the extension of a machine’s productive life, thus sustaining material resources and reducing energy consumption. These associated effects conserve nonrenewable fossil fuel resources over the working life of the machine. The report concluded that by effectively applying the existing, 1966, knowledge and techniques in tribology, UK industry could make annual savings in operating costs of £515 million at 1966 prices. 32.3.1.3 Working Lifetimes of Lubricants and Hydraulic Fluids The volume of lubricant and hydraulic fluids used worldwide depends not only upon the number of applications but also on the service life between service life refill of those machines. The determinants are twofold — the prescribed lifetime of the lubricant and hydraulic fluids and also their required replacement. The simple equation is that if the lubricant can be formulated to last longer, then overall demand decreases. Some additional energy is required to modify the base oil, as one example, by more extensive hydrotreatment, but if this doubles the service life of the lubricant then its demand is halved and also its
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environmental impact pro rata. Further, less used material must be disposed of. It is useful to compare lubricant consumption per major world region using the criterion of M$/GNP. Western Europe has always used the most highly formulated and sophisticated lubricants for its relatively demanding general automotive performance of high speeds using small engines. The use of lubricants decreased from 0.7 bls/GNP unit in 1990 to 0.6 bls/GNP unit in 1997 and has decreased further to 2004. For the same period, Eastern Europe has decreased from a much greater value 9.1 down to 6.0 units following its opening up to Western Europe. It shows every sign of moving rapidly toward the Western European average as it increasingly uses more modern lubricant formulations and service intervals. North America, with traditionally less sophisticated lubricant formulations because of its acceptance of more frequent/shorter lubricant changes, has declined from 1.7 to 1.1 units. This agrees with the general rule that U.S./North American service intervals are usually about half that of Europe. At the same time, Latin America has declined from 3.1 to 1.6 units, Australia/New Zealand down from 1.7 to 1.1 units, Asia from 1.9 to 1.4 units, Africa from 3.5 to 2.6 units and overall world consumption from 1.9 to 1.0 units. Whilst these are “broad brush” figures, they are self-consistent. They do not give the complete picture because GNP has increased overall and the values given above need to be translated into total consumption. Over ten years lubricant consumption has declined from ∼40 Mt/pa in 1989 to 37.6 Mt/pa in 1999 in the context of an increased total number of vehicles. This has not been a straight line decrease, fluctuations in demand reflect the worldwide economic climate, as also experienced by the fuel market. There are three major factors that are driving down lubricant demand: 1. A move toward more sophisticated lubricant formulations with longer service lives, particularly the case for Western Europe where 10,000 km service intervals for light vehicles have existed for the last 15 years and now extend well toward 20,000 km for both petrol/gasoline and diesel with some manufacturers offering vehicles with 50,000 km service intervals. Service interval limits are now determined by wear of other components of engine systems, such as spark plug erosion unless advanced designs are used. 2. Heavy diesel vehicle service intervals follow the same trends, a major heavy construction and off-road vehicle manufacturer has a service interval target of 400,000 km for the reasonably near future. The resulting used lubricants may be more heavily contaminated than earlier but their mass is considerably less and their resource requirements are also less. 3. A move toward service refill when the lubricant condition requires it rather than when predicted — “condition monitoring.” The actual patterns of use for identical vehicles usually varies between regular long distance, uninterrupted journeys to multiple short journeys with frequent stops, the latter often being more demanding on the condition of the lubricant! Larger engines under demanding operating cycles may consume sufficient lubricant, for example, 6 l/day “top-up” for a total lubricant charge of 180 l, that “top-up” by fresh lubricant is sufficient to maintain a suitably protective lubricant condition in the engine. Increasingly, vehicles have a form of “on-board” engine monitoring that integrates the various types of driving mode experienced by individual vehicles and eventually reaches a predetermined internal value which informs the operator that an engine service is due. This is not a measure of lubricant condition but an integration of the severity of driving modes. A feasible proposal, as yet to be offered commercially, envisaged a joint program between a vehicle manufacturer and a national fuel supplier where the manufacturer’s vehicles are fitted with transponders (“transmitter/responders”) that transmit the current vehicle condition when interrogated as it passes by a fuel service station garage. The information is collated by a central data base and when a service is deemed to be required, a message is sent to the operator/owner. Such a scheme does not require much development because the components are there, transponder technology is well developed in the aerospace industry, medium/larger contemporary vehicles have engine management systems that store performance data and mobile telephony is very well developed. All that is required is for these systems to be integrated and be shown to be useful. For lubricants, the overall demand would be reduced.
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Pistons, piston rings
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Cooling and exhaust
Rolling resistance Bearings
Pumping
Valve train
Mechanical losses
Air resistance
To wheels
Acceleration
Total power from fuel 32 kW (100%)
Power to the wheels 3.8 kW (12%)
Auxiliares Transmission Mechanical losses 4.9 kW (15%)
FIGURE 32.4 Fuel energy distribution for a passenger car during an urban cycle. (From B.S. Andersson Company Perspectives in Vehicle Tribology-Volvo-Proc. 17th Leeds-Lyon Symposium on Tribology-Vehicle Tribology, Elsevier, 1991, pp. 503–506.)
4. By reducing the internal friction coefficients of machines. Whilst automotive manufacturers are improving the overall thermodynamic efficiency of their engines, both petrol and diesel, it is equally recognized that the internal friction of machines can, and should, be reduced. For “machines,” the whole automotive power train must be considered, that is, engine, gearbox, transmission, and wheelbearings. Only a small percentage of the available chemical energy from the fuel is transformed into mechanical energy to propel a vehicle, Figure 32.4. A generous assessment would be 20 to 30% of mechanical energy for movement delivered to the wheels from the available thermal energy. However, the data within Figure 32.4 is for a light vehicle engine undergoing an EU emission test, much more representative of “normal” use than an engine tested at high power level in a cell. During the emission test, only 12% of the available input energy is propelling the vehicle through the driven tyres. The rest of the energy is emitted as thermal energy from the radiator/air cooling, in the exhaust, internal energy losses, and a very small amount of unburned fuel, noise, and vibration. Some energy is “parasitic,” using part of the engine energy to (increasingly!) drive electrical and air conditioning systems. Some rejected heat is used, advantageously, to heat the vehicle. The cost implications of these mechanical losses are startling, illustrated by an analysis at current U.K. prices of £0.80/litre: Mechanical Energy losses as a % of fuel energy are 15% Cost of wasted mechanical energy/litre @ £0.85/litre, is £0.148 UK total cost/year, 25 M vehicles, 15k miles/year, 6 mi/litre is £9.25Bn A 10% reduction in mechanical losses by using improved lubricant formulations would save £925M each year from transports costs
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This calculation can be repeated for different fuel costs, different vehicle use and currencies, and will still give very significant financial savings. The obvious objectives are to shift the balance between the amount of heat rejected from the engine and power train and the thermal efficiency of the engine. For engine thermal efficiency applications, simplistic Carnot cycle calculations show that higher thermodynamic efficiencies are achieved by, inter alia, higher cylinder peak operating temperatures. But the higher operating temperatures involved increases the oxidative stress upon the lubricant that must be countered with improved additive technology and base oils with enhanced antioxidancy, such as the extensively hydrotreated mineral base oils and synthetic base oils. A practical step is to move to lower viscosity base oils in lubricant formulations, provided that the engine design can use them without long-term damage. A major manufacturer has moved from 5W-30 lubricants to 5W-20 for both factory fill and also their proprietary garage brand across almost all of their product range and their associated marques, in Europe and the United States. The environmental benefits are that an estimated 105 tonnes per annum of carbon dioxide will not be released into the atmosphere in the United States alone by reducing fuel consumption by ∼80 million litres per annum of petrol. In the context of the total volume of fuel consumed by the U.S. automotive industry each year this is not much but every contribution counts. The problems for the formulator of lower viscosity lubricants are several. For conventional hydrocarbon base oils, lower viscosity is related to lower molecular weight, in turn to higher volatility. But if the engines operate at higher temperatures, then the volatile loss will be much greater. Therefore, base oils of different structures with lower volatilities and higher antioxidancy reserve are required, either as mineral oils extensively modified by further catalytic hydrogenation, or synthetic PAOs or esters. There is a choice between full synthetic ester base oils and extensively hydrotreated iso-paraffin mineral base oils for both reduced viscosity and volatility. The contribution of these base oils to reducing internal engine friction is that they can only work for hydrodynamic lubrication conditions. But considerable energy is absorbed at boundary lubrication conditions where the hydrodynamic film has broken down and the friction coefficient becomes independent of the lubricant viscosity. The boundary tribological condition depends upon the nature of the contacting surfaces. These surfaces can be modified by appropriate “surface active” additives. Many compounds have been proposed for this role, some are effective but the most effective are the molybdenum dithiocarbamates, as modified by various peripheral chemical groups, which degrade at the surfaces under severe physical conditions of temperature, pressure, and shear, and deliver molybdenum disulphide-type films onto surfaces with low energy shearable, or lamellar plane surfaces, which reduce friction between the surfaces. The problem is that the deposition reaction is irreversible and thus uses up the additive. The positive effect of the additive will decrease when its concentration in the lubricant formulation is used up. The degradation products have to move on and some are emitted in the exhaust — molybdenum is not a problem but the sulfur content can be a problem for its deleterious effect on the effectiveness of the catalyst. Surface active additive technology development has now moved away from metal/sulfur compounds toward derivatives of long chain fatty acid compounds that will give a lamellar effect through the formation of mono-, di-, tri, or multilayer films. The multilayer film effect is analogous to the lamellar structure of the molybdenum thiocarbamates, giving low surface energy structures that reduce friction by minimization of the adhesive forces between the metal surfaces in relative motion. This approach is successfully used in lubricity additives for diesel fuel in reducing injector pump wear. The friction reduction effects of surface active compounds must be carefully assessed using tribology. It is relatively easy to use a wear measurement device such as a “pin-on-disk” machine but at high rotation speeds hydrodynamic conditions apply and the determinant is the viscosity of the formulation. However, at low speeds the metal surfaces will contact each other and boundary conditions apply. Under the latter tribological conditions, the friction coefficient can be reduced by up to 20% by the use of surface active additives. However, this reduction is to but one part of the power train: the overall contribution to an increase in fuel efficiency might be of the order of 1 to 2%. To measure a friction reduction on a laboratory
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machine is promising. The effect must be translated into a long-lasting effect for the service life of the lubricant. The question is not only: • Does the additive “work” by producing an overall friction reduction effect? • But also, for how long does it work? This is a serious current issue for the first decades of the 21st century — there are many compounds that give an initial reduced friction effect. But far fewer give a consistent reduced friction performance over a longer time scale, such as the (extended) service life of a lubricant and not to lose that activity. Substantial development is undertaken into surface active additive compound systems with enhanced activity for the longer lubricant lifetimes, which are more resistant to oxidation. Satisfying all three is a substantial challenge!
32.4 Lubricant and Hydraulic Fluids as Wastes 32.4.1 Definitions of Waste The crucial issue in the life cycle analysis of lubricant and hydraulic fluids is how they are treated at the end of their initial useful life. It can be usefully demonstrable to take the used high quality lubricant from a demanding application such as a high speed express train diesel engine and use it in a much lower level of technical demand such as a shunting (switch) diesel locomotive. Then eventually from that application, to take that engine lubricant engine now containing high (∼10%) levels of soot and produce a coarse grease for the lubrication of railway points/switches. In reality there is always imbalance between the volumes of lubricant required at each level, in this case with too much waste lubricant produced at the top end by the high performance diesel engines and not enough demand for grease at the bottom end for points/switches. From this imbalance of demand and supply at various levels, there will always be used lubricant to deal with, as waste or for reclamation/recycling. A cornerstone of waste management legislation is the definition of “waste.” But the problem of dealing with waste is that, unfortunately, to different people the term “waste” means different things and this confusion means that there is no simple definition. Whether or not a substance is “waste” is crucial because a waste management licence is required to deposit, recover or dispose of waste. The answer is determined based on the facts of a particular situation and the pertinent law, irrespective of nation state. There is no specific legislation for lubricant and hydraulic fluid wastes, if they are to be regarded as waste then they are covered by the general description of waste. The EU legislation provides a Community perspective on what is regarded as waste, but national legislation enacts that legislation. The definition of waste in the United Kingdom has changed in recent years to reflect changes in EU definitions and people’s attitudes to the recycling and reuse of materials. The U.K. Environmental Protection Act 1990 (EPA 1990) and the Controlled Waste Regulations (1992) made under that Act provide definitions of waste and “controlled” waste for regulatory purposes [16]. Waste defined under these regulations includes effluents, scrap materials, unwanted surplus substances, as well as substances or articles that were broken, worn out, contaminated, or otherwise spoiled. This definition does not deal with materials that are to be recycled or reused nor with the increased public awareness of the problems of waste together with new management techniques such as environmental auditing and life cycle analysis. What is “waste” for one person is increasingly considered as a “raw material” for somebody else. The new definition of waste is: “Waste” means any substance or object in the categories set out in Schedule 2B to this Act (i.e., the EPA 1990)[16] which the holder discards, or intends to, or is required to discard. Schedule 2B is over-categorized and could be reduced considerably, some further definitions are necessary: • The “holder” is the producer of the waste or the person in possessing it • The “producer” is anyone producing waste, carrying out preprocessing, mixing, or other operations resulting in a change in the nature or composition of the waste
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• A service change lubricant is classified as waste if it has been discarded, disposed of, got rid of by the holder, or it is either intended or required to be so Another consideration is whether the substance/object is no longer part of the normal commercial cycle/chain of utility. Some items that may eventually be recycled are treated as waste because they need reprocessing before they can be reused.
32.4.2 Types of Waste There are various legal definitions of waste in most nation states. In the United Kingdom the EPA 1990 and the Controlled Waste Regulations 1992 made under it, divide waste into two categories: • Controlled Waste — meaning household, industrial, and commercial waste • Noncontrolled Waste — meaning agricultural, mine, and quarry waste “NonControlled” is a misleading term as agricultural, mine, and quarry wastes are controlled by separate, specific, legislation specific to their areas in most nation states. The Waste Management Licensing Regulations 1994 emphasizes “Directive Waste” (referring to the definition of waste in the EU Directive 91/156). Hazardous and Special Waste — terms often used for “hazardous wastes,” defined in the EU Hazardous Waste Directive (91/689) as referring to wastes that have hazardous characteristics such as corrosive, infectious, or ecotoxic substances. “Hazardous waste” is an EU term, the United Kingdom uses the term “Special Waste,” in general terms defined as waste that is dangerous to life and is controlled by the Special Waste Regulations 1996. The residues from recycling used lubricants are usually classified as “Special Wastes.”
32.4.3 Development of a U.K. National Waste Strategy, “Making Waste Work” The UK Government’s Strategy for Sustainable Waste Management in England and Wales, “Making Waste Work” December 1995 (“MWW”), had three key objectives aimed at achieving sustainable development: • Reduce the amount of waste produced • Make the best use of the waste produced • Choose waste management practices that minimized the risks of immediate and future environment pollution and harm to human health To achieve these objectives, waste management options are ranked as a hierarchy with the most preferred options at the top, Figure 32.5. The recovery of materials is preferred to the recovery of energy. This follows the same pattern as the EU Principles of Waste Management, which are the prevention of waste as the first priority through reduction and reuse, followed by recovery and then the safe disposal of waste. The U.K. targets for “Making Waste Work” include reduction of controlled waste going to landfill to 60% by 2005, to recover 40% of municipal waste by 2005.
32.4.4 Introduction to the “Waste Strategy 2000” In May 2000 the U.K. Government published its National Waste Strategy, “Waste Strategy 2000,” which had several purposes, a national waste strategy is a requirement of the EU Waste Framework Directive, to help ensure U.K. compliance with other EU legislation such as the Landfill Directive, the Hazardous Waste Directive, and the Packaging Waste Directive, and to help waste management contribute toward sustainable development in Britain. These are intended to deliver the benefits of increasing the diversion of waste away from landfill, encouraging the public to reuse and recycle waste, halting the rise in waste generation, using “the best practicable environmental option” as the basis for waste management decisions, increasing general awareness of the end-of-life impacts of products, and developing markets for recyclable materials. The Strategy sets a range of new targets, for achieving what the Government considers to be
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Reduction
Reuse Recovery Recycling—Composting—Energy
Disposal
FIGURE 32.5 The waste hierarchy.
“the appropriate balance between different waste management options” and this will be reviewed by a number of Government Departments and Agencies in 2005, 2010, and 2015, with the mid-point review being undertaken with particular thoroughness.
32.4.5 Waste Reduction The first step in controlling waste is to reduce the amount of waste produced in the first place. The EU recommends that the following techniques are used and developed for waste prevention: • • • • •
Promotion of “clean” technologies and products Reduce the hazardous nature of wastes Establish technical standards and rules to limit the presence of dangerous substances in products Promote and use recycling schemes The appropriate use of economic instruments, eco-balances, eco-audit schemes, life cycle analysis, and actions of consumer information and education as well as the development of the eco-label system
32.4.5.1 Minimizing Resource Use One of the fundamental tenets of the sustainability concept is to use less material for the same purpose, “less is more,” thus “conserving natural resources,” and general examples include: • • • • •
In 1970 an average plastic yoghurt pot was 11.8 g, in 1990, 5 g, a 58% decrease A plastic detergent bottle was 300 g in 1970, 100 gm by 1985, a 67% decrease Plastic film for the same application from 180 µm in 1985 to 80 µm now Plastic sacks from 300 to 165 µm, carrier bags from 45 (1990) to 15 µm Average weight of stretch wrap for wrapping pallets now 350 g, replacing 1400 g of shrink wrap used in 1990. Further, use of “soft” adhesives as applied “strips” to hold sacks or containers together could replace plastic wrapping altogether.
32.4.5.2 And Lubricants? Comparison of the lubricant requirements of a European“sports saloon”of the 1950s, with a contemporary light vehicle, shows that: • In 1950, the lubricant requirements for a 2.4 l, 100 bhp engine were 9 l to be changed every 1,000 mi of a monograde, SAE 40/50 in summer, which was changed to an SAE 30 in winter. Decarbonization of the combustion chamber, of valves and piston crowns was accepted as a necessity at ∼25,000 mi.
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A mechanical rebore of the cylinder block was expected and accepted at 40,000 /50,000 mi, with replacement oversize in-cylinder components. A vehicle lifetime of 100,000 mi was exceptional. The reclamation or recycling of the used lubricant was not contemplated. • In 2004, the lubricant requirements for an engine of similar output would 4 to 5 l, to last for a service interval change of at least 10,000, more probably 15,000 mi. Decarbonization and reboring of the in-cylinder components is almost unknown. Vehicle lifetimes of over 100,000 mi are common with over 200,000 mi not exceptional. The used lubricant is now expected to be recycled. The lubricant volume used per 10,000 mi has been reduced by at least 95% but for a much more refined product. After allowing for the small increased energy input to achieve the more refined product, there is still a drastic reduction in resource utilization and also energy input. Whilst reduced internal engine friction lubricants have an important part to play in reducing fuel consumption and extending the useful operating life of engines, the major contributions of lubricants to sustainability is extended operating service intervals to reduce the amount of material used and recycling to produce an acceptable quality base oil for new formulations.
32.4.6 Waste Recovery After waste minimization has been carried out and the opportunities for reuse explored, the next step is material and energy recovery. The EU Waste Directive recommends that, when it is environmentally sound, preference should be given to the recovery of materials over energy recovery, reflecting the greater effect on the prevention of waste produced by material recovery than by energy recovery. Recovery of materials includes recycling. 32.4.6.1 Material Recycling It is recognized that material recycling cannot always be done and is sometimes not economically feasible to be done. Used cling film, as an example, needs unraveling, washing, drying, and gentle ironing to be reusable. Or food preparation practices could be changed to reduce or remove the need for the use of cling film. But where circumstances are suitable, defined broadly as large arisings, clean material and of a single type, then recycling can lead to high quality products and good resource conservation. For lubricant (mainly) and hydraulic formulations, apparently similar fluids can contain different substances, such as different base stocks, base fluids, and additives. The different fluids to be considered for recycling do not have, nor carry, “identifiers” as some plastics have imprinted upon them. An external basis for recognizing the different types of lubricants and differentiating between them is not available. 32.4.6.2 The Collection Problem and the Entropy Model Before reclamation or recycling, the used lubricant and hydraulic fluids have to be collected to make a viable volume for processing. The collection problem of used lubricant can be compared to a High Entropy model seeking to move toward a Low Entropy solution, which requires an energy input. For after production at a small number of centralized lubricant blending plants, the packaged product is distributed through garages and retailers to individual vehicles, such as the 25M in the United Kingdom. In entropy terms, the system has gone from a low entropy, ordered system to one of high entropy, extensively disordered, system. To reverse this, energy must be expended, which is the energy of collection and part of the overall costing plan. Used lubricant is collected into garage collection tanks and Local Authority Waste Disposal Points from which it is tankered to central collection or treatment plants. Tankering itself requires an expenditure of energy, prior collection energies are not usually considered for taking used lubricants to a collection point. There is an increasing “Take Back” expectation of retailing and wholesale organizations, particularly in Western Europe.
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32.4.6.3 Lubricants as Waste, Fuels, or for Reclamation/Recycling There are various processes for reclamation or recycling that have been developed in response to environmental concerns and programs but the extent to which they are used depends on the process and resource economics for a particular intermediate involved. In the final analysis, the recycled feedstock must compete successfully on price and quality with virgin feedstock materials — this is part of the “Waste Paradigm.” Feedstock recycling for lubricants and hydraulic fluids involves separation of the base oil from the used, mixed, materials as the only really useful and recoverable material. The particulates, sooty material and used/unused additives have not been considered for recycling and must be disposed of. If used lubricants are regarded as a “Waste” for disposal, they are then subject to the Special Waste and Signage Regulations of the United Kingdom, or their equivalent in other countries. The additive package in the original virgin oil and its degradation products together with combustion-derived contaminants such as soot, water, and fuel in the used lubricants place them in the “Special Wastes” category. As such, their producers, transport contractors and ultimate disposers are subject to administrative measures designed to track their progress “from the cradle to the grave” in respect of their arising, transport, and final disposal, an expensive but necessary process. Therefore, it is advantageous to reclaim or recycle useful components from used lubricants, which are the major and relatively innocuous part, and to deal with the separated minor residues composed of spent additives and contaminants as “Special Wastes.” 32.4.6.4 Reclamation of Lubricants Some lubricants can be “reclaimed” and then reused, where “reclamation” means low levels of processing such as: • Straightforward filtration to remove particulates to a standard suitable for the further purpose of the reclaimed oil • Additive replenishment up to the original specification Filtering and replenishment is suitable for lubricant and hydraulic fluids that are not used in reciprocating i/c engines, nor exposed to thermal stress and are used in closed systems. The essential distinction is that the lubricants and hydraulic fluids are not exposed to fuel combustion processes and that the direct and indirect products of combustion do not accumulate in the lubricant over time. An excellent illustrative example of reclamation is the cleaning and replenishment of hydraulic oils in plastic injection molding machines. There are no combustion products and metal wear particles are minimal. The hydraulic oil actuates the clamping of the mold platens under high closure forces, drives the injection screw, and opens the platens at the end of the injection cycle to release the product. The service life of hydraulic oils used in injection molding machines is several years. Moreover, these oils can be “reclaimed” on site by a mobile filter apparatus. The additives are usually straightforward antiwear and antioxidants whose concentrations can be readily measured and replenished as required to the original specification. 32.4.6.5 Recycling of Lubricants The composition of used lubricants is that of the fresh lubricant base oil, modified by selective evaporation of the lighter fractions, plus some remaining original additive but usually much more spent additive plus contaminants of heavier fuel fractions, water, soot, and metallic wear particles. Another problem is the variable composition of the base oil component of waste lubricants, for example, Gp. I–III type mineral base oils, plus synthetic polyalphaolefins, PAOs, as Gp. IV and polyol esters as Gp. V. Recycling of lubricants requires separation processes to produce a “good” quality base oil for reuse in formulations with a minority residue of no use for disposal. Ideally separation processes should have low energy requirements, very good efficiencies of separation of base oil and contaminants to produce an innocuous residue, and a very good quality base oil product. It is very difficult to achieve all of these at the same time in the same process. Reverting to the entropy analogy, to achieve better separation requires
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a higher level of energy input. The practicable separation processes used to recycle used lubricants are described in another chapter but are recalled briefly as either: • The clay/acid treatment process, which produces ∼60% usable base oil product but with a dark color and burnt odor, disadvantageous in quality comparisons to virgin base oil. The clay/acid/additive/contaminant residue is acid and “tarry” that makes its disposal both very difficult and expensive. Capital investment can be low and small acid/clay treatment plants can be mounted on a trailer. The mobility of such small plants has unfortunately been used by criminals in clandestine operations that produce a low quality, very acid, product. This may be sold as a cheap lubricant or may be blended with lower hydrocarbons of lower taxation levels such as kerosene to produce a “diesel fuel” of appallingly low quality, which is sufficiently acid to corrode fuel systems and engines. This is not helpful to the image of recycled lubricants. • The solvent extraction process, which can produce up to ∼75% of usable product with a relatively innocuous residue, using liquid propane. The principle of operation is that “like dissolves like,” where the propane as a nonpolar, nonhydroxylic hydrocarbon dissolves the nonpolar hydrocarbons. The usually polar or high molecular weight additives and residues do not dissolve and are separated by filtration. The propane solution of hydrocarbons is then de-pressurized to volatilize the propane leaving a good quality base oil but darker in color than its original and with a burned odor. • The vacuum distillation, hydrotreatment, and further distillation processes, give a range of superior quality recycled base oils of significantly less color and odor with a recovery efficiency of up to 95% of that available in the used lubricant. To achieve the higher quality of recycled base oil requires a higher level of input energy to operate the thin film evaporator columns and hydrogenation. Whilst up to 95% of hydrocarbons may be recovered, some of this may be as lower molecular weight fractions that are recycled as fuels. There can be a surprising “carry-over” of trace metals and polymer breakdown products into the base oil, which require further treatment to improve final product quality. A solvent extraction process using dimethylformamide will remove residual sheared polymers and water. 32.4.6.6 Recycled Base Oils and the Waste Paradigm The Waste Paradigm relates the essential units of materials, from reserves through to wastes and is very instructive for the acceptability of recycled base oil into the product chain. It has the normal form of Figure 32.6. The crucial insight given by the Waste Paradigm is that if the recycled material is to be an effective contributor to material flows and can compete with fresh raw materials, in this case recycled base oils with virgin base oils, then the recycling process must produce recycled base oil that can compete at the Material stage in the scheme above, on grounds of : • • • •
Quality Cost Compatibility with additives Consistency of supply Resource → Reserve → Material → Product → Use
Recycling
FIGURE 32.6
The waste paradigm.
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The base oil product from the recycling of used lubricants must therefore be of good quality, relative to the original base oil. The recycled product to meet this challenge particularly comes from the vacuum distillation/hydrogenation process that has improved oxidation resistance and enhanced viscosity index, both desirable properties. A typical recycled base oil from this process can be classified as a “better than” Gp. I/“almost as good as” a Gp. II base oil. 32.4.6.7 Residues from Recycling The residues from the recycling processes are “special wastes” under all nation state regulations and need to be rendered innocuous. The acid treatment/clay/spent additive/soot/filtration process are extremely difficult to dispose of and are a substantial detriment to the environmental acceptability and economic viability of this process. Both the propane liquid/liquid solvent extraction process and the vacuum distillation/hydrogenation/further distillation processes give additive/ degraded additive/soot particulate sludges that can be disposed of by incineration, either in special incinerators or in authorized cement kilns. Other separation processes produce residual sludges that are difficult to dispose of and difficult to be acceptably incinerated. It is very expensive to send “Special Waste” materials to landfill.
32.4.7 Used/Waste Lubricants as Fuels A convenient route for used/waste lubricant disposal was, until recently, as a supplement for solid and liquid fuels to supplement their calorific values. This was both convenient and economically attractive with a fuel tax rebate of $40 per tonne. The EU Waste Directive removed this subsidy from January 1, 2004 with a complete ban on waste lubricant burning in power (utility) stations from January 1, 2006. After 2005, waste lubricants can only be burned in specially approved incinerators or equally approved cement kilns. The provisions of the EU Waste Directive in respect of the combustion of raw or partially refined waste lubricants as fuels combine financial and legislative incentives to recycle lubricant components. This cheap, simple, but nonsustainable route to waste lubricant disposal has been closed and recycling must now be seriously considered. The remaining alternatives of specialized incinerators or cement kiln fuels are controlled by, first, the formidably high costs of specialized incineration and second, the finite number of cement kilns that increasingly use other wastes such as mixed solvent laboratory slops or shredded tire rubber as fuel. Cement kilns are popular disposal sources of “difficult” substances, subject to the EA/HMIP (UK) “authorization” procedure and its essential equivalent in other nation states. The problem is that the capacity of cement kilns to use “wastes” of whatever form or composition is necessarily limited and are not panacea solutions of infinite capacity to the “special waste” disposal problem. A proposal to use “unusual” fuels in any major industrial process in any developed country requires regulatory authorization for that proposal to vary the fuel in that process.
32.4.8 Incineration Incinerators are designed to destroy the organic components of wastes by combustion. However, most waste has mixed content, whatever its nature and will therefore contain both combustible organics and noncombustible inorganics. By burning the organic fraction and converting it to carbon dioxide and water vapor, incineration reduces the waste volume and its threat to the environment. Incineration converts waste into an ash residue, known as the product solid and a variety of gases, known as the product gases. The hazardous nature of these products depends upon the type of waste and the conditions for its incineration. After incineration, the ash volume is considerably less than the original waste, is generally inert, and can be safely landfilled. The heat produced by the incineration process can often be recovered before noxious gases are removed and released to the atmosphere.
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32.4.8.1 Waste Incineration Directives Directive 94/67/EC on the Incineration of Hazardous Waste was adopted in 1994 and implemented by Member States by 1996. This Directive was implemented in the United Kingdom through the existing system of Integrated Pollution Control (IPC) under Part I of the EPA 1990. Its main effect was to introduce more stringent emission standards for operators of hazardous waste incinerators. Many of its provisions also applied to industrial combustion processes using hazardous waste as a fuel, such as cement production. The new EU Incinerator Directive approved in June 2000 [17] replaces 94/67/EC as well as Directives 89/369/EEC and 89/429/EEC on municipal waste. The new Directive introduces tighter emission and other operating standards for virtually all incinerators, aiming to prevent or minimize harm to the environment media of air, soil, surface water, and ground water from incinerating hazardous waste. The Directive also applies to co-incineration, occurring at which plants burn hazardous waste as a fuel which are not dedicated hazardous waste incinerators. Authorization may only be granted for a co-incineration process if the hazardous waste burners are located and the waste fed in such a way as to achieve the maximum possible level of incineration. The conditions must specify the minimum and maximum mass flows of hazardous waste and lowest and maximum calorific values of the wastes, the maximum concentrations of pollutants such as PCBs, halogens, sulfur, and toxic metals. The 2000 Waste Incineration Directive has radically raised emission standards that require high standards of plant and operation. In turn, this has sharply raised the costs of incinerating wastes. In an analogous way to the closing of the waste disposal as fuel option, as described in Section 32.4.7, the provisions of the EU Waste Incineration Directive have effectively closed off incineration of waste lubricants as a route for disposal, emphasizing recycling as the preferred route. Effectively, only the separated sludges from the recycling process can be incinerated.
32.4.9 Waste Disposal to Landfill Landfill is often regarded as the “last resort” waste management option. The EU recommends that uncontrolled landfill is an operational area that needs special and strong action. An engineered landfill is designed to contain waste and its decomposition products until they present no significant risks to health or the environment. The three main areas of legislative control relating to landfill development are as follows: • The planning system, which controls the development and use of land in the public interest • Pollution control legislation, incorporating waste management licensing, and measures for environmental protection from the effects of the landfill, such as protection of existing groundwater resources from landfill leachate • Regulatory and statutory controls to protect health and safety, and to ensure minimum standards for civil engineering construction 32.4.9.1 The EU Landfill Directive In April 1999 the EU Directive on the landfilling of waste was adopted [18] and had a significant impact on the European Waste Management industry, principally through the costs of landfill. Its main aims are: the harmonisation in all Member States of technical and environmental standards for landfill and to ensure a high level of protection for the environment, in particular of soil and ground water. Under the Directive, waste is classified according to its characteristics, in parallel landfill sites are also to be classified as for acceptance of these categories and certain waste types are no longer accepted for landfill. The last includes, amongst others, liquid wastes that are directly relevant to hydraulic and hydraulic fluid wastes. The Directive includes a requirement that all waste must be treated before being landfilled with the intention reducing the volume or the hazardous nature of the waste, to facilitate waste handling and increase recovery.
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32.4.9.2 The Landfill Tax In 1996, a landfill tax was introduced in the United Kingdom to ensure that landfill waste disposal is properly priced to promote greater efficiency in the waste management market and in the economy as a whole. It also applied the “polluter pays principle” so that those producing pollution should pay for its treatment/removal and not use the environment as a free waste disposal resource as well as to promote a more sustainable approach to waste management which produces less waste and reuses or recovers value from more waste. The tax complements and reinforces the general approach and policies of the U.K. Government and also the EU for sustainable waste management. The overall aim of these policies is to increase the proportion of waste managed by using options toward the top of the of waste management hierarchy, as in Figure 34.1. The tax allows waste disposal companies to pass the additional costs on to waste producers. In turn this means that waste producers are made aware of the true costs of their activities and so have the incentive to reduce and make better use of the waste they produce. Landfill tax was introduced at £2 per tonne for inert waste and £7 per tonne for active waste. The tax was increased to £10 per tonne for active waste in April 1999 and the Chancellor of the Exchequer announced an annual increase of £1 per year for five years. This charge has acted as a substantial disincentive for the landfill of wastes other than those that cannot disposed of by other means.
32.4.10 Directions in the Disposal of Used Lubricants From the examples given, the driving economic force in the EU is to reclaim/recycle used lubricants for their usable materials. The emphasis is upon not using used lubricants as fuels nor to incinerate or landfill them as wastes. This requires adjustments in the lubricant industry, which now has to include the concept of sustainability in its operations and accept recycled base oils into its formulations.
32.5 Environmental Pollution by Used Lubricants 32.5.1 The Polluting Effects of Used Lubricant When indiscriminately disposed of to land, the transport of used lubricant material depends upon the nature of the subsoil. If the subsoil is clay or clay-rich, then the materials strongly adsorb the lubricant oil mass and will have a very low annual rate of transport, particularly if there is a low rate of hydrological movement. If the subsoil is predominantly sand then the oil mass is less weakly adsorbed and any transporting force such as gravity or water flow will slowly move the oil deposit through the subsoil. The oil deposit will gradually deplete itself and its immediate vicinity of oxygen and becomes anaerobic, emitting unpleasant odors from the soil. If the waste lubricating oil is indiscriminately disposed of to water, either directly or by seepage from deposits on land, then surface slicks are formed on the water course that are aesthetically unacceptable. An immiscible surface film on the water course prevents oxygen transfer across the interface between air and water. Very quickly the water course becomes anoxic and odorous. Any waste lubricating deposited onto land will gradually leach original and degraded toxic substances into water and soils over time, dependent upon the nature of the soil. The more volatile components of the waste lubricant evaporate leaving behind the most viscous and heavy chemical components, often as a surface film. The environmental degradation reaction rate of the lubricant becomes diffusion controlled, for example, oxygen diffusion for oxidation becomes diffusion controlled, which will slow down the rate of oxidation. Whilst hydrocarbons will not readily emulsify, the presence of other compounds from the original additive package or arising from degradation in use will have emulsifying actions to produce sludges, usually unattractive to view. The environmental implications of the hydrocarbon base oils are that they are difficult to readily dispose of. They have a very high demand of environmental oxygen in the longer term and take a long time to decompose naturally. The hydrocarbon base oils and their intermediate decomposition products
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will emulsify with “worked” water, such as waves and surf, to produce an unpleasant “chocolate mousse” emulsion at shorelines. In addition to being unpleasant, the environmental implications of the hydrocarbon base oils are long-lived due to their low reactivity. They cause immense damage to marine life by coating organisms, birds, and animals with a relatively impervious (to oxygen diffusion) oily layer, which also destroys the waterproofing effect of bird’s feathers.
32.5.2 The Biodegradability of Lubricants Biodegradability is increasingly required for total loss engine lubricant systems is increasingly required in developed countries, such as chainsaws, outboard motors, recreational vehicles for off-road use, etc. As an example, two-stroke engines for outboards, portable machines, and apparatus are increasingly used; their total loss lubricant systems coat the leaves of vegetation with a fine film of lubricant and also deposit reactive additives into the environment “Biodegradability” is the ability of a substance to be degraded in the environment by natural events, be they solar radiation, oxidation, or biological degradation. These “events” or “conditions” are clearly very variable; therefore standard conditions must be established and strictly adhered to. The issues are what those test conditions should be, the extent of degradation, and over what time period the test should be conducted. An additional complication is that the experimental basis on which the biodegradability tests are based were for single substances. However, lubricants are both complex formulations and also variably degraded materials. There are at least six versions of biodegradability tests [19]. These arise for different conditions, different types of materials, and for different continental regions. As these tests are biological procedures, then variable data can result. Establishing a reliable biodegradability test program for a company is a very formidable undertaking in terms of resources. A cost-effective approach is to use an established, accredited, independent laboratory with a proven track record in biodegradability measurements once the choice of test and its parameters have been agreed.
32.6 The Environmental Future and Lubricants 32.6.1 Annual Statistics and Used Arisings in the United Kingdom The total mass of new lubricants produced and sold in the United Kingdom is around ∼1 Mt/pa, with the automotive and industrial ratios of approximately 3 : 1. The long-term consumption of all forms of lubricants is slowly diminishing as a result of longer vehicle service intervals and higher lubricant quality. The waste statistics for lubricants are “unreliable” with high, 50% of mass, “lost” in use by combustion, leaks, and seeps, inappropriate disposal and “tipping,” giving a maximum total U.K. waste lubricant mass of ∼500, 000 tonnes per annum. This percentage of waste is lower than other developed countries, probably because of the extensive “do-it-yourself ” culture in car maintenance in the United Kingdom compared with the “do-it-for-me” culture in countries such as Germany and North America.
32.6.2 Trends in Waste Lubricant Arisings The global trends in waste lubricant arisings are driven by: • Longer service intervals, up to 50K km for light vehicles, already leading to reduced lubricant consumption • Much longer for freight vehicles, 400K km envisaged in future (moving toward “fill-for-life?”) • Lower sump oil volumes, but • Higher levels of contamination by soot particulates, additive degradation products, and much higher levels of PolyNuclear Aromatic Hydrocarbons (PAHs), some of which are carcinogenic
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Environmental Implications and Sustainability Concepts for Lubricants
32-29
The main Health and Safety, etc., problems (in addition to the minor problems of fresh lubricant) are PAHs. Increased service intervals sharply increase PAHs at more than 6K (gasoline) miles. This has immediate implications for operators/mechanics dealing with extended use lubricants and “enhanced personal care protection” is required for anyone handling these used lubricants. PAHs are removed from used lubricants by the separation processes described previously, the vacuum distillation/hydrogenation process being the most effective at removing PAHs by far. An additional solvent extraction step can give up to >99.7% PAHs for the vacuum/hydrogenation/solvent extraction process. Whilst this is beneficial for the base oil product quality, PAHs do not disappear and collect in the residue and are part of the disposal problem. The main disposal route up until 2003/5 is the “soft option” of “clean-up” (to various degrees), then either: • Use as “cleaner” fuel in power stations/boilers • Use as “dirtier” fuel in cement production Separation processes can produce good quality base oils for recycling into new formulations but also necessarily produce toxic residual sludges which are difficult to dispose of, difficult to be acceptably incinerated, or sent to landfill as “Special Waste,” which is increasingly very expensive.
32.6.3 The Environmental Business Economics of Recycling Lubricants The business economics of recycling used lubricants depend upon various factors such as the internal economics of lubricant production, the external economics of dealing with the environmental costs of used lubricants, the acceptability of recycled materials as lubricants, and the relative cost of virgin base oils. Recycling used lubricants is currently a marginally viable “business” overall which is: • Good for large sources such as bus, freight, and train companies using single formulations of lubricants. • Not good for recycling collection from individual vehicles using a range of lubricants unless a recycling surcharge/levy is applied. What form that might take is very arguable but a useful analogous example might be the beverage container tax used in many States of the Union; this has the benefit of being fiscally neutral if the user acts in an environmentally responsible way. • At what level of administration and by which authority should the responsibilities for disposal rest? At present it is the “local authority” if placed within its recycling facilities. Otherwise, the general “Duty of Care” for “Special Wastes” applies if the used lubricant is treated as a waste. The emerging responsibility is moving toward the “vendor” for the “end-of-life” disposal of lubricants, which requires the retailer to take back an equivalent volume of used lubricant when purchasing fresh lubricant, already an emerging issue in the United Kingdom, established in some nation states of the EU. The Packaging Regulations (EA, 95, etc., made under an EU Directive) cover the current packaging of lubricants. From any consideration of lubricant packaging it is clearly evident that it is suitable for purpose, is not overpackaged, is secure and safe. Therefore, the current packaging of lubricants will probably not be affected by the enhanced requirements of the Packaging Regulations. But in future the reuse of packaging through collection and return of containers for refilling will become more important. The quality of the content delivered then becomes an issue. The EU is moving toward the manufacturer’s responsibility for disposal of all contents of vehicle when scrapped, which includes hydraulic fluids and engine lubricants.
32.6.4 Barriers to Acceptance of Recycled Base Oils Whereas reduced friction lubricant and hydraulic formulations contribute to energy efficiency, their major contribution to sustainability is through extended drain and use periods and their successful recycling to produce good quality base oils for reformulation.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
A problem in certain countries is a resistance to the acceptance of recycled products of all forms, regarded as being “inferior” in some way. However, in Western and Central Europe, particularly Germany, recycled materials of proven quality are accepted. Two major manufacturers accept the use of recycled base oils in lubricant formulations for use in their vehicles, subject to low PAH levels in the final formulation. In the final analysis, the vehicle manufacturers carry the responsibility of the vehicle warranty. The barriers to acceptance of oil products containing recycled base oils can be overcome by consumer information and education, to the level of acceptance in Germany.
32.6.5 Relative Costs Ultimately, if environmental/quality/safety issues can be met, then the relative costs of recycled/virgin base oils is the arbiter. The recent rise in crude oil price, stabilizing at much higher level, 40$+/bl (early 2005) for light sulphur crude petroleum is due to: • • • •
Continuing unrest in producing countries Business/political chaos in Russia Industrial action in some countries and political uncertainty in others The extraordinary economic development of China and India, drawing in vast amounts of oil products
The net result is that the crude oil production/consumption balance is now so finely balanced that uncertainty raises crude prices further, which quickly passes through into yet further oil product prices. This gives an increased demand for Gp. I oils but there is a global shortage, which is ideal for recycled lubricant base oils to supply.
32.6.6 Conclusions This chapter has concentrated upon the environmental developments relating to lubricants in the EU and in the United Kingdom as an example of how legislation is framed for effective implementation in individual nation states. But the general direction is the same for all countries, more so for developed countries and less for others. Comparison of environmental law across continents shows how fundamentally similar the underlying purposes of these laws are, even if their procedures of enforcement may appear to differ according to the legal processes of each country. The message is clear — the unlicensed disposal of used lubricants is an increasingly severe offense and that recycling of good quality base oils is now a required operation. This will become an important contributor to the formulation of new lubricant products.
References [1] 1987—Report of the World Commission on Environment and Development, “Our Common Future,” chaired by the ex-prime minister of Norway, GroHaarlem Brundtland. [2] 1972 — The United Nations Conference on the Human Environment (UNCHE), Stockholm, Sweden. [3] 1992 — the“Earth Summit”(the UN Conference on Environment and Development) in Rio de Janeiro, Brazil. [4] 1997 — A special UN conference reviewed implementation of Agenda 21 (as Rio+5). [5] 2002 — Ten years after the Earth Summit in Rio in 1992, the 2002 Johannesburg Summit. [6] “Towards Sustainability,” EU Fifth Environmental Action Programme, 1993. [7] The EU Treaty of Amsterdam, 1999. [8] “Environment 2010: Our Future, Our Choice,” EU Sixth Environmental Action Programme (2001– 2010), 2001. [9] “A Better Quality of Life,” the U.K. Government Strategy for Sustainable Development, 1999.
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Environmental Implications and Sustainability Concepts for Lubricants
32-31
[10] The U.K. Local Government Act, 2000. [11] Sulphur Content of U.K. Fuel Oils, “Pollution Handbook 2004,” National Society for Clean Air, U.K., ISBN 0903 474 58 7. [12] CEC/93/EF13. [13] Daimler-Chrysler “HighTech Report,” 2/2003, p. 35. [14] “Responsible Care,” the U.K. Chemical Industries Association (http://www.cia.org.uk/industry/ care.htm), 1989. [15] “The Jost Report,” a U.K. Government committee report, chaired by Peter Jost, HMSO, Lubrication, Tribology, Education and Research, DES Report, London, U.K., 1966. [16] Definitions of “waste” and “controlled” waste for regulatory purposes, the U.K. Environmental Protection Act 1990 (EPA 1990) and the Controlled Waste Regulations (1992) made under that Act, “Pollution Handbook 2004,” National Society for Clean Air, UK, ISBN 0903 474 58 7. [17] EU Waste Incineration Directive 2000/76/EC. [18] EU Council Directive on the Landfilling of Waste, 99/31/EC. [19] See, for example, S.13.8 by C.I. Betton in “The Chemistry and Technology of Lubricants,” R.M. Mortier and S.T. Orszulik, Eds, Blackie Academic and Professional (VCH Publishers in USA and Canada), 1st ed., 1992, ISBN 0 7514 0117 X, 1 56081 594 9 in USA and Canada.
© 2006 by Taylor & Francis Group, LLC
33 Lubrication Program Development and Scheduling 33.1 33.2 33.3 33.4 33.5
Mike Johnson Noria Corporation
Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Machine Criticality and Operating Environment . . . Data Collection Strategies . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Machine Data Collection Criteria . . . . . . . . . . . . . . . . . . . . Lubricant Type, Quantity, Frequency, Application Method, and Time Stamp Decisions . . . . . . . . . . . . . . . . . 33.6 Contamination Control Requirements . . . . . . . . . . . . . . . 33.7 Oil Analysis Requirements . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33.8 Activity Sequencing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33.9 Planning and Scheduling Management . . . . . . . . . . . . . . 33.10 Conclusion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
33-1 33-2 33-3 33-5 33-6 33-18 33-22 33-25 33-28 33-28 33-29
33.1 Introduction Effective machine relubrication practices are as critical to the practice of reliability engineering as proper shaft alignment and component balancing. However, while misalignment and imbalance may reveal themselves through the outward symptoms of elevated temperatures, elevated vibration, and loud noises, the symptoms of poor lubrication are often imperceptible. For many years the industry contended that if one provided enough of the right product at a reasonable frequency that the lubricated components would be sufficiently protected. While this may suffice for some low intensity operating environments, a “Best Practice” is justified for highly competitive businesses like Steel, Cement, Paper, and Automotive production. A properly devised “best practice,” will incorporate operation and machine specific requirements including: • Machine criticality and operating environment • Data collection strategies • Machine data collection criteria 33-1
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33-2
• • • • •
Handbook of Lubrication and Tribology
Lubricant type, quantity, and frequency requirements Contamination control requirements Oil analysis requirements Activity Sequencing Planning and Scheduling Management
This chapter will review the requirements associated with effective machine lubrication practice development.
33.2 Machine Criticality and Operating Environment Within any given production facility there are variety of machines with a variety of responsibilities. Some of those machines will be built to withstand physical stresses several times greater than the actual production processes will impose. The extent of extra capacity as gauged by the machine designer would be characterized as “service factor.” There are differing philosophies as to how much “extra capacity” should be incorporated into a machine design, but it is generally accepted that the greater the “service factor,” the longer the machine would be expected to last between required rebuilds. Of course, higher service factors required greater capital investment. As financial managers became more sophisticated in measuring capital usefulness, engineering departments began to consider whether the extra capital invested would ever be released in the form of higher productivity, and then began to squeeze the “service factors” toward a minimum acceptable level. As service factors fall, the relative care and attention that should be applied to the machine will necessarily rise in order to sustain equivalent production. The amount of resources dedicated to machine care must be allocated according to that machine’s importance to the production process, and as the importance ranking increases the resource allocation should also increase, as illustrated in Graph 33.1. The process for grading the machine’s importance to the production environment is called Criticality Assessment. An effective criticality assessment will consider various measurable parameters for each machine. Assessment parameters could include several factors, such as: • • • • •
Machine Function Hourly value of machine function Machine failure risk to employees Machine failure risk to the environment Machine failure risk to production quality
Importance to production
1000 800 600 400 200 0 $ 100
$ 350
$ 600
$ 850
$ 1,100
Annual resource allocation per asset
GRAPH 33.1 Criticality ranking and resource allocation.
© 2006 by Taylor & Francis Group, LLC
Lubrication Program Development and Scheduling
• • • •
33-3
Machine durability (Rate of failure, Mean time between failures, Standard deviation, etc.) Speed of failure Cost of machine repair Operating environment severity
The operating environment will also influence the degree of detail that the relubrication procedure should entail. Machines operating in a benign environment, such as the environment in an automotive production facility paint shop, will require less attention for contamination control features than would be necessary for equipment operating a steel mill melt shop. Operating environment factors that should influence lubrication plan development include things such as: • • • • •
Atmospheric temperature Atmospheric humidity Atmospheric dust and dirt exposure Machine direct exposure to production chemicals, moisture, or solid contaminants Machine stability (vibration from external sources)
If a machine has both a high criticality ranking and a high environmental influence factor, then the relubrication practice will need to incorporate measures to adjust frequencies, volumes, and application methods, as well as measures and modification plans to minimize the risk that the environment provides to machine health. A thorough criticality assessment method should incorporate a factor for operating environment. The site reliability engineer will be able to provide guidance for the amount of detail that each machine requires based on the current criticality standard.
33.3 Data Collection Strategies Once the list of machines has been established and prioritized from most to least critical, the MLT (Machine Lubrication Technician) may begin to focus on the task of collecting the myriad of details that are required to devise accurate practices. From the outside of the facility it may seem to be a whale-sized task, but with the help of an orderly and systematic approach, the whale can be eaten — one bite at a time. Each asset must be observed, and the technical details of each characteristic must be recorded as completely as possible. Some of the details may reside in a computerized maintenance management system (CMMS) and may be quickly accessed. It is highly likely that regardless of the state of CMMS development the details required for lubrication practice development will be incomplete. Common details may reside in the CMMS, or in an Original Equipment Manufacturer’s (OEM) operations and maintenance records. For instance, a machine’s motor builder, size, type, horsepower ratings, and other similar details may be reflected in these records. Over time, however, motors are repaired, or moved to other service areas, and the original record may not reflect the “as is” state. With that expectation in mind, it is necessary to conduct a physical review of each asset to verify existing information and to supplement with (as built) details. There could be any number of ways to organize the data collection sequence. Batch process: Environments that are batch oriented sometimes expand in a piecemeal fashion. Figure 33.1 depicts a production setting where the layout location of the production process is not sequential. In this environment one can collect drive train data using location markers in the plant as a sort of geographical reference point. Continuous process: One approach to organize the data collection exercise is to follow the flow of raw materials through the conversion process. Figure 33.2 shows a sequential view of the major processes for manufacturing Portland cement. Each individual section will contain drive trains operating in tandem or in sequence to the other drive trains. Capturing drive train lube sump details by following the material conversion process also helps with the organization of lube points for a route plan.
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Handbook of Lubrication and Tribology
1 Office area
2 A
G 3
B
F
D
C
H
E
4 5
Raw materials storage
FIGURE 33.1
Raw materials storage
Raw materials storage
Batch plant layout diagram.
Kiln Cooler
Raw materials storage
Raw mill Finish mill #1 Shipping Finish mill #2 Rotary dryer
Compressor room
Crusher and screen
FIGURE 33.2
Portland cement mill major process overview.
Alternatively one may visit each product cell wherever it may exist in the plant layout until all of the drive trains have been reviewed. This process will certainly work, but it does require a higher level of organization to keep the records straight, and will likely require revisiting each machine when it is time to prepare efficient relubrication routes. Planned availability (production demand flow): Where there is a clearly defined risk to personnel if the machine is operating, it may be necessary to either catch the machine when it is idle, or idle the machine at the point that the surveyor is ready to review the compliment of components. An electromechanical robot would fall into this category.
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Lubrication Program Development and Scheduling TABLE 33.1
33-5
Components that are Common to a Simple Conveyor
Items
Make
Model
No.
Components
Type
Drive train — conveyor A1 Motor GE
5KS511SN3260HB
1
2
Bearing
SKF
90BCO3JP3 75BCO3JP3
Coupling Gear reducer
Falk Falk
1080T 2145 Y2B
1 1
2
Seal Backstop
Falk Falk
Type K PRT 65
Coupling Head pulley
Falk
1140T
1 2
Bearing
SKF
2 2
Bearing Bearing
SKF SKF
3 2
Bearing Bearing
Timken Koyo
22234CCK/W33 22234CCK/W34 22230C/W35 22230C/W36 22230C/W38 LM 11949 UCF208-24
Snub pulley Tail pulley Trophing rollers Return roller bearings
Cont. Conv Cont. Conv
78AH650-66G 78AH650-66G
192 48
Make
Model
Note: Each individual lubricated component should be accounted for.
TABLE 33.2
Type of Details that Should be Collected Whenever Possible
Component
Make
Model
Gear sump
Falk
2145 Y2B
Details Type of reductions Number of reductions Input speed Reduction ratio Bearing types Bearing sizes Stated AGMA lubricant grade Stated AGMA lubricant type Sump temperature Oil distribution method (bath or circulation) Presence of filtration Presence of vent breather Number and size of drain port openings
33.4 Machine Data Collection Criteria 1. The lubrication survey entails collection of information on every lubricated component on every drive train within the production system. The specific detail that must be collected depends on whether the machine is oil or grease lubricated. Table 33.1 describes the type of component by component detail that is required to build an appropriate practice. The specific detail that must be collected depends on whether the lubricated component requires oil or grease. 2. Oil based lubrication denotes a fluid sump. The sump may or may not require sophisticated sealing materials to exclude contaminants, fittings to accommodate oil sample collection, and fittings to accommodate contamination control requirements. Two identical machines may have appreciably different relubrication requirements at the time of data collection based on machine criticality, as previously discussed, but that criticality requirement may change. Table 33.2 offers a suggested level of detail for the collection of gearbox sumps. A similar level of system specific details is warranted for bearing, coupling, hydraulic, or circulation sumps. Over time increased product demand can influence the criticality assessment, causing the criticality factor to both rise and fall. If the details collected are only sufficiently detailed enough to cover the
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33-6
Handbook of Lubrication and Tribology TABLE 33.3 Grease Lubricated Sumps Require Collection of Additional Details Component
Make
Model
Bearing
SKF
22234CCKW33
Details Component type Component size Shaft speed Static rating (C/P — for bearings) Oil or grease lubricated Degree of seal (shielded, sealed, open) Seal type Relubrication method (manual vs. automatic) OEM designated lubricant OEM original fill lubricant type if any OEM designated lubricant volume Operating temperature Operating atmosphere — moisture Operating atmosphere — temperature Operating atmosphere — dust or dirt Operating atmosphere — process chemicals
immediate need then the process will unnecessarily have to be conducted again at a later date. Sufficient detail should be collected to enable the development of a “best practice” even if all of the details are not used immediately. 3. Grease based lubrication denotes a sump that requires systematic refreshment of the lubricant in the sump. The relubrication frequency may be long or may be short, depending on how the machine is operated. As the speed and load increases and the ambient environment becomes more severe the calculated relubrication interval (time between relubrication events) decreases. Relubrication of greased components requires a greater dedication to detail than is typical for oil lubricated components. Table 33.3 offers a suggested level of detail that should be captured for bearings that are either oil or grease lubricated. Similar details should be captured for other grease lubricated components. Table 33.4 proposes general guidelines to follow when collecting information on other types of components. It is not possible in the pages of this chapter to redress data collection parameters for the myriad of types of machines and lubricated components, so the general advice offered is intended to stimulate ideas for a creative pursuit of the maximum level of detail that can be collected in a timely matter.
33.5 Lubricant Type, Quantity, Frequency, Application Method, and Time Stamp Decisions A wide variety of operational factors will influence the final product selection decision. The factors will be different from industry to industry, and within any given industry from plant to plant, and within the plant, from department to department. Environmental factors and influences should be addressed following review of the fundamental component requirements, assuming that each component was properly sized and constructed from appropriate materials for the intended application. Product selection will also influence necessary relubrication volume and frequency decisions. Again, there can be many variables. A safe place to begin with any product selection decision process is with the fundamental engineering units as denoted by the lubricated component manufacturer. While there may be misgivings about the quality of some OEM guidance, this is none the less, the best place to begin this process. Content in the previous section covers lubricant selection, quantities, and application methods for various lubricated machines and mechanical components. It would be impractical to rehash those details
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Lubrication Program Development and Scheduling TABLE 33.4 Component
33-7
Miscellaneous Factors for Consideration Make
Model
Details
Sumps
Oil or grease lubricated Degree of seal or shielding Seal type if known Relubrication method (manual vs. automatic) OEM designated lubricant OEM original fill lubricant type if any OEM designated lubricant volume Operating temperature Operating atmosphere — moisture Operating atmosphere — temperature Operating atmosphere — dust or dirt Operating atmosphere — process chemicals Type of lubricated components Size of lubricated components Rotating speed of lubricated components, if any
Surface coated components
Contact surface length Contact surface width Surface area Surface velocity (speed of surface movement) Application method Operating temperature Operating atmosphere — moisture Operating atmosphere — temperature Operating atmosphere — dust or dirt Operating atmosphere — process chemicals
in this short section, but the thought process should be reviewed to see how the decisions will be used in the course of developing an effective machine lubrication practice. This section provides very specific advice on plain and element bearings, and only general guidance on where to find the information covering other component types, including element and plain bearings, gearing, couplings chains, and cables. These individual components may be assembled into a wide variety of operating systems, but the fundamental questions regarding oil film thickness and film type will always lie at the heart of the product selection decision. Bearings: All machines that have moving parts will have bearings of some type. The bearings may be as simple as flat surfaces mating with flat surfaces, such as the slideway in a machine tool, or may have sophisticated geometries, such as is the case with a ball screw. Plain and element bearings supporting rotating shafts are most common, and are found in nearly every machine one might imagine. 1. Product type selection for plain and element bearings. (a) Plain bearings, also called journal bearings, as shown in Figure 33.3 support and constrain the motion of a rotating shaft. Figure 33.4 shows different styles of plain bearings that may exist for both grease and oil lubricated conditions. The principal forces applied to the shaft are typically radial (perpendicular to the axis of the shaft) but may also be axial (in the same direction as the axis of the shaft). Oil film formation is said to be “hydro-dynamic” in nature, and film formation is achieved when the oil accumulates at the contact point between the shaft and the bearing, forcing the shaft to float on the accumulated oil. Since there is no expectation of metallic contact during normal operating conditions the OEM and equipment supplier would likely recommend R&O (Rust and Oxidation inhibited) circulating oils for oil bath and circulation systems. Some
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33-8
Handbook of Lubrication and Tribology Oil inlet
Housing
Bearing liner
Clearance
Split type (some)
FIGURE 33.3
Housing Bearing liner Segment (split type) Oil inlet Drain Journal
Common journal bearing components.
• Journal bearings support and constrain rotating motion subject to radial loading • The principle force acts perpendicular to the axis of the shaft • The journal is the section of the shaft that rests on the bearing • Sufficient oil viscosity and shaft speed needed to maintain hydrodynamic oil film
FIGURE 33.4
Journal or shaft
Common journal bearing components
Journal bearing design Sleeve or bushings
Multi-part
Split Half bearing
(Noria Corporation. Machinery Lubrication Level II Training Seminar, Slide G2103)
manufacturers suggest the use of “compounded” oils that are fortified with metal wetting additives to enhance the lubricants tenacity and surface protection capacity. Plain bearings are typically oil lubricated, but where slow turning shafts and heavy unit loads create conditions where the pressure wedge may not remain stable, grease may be selected. Guidance provided by Lansdown suggests limits for grease lubrication of plain bearings is based on a limit of 2 m/sec or 400 ft/min of linear shaft speed based on the surface speed of the shaft. This equates to 400 RPM for a 5 cm bearing; less than 50 RPM for an 80 cm bearing. This is to limit churning of the grease that would ultimately lead to grease and bearing failure [1]. Extreme Pressure (EP) agents are not typically specified for grease lubrication of plain bearings. Solid film antiwear (AW) additives, such as graphite and molybdenum, are particularly useful for these conditions, and should be considered to help to minimize the extent of wear given the higher rate of metal contact. Most greased bearings are once-through applications. As such, there is little risk that the grease will be subject to long-term oxidation stress that oil lubricants will face. Strong EP and AW chemical additives are not considered to be particularly helpful where soft or “yellow” metals (bronze or babbitt) are in use. A given amount of heat energy must be generated from friction to initiate the reaction between the additive and the metal surface. With steel on bronze, or steel on other soft metals, scoring occurs rapidly enough that the chemical EP and AW additive cannot effectively perform the assigned task. Additionally, these additives tend to generate acids during decomposition that may be cause corrosion. The same consideration applies to lubricant selection for oil lubricated bearings.
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Lubrication Program Development and Scheduling
33-9
Life adjustment factors relating to operating conditions and bearing material
A23
(b) Product type selection for element bearings is fairly well defined by bearing manufactures, for both oil and grease selection. The dominant lubrication mode for element bearings is elasto-hydrodynamic (EHD), or “thin film” condition. EHD conditions provide for complete separation of interfacing surfaces, but with separation ranging from .5 to 3 µm for ball and roller type bearings. Given that the surfaces are intended to be separated, albeit with a very thin layer of oil, the manufacturers suggest the use of R&O inhibited mineral oils for lightly loaded bearings and oils with surface reactive (AW and EP) additives for those applications where normal operation carries a high risk of film collapse. It is possible to closely estimate (calculate) whether a given bearing in a given set of conditions requires an EP fortified oil or grease through the use of a film thickness ratio called K or Kappa factor (K ). This is the ratio of actual viscosity at operating temperature divided by the bearing manufacturers required viscosity at operating temperature. Graph 33.2 shows a viscosity ratio range for which EP additives are highly recommended. Table 33.5 offers recommendations for oil types based on bearing manufacturer general guidelines for meeting film thickness requirements, and based on practical experience for maintaining lubricant health.
K=
5 Possible need of EP oils when Kappa is less than 1
2 1 0.5
Maximum bearing life at Kappa of 4
0.2 0.1 0.05 0.05
0.1
0.2
0.5
1
2
5 10 K = v/v1
Viscosity at operating temperature Required viscosity for the bearing
Kappa of 1 – 2.5 is desired
GRAPH 33.2 Noria Corporation. Machinery Lubrication Level II Training Seminar, Slide 1561. Source: Lubricants in Operation, U.J. Moller and U. Boor. 1996. © John Wiley & Sons Limited. Reproduced with permission. TABLE 33.5 General Guidance for Selection of Lubricant Type for Element Bearings Proposed oil type
Element bearing P/C
Viscosity Kappa
Operating temperature ◦C
I, MO I, AW, MO I, AW, S EP, MO EP, S
<.15 <.15 <.15 >.15 >.15
>4.0 1–4 1–4 <1.0 <1.0
10–80 10–80 <10a,b 10–80 >70a
Note: a Sustained Operating Temperature. b Risk of startup at or below the low level is high. I = Inhibited, AW = Anitwear, EP = Extreme pressure, MO = Mineral oil, S = Synthetic, P = Actual radial load, C = Bearing load rating.
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Pressure (psi)
1400
PM
Calculation of journal bearing pressure:
60 0R
1600
1800
1800
RPM 1200 RPM 900 RP M
Handbook of Lubrication and Tribology
3600 RPM
33-10
Pressure (psi) =
W LD
where:
1200 1000
0
30
800
w = Load on shaft (lbs) l = Axial length of bearing d = Diameter of shaft (inches)
M RP
600 400
100
RPM
Calculation of thrust bearing pressure: Pressure (psi) =
200
LD
where:
0 Viscosity (cP) at operation temperature
0.4W
20
40
60 cSt =
80 100 120 140 160 cP Oil specific gravity
w = Thrust load (lbs) l = Width of bearing ring (inches) d = Average pad diameter (inches)
GRAPH 33.3 Noria Corporation. ML II, Slide 1563. Source: Tribology Handbook, 2nd Edn Neale, Michael, Page C7.2. Reproduced wih permission. TABLE 33.6 Viscosity Extract from ISO Viscosity Classificationa (ISO 3448) for Sliding Bearings at a Working Temperature Range of 15–60◦ C Bearing pressureb Rotation speed range (min−1 ) 5000–10,000 2000–5000 1000–2000 500–1000 300–500 100–300 50–100 <50
Light (<7 bar) (<70 N/cm2 ) ISO VG: 10 ISO VG: 15 ISO VG: 22 ISO VG: 32,46 ISO VG: 68,100 ISO VG: 100,150 ISO VG: 150,220 ISO VG: 220,320
Medium (7–17 bar) (70–170 N/cm2 )
ISO VG: 32,48 ISO VG: 68,100 ISO VG: 100,150 ISO VG: 150,220 ISO VG: 220,320 ISO VG: 320,460
Heavy (>17 bar) (>170N/cm2 )
ISO VG: 320,460 ISO VG: 460,680 ISO VG: 460,680
Note: These recommendations are for circulating lubrication; thicker oils are needed for loss lubrication. a ISO VG: kinematic viscosity in mm2 /sec at 40◦ C. b Bearing pressure p = F /b · d. N Source: Lubricants in Operation, U.J. Moller and U. Boor. 1996. © John Wiley & Sons Limited. Reproduced with permission.
2. Viscosity selection for plain and element bearings. Correct viscosity selection is the single most important factor in correct lubricant selection and program design. For all types of bearings and lubricated mechanical components, for both oil and grease lubricated applications, the machine’s actual operating temperature will dictate which viscosity grade of the selected product will be necessary. (a) For plain bearing applications Graph 33.3 shows the relationship between shaft speed, oil viscosity, and shaft loading (PSI) that is useful to determine the minimum oil viscosity at operating temperature. Once the viscosity target is identified there are two approaches that the
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Lubrication Program Development and Scheduling
33-11
GRAPH 33.4 Viscosity at Operating Temperature. Source: SKF General Catalog 5000F. 2003. Reproduced with permission.
practitioner could use to determine the correct starting point viscosity selection as measured at 40◦ C [2]. Table 33.6 shows an alternate view of the same relationship [3]. Observing the following steps, the practitioner may use Graph 33.4 to determine the viscosity starting point (which viscosity at 40◦ C to select) [4]. Step 1: Identify the required viscosity point on the vehicle axis and draw a line from left to right across the graph. Step 2: Identify the operating temperature on the horizontal axis and draw a line from bottom to top across the graph. Step 3: Find the point where the two lines cross. Step 4: Identify the next higher viscosity point as the minimum viscosity starting point for the select application. This chart is designated for paraffinic mineral oils since it reflects viscosities with VI values at 100. It would be simple to create another chart reflecting the viscosities for either higher or lower VI values. (b) Viscosity Selection for element bearings. There are sophisticated equations that may be used to calculate the actual film thickness for a given set of design parameters. These equations are not easily adapted to actual conditions, and as such bearing manufacturers have provided short-cuts that enable accurate selection with easy-to-follow guidance.
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33-12
Handbook of Lubrication and Tribology Element Bearing Viscosity Selection Criteria
n1 1000 mm2/sec (cS )
4600 2
500
2300
5 10
Viscosity (mm2/sec)
n1 SUS
20
200
930
50
460
100 10
0
20
0
50
230
50
0R
PM
10 15 00 20 00 30 00 0 50 0 00 10 00 0 20 00 0 I
20
10
50
5
100
II
60
40
00
10
Approximate saybolt universal seconds (SUS)
TABLE 33.7
0
00
3 10
00
20
35 50
100
200
500
Pitch diameter (mm)
1000 dmmm
This is a three step process, as follows: Step 1. Use Formula 33.1 to estimate bearing Pitch Diameter, dm , dm = (QD + ID)/2
(33.1)
where QD is bearing outer diameter and ID is bearing bore. Step 2. Determine shaft rotation speed. Step 3. Using Table 33.7, locate the intersection between the shaft speed and a line intersection with the calculated dm value. Step 4. Draw a line to the Y axis to read the required viscosity in centistokes (mm2 /sec). Step 5. Determine the correct viscosity at 40◦ C in the same manner as noted above. The viscosity estimated in the process noted above is the minimum acceptable viscosity for the given conditions to maintain a “fat” oil film in an element bearing. Bearing manufacturers propose that any time the film ratio falls below this level then wear resistant and seizure resistant additives be incorporated to protect surfaces [5]. It is advisable to provide a cushion when making the final oil selection given that change in mechanical, environmental, or production conditions may cause shock loading. The degree of cushion selected should be determined by the type and operation of the equipment,
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Lubrication Program Development and Scheduling TABLE 33.8
33-13
Maximum Bearing Speed nDm Factors
Bearing type
Oil lubricated
Grease lubricated
500,000 500,000 290,000 280,000
340,000 300,000 145,000 140,000
Radial ball bearings Cylindrical roller bearings Spherical roller bearings Thrust-ball and roller bearings
Source: Lubrication Fundamentals, Wills George. 1980. © Mobil Oil Corporation. Reproduced with permission.
by the general speed of the equipment, and by consideration of other lubricants that may already be available in the work area. A manufacturer’s general purpose (GP) greases commonly have viscosities ranging between 100 and 220 cSt, depending on these factors. Keep in mind that thicker oils, and thicker grease consistencies, tend to churn, generate heat, and consume energy. (c) Special consideration for grease lubricated bearings. There are speed limits to the effective grease relubrication of all element bearings. The factor used to gauge the decision to use either grease or oil is called the nDm, which is shown in Formula 33.2. nDm =
(ID + OD) × N 2
(33.2)
where ID is bearing bore, OD bearing outside diameter, and N is shaft speed. Large bearing sizes and high shaft speeds create high nDm values. As the nDm value increases the extent of churning and thickener degradation increase relative to bearing element type. Since each grease will respond differently to the “working” effect that the element produces, and since some greases will soften and some will harden, it is impossible to determine short of conducting laboratory tests for each grease type, just how well any given grease will withstand work shear and consistency changes. Bearing element type has an influence on grease life, as shown in Table 33.8. As the calculated nDm value approaches the limits proposed by Table 33.8, the practitioner will need to increase the relubrication interval and be particularly cautious to follow rigorously defined feed rates to prevent churning and lubricant destruction [6]. 3. Lubricant supply volume for plain and element bearings. Following selection of the type of lubricant at the appropriate viscometric range, the next exercise is to select the correct volume of lubricant for replacement. As this section deals with developing relubrication practices, the focus will be on replenishment rather than design criteria for sump capacities and distribution system flows. (a) Oil replenishment volumes are entirely dependent on oil loss through operation, leakage, contamination, or normally scheduled “bleed and feed.” Replenishment volumes are dictated by sump capacity requirements. Bearing oil sumps have some mechanism for indicating the normal sump capacity. In most cases there is an external indicator such as an oil bowl, level indicator or gauge, or dipstick. The lube technician should specify the nature and location of the indicator, and unit of measure if it is not evident from physical observation, but since oil loss is machine dependent, and since machines are not designed to relieve a given quantity of oil through normal use, it is not possible to project before the fact the quantity of oil to be used in replacement. (b) Grease replenishment volumes are an altogether different story. Grease lubricated applications are nearly always expected to be“continuous loss”systems. As such, planned, consistent, systematic replenishment is necessary to protect the grease in the machine and the machine itself. • Grease replacement volume for plain bearings is based on the required resupply per hour of operation.
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Handbook of Lubrication and Tribology TABLE 33.9 kg Factor Based on Shaft Speed Shaft speed rev/min
kg
Up to 100 101–250 251–500 501–1000
0.1 0.2 0.4 1
TABLE 33.10 Grease Relubrication Interval Factors for Changing Environmental Conditions Condition
Average operating range
Correction factor
Temperature Ft
Housing below 150◦ F 150 to 175◦ F 175 to 200◦ F Above 200◦ F
1.0 0.5 0.2 0.1
Contamination Fc
Light, nonabrasive dust Heavy, nonabrasive dust Light, abrasive dust Heavy, abrasive dust
1.0 0.7 0.4 0.2
Moisture Fm
Humidity mostly below 80% Humidity between 80 and 90% Occasional condensation Occasional water on housing
1.0 0.7 0.4 0.1
Vibration Fv
Less than 0.2 ips velocity, peak 0.2 to 0.4 ips Above 0.4 (see note)
1.0 0.6 0.3
Position Fp
Horizontal bore centerline 45◦ bore centerline Vertical centerline
1.0 0.5 0.3
Bearing Design Fd
Ball bearings Cylindrical and needle roller bearings Tapered and spherical roller bearings
10 5.0 1.0
Formula 33.3 used in conjunction with Table 33.9 may be used to calculate the grease replacement volume, as follows [7,8]: Qg = kg × Cd × π × d × b
(33.3)
where Qg is grease volume per hour, kg is rotation speed factor (Table 33.9), Cd is diametrical clearance, π = 3.14, d is shaft diameter, and b is bearing width. Formula 33.3 should be considered as a reasonable quantitative starting point, but operating conditions should influence actual feed rates. The service factors noted on Table 33.10 could also be used to adjust the time (1 h) for the calculated quantity by multiplying the time value in hours by the actual operating condition factors [9]. This volume should be uniformly distributed during the course of the final time cycle, or to the extent that program management can allow. If the calculated quantity was 6 g/h, then 1 g per each 10 min would be better than 6 g per each 60 min. Continuous feed is desirable.
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Lubrication Program Development and Scheduling
33-15
• Grease Replacement Volumes for Element Bearings. Estimating replacement volumes for element bearings is somewhat simpler to estimate. It should be restated again that the calculated estimates are based on sound design engineering principles, and must be subject to adjustment for “as built” design and operating conditions. An SKF formula, Formula 33.4, is a simple and convenient starting point for estimating grease replenishment quantities for element bearings [10]. Qg = D × B × 0.005
(33.4)
where Qg is grease quantity in grams per interval, D is bearing outside diameter (mm), and B is bearing width (mm) (0.114 may be use as the multiplier with english units to provide output Q in ounces.) Where actual bearing dimensions are not known, a close proximity to the actual suggested value could be estimated by using housing dimensions and factoring again by 0.33. This does not provide exactly the same value for all bearing types given bearing element and construction differences, but it is generally close. It is suggested that some comparisons should be done per bearing type per bearing manufacturer before proceeding uniformly with this factoring method. Lubriquip, Incorporated, Technical Bulletin #20115 provides formulas and concise direction for determining lubrication volume requirements for bearings and various other types of lubricated components [11]. • Relubrication Frequency. The basis for frequency for grease relubricated plain bearings is shown in Formula 33.3 which provides recommended nominal volume per hour of operation. Formula 33.5 provides a baseline that incorporates operating factors from Table 33.10 for vibration, solid and moisture contaminants, heat, bearing type, and shaft axis, each of which can have an negative impact on the lifecycle and effectiveness of the lubricant in use [12]. T =K×
14, 000, 000 n × (d 0.5 )
−4×d
(33.5)
where T is time until next relubrication (h), K is product of all correction factors Ft ×Fc ×Fm ×Fv ×Fp ×Fd (see Table 33.10), n is speed (rpm), and d is bore diameter (mm). Note: ips = in./sec, 0.2 in./sec = 0.5 mm/sec. SKF, Timken, and FAG Lubrication Guideline publications provide alternate quantitative approaches that are also valid, and could be considered as a strong reference starting point [13–15]. • Lubricant Application. There are various methods that may be selected to add the required quantity to the designated component. The high volume and short frequency applications will benefit most from some form of automatic lubrication supply. Semiautomatic (single point automatic lubricator), such as seen in Figure 33.5 or fully automatic (multi-point automatic lubrication system), such as seen in Figure 33.6, provide feed conditions that best meet reliability goals, particularly for high volume and short interval requirements. Most bearings, both element and plain type, can be covered to meet nominal reliability requirements with manual (grease gun) relubrication methods. 4. Time Stamp. A nominal value should be assigned to each practice type based on some type of time study. A detailed time study would consist of breaking the whole of the process completely down into individual movement, assigning a typical value for the movement from a standardized table, and the agglomerating the scores. This is useful for highly repetitive tasks. Machine relubrication tasks are repetitive in the sense that the same activities are repeated, but the scale of the repeated
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33-16
Handbook of Lubrication and Tribology
Transparent polycarbonate housing for viewing lubricant level Replaceable lubricant cartridge and battery pack
Positive displacement pump ejects lubricant at maximum 350 psi Solid-state programmable controller times output cycles
The MEMO contains the program which controls lubricant ejection
FIGURE 33.5 Cross-section of a programmable positive displacement single point lubricator. Source: Getting the Most from Single-Point Lubricators Luis F. Rizo, Elfeo Inc. Reproduced with permission.
Pumping station (pressurizes line intermittently)
Pressure gauge
Injectors Single-line injectors
FIGURE 33.6 Overview of a single line series automatic lubrication system. Source: Automated Lubrication-Benefit & Design Options, Way ne Mitchell/Machinery Lubrications Magazine. 2001. © Lincoln Corporation. Reproduced with permission.
activities makes this type of reference prone to errors. A more generalized time assessment can provide sufficient detail to support labor balancing and allocation by department. (a) Activities that are conducted as stand alone work orders, such as flushing a hydraulic system, should reflect the amount of time required to prepare for the activity, travel to the asset or group of assets, perform the task, travel back, and store materials for reuse at a later date. (b) Activities intended to be done in sequence with other similar activities, such as bearing relubrication, or topping reservoirs should have an allocation for preparation, travel to and from, and storing gear, but the allocation would be distributed into a series of individual tasks. Table 33.11 offers guidance for estimated fixed and variable time per procedure type. The fixed value suggests that any time this type of activity is scheduled that this amount of time should be allocated, regardless of the number of the type of activity. This could be considered a one-time charge any time that particular activity is scheduled. The “Time Per Task” value is a variable for each activity to be scheduled in a sequence in addition to the one time charge as noted above.
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Lubrication Program Development and Scheduling TABLE 33.11
33-17
Incremental Time Measurements for Common Lubrication Lubrication task time estimates
Task Top-up and inspection Sample (vacuum w/minimess) General bearing regrease Motor regrease Coupling regrease Small to medium sump drain and fill Small to medium sump drain and fill (w/ filter cart) Chain lubrication (aerosol) Open Gear Lubrication (aerosol)
Fixed time (min)
Time per task (min)
15 15 15 15 15 15 15 15 15
7.5 6 3 15 60 40 20 5 10
If the route required 2 bearing relubrication (greasing) and 1 reservoir change then the time tally for the route would be: Bearing time allotment = (1 × 15 min) + (2 × 3 min) = 21 min Small reservoir change allotment = (1 × 15 min) + (1 × 40 min) = 55 min Once the routes are rolled together these time stamps will play an important role in work balancing to assure coverage. Following the example for bearings, lubricant selection by type (R&O, AW, EP, Synthetic or Mineral Oil), viscosity (ISO Grade) and/or stiffness (NLGI Grade), and application rate, gear relubrication must receive a similar treatment in the process of designating type, thickness, quantity, and application method in order to develop a precise machine relubrication procedure. The following sections provide reference points that can be used to find the type of information that will be necessary to develop the requisite level of detail, but will not try to elaborate on the specific details themselves. Gears: 1. Enclosed Gearing. Excellent advice for selection of lubricant by base oil and additive type, and viscosity is provided by the American Gear Manufacturers Association (AGMA) Technical Bulletins ANSI AGMA 9005 E02 and AGMA 925-A03. These documents provide background and detailed knowledge for viscosity selection under various operating conditions. 2. Open Gearing. AGMA Technical Bulletin ANSI AGMA 9005 E02 provides background and detailed knowledge for product selection, product performance measurement methods, and volumetric application recommendations for open gear applications [18]. Couplings: 1. Excellent advice is available from most coupling manufacturers, including product selection, viscosity, product volume, and application methods per type of coupling. Again, AGMA provides very specific detail on coupling lubrication through Technical Bulletin ANSI/AGMA 9001 B97. This bulletin may be found at web reference [19, 20]. Chains: 1. Helpful chain lubrication information may be found at the American Chain Association document“Identification, Installation, Lubrication and Maintenance of Power Transmission Roller Chains in ANSI B29.1 and ANSI B29.3,” and “Fundamentals of Chain Lubrication” [21, 22]. Hydraulic Systems: 1. Hydraulic systems are composed of several subsystems that each require specialized consideration for appropriate product selection. The subsystems include the pump, valves, motors, cylinders and other working components, piping and sump. Each of these system components may have specialized lubrication requirements that may compete with other systems components based on the design, function, environment, and operation of the hydraulic system. Since the subsystems must function from within the broader context of the whole, some sacrifices may have to be made.
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33-18
Handbook of Lubrication and Tribology Causes of bearing failures
4.5
ve rlo 4. adi 2 ng
Corrosion 4.2
er Oth
O
Particle contamination 44.9
Insufficient lubrication 10.8
Disassembly 13.4
ent
nm alig Mis 12.7
GRAPH 33.5 Bearing failures by failure category.
Product type selections can be challenging, but typically resolve to the degree of wear resistance protection dictated by the main system pump, which typically require AW additives, and viscosity conditions dictated by the pump and working components. Additional attention must be provided for fluid filterability properties in the event that the control system operates with proportional servo control mechanisms. Particular concern must be given to the dynamic temperature range, particularly if the system operates intermittently and outdoors. There are a variety of useful, detailed technical resources available in either print or via web based electronic print, from a simple overview of system requirements to a sophisticated component by component and step by step method for evaluating fluid selection [23–25].
33.6 Contamination Control Requirements Contamination control may be the single largest area for improvement and financial reward in the field of maintenance and lubrication program management. Contamination control interests and technical contribution derive from interests in both the field of hydraulics, the early area of concern for fluid contamination, and tribology. Hydraulic system contaminants are one of the, if not the single most, biggest contributors to system failure. This position has been widely understood by industry experts for many years. Recently, with the advent of “Proactive” maintenance focus, a similar level of interest for contamination control has developed for circulating and noncirculating sumps for other types of mechanical components. The lubricant supplier strictly speaking has not had a voice in the discussion about fluid contamination control since this specialty has been the domain of hydraulics system component and hydraulic filter component manufacturers. It should none the less be part of the fundamental design of an effective machine lubrication care specification. Abundant evidence exists from leading component manufacturers that shows the relationship between contamination levels and lubricated component failures. Graph 33.5 (SKF Pie Chart) illustrates the relationship between bearing life and contamination control. Accordingly, the lubrication specialist should be involved in decisions for setting basic contamination control requirements for all systems, even those traditionally not considered candidates for advanced contamination control [26]. Aspects of contamination control improvement considerations fall into two categories: contaminant exclusion and contaminant removal. Actions associated with either will be driven by the reliability ranking
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Lubrication Program Development and Scheduling
Open vents
33-19
Poor, ingression prone vents
Good, protective ventilation Expansion chambers
Spin-on filters
Breather filler caps
Filter Vent plugs
Desiccating breathers Desiccant
FIGURE 33.7 A single line parallel system, including the controller, reservoir, pump, feed line, and individual injectors.
of the asset. A high machine reliability priority will justify more extensive development and machine improvement for contamination control. Contaminant exclusion: The three primary avenues through which contaminants enter tribological systems, whether oil circulation or grease sumps, are breather vent ports, seals and through lubricant replenishment. 1. Vent Ports. Oil lubricated circulating and noncirculating sumps must have vent ports to function properly. The vents are designed to allow easy passage of air across a sump opening while minimizing risk of gross, bulk contamination (large particles, process chemicals, rain drops, etc.). Most vent ports have threaded openings for easy installation or removal of threaded attachments. Figure 33.7 illustrates a traditional vent port configurations designed for use on a hydraulic or oil circulation reservoirs. These types of vents allow for significant ingression of various airborne contaminants, including moisture, and should be upgraded to types shown in the chart. The tighter the contaminant control target (cleanliness target) the higher the quality of the replacement configuration, leading up to Beta designated breather filter elements for the highest priority systems. 2. Shaft and cylinder seal points. Another high ingression point that should be reviewed and modified is the shaft seal or hydraulic piston seal point. Figure 33.8 represents a typical lip seal (top left) configuration contrasted against a typical bearing isolator (bottom right). Lip seals are more useful for containing lubricants than excluding contaminants and only then until the seal condition is degraded due to surface contact with the shaft or three body abrasive wear from contamination. Hydraulic cylinder seal and wipers are equipped with similar quality of materials that degrade rapidly once put into service. It should be noted that there are many designs and a great disparity in the quality of materials that may be found on OEM equipped hydraulic system components, based on the quality philosophy of the manufacturer. Performance for these components may be improved through the use of cylinder isolators as shown in Figure 33.9. Lubricant handling is another area where minor improvements may provide large dividends. These issues are part of effective lubrication program design, but should be addressed as an item separate from machine relubrication methods development.
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33-20
Handbook of Lubrication and Tribology Radial lip-type seal preventing impurities from getting in Rubber lip turned inward Rubber lip turned outward
Radial lip-type seal preventing grease from leaking out
Multi-stage labyrinth seal
FIGURE 33.8 Comparison of two shaft seals. On the left is a shaft “Lip” seal and on the right is a shaft multi-stage labyrinth seal. Keep wiper seals in good repair
Wiper
Rod seal
Oil system
Atmosphere
Use rod boots where possible
FIGURE 33.9
Hydraulic cylinder seal options, including the rod boot and the traditional shaft seal and wiper band.
3. Passive shields can be installed between lubricated components and sumps and high heat sources as a means for controlling radiant heat transfer. Additionally, passive shields may be used as weather brakes to minimize the effects of direct sunlight, cold, precipitation, or airborne solid debris. Contamination removal: Filtration has been addressed separately in this text, but should be identified again as a key aspect of precision lubrication care. Selection of filter media follows establishment of contamination control targets and alarm settings, and is an important aspect of the condition control loop. Once the target control values are established the lubrication technician should review each system for the existence of moisture, air, heat, and solid contaminant removal systems, and add these according to the level of machine risk. 1. Thermal control. Fluid flow around heat exchangers should be reviewed to assure that the area of immediate exposure is not in a dead flow zone and to assure that the heat rate per surface area is low enough to prevent burning of the fluid. Coolers should be reviewed for correct sizing and/or coolant flow conditions. Shell and tube heat exchangers should also be checked periodically for seal and sidewall integrity. 2. Moisture control. Reservoirs should be reviewed for evidence of condensation on the tops or sides of the lube oil tanks, and for evidence of moisture accumulation at the bottom of tanks.
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Lubrication Program Development and Scheduling
33-21
Tank options 1
Return diffuser
2
Baffles and residence time
Inlet Outlet Baffle Inlet (Top view)
Plate
Baffle Inlet
Outlet
Outlet
3
Coalesence of air bubbles
60 Mesh screen
Inlet Outlet
FIGURE 33.10 Options for controlling aeration in hydraulic tanks and other lubricating reservoirs.
Super-saturation in the headspace can be controlled through force draft air flow across the head space for large tanks, or through the use of desiccant breathers for small systems. 3. Air entrainment control. Air entrainment, or air saturation, is a problem primarily in systems that have high flows, low fluid dwell times, and where the fluid plunges into the reservoir at the return as suggested in Figure 33.10. These are all issues of system design, and should be addressed as systematic improvement. Symptoms of air entrainment should be noted for consideration during fluid selection and permanent system modifications. Vacuum distillation is useful for de-gassing system with high entraining levels and may be considered as a useful intermediate step. 4. Solid contamination control. Where systems have in-place continuous filtration the lubrication technician should evaluate the capability of the existing system and either accept or make recommendations for improvements. Where the system, or sump, does not have continuous filtration the technician must make a decision to either recommend no change, the addition of side-stream intermittent filtration or the addition of hard piped continuous filtration. The decision will be driven by the rate of ingression and the system requirements for cleanliness. (a) A common condition exists where hydraulic and circulation system designers working independently or alongside the OEM will install filtration capacity that meets minimum component manufacturer operating requirements, but does not meet the owner’s reliability requirements. These scenarios are very common. The lubrication technician must evaluate each system pressure and return side flow circuits for the presence of full flow element and for the quality of those elements. As noted in the chapter on contamination control, the filter industry relies heavily on statistical quality testing to measure the capture efficiency of individual element designs. Each element should have a Beta (β) efficiency rating for the sake of element quality comparison. The lubrication technician must verify that the element is β rated at a performance level that meets the expected OEM ISO cleanliness code. Although the β rating does not necessarily guarantee performance (due to the many variables that may exist at any given time for a hydraulic circuit), it is certainly a useful criteria to begin a quality comparison. (b) Many sumps exist with cooling or heating systems but no filtration system. If the criticality of these systems is high enough these systems should be considered for the addition of a low pressure, constant flow filtration circuit either following or in front of the thermal control
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33-22
Handbook of Lubrication and Tribology
Offline filtration, repairs and equipment rebuild flushing and flushing during equipment commissioning
FIGURE 33.11 operations.
Example of a portable filter cart, or “kidney loop” filter system used to filter a reservoir during
circuit. Again, any additions should be sized to meet cleanliness requirements with rated (tested) filter designs. (c) Finally, many sumps exist that lack any thermal control or filtration circuits. Bearing sumps and gear sumps commonly do not have these features. Often these sumps have proven their respective toughness through sustained years of production without failure. For these “untreated” sumps, the lubrication technician and reliability engineer should evaluate the system for long-term reliability, and for future production requirements, and decide as to whether the addition of hard piped filter system should be installed or whether the system may benefit sufficiently from the use of intermittent filtration through a side stream portable filter system, as depicted by Figure 33.11. A system modification plan should be developed for either the use of fluid quick disconnects or the installation of a continuous flow system, and in both cases the element quality should be evaluated for ability to meet defined program objectives. (d) Where the components are capable of withstanding appreciable contaminant induced abrasive wear, and system reliability is good relative to the production goals, then a clear option is to take no action.
33.7 Oil Analysis Requirements Oil analysis has been used to help equipment operators monitor oil health, machine wear, and contamination levels. Traditionally oil analysis was conducted by the lubricant supplier to help guide the customer on lubricant change cycles. While this is a worthwhile endeavor itself, there is appreciable value that may also be derived by using oil analysis as a primary control loop for management of the lubricant “in-situ.” The lubrication practitioner should incorporate oil analysis into the set of machine specific lubrication practices, and integrate these results into the long term plan for in-situ sump and lubricant management. There are four key areas that the practitioner must address, as follows: 1. 2. 3. 4.
Asset selection Test slate selection Alarm types and limits Analysis frequency decisions
Practical guidance for each of these parameters is addressed in the section on used lubricant analysis. A brief review will be provided on each criterion strictly as a means for establishing limits for lubrication practice development. Asset selection: Oil analysis can be a power tool when applied selectively to assets that provide high production value or cost control value. Early in the exercise the practitioner was directed to review the site’s machine criticality assessment to reach an understanding of which of the assets should be viewed more closely, or should receive a disproportionate share of attention for improvements or upgrades.
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Lubrication Program Development and Scheduling TABLE 33.12
33-23
Gear Lubricants Typical Oil Analysis Test State
Test
Enclosed gears
Open gears spray/splash lubricated
Viscosity Atomic emission spectroscopy (AES) Water Crackle or FTIR Karl Fischer Particle count Large wear debrisa Rotrode filter spectroscopy (RDE)a X-ray fluorescence (XRF) DR ferrography Analytical ferrography FTIR oxidation Acid number (AN)
Routine Routine Exception (3a)
Routine Routine Exception (3a)
Routine Recommended
Not required Recommended
Exception (4, 5) Routine Exception (7)
Exception (4, 5) Routine Exception (7)
a Either RDE, DRF, or XRF should be considered routine to look for large wear particles greater than 10 µm in size.
Lubricant analysis is a tool that can provide a high return on investment when used properly, providing returns on the order of five to one and higher [27]. None the less, the relative effectiveness of every dollar spent should be considered fully before recommending an oil analysis program. Assets with high reliability criticality scores, that function in aggressive environments (moisture, atmospheric debris, chemical saturation, ambient temperature extremes), that function within operational extremes (heat, cold, constant or intermittent load, isolation) and that have traditionally low MTBF (mean time between failure) scores nearly always warrant the application of an oil analysis routine, regardless of the sump size or preexisting oil sump change intervals. As the criticality score declines, environmental and operating conditions moderate, and overall reliability stabilizes, the relative economic benefit of each dollar applied to oil analysis may be replaced by comparatively higher benefit from other condition assessment methods. Test slate selection: Following decisions regarding about which assets to apply systematic oil analysis, the practitioner next would determine the types of test to conduct for each asset. The final selection of tests is referred to as a test slate. The test slate will comprise test methods for primary and secondary tests, as was described in the chapter on oil analysis. A Gear Lubricant test slate is presented in Table 33.12 [28]. Table 33.13 depicts various primary and secondary tests that may be selected depending on the types of measurements that are desired, and the relative strength of the test type for providing useful detail [29]. The most critical assets will receive a broader test slate than the less critical, and will likely receive an additional layer of superficial on-site screens (viscosity, moisture, particle count) on a very frequent basis in support of a more detailed laboratory analysis conducted on a less frequent basis. Each test type may have a variety of test procedures that could, or should, be considered in arriving at the final definition of test slate methods. The select oil analysis laboratory will likely have constructive feedback to offer on the final choice of ASTM test methods to assign to each parameter in the test slate. These decisions are explored in greater detail in the chapter on oil analysis. Alarm types and limits: Following the selection of the assets to trend, and the types of tests to conduct, the technician would assign score limits for normal, alert, and alarm status that would be applied to each test. Absolute and statistical mathematical methods are used to calculate alert and alarm score limits. A pass-fail evaluation is generally applied to on-site screens, but depending on the sophistication of the screen statistical and absolute alert and alarm limits could be applied. Table 33.14 provides an overview of a system of general industrial and automotive alert and alarm limits.
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TABLE 33.13
Generic Test State for Various Machine Types and Production Environments
Test procedure
Drive train
Hydraulics (mobileindustrial)
Paper machine oils
Industrial gear
Motor and pump
Steam turbines
Gas turbines
Air/gas compressor
R R R
R R R
R R R
R R R
R R R
R R R8
R R R8
R R R8
R R R
O1 R
R O4
R O4
R O4
R O4
R O4
R O4
R O4
R O4
R — — O2 E(2) R E(1ii) — O1 — E(1i,4) E(1i,4)
R E(3i) — R — — — R — — E(1i,4) E(1i,4)
R E(3i) — R — — — R5 — — E(1i,4) E(1i,4)
— R R R — — — (R4 ) — R5 E(1i,4) E(1i,4)
R E(3i) — R — — — O7 — — E(1i,4) E(1i,4)
R E(3i) — R — — — O7 — — E(1i,4) E(1i,4)
R E(3i) R R
R E(3i) R R O9 — — R — R E(1i,4) E(1i,4)
R E(3i) R R O9 — — R — R E(1i,4) E(1i,4)
— — R — R E(1i,4) E(1i,4)
Note: R = Routine testing. E = Exception testing (exception test performed based on positive result from test number in parenthesis). O = Test recommended for other reasons (see explanation for details). 1 Required to differentiate straight vs. multigrade engine oils. 2 Useful in evaluating filtration performance. 3 Should be considered routine for extended drains. 4 Useful in determining the viscosity at operating temperatures. 5 Required for condition-based oil changes of large volume systems. 6 Large volumes and typically high make-up oil add rates make rotating pressure vessel oxidation test (RPVOT) a better condition-based change test procedure. 7 Useful for condition-based changes of large oil volumes. 8 Even though typical lubricant may not contain organo-metallic additives, elemental analysis is useful in determining if wrong oil type has been added. 9 Useful in determining the presence of dissolved flammable gases. © 2006 by Taylor & Francis Group, LLC
Handbook of Lubrication and Tribology
Elemental analysis (ASTM D4951, D5185, or D6595) Wear metals (Fe, Cu, Pb, Sn, etc.) Contaminants (Si, Na, K) Additives (Zn, P, Ca, Ba, Mg, B, etc.) Kinematic viscosity (ASTM D445) 40◦ C 100◦ C Water Crackle test Karl Fischer moisture (ASTM D6304) Demulsibility (ASTM D1401) Particle count (ISO 11500) Flash point (ASTM D92 or D93) FTIR soot (ASTM E168) Glycol (ASTM D2982, D4291, D5185, or D6595) Acid number (ASTM D664 or D974) Base number (ASTM D2896, D974, or D4739) RPVOT (ASTM D2272) Ferrous density Analytical ferrography
Diesel engines
Lubrication Program Development and Scheduling TABLE 33.14
33-25
Generic Test State Alarm Limits and Alarm Types
Parameter Chemical index Contamination index Ferrous index DV at 40◦ C FW index ICP — additive ICP — wear debris FTIR — Ox , Ni, S FTIR — additives Fe density (DR, FW, PQ) Acid number Analytical ferro.
Units
Critical
Caution
Normal
Index Index Index % Stat % Stat % % Stat Absolute Qualitative
+2σ +2σ +2σ ±15 +2σ ±50 +2σ ±5 ±50 +2σ 1.0 >base NA
+1σ +1σ +1σ ±10 +1σ ±25 +1σ ±25 ±25 +1σ 0.2 >Inflection NA
< +1σ < +1σ < +1σ Baseline < +1σ Baseline < +1σ Baseline Baseline Average Inflection point —
Source: M. Johnson, and M. Spurlock. Noria Corporation. 2004.
Analysis frequency decisions: Oil analysis frequency selection, like other aspects of oil analysis program decisions, is driven by machine operational factors. Intervals are shorter where criticality is high and the potential for contaminant level changes is high, particularly as it pertains to solid and liquid contaminants. Table 33.15 provides a basis for making systematic sample interval decisions for most common industrial and automotive machine types. The technician should strive to be as consistent as possible when making adjustments from machine to machine to align work schedules.
33.8 Activity Sequencing Relubrication scheduling programs that include large numbers of assets can become unwieldy unless the whole of the program is divided into manageable parts. How those parts are best managed will vary depending on the size of the production site, the tools available for moving materials between discrete jobs, the timing and logistics associated with relubrication intervals, and the scope of the activity to be conducted. A scheduling approach that may constitute the most efficient method at one site may be a poor choice at another site. Relubrication activities fundamentally divided into tasks that are operating condition dependent and tasks that can be conducted without concern for the operating state. Downtime dependent tasks: Tasks that must be conducted when the machine is in a nonoperating state must be uniquely defined accordingly during the survey process. These tasks will be scheduled around other “downtime” activities for this asset as timing allows. The technician will bundle the identified materials and tools required for the job much like that of repair work order activities. Scheduling is complicated by the fact that the required task, which might otherwise be done based on a carefully selected schedule, will have to be done around nonscheduled or emergency availability. An example of this is a cooling tower drive sample collection activity. While ideally the machine might be modified to facilitate relubrication regardless of the operating state, until the necessary modifications are made the work must be conducted when the technician can physically access the sump. Runtime dependent tasks: It is preferable to conduct routine relubrication activities on machines during a running state. This may be considered the default state. If the task may be conducted in either state then the technician should note that the procedure may be conducted in either state. The majority of activities will be conducted while the machine is in operation. An ultrasonic assisted bearing relubrication activity would fall into this category. It would be impossible to conduct this type of activity, which is preferable for electric motor relubrication, if the machine is not operating. The oil sample collected from a positive pressure supply line must of course be
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Handbook of Lubrication and Tribology TABLE 33.15
Systematic Method to Select a Sample Frequency Interval
conducted during a run state. An oil sample from a static line, or a fixed sample port on a reservoir, may be conducted in either state, even though it is preferable to collect oil samples from operating machines. For machines designated to be lubricated during a running state, the practitioner must next decide whether to schedule activities based on a “whole” machine practice, or based on the nature of the practice to be conducted. Again, the type of machines to be covered will help influence these decisions. 1. Whole machine scheduling. Where the machine might be scheduled out of production for a period of time to receive routine care, including relubrication, the sum of the tasks to be conducted may be scheduled to be done all at the same time. Industrial vehicle inspection and relubrication activities are conducted based on miles or hours of operation, depending on the type of vehicle. A mine haul truck would likely have schedules
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Lubrication Program Development and Scheduling
33-27
3 2
4
1
5 6
11
7
10
8
9
Conveyor A
Conveyor B
11
10
9
8
6
7
5 1 4
2 3
FIGURE 33.12 Top-view of a typical drive train relubrication sequence.
based on hours of operation. If the truck has a diesel engine the schedule might be based on 250 to 500 h operation intervals, which would coincide with suggested oil change intervals. With each passing unit of hours, additional tasks would be added, leading up to comprehensive depot level overhauls. During these planned events the practitioner would prepare to cover all the required lubrication related tasks at one time. 2. Practice type scheduling. Contrast this schedule method to a site utilizing conveyors to transport materials rather than haul trucks. Earlier in the process each practice type was defined, the specific lubricant type was defined, and the frequency was defined, all based on sound engineering principles. Once the data has been collected and tabulated, the data can be sorted and ordered in an efficiently work flow based on the products, type of activities, and the frequencies that have been defined as appropriate. For a conveyor drive train the practitioner faces about seven fundamental types of tasks to provide relubrication coverage on a conveyor, including motor lubrication, coupling lubrication, bearing lubrication, routine level check, sampling, filtration, and oil change. In this example, the belt pulley bearing lubrication activities could be grouped into a route. They are more similar than dissimilar. They will likely require a single type of product for a large number of components. When possible the activities should be bundled, and then set into a linear sequence, and this sequence would the be set into a route, or a series of routes, to minimize the number of products carried and the number of steps walked to fulfill the required task. Figure 33.12 shows a top view of two conveyors sitting adjacent in a material transfer process. Each number depicts an identification of a lubricated component in the drive train, and the RED ARROWS depict the preferred sequence to conduct the collective tasks. It would be possible to lubricate the conveyors based on a sequence from A1 through the order of numbered components, but this sequence would create unnecessary steps and consume valuable labor resources. Similarly, all common component tasks should be delineated such that common functions with common materials on common frequencies may be grouped together and executed
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Handbook of Lubrication and Tribology
using the least amount of time. Examples of routes by common “ingredients” could include: • 1 month grease routes, grease Z • 1 month grease routes, grease Y • 1 week gear drive level check and oil top-up route, Gear Oil 220 • 1 week hydraulic level check and top-up route, Hydraulic Oil AW68 • 6 month Ultrasonic motor relubrication route, Grease W • 3 month sample route
33.9 Planning and Scheduling Management Finally, once the routes have been set, walked down, and confirmed, the lubrication technician is ready to begin scheduling the necessary work. Key concerns with lubrication planning and scheduling are completeness, work balancing, and data management. Design completeness: The lubrication program is not complete until every lubricated component is accounted for, regardless of its relative criticality factor. A cursory check of each asset in the completed database is a minimum requirement. A detailed check should be in order for the high criticality items. Work balancing: The time stamps for each activity can now be rolled together to reflect the total estimated time for a series of activities. These “route” time stamps can then be used for effective labor allocation and balancing within a work cell or for the entire facility. Work schedule and work order generation: Modern industrial complexes use computerized maintenance management system (CMMS) to track work activities by asset type where possible. The CMMS must provide sufficient data (hierarchal) structure that the individual component can be grouped or tracked. This is often not possible without significant reordering of the CMMS structure, which is not likely to happen. There are several stand alone lubrication scheduling programs that could be used to manage the work process (open, issue, verify, close routes) to the detail level following scheduling orders from the CMMS programs. This does create added complexity to data logging and tracking, and the decision to run two programs will have to be considered on a case by case basis. Work verification: An important final step in assuring effective lubrication program management is verification and tracking of the activities. A lubrication program lead or supervisor should periodically spot check to verify completeness of all scheduled activities, and follow up on notes provided by the lubrication technicians on the printed work orders. Where a collaborative process exists the skilled lubrication technician becomes the eyes and ears of mechanical supervisors, and can provide direction that will help eliminate “sunk” costs associated with poor equipment configuration and maintenance hostile machine designs.
33.10 Conclusion The lubrication engineer or technician has many responsibilities that must be covered during the course of devising a machine care program that may be considered best practice. Traditionally, the lubrication engineer has focused the majority of his/her effort on matching the operating requirements of the machine with the appropriate lubricant viscosity, the appropriate grease stiffness, and the appropriate lubricant lifecycle qualities relative to the actual operating characteristics. Recently additional focus has been applied to developing a comprehensive volume, frequency, contamination control, oil analysis test slate, and thorough but superficial inspection criteria, all designed to promote improved reliability, production quality, and long-term utility. The lubrication engineer must begin the exercise with broadly applicable quantitative tools, and then refine all recommendations based on experience and product capabilities knowledge.
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Lubrication Program Development and Scheduling
33-29
A properly devised “best practice,” will incorporate operation and machine specific requirements including: • • • • • • • •
Machine criticality and operating environment Data collection strategies Machine data collection criteria Lubricant type, quantity, and frequency requirements Contamination control requirements Oil analysis requirements Activity Sequencing Planning and Scheduling Management
Competitive advantage is not limited to acquiring quality production machines or quality maintenance materials. Knowledge will pay a vital role for the lubrication program manager and the designated lubrication technicians in achieving maximum value from each dollar invested in the plant tribology program. Leadership and project management skills would also serve the modern lubrication technician. Business planning, skills development, and project management are becoming increasingly vital as well as more and more sophisticated production machines, operating at higher loads and higher speeds, are replacing older, less reliable and less productive assets.
References [1] Lansdown, A.R. “Lubrication and Lubricant Selection: A Practical Guide.” Antony Rowe Ltd., Chippenham, Wiltshire, 1996. [2] Neale, Michael. Tribology Handbook, 2nd ed., page C7.2. [3] U.J. Moller and U. Boor, Lubricants in Operation, Table 3.2, Page 111. Mechanical Engineering Publication Limited. Revised English Version, 1996. [4] Moller, Boor, Lubricants in Operation, Page 116. (PERMISSION REQUIRED) [5] SKF Corporation Bearing Maintenance and Installation Guide, Page 29. February 1992. (PERMISSION REQUIRED) [6] Booser, Bloch — (Secondary Reference: MLII, Slide #1557.) (PERMISSION REQUIRED) [7] Neale, Michael, Tribology Handbook, 2nd ed. Page A7.5 (PERMISSION REQUIRED) [8] Neal, Michael, Tribology Handbook, 2nd ed. Page A7.5 (PERMISSION REQUIRED) [9] Luegner, Tex, Practical Handbook of Machinery Lubrication, 2nd ed. (PERMISSION REQUIRED) [10] SKF Corporation Bearing Maintenance and Installation Guide, Page 29. February 1992. (PERMISSION REQUIRED) [11] Lubriquip Technical Bulletin #20115 (http://www.lubriquip.com/pdf/20115.pdf ). [12] Luegner, Tex, Practical Handbook of Machinery Lubrication, 2nd ed. (PERMISSION REQUIRED) [13] FAG Roller Bearing Lubrication Guideline WL81115E. http://www.fag-industrial-services.com/gen/ download/1/15/40/37/FAG_Rolling_Bearing_Lubrication_WL81115E.pdf. [14] Timken Bearing Company. http://www.timken.com/industries/torrington/catalog/pdf/general/ form640.pdf. [15] SKF Bearing Company. http://mapro.skf.com. [16] Perma Corporation Perma Star Single Point Lubricator. (http://www.permausa.com/sv.htm) (PERMISSION REQUIRED) [17] Noria Corporation Machinery Lubrication Seminar ML I (Lincoln Industrial Corporation. http://www.lincolnindustrial.com). [18] http://www.agma.org/site_search_results.asp(Select “publications,” then “ANSI-AGMA Publication Standards,” then search for “9005-E02”). The same web reference provides links to the AGMA 925 standard as well. [19] Technical Bulletin ANSI/AGMA 9001 B97. http://www.agma.org/site_search_results.asp (Select “publications,” then “ANSI-AGMA Publication Standards,” then search for “9001 B97”).
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[20] Falk Corporation. Installation and Maintenance of Double and Single Engagement Gear Couplings Technical Bulletin 458-110, http://www.falkcorp.com/dist-info/main-listofpublications.asp. [21] http://www.americanchainassn.org/ACAPubs.htm. [22] Wright, John L., “Fundamentals of Chain Lubrication.” Machinery Lubrication Magazine. (PERMISSION REQUIRED) http://www.machinerylubrication.com/article_detail.asp?articleid= 316&relatedbookgroup=Lubrication. [23] http://www.machinerylubrication.com/article_detail.asp?articleid=277&relatedbookgroup= Hydraulics. [24] http://www.maintenanceresources.com/Bookstore/FluidPower/HandbookHydFluid.htm. [25] http://www.insidersecretstohydraulics.com/index2.html. [26] ML I. Slide # 1218a. Ref.: SKF, Pioneer, RP, Idcon, others. [27] Evans, John and Wearcheck, S.A. How to Calculate the Effect of Oil Analysis on the Bottom Line. Practicing Oil Analysis, July 2004. (PERMISSION REQUIRED) [28] Walsh, Dan, “National Tribology Services. Gear Lube Test Slate Selection.” Practicing Oil Analysis Magazine. March 2002 (PERMISSION REQUIRED) [29] Barnes, Mark, Ph.D., Noria Corporation, “Reality Check: Time to Retool Your Test Slate.” Practicing Oil Analysis Magazine. January 2003 (PERMISSION REQUIRED)
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34 Lubricant Storage, Handling, and Dispensing 34.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 34.2 Receiving New Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
34-1 34-2
Lubricant Manufacturing and Delivery • Lubricant Packaging • Storage Stability and Inventory Control
34.3 Storing Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
34-5
Bulk Tanks • Totes • Barrels • One-Shot Containers • Grease Storage
Mark Barnes Noria Corporation
34.4 Lubricant Storeroom Design . . . . . . . . . . . . . . . . . . . . . . . . . . 34.5 Dispensing Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 34.6 Insuring Product Integrity . . . . . . . . . . . . . . . . . . . . . . . . . . . . 34.7 Conclusion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
34-9 34-9 34-12 34-15 34-15
34.1 Introduction The impact that inappropriate storage of new lubricants can have on equipment reliability and longevity cannot be overstated. Put simply, an inability to control the quality of a lubricant while in storage, or allowing contamination ingress due to poor storage practices can result in a diminished life expectancy for all equipment, from the smallest grease lubricated bearings to the largest steam or gas turbine system. Likewise, effective handling and dispensing of lubricant in the plant has an equally vital role to play on the impact that in-service lubricants can have on component life. Exposing lubricants to environmental contaminants such as water and particles during basic lubrication procedures such as oil top-offs or regreasing can compromise both the effectiveness of the lubricant itself and insure that those contaminants are introduced into the oil wetted path, resulting in many commonly encountered failure mechanisms such as abrasion, erosion, surface fatigue, and corrosion. In this chapter, we explore those factors that must be considered in order to insure that new lubricants retain their ability to provide effective lubrication from the initial receipt of a lubricant, to putting that lubricant into service.
34-1
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Handbook of Lubrication and Tribology
34.2 Receiving New Lubricants 34.2.1 Lubricant Manufacturing and Delivery The steps involved in the formulation and ultimate delivery of a lubricant to any industrial facility are illustrated in Figure 34.1. In general, lubricating fluids are manufactured by taking appropriately selected base oils and blending with appropriate additives before delivering to the end user in bulk (as illustrated in the figure), by tote, drum, or individual containers such as plastic quart containers commonly used for passenger vehicle oils. Contamination and base oil/additive degradation can occur at any step of this process. While the lubricant manufacturer is charged with insuring that the selection of base oils and additive package meets the required performance characteristics of the lubricant in question, the manufacturer is also responsible for insuring an equally important characteristic — that the formulated lubricant is delivered to the end user without levels of contamination that might compromise the integrity of the fluid before it even goes into service. Even before delivery to the end user, significant levels of contaminants, including heat, moisture, and particles can cause premature lubricant degradation in storage, at the manufacturing facility, at the blend plant, or at the lubricant distribution warehouse. Upon delivery to the end-user, the degree to which these contaminants should be excluded will be critically dependent on application. For example, a highly critical hydraulic application, with very tight tolerances and high pressures will obviously require a significantly higher degree of contamination exclusion than a high viscosity open gear oil, to be used on a noncritical application. Nevertheless, whatever the application, it is still incumbent upon the manufacturer to provide some degree of contamination exclusion. The onus for new fluid cleanliness cannot however be put squarely on the shoulders on the lubricant supplier. Even if the supplier delivers lubricants below target levels of particle or moisture contamination,
Base oil storage tank
Pump Oil from refinery
Filter
Dedicated clean oil delivery truck
Precleaned hydraulic reservoir
FIGURE 34.1
Typical process for lubricant formulation and bulk transfer.
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Additive blending vessel
Custom cleaner oil storage
Lubricant Storage, Handling, and Dispensing
34-3
oftentimes poor practices at the plant site can compromise the efforts of the lubricant supplier to insure product integrity. Likewise, the cost to control contaminants throughout the manufacturing and delivery process must also be considered. While filtering oil to achieve a desired fluid cleanliness level is always possible, the cost to achieve this level of cleanliness, only to be compromised by poor practices once onsite may render this approach less than optimal. As an example, the typical cost associated with delivering a hydraulic fluid to a cleanliness rating of ISO 16/14/11 (per ISO 4406:99 [1]) is typically around $0.25 per gallon over and above the cost of the lubricant itself. In order to balance cost vs. benefit, it is generally good practice to set two levels of fluid cleanliness, one for new oil deliveries and one for in-service oils. Typically the new oil cleanliness requirements should be set 1–2 ISO range codes higher than the in-service targets as a compromise between cost and acceptable new fluid cleanliness levels. Water is perhaps a more harmful contaminant for stored lubricants than particles. While it is true that certain solid contaminants can cause adsorptive additive depletion along with promoting catalytic degradation at higher temperatures, the effects of water washing of additives, along with hydrolytic effects make water a far bigger threat for stored lubricants. It is generally recommended to set maximum permissible water levels below the saturation point of the fluid at all ambient temperatures to which the fluid will be exposed during manufacture, storage, transportation, and dispensing. For some lightly additized oils such as turbine oils, this may mean less than 100 ppm or 0.01% (v/v).
34.2.2 Lubricant Packaging New lubricant deliveries can be separated into two basic categories — bulk or prepackaged. While bulk lubricant deliveries are typically provided via either tanker truck or rail car, often thousands of gallons at a time, prepackaged lubricants are supplied in a host of containers from large 250 to 300 gallon totes, to barrels (either steel or plastic), or in single use containers such as the plastic quart cans commonly used to sell passenger vehicle engine oils and other fluids. The choice between bulk vs. prepackaged delivery will to a large extent be governed by cost and usage patterns. For example, while the cost of bulk delivery may range between $0.20 and $0.35/gallon, the associated cost of tote or barrel delivery may range anywhere from $0.35 to 0.55/gallon, depending on the volume of oil used per year. Beyond cost, other deciding factors include cleanliness requirements, storage stability and environmental concerns and achieving appropriate inventory stock rotation as indicated below. Table 34.1 illustrates the relative advantages and disadvantages of bulk lubricants vs. prepackaged products.
34.2.3 Storage Stability and Inventory Control The life of any lubricant is strongly influenced by the ambient environmental conditions under which it is stored. Stored outside where the potential for moisture to enter the oil via rain, snow, or humidity, or where high or low summer or winter temperatures are experienced can reduce the life expectancy of
TABLE 34.1
Comparison of Bulk vs. Pre-Packaged Lubricants
Risk of supplier cross contamination Risk of contamination ingress during handling Storage stability Safety risk of handling Environmental spill risk Distributor inventory aging risk Handling cost
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Bulk
Tote
Barrel
One-shot container
High Medium Medium Low High Medium Low
Medium Medium Medium Medium Medium Medium Medium
Low Medium Medium Medium Medium High Medium
Low Low Medium Low low High High
34-4
Handbook of Lubrication and Tribology TABLE 34.2
Factors Affecting Storage Life of Lubricants
Variable
Increase storage life
Shorten storage life
Base oil
Lower-grade mineral oils and inorganic esters
Additives Thickener Storage temperature Temperature variability Container
Highly refined mineral oils, synthetic hydrocarbons, and inert synthetics like silicon-based oils R&O additives No Low Low Plastic containers or liners
Humidity Agitation Outdoor storage
Low Low No
EP additives Yes High High Metal drums, especially poorly conditioned ones High High Yes
any lubricant to a matter of months. By contrast, highly stable oils such as turbine oils, stored in climate controlled conditions can be expected to last literarily for years. While little research has been conducted to determine the absolute effects of poor storage on lubricating oils and greases, it is generally acknowledged that extreme temperatures, particles, and moisture all contribute to premature oil degradation. Most at risk are more highly additized oils such as engine oils, transmission fluids, and tractor oils. These oils contain such high concentrations of additives that any slight change in temperature — particularly low winter temperatures — can cause additives to drop out of suspension. Table 34.2 provides a general description of factors that shorten the life of stored lubricants. Water is perhaps the most deleterious of all contaminants. Certain additives such as zinc-based antiwear additives and organo-phosphates are prone to a hydrolytic reaction rendering these additives ineffective. While this effect is unlikely to occur at lower temperatures, storing oils in highly humid environments above 110◦ F can lead to significant degradation of lubricating oils. Other additives are also affected by the presence of water, due to the effects of water washing. Water washing involves the dissolution of the additives preferentially in water compared to the oil. Under these circumstances, these additives, which include organometallic additives such as calcium and magnesiumbased detergents and borate-type extreme pressure (EP) additives can be stripped from the new lubricant. Some additives in new formulations are not completely dissolved in the oil. When the oil reaches service temperatures these additives may finally dissolve, a process known as “bedding in.” Other additives by design will never dissolve. For example, some gear oils may be formulated with solid additive suspensions such as graphite, molybdenum disulfide, or borates. These oils should not be stored for prolonged periods because the solid additives are prone to settling and should always be agitated before being put into service. Even oils that are not prone to chemical degradation or settling can also be compromised by prolonged exposure to heat or water. Stored in steel drum or tanks, the formation of rust can cause rust scale to form, resulting in particle contamination in the oil. Under extreme levels of water contamination, other effects including bacterial growth can also occur, resulting in premature oil degradation and reduced lubricant life once put into service. In general, long-term storage at moderate temperatures and low humidity has little effect on most premium lubricating oils, hydraulic fluids, and process oils. However, some products may deteriorate and become unsuitable for use if stored longer than three months to a year from the date of manufacture. Table 34.3 provides general guidelines for the maximum amount of time a lubricant should be stored to avoid performance degradation under normal conditions (clean and dry) and moderate temperatures (60F to 80F). If a product exceeds its maximum recommended storage time, it should be sampled and tested to confirm it is fit-for-purpose, using the appropriate performance tests outlined below. Even if an oil is stored appropriately and is deemed to be relatively “storage stable” based on the preceding information, it makes sense to control inventory levels such that excessive quantities are being stored for prolonged periods of time. Likewise, for those products that are constantly being consumed and
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Lubricant Storage, Handling, and Dispensing TABLE 34.3
34-5
Typically Recommended Storage Life for Lubricants
Product Lithium, lithium complex, and polyurea greases Calcium complex grease Motor oils, gear oils Fluids or lubricants with solid additives Turbine oils, hydraulic fluids, R&O oils Emulsion-type hydraulic fluids Soluble oils Custom blended soluble oils
Recommended maximum storage time, months 12 6 6 3 18 6 6 3
replenished, a first-in-first-out inventory control procedure should be adopted so that the oldest stored product is used first. Concurrent with this, all stored lubricants should be labeled with salient information such as date of manufacture, date of receipt, and a “use-by” date (based on Table 34.3), as well as product information including manufacturer, brand and grade and inventory control levels (so-called max/mins).
34.3 Storing Lubricants 34.3.1 Bulk Tanks Bulk tank offer perhaps the most convenient storage of large quantities of lubricants. Properly designed and maintained, and used with appropriate transfer containers, there is no reason why bulk tanks cannot provide the same degree of fluid handling precision and cleanliness that prepackaged containers provide. Several American Petroleum Institute (API) guidelines exist that relate to the design, fabrication, and commissioning of bulk storage tanks [2,3]. Tanks can be constructed from a number of materials, including stainless steel, mild steel plate, and anodized aluminum. While stainless steel and anodized aluminum are typically higher in price, they typically offer significantly lower maintenance cost over the long term. Mild steel tanks, though the least expensive are prone to rust and corrosion. As such, it is advisable to treat the inside with a rust preventative or other corrosion prevention coating such as plastic or epoxy. All storage vessels need appropriate vents or breathers to allow air to enter or exit the tank during oil filling/draining. While it is important that these breathers have a high air flow capacity, it is equally important that they offer a fine enough particle removal size rating to maintain appropriate levels of fluid cleanliness. Quality tank and vent breathers typically offer in excess of 98% particle capture efficiency at 3 µm. Where bulk tanks will be housed outside, or in highly humid environment, any vent or breather should also containing a desiccating material such as silica gel, to help remove any trace of water vapor from the air as it enters the tank. All tanks need to be installed with environmental considerations in mind. In the United States, these guidelines can be found under EPA guidelines CFR280 [4]. Perhaps the most fundamental environmental requirement is the need to have sufficient spill containment to house the complete volume of lubricant that can be stored in the tank. Common practice is to line the bottom of any spill containment with an absorbent such as sand (Figure 34.2). In order to maintain appropriate levels of fluid cleanliness both inside the tank and during usage, oil should be filtered into and out of the tank. The filter should be rated to maintain appropriate levels of fluid cleanliness based on the application to which the fluid is to be supplied. Tanks should be configured with quick connect to allow for filling and dispensing without exposing the lubricant to the outside environment and any transfer pipes or hoses capped immediately after use. In addition to contamination control measures, the tank should have an appropriate oil level gauge, a drain to allow for periodic flushing, inspection hatches to allow for entry and cleaning (note: confined
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Handbook of Lubrication and Tribology
Sand minimum 3” thick
Containment capacity must equal or exceed maximum volume of tank
FIGURE 34.2
Spill containment recommendations for bulk storage.
space entry rules apply here) and an appropriate oil sample valve. For very deep tanks particularly whether lubricants will be stored for extended periods a series of sample valves at different heights can also be useful to test for signs of additive stratification. Detailed information on best practice for bulk lubricant storage and handling can be found in References 4 and 5.
34.3.2 Totes Intermediate bulk storage containers (Figure 34.3), often referred to as “totes” offer many of the advantages of barrels, but with convenience of receiving large shipments direct from the manufacturer. These containers typically hold 250 to 300 gallons of lubricants and are typically fabricated from stainless steel, plastic, or disposable cardboard. Many tote bins come with liners insuring appropriate contamination control. The use of liners also negates the need for secondary external spill containment since the vessel itself provides adequate protection. Many of the comments regarding bulk storage tanks such as the use of filters to dispense lubricants, appropriately rated vents and breathers, and level gauges apply to tote bins, which are often connected via hard piping, serving almost as a semipermanent bulk tank.
34.3.3 Barrels Barrel storage is perhaps the most commonly used lubricant storage method. They are available in either steel or plastic, typically holding 55 gal or 205 l. Because of the weight of the barrel and fluid, a typical barrel may weigh in excess of 450 lbs, making safe handling and lifting a priority. Steel drums are the most commonly used. With care, they are reusable, convenient, and inexpensive though they are prone to impact damage (e.g., from fork-lifts or dropping) and internal corrosion. Several manufacturers supply lubricants in plastic barrels. These are typically cleaner when new, do not rust, and are somewhat lighter than steel. Because they are less rigid than steel, plastic drums can be difficult to handle when full. Caution should also be exercised when pumping oil from plastic barrels since the flowing fluid can cause static charges to build-up, resulting in potential fire and explosive conditions. Grounding straps should always be used when dispensing fluids via a transfer pump from a plastic drum. Storage of drums and barrels outside should be avoided at all costs. The main reason for this is due to the effects of change in ambient temperature and humidity. Even when sealed, barrels have a tendency to breathe (Figure 34.4). During the daytime, heating of the oil and air in the head space of the barrel cause air inside the barrel to escape while at night when the temperature drops, air is drawn back into the barrel. This “breathing” effect causes airborne contaminants, particularly moisture to be drawn inside the barrel, causing premature lubricant contamination and potential degradation.
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Lubricant Storage, Handling, and Dispensing
34-7
FIGURE 34.3 Example of a tote storage tank.
Rain
Water
Air escapes
Water
Air space COOL
Air space reduced
Clean oil as delivered
WARM Oil and air in barrel expand when warm. Some of air above oil escapes
Powerful suction created
COOL Water drawn in when oil and air contract when cooling
Water
FIGURE 34.4 Breathing of barrels.
Where outdoor storage cannot be avoided, barrels should be either covered with plastic drum covers, in a custom made free-standing enclosure such as that shown in Figure 34.5, or on their side, with the bungs at 3 o’clock and 9 o’clock so that the air in the head space is at the top of the barrel, and air is unable to escape (Figure 34.6).
34.3.4 One-Shot Containers Small so-called one-shot containers such as those used for passenger vehicle engine oils and other fluids offer the greatest flexibility in terms of being able to take a sealed, new lubricant to the point of application.
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Outdoor storage
Holds 55 gallons Spill pallet
FIGURE 34.5
Free standing storage enclosure for 55 gallon drums.
Drums
Vinyl or canvas cover
Notched planking
FIGURE 34.6
Temporary storage of barrels. (Note the location of the bungs at 3 and 9 o’clock.)
This insures that the opportunities for contamination ingression are limited. Likewise, the typically new oil cleanliness levels are much better with these containers while the potential for cross contamination is minimal. While the volume of lubricants used for many industrial applications preclude the use of such containers, wherever feasible, one-shot containers are perhaps the best option.
34.3.5 Grease Storage Greases come supplied in drum, pails, and small disposable cartridges. Just like lubricating oil, storage stability, product integrity, and cleanliness are the most important considerations. Beside contamination and environmental stresses outlined earlier, one other major factor to consider with greases is separation of the base oil from the grease thickener while in storage. Under normal conditions (average temperature and humidity) most grease will shed some oil. Whenever this is observed on top of a pail or barrel, it is common and good practice to stir this oil back into the grease. Excessive separation should be considered a cause for concern and should be thoroughly invested with specific ASTM performance tests. Contamination is a particular concern with greases. Unlike lubricating oils where contaminants can be readily removed via filtration, once a grease is contaminated, it is a virtual certainty that these contaminants will be introduced into equipment if the grease is used. This can result in premature equipment failure due to abrasion and fatigue.
© 2006 by Taylor & Francis Group, LLC
Lubricant Storage, Handling, and Dispensing
FIGURE 34.7
34-9
Typical layout of a lubricant storage room.
Similarly cross contamination of different greases through using transfer tools, or grease guns for two different products should be avoided. Many different grease thickeners are considered to be incompatible and can lead to excessive softening of thickening once the cross contaminated grease is put into service. Just like lubricating oils, greases should be stored with pertinent information such as date of manufacture, date of receipt, “used-by” date (based on Table 34.3), and product data including manufacturer and brand.
34.4 Lubricant Storeroom Design Ideally, all lubricants should be stored inside in a climate (temperature and humidity) controlled environment. In addition to climate control, the design and layout of the store room should be geared for efficient and ergonomic work flow, with first-in first-out principles in mind. Figure 34.7 shows an ideal layout for a lubricant storage room, while Figure 34.8 shows the main features to be considered when designing an ideal storage room. Any lubricant storage room should be designed with enough working space, as well as storage space for as many commonly used lubricants as necessary. In addition, the room should be designed to provide for spill contamination and to comply with any local, state, or federal environmental regulations. Safety considerations, such as explosion proof lighting and a fire suppression system should also be factored into the design. In some cases, the provision of small self-contained storage tanks is warranted (Figure 34.9). In this case, new lubricants are pumped from a 55 gal drum into the storage tank. In this case, the oil should be filtered both into the storage tank and from the tank during dispensing using an appropriately rated filter. To avoid cross contamination, dedicated transfer lines, pumps, and filters should be used for each lubricant type and grade, while each tank should have an appropriate desiccant and particle removing breather, either dedicated to each tank (preferred), or gang mounted as shown in Figure 34.9.
34.5 Dispensing Lubricants The final step in insuring the integrity of new lubricants before they go into service is to insure that the lubricant quality and cleanliness is maintained during dispensing and application. In general, it is good
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Handbook of Lubrication and Tribology Ventilation with filter Automatic sprinkler system
Explosion-proof lighting and switches
Sealed or one-shot lubricant dispenser Cabinet or locker
Fire-proof construction
Clean rags, funnels, sampling hardware, etc. Vent filters Ground
Automatic grease dispensing systems (avoid opening grease pails)
Covered waste can Dispense oil through filters
FIGURE 34.8
Key components of a lubricant storage room.
Color-coded labels
Ganged desiccant breathers
Posted procedures Fluid level gauges
Pumped dispensing through filters, dedicated hoses, pumps, etc.
FIGURE 34.9
Ideal configuration of bulk storage containers.
© 2006 by Taylor & Francis Group, LLC
Drip pan for auxiliary pumps, etc
Lubricant Storage, Handling, and Dispensing
34-11
FIGURE 34.10 Skid mounted drum carrier and integral filter transfer cart.
practice to dispense lubricants using a filter cart dedicated to the type and grade of lubricant in use. The filter cart should be equipped with a suitable particle removing filter in order to achieve and exceed the desired ISO solid contamination code for any specific application. A variety of appropriate transfer containers and skid mounted systems such as those shown in Figure 34.10 are commercially available, which allow for dispensing fluid using a filter transfer cart. In some applications such as large volume circulating systems (e.g., turbine), it makes more sense to dispense directly from a bulk storage tank. Under these circumstances, provision should be made to hard pipe to supply lines, again including an appropriately rated filter to insure fluid cleanliness. In either case, as stated earlier, the bulk tank or drum should be equipped with an appropriate desiccant and particle removing breather to insure the cleanliness and integrity of the bulk stored fluid, as air enters the storage vessel and the fluid is dispensed. For applications where smaller oil volumes are required, for example, for equipment top-offs, or where accessibility is an issue, dispensing from a barrel or from bulk may not be feasible. Under these circumstances, top-off or intermediate transfer containers will undoubtedly need to be used. Historically, many transfer containers in use are made from galvanized steel. While simple and inexpensive, the use of galvanized containers and other dispensing equipment such as funnels should be avoided since some lubricants, most notably those containing zinc-based additives have a tendency to react with galvanizing. Instead, rigid plastic containers, such as those shown in Figure 34.11, which include O-ring seals on both the lid and fill spout should be used. Inevitably, these top-off containers need to be filled either from a bulk dispensing station such as that shown in Figure 34.9, or directly from the barrel. When filling directly from barrels, the barrel oil should be dispensed either using a filter cart or for lower viscosity fluids using gravity through an appropriate filter as shown in Figure 34.12. Again, continued cleanliness of the oil in the barrel is insured by the use of an appropriate vent breather. When using top-off containers and other transfer equipment, preventing cross-contamination is important in insuring equipment reliability. It is advisable to designate filter carts, top-off containers, funnels, and other transfer equipment to one class and grade of lubricant. To prevent accidental cross contamination, it is recommended to label each storage container or barrel with a specific color- and
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FIGURE 34.11
Handbook of Lubrication and Tribology
Example of high quality oil top-off containers.
Desiccant breather
Standard pipe fittings
Fluid drums Level gauge
Spin-on filter Male quickconnect
FIGURE 34.12
Valve
Dispensing pumps are required for higher viscosity fluids
Bulk barrel dispensing station.
shape-coded tag to avoid accidental mixing as shown in Figure 34.13. In this case, the color is used as the primary indicator of fluid type, while the shape can be useful for individuals who may be color blind.
34.6 Insuring Product Integrity While lubricant manufactures and distributors typically make every effort to insure new lubricants meet appropriate performance criteria, mistakes can and do happen. Likewise, most distributors are unable or
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34-13
FIGURE 34.13 Example of color and shape coded transfer equipment. TABLE 34.4
Common New Oil Performance Properties Tests ASTM/ISO procedure
Turbine oils
Paper machine oils
Gear oils
Engine oils
Hydraulic fluids
D445 D5185 D6304 Crackle ISO 4406:99 D664 D4739 (or D896) D1401 (or D2711) D892 D2272 D665 D130 D3427
• •a • — • • — • • • • • •
• • • — • • — • • —• • •
• • — • • • — • • — • • —
• • — • — — • — —— — — —
• • • — • • — • • —• • •
Test Kinematic viscosity Elemental analysis Water Free water Particle count Acid number Base number Demulsibility Foam stability/tendency RPVOT Rust inhibition Copper strip corrosion Air release
a Turbine oils typically do not contain organometallic additives, though this test is still recommended to warn of cross
contamination with other products.
wiling to offer any degree of certainty as to levels of contamination that may be present in new lubricant deliveries, particularly with respect to particles and moisture. Because of this, it is advisable particularly for critical applications to conduct certain quality control (QC) tests to insure that both the lubricant performance properties and cleanliness have not been compromised during manufacture, or due to distributor storage and handling practices. Selecting which QC tests should be performed should be based on product type and the anticipated performance properties of significance. Table 34.4, though not an exhaustive list of all possible tests, provides a list of general guidelines of what tests should be performed. For bulk deliveries, it is both feasible and generally advisable to perform these quality assurance tests on a batch-by-batch basis. For tote or barrel delivery, it is unrealistic to expect what can amount to several hundred dollars of tests on each barrel. Instead, it is advisable to adopt a more pragmatic approach by randomly sampling (say every 10 or 20 barrels received) to, at a minimum, provide at least some accountability to the lubricant supplier. Some tests can be performed fairly inexpensively meaning that it may be justifiable to run these tests on each new barrel — particularly where the lubricant will be used in critical applications. Under these circumstances, tests such as viscosity, water content, particle count, and elemental additive content should
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34-14
Handbook of Lubrication and Tribology TABLE 34.5
Common Oil Analysis Field Tests
Test for Soot Particle contamination (patch) Fuel dilution Moisture contamination
Blotter spot Patch test kit, sediment Odor, blotter spot, viscosity comparitor Crackle, opacity, static sit, calcium hydride tester
Freon/refrigerant in oil Glycol Viscosity change AN/BN (corrosion potential)
Crackle test and visgage Blotter spot, Schiff ’s reagent Viscosity comparitor Field kits
Ferrous wear debris
Filter inspection, magnetic plug/probe, magnet patch test Filter inspection, patch test Blotter spot Blotter spot, color, odor, AN kits, viscosity comparitor
Nonferrous wear debris Dispersancy Oxidation Demulsibility Antifoam Additive depletion (overbase and antioxidant additives)
FIGURE 34.14
Field tests
Blender comparitor Blender comparitor AN kits
Falling ball viscometer — “visgage” (courtesy Louis C. Eitzen Co. Inc.)
all be conducted routinely, not just as a QC check, but also to provide a baseline for subsequent in-service oil analysis. In many instances, it is desirable to be able to perform QC checks onsite, allowing immediate feedback and the ability to “accept” or “reject” batches or barrels without having to wait for data to be returned from the lab. While many more sophisticated users have stared using high quality onsite oil analysis instruments such as particle counters and viscometers, simple QC checks can be performed with little or no costs and just a small initial investment. Table 34.5 illustrates which test can be performed and how these can be achieved onsite. Perhaps the most important property of any lubricating oil is viscosity. Insuring that any received lubricant is “in-grade” is easy to accomplish using a simple hand-held device known as a falling ball viscosity comparator (Figure 34.14). Though not as accurate as ASTM D445, this quick, simple field test
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Lubricant Storage, Handling, and Dispensing
34-15
is usually sufficient to determine whether new lubricant batches fall within their designated ISO or SAE viscosity grade. Similarly, the presence of free water can be easily detected using a standard hot-plate crackle test. In this procedure, oil is dropped onto a hot plate around 325◦ F. Any significant quantities of free water, cause water vapor bubbles to form, resulting in a scintillation or “crackle” effect much like placing wet food into a hot fry pan. Again, while not as accurate ASTM procedures such as D6304 (Karl Fischer Moisture) or D95 (Dean and Stark distillation), it is a simple, easy-to-perform test that provides immediate feedback. Determining the presence of particles in new oil can be a little more difficult. While the use of onsite electronic optical particles counters is becoming more and more prevalent, a simple field test involving a patch test kit can oftentimes be as effective for assessing the cleanliness of new oils. In this procedure, oil is passed through a filter patch (typically a 2 µm or smaller patch pore size is recommended for new oils). After the patch is solvent-rinsed to remove excess oil, the particle types and concentration can be examined with a small field microscope. Using a particle size and distribution comparator chart can also aid in providing a quantitative assessment of the degree of particle contamination, again allowing a “pass/fail” assessment to be made for any new lubricant batch received.
34.7 Conclusion Of all the factors associated with lubrication, insuring that new lubricants meet minimum performance and cleanliness requirements is perhaps the most fundamental in insuring equipment reliability and longevity. While lubricant suppliers and manufacturers have a vital role to play in this process, deploying the practices outlined in this chapter should help insure that end-users of lubricants can achieve best practice in lubricant storage and handling and in turn insure their equipment achieves optimum levels of availability and reliability.
References [1] [2] [3] [4]
ISO 4406:99 reference. API Standard 620 — Design and Construction of Large, Welded, Low-Pressure Storage Tanks. API 650 — Welded Steel Tanks for Oil Storage. EPA 40 CFR PART 280 — Technical Standards and Corrective Action Requirements for Owners and Operators of Underground Storage Tanks. [5] “Best Practices in Bulk Lubricant Storage and Handling,” Matthew Dinslage, Jim C. Fitch, and Sabrin Khaled Gebarin, Lubrication Excellence 2004 Conference Proceedings, Noria Corporation Publishing.
© 2006 by Taylor & Francis Group, LLC
35 Conservation of Lubricants and Energy 35.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35.2 Conservation of Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . .
35-1 35-2
Trends • Improved Manufacturing and Formulation • Packaging and Handling
35.3 Lubricant Utilization . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
35-4
Extended Life Lubricants • Synthetic Lubricants and Functional Fluids • Greases • Solid Lubricants
35.4 Conservation of Energy . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35.5 Energy-Conserving Fluid Properties . . . . . . . . . . . . . . . . .
35-5 35-6
Kinematic Viscosity • Viscosity Index • Non-Newtonian Properties • High-Temperature Shear Stability • Pressure–Viscosity Coefficient • Bulk Modulus • Boundary Film-Strength Properties • Grease Consistency • Grease-Channeling Properties
35.6 Wear . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35.7 The Surprising Role of Particle Contamination on Fuel Economy . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
35-7 35-7
Antiwear Additive Depletion • Combustion Efficiency Losses • Frictional Losses • Viscosity Churning Losses • Stiction Losses
Robert L. Johnson and James C. Fitch Noria Corporation
35.8 Role of Lubrication Practices . . . . . . . . . . . . . . . . . . . . . . . . . 35.9 Role of Machine Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35.10 Environmental Stewardship . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
35-9 35-9 35-10 35-10
35.1 Introduction Many of the materials essential to global industrial markets have been identified as being in potential short supply. A substantial number of those materials originate from regions of the world where interrupted supply is a real and present risk. Petroleum is a finite world resource with continuing supply and economic problems, with more special concerns for fuels than for petroleum based lubricants. It is the objective of this chapter to present specific conservation practices for lubricants and functional (hydraulics, coolants, etc.) fluids used in tribological components and for energy. The treatment will necessarily be brief, but references will point to more detail information. 35-1
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35.2 Conservation of Lubricants 35.2.1 Trends Lubricants and hydraulic fluids are most commonly derived from petroleum sources. Petroleum serves as the base-stock in the majority of liquid lubricants and greases and very often is a raw material in the synthesis of unique lubricants (typically referred to as synthetics). Details of the distribution of the industrial and automotive lubricant markets in the United States are shown in Figure 35.1. A five-year trend of industrial and automotive lubricants sales is shown in Figure 35.2.
35.2.2 Improved Manufacturing and Formulation The manufacturing methods used in refining lubricants are significant to material conservation and energy. Modern hydrogen processed bases oils (API Groups II and III) are popular because they (1) minimize
6%
Industrial engine oils Metalworking oils
Automotive
2%
Industrial 13%
General industrial oils
20%
Process oils
42%
Multigrade crankcase oils Automotive transmission and hydraulic fluids Other automotive oils
11% 2%
Monograde crankcase oils
4%
FIGURE 35.1 2002 sales of lubricants (total: 2.4 billion gal). (Taken from NPRA, Lubes‘n’Greases magazine. With permission.) 1600
Sales (Millions of Gallons)
1400 1998 1999 2000 2001 2002
1200 1000 800 600 400 200 0 Automotive oils
Industrial oils
Grease
FIGURE 35.2 Five-year trend of total reported sales of lubricants. (Taken from NPRA, Lubes‘n’Greases magazine. With permission.)
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Conservation of Lubricants and Energy
35-3
1.8
Price (U.S.$/gal)
1.5
Group II base oil∗ Group I base oil∗ West Texas intermediate crude (monthly average) ∗Base oil prices are lowest U.S. postings for each category.
1.2 0.9 0.6
. ov . D ec Ja . n. 04 Fe b. M ar . Ap r. M ay Ju ne Ju ly Au g. Se pt .
ct
N
O
Se p
t.
0.3
FIGURE 35.3 Base oil and crude prices. (Copyright 2004, Lubes‘n’Greases magazine. With permission.) 80
Naphthenic
70 Volume (millions of barrels)
48.2 60
49
44.4 45.5
48.7
Paraffinic
51.6 53.2 53.8 51.6 51.7
51
51
50 40 30 20 10 0
1992 1993 1994 1995 1996 1997 1998 1999 2000 2001 2002 2003
FIGURE 35.4 United States refinery production volumes. (Taken from NPRA, Lubes‘n’Greases magazine. With permission.)
solvent, acid, and clay, (2) reduce by-product disposal problems, (3) increase yield, (4) lower costs, (5) permit use of wide range of crudes, (6) improve color, and (7) give higher viscosity index (VI). Figure 35.3 shows a 12-year base oil production trend for paraffinic and naphthenic stocks. Pricing trends in 2003–2004 for Group I mineral oils and Group II hydroprocessed oils is shown in Figure 35.4.
35.2.3 Packaging and Handling Packaging and handling practices as detailed in Chapter 28 have a significant contribution to conservation of lubricants. Contamination commonly occurs when containers are left open in point-of-use and storage areas. The presence of moisture and particulates degrade the effectiveness of all types of lubricants, in many cases requiring that the lubricants be discarded. Recommended practice often requires packaging in sealed containers of the size, or fraction thereof, usually needed for the system. Where drum-sized
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or tote-bin containers are used, they should not be left open and storage position should be such that contamination cannot accumulate at the vent, drainage, or fill holes. Usually a horizontal drum position is indicated and such a simple consideration can conserve lubricants by minimizing discards and assuring expected performance.
35.3 Lubricant Utilization 35.3.1 Extended Life Lubricants Extended life cycle use of lubricants is clearly conservation effective. The limiting life of many types of lubricants is determined by either intolerable contamination or additive depletion. Formulators, recognizing the importance of conservation and cost effectiveness, are providing additive packages for extended service. With sufficient life from improved additives and more robust base stocks, the primary importance is contamination and that may be controlled by vigilant contaminant exclusion and filtration. In considering the potential for extending the lubricant life cycle, the experience of airlines with turbine engine lubricants is of interest. Using synthetic lubricants, the airlines do not change lubricants between major engine maintenance events; for several thousand hours operation only minor makeup oil is added to compensate for leakage and consumption. That is the equivalent of millions of miles per oil change. Another illustration of extended life cycle lubrication is in hermetically sealed household refrigerators where continuous operation for over two decades is common. Such hermetically sealed systems are ideal for extended lubricant service life, emphasizing again the importance of contamination control.
35.3.2 Synthetic Lubricants and Functional Fluids Conservation practices important to petroleum lubricants are also generally relevant to synthetic lubricants and functional fluids. Because synthetics can have greater oxidation stability than mineral-based lubricants, the possibilities for extended service life have even greater potential. The cost of synthetics is higher than refined petroleum products and the applications are more specialized and controlled than for petroleum lubricants. Airlines have used reprocessed phosphate ester hydraulic fluids; polyphenyl ether lubricants have also been recycled for some military use. In those cases, low initial cost as well as favorable collection and processing circumstances allowed recycling to be cost effective.
35.3.3 Greases Most greases are compounded from mineral-based lubricants. Conservation and economics can be achieved by many of the same measures cited for petroleum oils. Extended operating life is gained through product improvement and contamination control. Improved sealing practices in applications offer major gains in conserving greases by minimizing contamination and leakage. More stable grease thickeners, as well as new and improved lubricant additives, have allowed significantly extended regreasing periods for mechanical components. Products with capabilities for high operating temperatures and others with resistance to displacement by water impingement are examples of lubricants that have extended relubrication intervals. For example, wear life of components have markedly improved while the incidence of mechanical failures have markedly decreased by modern thickeners, base oils, and additives.
35.3.4 Solid Lubricants Solid lubricants are often used in greases, in slurries with liquid carriers, in bonded surface films, and in self-lubricating composite materials. Optimum concentrations have been determined in each application system. The extended use of solid lubricants in many applications has significant impacts on reliability and energy uses. Although the total volume of solid lubricants is less than one might consider in view of the many applications, even lesser quantities can often be used. The usual thickness of bonded films is in the range of 0.0002 to 0.0005 in.; a thickness dictated by production control limitations of the
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Conservation of Lubricants and Energy
35-5
coating processes. Vacuum processes such as sputter-ion deposition allows improved film uniformity and, therefore, thinner films are used. Similarly, with self-lubricating bulk solids, optimizing the required functions can conserve lubricating materials. In addition to friction and wear considerations in selection of self-lubricating solids, resistance to environment, mechanical stress, and thermal stress must be anticipated. Composite systems and specific function coatings can greatly extend the lives and hence conserve such materials.
35.4 Conservation of Energy It seems counterintuitive that lubricants selected to optimize wear control may not be optimum when it comes to energy conservation. In fact, in view of today’s growing pressure to reduce demand on nonrenewable energy resources and increase operating profits, we are increasingly facing a shift of emphases from past lubrication objectives. Energy-conserving lubrication offers motivation on several fronts. Consider the following: 1. When energy consumption is economized, equipment operating costs come down, translating to a boost in business profits, regardless of whether the energy source is renewable (hydro, solar, wind) or nonrenewable (coal or petroleum). For many industries, the cost of energy far exceeds the cost of maintenance, machine repair, and even downtime. A small percentage of reduction in energy consumption can translate into large returns. 2. Reduced demand on nonrenewable fossil fuels means cleaner air, reduced greenhouse gas emissions and a healthier environment (of growing political and social importance in view of the Kyoto Protocol on global warming, ISO 14001, Clear Air Act, etc.). When fuels don’t burn, there is no waste stream (smoke stack, tail pipe, etc.) and the risk of pollutants from emissions such as nitrogen oxides (the principle component of smog), sulfates, CO2 , and unburned hydrocarbons is reduced proportionally. 3. With few exceptions, lubricants and lubrication methods that reduce energy consumption will also reduce heat and wear debris generation; however, the reverse may not hold true. When heat and wear debris are reduced, less stress is imposed on additives and the base oil. The result will be longer thermal and oxidative stability, and in turn, longer oil drains, lower oil consumption, and the ancillary costs associated with oil changes (as much as 40 times the cost of the lubricant itself!). 4. When lubricant consumption is reduced, so too is the disposal of environmentally polluting waste oil and certain suspended contaminants, some of which may be hazardous and toxic; ethylene glycol (antifreeze), for example. 5. When there is better economy in the consumption of both petroleum fuels and mineral-based lube oil, there is reduced dependence on foreign sources of crude oil, including those from politically unstable countries. 6. In certain countries, including European Union nations, reductions in the consumption of nonrenewable fuels can avert energy tax penalties such as the Climate Change Levy in the United Kingdom. In recent years, there has been growing interest in energy-conserving lubricants and energy-conserving lubrication. Note, energy-conserving lubricants relate to formulation (base stocks and additives) and their selection for machine application. In contrast, energy-conserving lubrication includes the use and application of lubricants (change intervals, delivery methods, lube volume, etc.). Both can have a marked impact on energy conservation. Energy economy and wear control do not necessarily go hand-in-hand. In certain cases, they may be conflicting objectives. For many organizations, environmental factors and energy costs fall low on the list of priorities compared to productivity and machine reliability. In such cases, the principle objective of the practice of lubrication is to reduce wear and maximize reliability.
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35.5 Energy-Conserving Fluid Properties When formulating or selecting lubricants, the following properties are important in reducing friction and energy consumption.
35.5.1 Kinematic Viscosity When it comes to energy economy, viscosity can be both an inhibitor and an enabler. Recalling the well-known Stribeck curve, the oil film produced by hydrodynamic lubrication is directly influenced by viscosity. However, too much viscosity causes churning losses (excessive internal oil friction) and heat production, especially in engines, gears, bearings, and hydraulics. In addition to energy losses, this increased heat can more rapidly breakdown the oil and its additives.
35.5.2 Viscosity Index Kinematic viscosity by itself defines an oil’s resistance only to flow and shear at a single temperature, typically 40 or 100◦ C. However, in normal operation, lubricating oils transition through a wide range of temperatures. As such, it is the oil’s VI combined with kinematic viscosity that defines what the viscosity will be at a specific operating temperature. Will it be too high when ambient start-up temperatures are low and too low when operating temperatures are high? Likewise, what will be the time-weighted average viscosity of the lubricating oil during the machine’s service life? It is this average viscosity that defines energy consumption, not the occasional temperature-based viscosity excursions that may have a greater impact on wear (cold starts for instance). In general, the significance of VI on energy conservation and wear is often sharply underestimated.
35.5.3 Non-Newtonian Properties Fluids that exhibit shear-dependent viscosity changes (known as the non-Newtonian fluids) are known to reduce energy consumption in many machines. Good examples are VI-improved motor oils (multigrades) and many all-season hydraulic fluids. As fluid movement increases (shearing) during service, the oil’s effective viscosity self-regulates slightly downward, along with energy consumption. This, in part, explains why high-VI, multigrade motor oils are generally those that are designated energy-conserving by the API.
35.5.4 High-Temperature Shear Stability Synthetics and other high-VI base oils perform best here, as do multigrade formulations with low-VI concentrations. Temperature and viscosity shear-back at high temperatures can lead to loss of critical lubricant film strength, leading to power losses and wear. However, temporary shear thinning can also reduce parasitic viscous drag in crankshaft bearings.
35.5.5 Pressure–Viscosity Coefficient The role of pressure–viscosity (PV) coefficient on energy consumption is not well defined in the literature. However, it is widely understood that many base oils exhibit a sharp increase in viscosity as pressure rises; a necessary quality of lubricants in achieving effective elastohydrodynamic lubrication (EHD). Some oils, such as mineral oils and PAOs (Polyalphaolefins), have higher PV coefficients than others, such as esterbased synthetics and water-based fluids. While high PV coefficients may be important at reducing contact fatigue wear, in some cases, this property may contribute to lower fuel economy. The high pressureinduced viscosity in sliding frictional zones and in hydraulic systems could result in exceedingly high viscous drag energy losses.
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Conservation of Lubricants and Energy
35-7
35.5.6 Bulk Modulus A fluid that is sponge-like and easily compressed has low bulk modulus of elasticity. The more compressible a lubricant is, the more potential for lost energy and heat production. This is especially true in hydraulic and lube oil circulating systems.
35.5.7 Boundary Film-Strength Properties Many lubricants and hydraulic fluids can gain considerable film strength under boundary and mixed-film lubrication from the base oil, without the need for additives. A phosphate ester synthetic is an example of a fluid with intrinsic lubricity. Most other lubricants rely on additives such as friction modifiers, antiwear agents, extreme pressure (antiscuff), solid lubricants, and fatty acids. The effectiveness of these additives at reducing wear, friction, and energy consumption can vary considerably between the different additive types employed. The performance of these additives also varies by machine and application (load, speed, metallurgy, temperature, and contact geometry).
35.5.8 Grease Consistency The consistency of grease can have an impact on energy consumption in ways similar to viscosity. The energy needed to move grease in frictional zones and in adjacent cavities by moving machine elements is affected by its consistency and shear rate (grease is non-Newtonian). So too, energy is required in some applications to pump grease to bearings and gears. Pumping energy losses is influenced, in part, by grease consistency and thickener type.
35.5.9 Grease-Channeling Properties A grease that has good channeling characteristics helps keep the bulk lubricant away from moving elements, avoiding excessive churning and drag losses. Poor channeling characteristics may lead to increased energy consumption, heat production, and base oil oxidation.
35.6 Wear Wear from boundary friction can have a near-term adverse effect on fuel economy and can generate heat. Wear is often the result of such things as lubricant starvation, low viscosity, poor or degraded antiwear additive performance, dirt and other contaminants, deposits (e.g., ring grooves), etc. One researcher identified an 8◦ C (14.4◦ F) increase in bearing oil temperature, which he attributed to solid contamination of the oil. However, there is also a long-term effect especially in engines. Over a period of time, an engine loses so much metal (rings, cylinder walls, cam follower, cam lobe, etc.) that combustion efficiency is severely impaired. (This is discussed in greater detail in the next section.) Loss of combustion efficiency directly impacts fuel economy and tailpipe emissions. In this respect, the average fuel economy performance of a motor oil over a period of 100,000 miles or more is a better assessment of its life-cycle performance as opposed to snapshot energy consumption assessments of new engines and new oils.
35.7 The Surprising Role of Particle Contamination on Fuel Economy When a lubricant degrades, it forms reaction products that become insoluble and corrosive. So too, the original properties of lubricity and dispersancy can become impaired as the lubricant ages and additives deplete. Much has been published about the risks associated with overextended oil drains and the buildup of carbon insolubles from combustion blow-by, especially in diesel engines.
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There have been surprisingly few studies published on the impact of fine abrasives in a motor oil as it relates to fuel economy over the engine’s life. Yet, it is not hard to imagine numerous scenarios in which solid abrasives suspended in the oil could diminish optimum energy performance. Below is a list of several scenarios.
35.7.1 Antiwear Additive Depletion High soot load of crankcase lubricants has been reported to induce abrasive wear and impair the performance of zinc dialkyldithiophosphate (ZDDP) antiwear additives. The problem is more pronounced in diesel engines. Some researchers believe that soot and dust particles exhibit polar absorbencies, and as such, can tie-up the antiwear additive and diminish its ability to control friction in boundary contacts (cam nose, ring/cylinder walls, etc.). However, there appears to be greater evidence that soot itself is highly abrasive in frictional zones where dynamic clearances are 1 µm. These include cam/follower and ring reversal areas on cylinder walls.
35.7.2 Combustion Efficiency Losses Sooner or later, wear from abrasive particles and deposits from carbon and oxide insolubles will interfere with efficient combustion in an engine. Valve train wear (cams, valve guides, etc.) can impact timing and valve movement. Wear of rings, pistons, and cylinder walls influences volumetric compression efficiency and combustion blow-by resulting in power loss. Particle-induced wear is greatest when the particle sizes are in the same range as the oil film thickness (Figure 35.5). For diesel and gasoline engines, there are a surprising number of laboratories and field studies that report the need to control particles below 10 µm. One such study by General Motors concluded that, “controlling particles in the 3 µm to 10 µm range had the greatest impact on wear rates and that engine wear rates correlated directly to the dust concentration levels in the sump.”
35.7.3 Frictional Losses When hard clearance-size particles disrupt oil films, including boundary chemical films, increased friction and wear will occur. One researcher reports that 40 to 50% of the friction losses of an engine are attributable to the ring/cylinder contacts, with two-thirds of the loss assigned to the upper compression ring. It has been documented that there is an extremely high level of sensitivity at the ring-to-cylinder zone of the engine to both oil- and air-borne contaminants. Hence, abrasive wear in an engine’s ring/cylinder area translates directly to increased friction, blow-by, compression losses, and reduced fuel economy.
35.7.4 Viscosity Churning Losses Wear particles accelerate the oxidative thickening of aged oil. High soot load and lack of soot dispersancy can also have a large impact on oil viscosity increases. Viscosity-related internal fluid friction not only increases fuel consumption, but also generates more heat, which can lead to premature degradation of additives and base oil oxidation.
35.7.5 Stiction Losses Deposits in the combustion chamber and valve train can lead to restricted movements in rings and valve control. When hard particle contamination agglomerates with soot and sludge to form adherent deposits between valves and guides, a tenacious interference, called stiction, results. Stiction causes power loss and engine knock. It causes the timing of the port openings and closings to vary, leading to incomplete combustion and risk of backfiring. Advanced phases of this problem can lead to a burned valve seat.
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Conservation of Lubricants and Energy
35-9
Piston ring
30 m/sec Cylinder wall
Cylinder wall wear (mm)
Oil film 3
10–2
10–3
Wear highest where particle size equals film thickness
10–4 Babbit
Copper-lead
0.1 1 10 Minimum film thickness Particle diameter Diesel engine oil film thickness Component
Oil film thickness (mm)
Ring-to-cylinder
3.0–7
Rod bearings
0.5–20
Main shaft bearings
0.8–50
Turbocharger bearings
0.5–20
Piston pin bushing
0.5–15
Valve train
0–1.0
Gearing
0–1.5
FIGURE 35.5 Common oil film thickness. (Taken from Noria Corporation. With permission.)
35.8 Role of Lubrication Practices While lubricant formulation and selection are important, energy conservation is also influenced by machine design and lubricant application factors. A superior lubricant cannot offer redemptive relief for poor lubrication practices and machine design. Even the very best lubricants cannot protect against destruction caused by dirt and water contamination. Overgreasing of bearings is known to increase frictional losses and raise bearing temperature. The same is true for bearings that are underlubricated. For bath lubricated bearings and splash lubricated gears, a change in oil level by as little as one-half inch (1.3 cm) can increase temperature by more than 10◦ C. This, of course, translates to greater energy consumption, shorter oil life, and increased wear. Excessively aerated oils due to worn seals and wrong oil levels can have similar effects (loss of bulk modulus). There have also been studies showing the negative effects of overextended oil change interval on fuel economy in diesel engines. Additionally, overextended filter changes cause excessive flow resistance and fluid bypass. Both can often be corrected by the frequent and proper use of oil analysis in selecting the optimum oil and filter change interval, tailored to equipment type and its application.
35.9 Role of Machine Design A machine’s design and the quality of its manufacture can also impact energy economy. Together with operating load and speed, machine design influences the type of lubricant that must be employed for
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Handbook of Lubrication and Tribology
wear protection and energy efficiency. Already mentioned is the importance of viscosity films produced by hydrodynamic and elastohydrodynamic lubrication as well as boundary film strength from additives and polar base oil chemistry. These lubrication regimes relate to the contact dynamics associated with a machine’s design and operating conditions. Additionally, specific film thickness, also known as lambda, brings into the picture the influence of surface roughness and shaft alignment. Many users and suppliers have reported energy savings from total-loss lubricant delivery technologies such as oil mist and centralized lubrication systems. The amount of fluid that a machine uses to lubricate frictional surfaces at any moment is extremely small compared to the amount of fluid some machines must keep in continuous motion. The advantage of some total-loss lubrication systems is that there is minimal loss of energy from constant fluid churning and flow resistance of lubricants moving through lines. An example of internal fluid friction is observed when an oil is placed in a bottle and then shaken. The oil’s temperature will rise. In addition, bath, splash, and recalculating lubrication systems use the same oil over and over. As we all know, this reused oil over time can become impaired by loss of additives, base oil oxidation, and rising concentrations of contaminants. In contrast, when well engineered and in the right application, oil mist and other certain total-loss systems can provide a continuous supply of fresh, clean, and dry new oil. Energy consumption is also influenced by the size and type of fittings, oil lines, and filters.
35.10 Environmental Stewardship In summary, lubricants, lubrication, and contamination play no small role in reducing energy consumption and the general wasteful use of petroleum products, including lubricants. Increasingly the selection and use of lubricants is going to stress greater importance on energy and environmental impact. At the same time, we will not lose sight of other vital objectives including machine reliability and safety.
References [1] Anon, Technical Options for Conservation of Metals, Library of Congress Catalog Card Number 79-600172, Congress of the United States, Office of Technology Assessment, Washington, D.C., 1979. [2] Barnett, R.S., Molybdenum Disulfide as an Additive for Lubricating Greases — A Summary 1973, Climax Molybdenum Company, Greenwich, Conn., 1973. [3] ASLE, Proceedings of the 2nd International Conference on Solid Lubrication, SP-6, 3978 Library of Congress Catalog Card Number 78-67090, American Society of Lubrication Engineers, Parkridge, Ill., 1978. [4] 2004–2005 Lubricants Industry Sourcebook, Lubes’n’Greases, Falls Church, VA, 2004, pp. 4–10. [5] Addison, J. and Needelman, W., Diesel Engine Lubricant Contamination and Wear, East Hills, NY, Pall Corp, 1986. [6] Andrews, G.E., Hall, J., Jones, M.H., Li, H., Rahman, A.A., and Saydali, S., The Influence of an Oil Recycler on Lubricating Oil Quality with Oil Age for a Bus Using In-Service Testing, Presented at SAE 2000 World Congress (SAE Paper 2000-01-0234), 2000. [7] Ballentine, B., Motor Oils — Fuel Economy vs. Wear, Machinery Lubrication, 2003. [8] Barris, M.A., Total Filtration: The Influence of Filter Selection on Engine Wear, Emissions and Performance, Presented at SAE Fuels and Lubricants Meeting (SAE paper 952557), 1995. [9] Feldhaus, L.B. and Hudgens, R.D., Diesel Engine Lube Filter Life Related to Oil Chemistry, Presented at SAE International Fuels and Lubricants Meeting (SAE Paper 780974), 1978. [10] Fitch, J.C., Troubleshooting Viscosity Excursions, Practicing Oil Analysis, May–June 2001. [11] Foder, J. and Ling, F.F., Friction Reduction in an IC Engine Through Improved Filtration and a New Lubricant Additive, Lubrication Engineering, October 1985.
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Conservation of Lubricants and Energy
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[12] Madhaven, P.V. and Needelman, W.M., Review of Lubricant Contamination and Diesel Engine Wear, Presented at SAE Truck and Bus Meeting and Exposition (SAE Paper 881827), 1988. [13] McGeehan, J. Uncovering the Problems with Extended Oil Drains, Machinery Lubrication, September–October 2001. [14] Staley, D.R., Correlating Lube Oil Filtration Efficiencies with Engine Wear, Presented at SAE Truck and Bus Meeting and Exposition (SAE Paper 881825). [15] Troyer, D., Consider This (Sidebar to From under the Hood — Multigrade oil — To Use or Not to Use), Machinery Lubrication, July–August, 2001. [16] Fitch, J.C., Energy-conserving lubrication — the endless debate, in How to Select a Motor Oil and Filter for Your Car or Truck, 2nd ed., Fitch, J.C. Ed., Noria Corp., Tulsa, OK, 2003, Chap. 7.
© 2006 by Taylor & Francis Group, LLC
36 Centralized Lubrication Systems — Theory and Practice 36.1 The Philosophy of Lubrication . . . . . . . . . . . . . . . . . . . . . . . 36.2 Using the Correct Grease . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 36.3 Centralized Grease Lubrication Systems . . . . . . . . . . . . .
Paul Conley and Ayzik Grach Lincoln Industrial
36-1 36-1 36-5
The Series Progressive System • Single Line Parallel Systems • Dual-Line Systems • Types of Two Line Systems
36.4 Pumping of Grease and Viscous Materials . . . . . . . . . . .
36-30
36.1 The Philosophy of Lubrication The purpose of lubrication is to prevent metal to metal contact between two moving members, reduce friction, remove heat, and flush out contaminants. The philosophy of an automatic lubrication system is to deliver the right amount of lubricant at the right time. The method to do this is to deliver small quantity of lubricant to a bearing often. The lubricant is delivered in small portions as the bearing consumes it. Applying more lubricant than the bearing can consume in a short period of time is often thrown off the moving components, causing housekeeping and environmental safety hazards. A properly designed automatic centralized system can deliver just the right amount of lubricant the bearing needs. To illustrate this fact, Figure 36.1 shows the effect of manual lubrication in comparison to automatic lubrication. With manual lubrication, lubrication of the bearings is normally performed when the machine is not running. The even distribution of grease between the shaft and journal cannot be assured. With automatic lubrication, the distribution of grease between the shaft and journal is assured (see Figure 36.2.)
36.2 Using the Correct Grease When designing an automatic grease lubrication system, the first priority is to use the correct lubricant. The lubricant must have the properties consistent for the application. For applications requiring grease, 36-1
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36-2
Handbook of Lubrication and Tribology Manual lubrication cycles Too much grease
Bearing seal breached
Max. bearing capacity Right amount grease
Too little grease 0
1
2
3
Missed lube event
Contamination begins
Bearing severe
4 Starved wear
Time between lubrication events days/weeks/months
Optimum lubrication amt. Maximum bearing lubricant capacity
Manual method.
Too much
Amount of lubricant dispensed
FIGURE 36.1
Extreme over/under lubrication Over/under lubrication
Max. bearing capacity Right amount
Too little
0
2
4
6
8
Time between lubrication events in hours
10 Extreme over/under lubrication Over/under lubrication Optimum lubrication amt. Maximum bearing lubricant capacity
FIGURE 36.2 Automated method.
the base oil lubricant properties and the thickener must meet the application needs to include it in an automatic system. In a perfect world, the best grease to use is one that contains the proper base oil lubrication properties with the heaviest or stiffest thickener soap. The purpose of grease over oil is to make the lubricant stay put in the bearing. Grease is normally used in applications where there are heavy loads and slow relative motion between the shaft and bearing. Grease should also be used when strong shock loading is present, frequent starting and stopping, and when there is insufficient or no bearing seal.
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Centralized Lubrication Systems — Theory and Practice
36-3
Using oil in these applications is not advisable. The oil would run off or out of the bearing causing a drop-off in film thickness to prevent metal to metal contact. The use of grease prevents this from occurring. It is important to know that the soap thickener in the grease does not provide any lubrication. Oil is suspended in the thickener, providing the necessary lubrication. For high load situations, grease can suspend solid lubricant and extreme pressure (EP) additives in the base oil to provide additional protection against metal to metal contact. Some of the desired characteristics of grease for use in automatic lubrication systems are: • Good shear and mechanical stability (reference ASTM D217) • Does not readily allow the base oil to separate out (reference ASTM D4425) • Good reversibility defined as the characteristic of the grease to recapture the base oil should separation of oil from the thickener occur. Separation of oil may be induced under high load and pressure conditions in the bearing. A grease with good reversibility will return to its original consistency after the high load is removed • Possesses good washout resistance • Protects against corrosion • Good pumpability and ventability With the development of complex greases, the above characteristics can be achieved. With the use of synthetic base oils, the temperature ranges at which grease can operate have been expanded. For grease used in automatic systems, one very important property to be considered is the measure of its pumpability or ventability. It would not be practical to produce a grease that is so stiff that it cannot be pumped or which produces such a resistance to flow that it cannot be used in an automatic system. The best combination is to use the thickest grease that can be pumped effectively. If the grease cannot be pumped, then it has no practical use; its resistance to flow is so great such that the pressure produced in the supply lines would be excessive and thus prohibitive. When centralized grease lubrication was introduced in the mid-1930s, the manufacturers of these centralized grease lubrication systems needed to understand the properties and flow characteristic of grease. Grease is classified as a non-Newtonian thixotropic pseudoplastic (viscosity decreased as the shear rate is increased) fluid. Most fluids such as water and oil are Newtonian fluids. That means the viscosity is constant as the shear rate changes. Figure 36.3 illustrates three types of viscosities. Many grease manufacturers rate the grease by its National Lubricating Grease Institute (NLGI) rating. This is a method that only measures the general stiffness at one temperature. According to the ASTM test method, the grease stiffness is measured at 77◦ F. The higher the NLGI rating of the grease, the stiffer the grease is and, in general, the higher apparent viscosity. The NLGI rating alone is not sufficient to determine its appropriate use in an automatic system. The user or designer of an automatic system must know the flow properties of the grease at the lowest temperature of the application. For example, a grease with an NLGI rating of 2 at 77◦ F will
Dilatant
Newtonian
Viscosity
Grease Pseudoplastic Plastic Shear rate (velocity)
FIGURE 36.3 Flow behavior of different types of fluid.
© 2006 by Taylor & Francis Group, LLC
36-4
FIGURE 36.4
Handbook of Lubrication and Tribology
Lincoln Ventmeter.
have an NLGI rating of 3 at 50◦ F. Another grease with an NLGI rating of 2 at 77◦ F may have an NLGI rating of 3 at 30◦ F. An NLGI number 1 grease will often behave like an NLGI number 2 grease at 30◦ F. For this reason, the best indication of a grease’s suitability for use in an automatic lubrication system is the Lincoln Ventmeter. Manufacturers and designers of automatic grease systems often use the Lincoln Ventmeter viscosity as the standard for selecting the correct grease for use in an automatic lubrication system. To provide an understanding of the usefulness of the Lincoln Ventmeter, let us take a look at the Lincoln Ventmeter (Figure 36.4). The Lincoln Ventmeter was developed as an instrument to measure the flow limits of grease. It is more precise than the classic NLGI number rating. It was developed in the early 1950s, and since 1965 has been used extensively in determining acceptable performance out of a single line injector type centralized lubrication systems. Even today, samples of grease are tested and evaluated using the Lincoln Ventmeter by grease manufacturers and designers of centralized lubrication systems. By measuring the flow ability of grease, an application engineer/technician or grease manufacturer can select the pump and line size to ensure good performance of the centralized grease lubrication system. The usefulness of the Lincoln Ventmeter is most noted in the following three ways: 1. What types of grease, according to consistency, can be used in a given grease supply line so that the pressure in the system will vent down sufficiently to successfully operate injectors? 2. What supply line length and diameter should be used for a specific type of grease in a centralized lubrication system? 3. When to utilize a lighter NLGI grade grease product so that the system will continue to operate correctly during colder temperatures? The ability of the grease to vent is important for proper operation of the system. It is also good that the grease be vented during the off time of the system so that the oil does tend to separate from the soap. Figure 36.5 shows a schematic of the Lincoln Ventmeter and how the test is conducted. The yield pressure of the grease is obtained in the following manner: With valve 1 closed and valve 2 open, the sample of grease is filled with grease using a pump or grease gun. Valve 2 is then closed. Using a grease pump or a grease gun, the Ventmeter is charged to a pressure of 1800 psig, relief valve number 1 is opened and the grease is discharged. The discharge of the grease will reduce the pressure. The pressure left in the system is read 30 sec after the relief valve is opened. Thirty seconds is a sufficient time for the grease to stabilize. The gauge reading is the Ventmeter viscosity and is measured in pressure in pounds per square
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Centralized Lubrication Systems — Theory and Practice
36-5 25 ft. ∅ 1/4 in Coiled tubing
Pressure guage
Check valve
Pump (lever gun) Valve 1
Valve 2
FIGURE 36.5 Lincoln Ventmeter schematic.
inch. The test is conducted normally at ambient, 30 and 0◦ F. This way the yield pressure is obtained at various temperature conditions. Tests are often done at progressively lower temperatures to establish a value when the grease ceases to flow. If the gauge reading goes to zero within the 30 sec, this is an indication that the grease has effectively no yield limit. Very light viscous grease behaves like oil and can be considered Newtonian. With lower temperatures or stiffer greases, the gauge reading will be some value other than zero. Supply line charts have been developed according to the Lincoln Ventmeter for three types of injectors (see Tables 36.1, 36.2, and 36.3).
36.3 Centralized Grease Lubrication Systems What do they do? The purpose of any automatic centralized lubrication system is to provide the correct amount of grease to the bearing at the right time. The method to do this is for a controller to turn on a pump that supplies grease to a positive displacement valves that will dispense a predetermined amount of grease to the bearing. There are essentially three types of lubrication systems used in industry. The oldest of these are the progressive and dual line systems developed in the late 1800s during the industrial revolution. Single line systems were developed in the late 1930s as an enhancement to the progressive and dual line systems. The main difference between one system and the other is the type of positive displacement valve. Each system is discussed to include principle of operations, features, design considerations, and strengths and weaknesses. The importance of monitoring system performance cannot be understated. Each and every type of lubrication has some provisions for monitoring. The degrees in which the systems can be monitored vary from visual indicators to full electronic transducer feedback monitoring.
36.3.1 The Series Progressive System The system gets its name from the serial and progressive nature in which the valves operate. Basic components of the system: • Pump • Controller • Progressive metering valve
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36-6 TABLE 36.1
Handbook of Lubrication and Tribology Supply Line Chart with Injectors that Require a Venting Pressure of 600 psi or Greater SL-1 and SL-11 supply line chart
NLGI grease
Lincoln ventmeter reading (psi)
SL-1 and SL-11 supply line chart
Nominal pipe size or ID of tube or hose (in.)
Max. supply line length (Ft)
NLGI grease
Lincoln ventmeter reading (psi)
Nominal pipe size or ID of tube or hose (in.)
Max. supply line length (Ft)
#0
0–100 0–100 0–100 0–100 0–100 0–100 0–100 0–100
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
1100 875 700 575 430 270 200 130
#2
300–400 300–400 300–400 300–400 300–400 300–400 300–400 300–400
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
280 200 175 140 100 70 50 30
#1
100–200 100–200 100–200 100–200 100–200 100–200 100–200 100–200
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
500 400 350 270 210 140 110 50
#3
400–500 400–500 400–500 400–500 400–500 400–500 400–500 400–500
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
230 170 140 100 80 55 40 30
#2
200–300 200–300 200–300 200–300 200–300 200–300 200–300 200–300
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
350 280 235 180 140 90 70 40
#3
500–600 500–600 500–600 500–600 500–600 500–600 500–600 500–600
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
190 140 120 90 65 45 36 15
The operation of a progressive system is straightforward. When the time for a lubrication event is necessary, a controller turns on a pump that sends lubricant to the progressive valves, which in turn meter the lubricant to the bearing. The output volume provided by the displacement of the metering piston determines the amount of lubrication to the bearing. The valves can be cycled once or multiple times to deliver an appropriate amount of grease to the bearing. The on-time selection and the pump output capacity determines the total amount of lubrication provided to the bearings. Figure 36.6 is a schematic of a typical series progressive automatic lubrication system. Figure 36.7 is a schematic and the operation of progressive valves and how they work. More than a drilled manifold block, the valve incorporates a series of metering valves, which accurately dispense lubricant from each outlet, overcoming back pressure of up to 1000 psi. Visual monitoring is provided with an indicator pin, which confirms a valve has completed a full cycle. Progressive divider valves are available for grease or oil applications, and in carbon steel and 303 stainless steel for corrosive environments. The inlet passageway is connected to all piston chambers at all times with only one piston free to move at any time. 1. With all pistons at the far right, lubricant from the inlet flows against the right end of the piston A (Illustration 1). 2. Lubricant flow shifts piston A from right to left, dispensing lubricant through connection passages to outlet 2 (Illustration 2). 3. Piston B shifts from right to left, dispensing lubricant through outlet 7. Lubricant flow is directed against the right side of piston C (Illustration 3).
© 2006 by Taylor & Francis Group, LLC
Centralized Lubrication Systems — Theory and Practice TABLE 36.2
Supply Line Chart with Injectors that Require a Venting Pressure of 200 psi SL-32 and SL-33 supply line chart
NLGI grease
36-7
Lincoln ventmeter reading (psi)
SL-32 and SL-33 supply line chart
Nominal pipe size or ID of tube or hose (in.)
Max. supply line length (Ft)
NLGI grease
Lincoln ventmeter reading (psi)
Nominal pipe size or ID of tube or hose (in.)
Max. supply line length (Ft)
#0
0–100 0–100 0–100 0–100 0–100 0–100 0–100 0–100
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
400 300 250 200 150 100 75 50
#2
300–400 300–400 300–400 300–400 300–400 300–400 300–400 300–400
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
100 75 63 50 37 25 18 12
#1
100–200 100–200 100–200 100–200 100–200 100–200 100–200 100–200
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
200 150 120 100 75 50 37 25
#3
400–500 400–500 400–500 400–500 400–500 400–500 400–500 400–500
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
80 60 50 40 30 20 15 10
#2
200–300 200–300 200–300 200–300 200–300 200–300 200–300 200–300
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
135 100 80 66 50 32 25 16
#3
500–600 500–600 500–600 500–600 500–600 500–600 500–600 500–600
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
65 50 40 32 24 NR NR NR
4. Piston C shifts from right to left, dispensing lubricant through outlet 5. Lubricant flow is directed against the right side of piston D (Illustration 4). 5. Piston D shifts from right to left, dispensing lubricant through outlet 3. Piston D’s shift directs lubricant through a connecting passage to the left side of piston A (Illustration 4).
Lubricant flow against the left side of piston A begins the second half-cycle, which shifts pistons from left to right, dispensing lubricant through outlets 1, 8, and 4 of the divider valve.
36.3.1.1 Crossporting a Divider Valve Outputs from adjacent outlets may be combined by installing a closure plug in one or more outlets. Lubricant from a plugged outlet is redirected to the next adjacent outlet in descending numerical order. Outlets 1 and 2 must not be plugged since they have no crossport passage to the next adjacent outlet. In Figure 36.8, outlets 5 and 3 are crossported and directed through outlet 1. In this example, outlet 1 will dispense three time as much lubricant as outlet 7. The tube ferrules in outlets 1 and 7 block the crossport passage so that lubricant flow is only directed through outlets. Typical outlets per cycle can range from as low as 0.0037 in.3 (0.06 cm3 ) to 0.012 in.3 (0.20 cm3 ) for nonmodular progressive valves. To increase the lube amount to the bearing, the valve will have to complete another cycle.
© 2006 by Taylor & Francis Group, LLC
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Handbook of Lubrication and Tribology
TABLE 36.3
Supply Line Chart with Quick Venting Injectors that Have a Venting Pressure of 1000 psi SL-V supply line chart
NLGI grease
Lincoln ventmeter reading (psi)
SL-V supply line chart
Nominal pipe size or ID of tube or hose (in.)
Max. supply line length (Ft)
NLGI grease
Lincoln ventmeter reading (psi)
Nominal pipe size or ID of tube or hose (in.)
Max. supply line length (Ft)
#0
0–100 0–100 0–100 0–100 0–100 0–100 0–100 0–100
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
2000 1500 1250 1000 750 500 375 250
#2
300–400 300–400 300–400 300–400 300–400 300–400 300–400 300–400
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
500 375 310 250 185 125 90 60
#1
100–200 100–200 100–200 100–200 100–200 100–200 100–200 100–200
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
1000 750 625 500 375 250 185 125
#3
400–500 400–500 400–500 400–500 400–500 400–500 400–500 400–500
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
400 300 250 200 150 100 75 50
#2
200–300 200–300 200–300 200–300 200–300 200–300 200–300 200–300
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
660 500 410 330 250 160 125 80
#3
500–600 500–600 500–600 500–600 500–600 500–600 500–600 500–600
2.00 1.50 1.25 1.00 0.75 0.50 0.38 0.25
330 250 200 160 125 80 60 40
Pump station Divider valve Indicator pin
Gauge
Timer Bearings
FIGURE 36.6
Typical progressive lubrication system.
36.3.1.2 The Modular Lubrication System The modular blocks are series progress valves that contain separate valve sections. With nonmodular valves, the measuring pistons are integral to the whole block. The separate valve sections provide more
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Centralized Lubrication Systems — Theory and Practice
36-9
FIGURE 36.7 Valve sequence illustration.
FIGURE 36.8
Crossport of divider valve.
flexibility in selecting output volumes. This is different from a nonmodular valve. Modular valves are more complex and are more difficult to manufacture. For modular systems, individual outlets can be configured to vary in the amount of lubricant volume to the bearing. The modular blocks can be stacked to provide larger outputs per cycle of the valves to larger bearing and smaller outputs for bearing with smaller grease lubrication requirements. To illustrate this point see Figure 36.9 and Figure 36.10 of a modular type progressive valve.
© 2006 by Taylor & Francis Group, LLC
36-10
Handbook of Lubrication and Tribology 1/4⬙ NPSF (F) lubricant outlet
Inlet section Intermediate section Divider valve
1/8⬙ NPSF (F) lubricant outlets
5/16⬙-24 UNF (F) alternate oulet/indicator ports
Tie rod
End section
FIGURE 36.9
FIGURE 36.10
Modular valve construction.
Modular valve.
The construction of modular type progressive valves in sections are shown in Figure 36.9. There is an inlet section, intermediate section, end section all positioned to accept the divider valve. Modular progressive valves can be rated for high pressure up to 7500 psi and for low pressure up to 3500 psi. High pressure valves contain very high manufacturing tolerances between the piston and cylinder to provide adequate sealing.
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Centralized Lubrication Systems — Theory and Practice
36-11
Additionally, individual valves can be configured with a single or twin designation. The internal passages determine if the valve will function as either single or twin. For a single valve section, the displacement of the piston from both ends will deliver the lubricant to one outlet. In effect, the single outlet is capable of delivering twice as much lubricant as a valve configured as a twin; as the output from both sides of the piston is dispensed through one outlet. The operation of a twin valve configuration will deliver the lubricant to separate outlets. The face of the valve is usually marked with an “S” or “T” to designate if the valve section is single or twin. Modular valves can also be configured to increase output delivered in a single cycle of the metering valve by crossporting one valve section to an adjacent valve section. Single valve sections can be crossported only on one side. Twin valve sections can be crossported on both sides. Manufacturers of these types of valves produce kits and design guides that show how and when the use of crossporting can be done. 36.3.1.3 Design Considerations for Progressive System As with all systems, knowing the grease flow characteristics is very important at the lowest temperature the systems is to work under. The grease flow characteristics can be obtained by the Lincoln Ventmeter. The amount of solid additives is important because with metal to metal fits, grease with high concentrations of solid lubricants such as low grade molybdenum disulfide, graphite, or copper antiseize additives can cause the progressive valves to lock up. Applications where a cluster of bearings are grouped together and where there are relative short distances from pump to bearing point make this type of system most effective. The number of lubrication points can be extended by using a main supply block, called a master block or primary valve, to feed a secondary block. The lubricant is divided and distributed through out the primary and secondary blocks to the bearing. With most progressive valves, the maximum number of outlets from one block is 18. With more than 18 lubrication points, secondary blocks will need to be added to cover the additional bearing points. With progressive metering valves, the number of bearing points often does not match the number of outlets. For example, for a 7-point lubrication system, an 8-outlet valve minimally would be used. One outlet would have to be crossported to distribute the lubricant to 7 points. One bearing would then be getting twice as much lubricant as the other 6. All progressive valves are configured to have an even number of outlets with a minimum of 6. Progressive valves are usually available in 6, 8, 12, and 18 outlets. Low output progressive blocks are not typically used for higher outputs of more than 0.10 in.3 /min for nonmodular valves and 5 in.3 /min of modular valves. The reason is that orifices in these valves are typically small and it is hard for grease to flow through. The orifices are usually kept small so that the amount of the piston shift can be controlled, thus keeping the accuracy of the grease volume within reasonable ranges. Three valves at a minimum are necessary to complete the hydraulic logic and for the valves to operate. This is the reason why there can be a minimum of 6 outlets. Never plug an output line in a series progressive valve. Plugging the outlet of a series progressive valve would be similar to causing a line blockage. With a progressive type valve, if one outlet is blocked, jammed, or prevented from moving, then the all pistons will be prevented from working and the metering block would cease to function. The dependency of one piston to another causes this to happen. If one goes then they all go. Crossporting can be done to match outlets with the number of bearings. This is done at the expense of some bearings getting more lubrication than the others. 36.3.1.4 System Monitoring Because of the dependency of the proper operation of all pistons, the need for monitoring is more critical for a progressive system. Visual monitoring can be done by detecting movement of an indicator pin connected to the piston extended out through the valve body or through a sensor that can electronically detect the movement of the indicator pin. A failure of the progressive valve can be detected and an alarm can be signaled that the bearing is not getting lubricated. As with all three types of systems, the pump can be driven pneumatically, hydraulically, or electrically. With grease systems, shovel type positive displacement reciprocating pumps are often used. (See
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Handbook of Lubrication and Tribology
Section 36.4 on pumping systems.) For progressive systems, smaller pumps that produce less output are normally used. 36.3.1.5 Strengths and Weaknesses The weaknesses of a progressive system are: 1. Blockage of one outlet disables the whole system. 2. To use this system in large systems requires complex piping systems. 3. Progressive systems are not flexible to changes in the number of bearing lubrication points once the initial system is set up. When adding or removing lubrication points, the need to relay out the piping is necessary. 4. There is no easy and practical way to adjust the lubricant output to a bearing once the system is set up. 5. The output grease setting to a bearing is in multiples of the outlet volume of measuring piston. Discrete and individual bearing setting can be adjusted once the valve is installed. 6. The amount of grease flow through one valve is limited. 7. The use of close tolerance metal to metal fits makes the valves susceptible to malfunctioning when contaminants exist in the system. The use of grease with solid additives such as molybdenum disulfide or graphite is limited or not recommended. 8. The valves have to make complete cycles to distribute grease to any bearing that requires more grease than one outlet can provide in one cycle. The strengths of the progressive system can be summarized as follows: 1. One valve instead of individual valves can be used to lubricant bearings. One valve can provide metered lubricant to a number of lubricant points. 2. System monitoring can detect a fault for every lube point should one valve or outlet be blocked. 3. There is no need for venting the system. 4. Elastomer seals are not used, which can sometime fail prematurely.
36.3.2 Single Line Parallel Systems The single line parallel gets its name because a single supply line is required and that the measuring valves called injectors can operate independently. The heart of the system is the injector. Basic components of the system: • • • • •
Pump Controller Injectors Vent valve Pressure switch
The operation of a single line lubrication system is straightforward. When lubricant is needed, the controller opens an air solenoid to turn on the pump. The pump produces flow and builds up pressure in the line. When the pressure reaches 1800 psig, the injectors operate and meter a predetermined amount of lubricant to a bearing. A pressure switch usually located farthest away from the pump senses when the pressure has reached 1800 psig. Once reached, the pressure switch sends a signal to the controller indicating that the system pressure was achieved. The controller then turns off the air solenoid valve and thus the air supply to the pump. For electrically or hydraulically operated pump, the controller will shut off electric or the flow of hydraulic fluid respectively. For pneumatically or hydraulically operated pumps,
© 2006 by Taylor & Francis Group, LLC
Centralized Lubrication Systems — Theory and Practice
36-13
Controller
Air inlet Supply line
Injector
Pressure switch
Pump Feedline
3-way vent valve
Bearing
FIGURE 36.11
Single line lubrication system.
when the air/hydraulic supply is turned off, a 3-way valve is activated, which directs any excess grease due to line expansion directly back to the reservoir. For electrical operation, the controller will shut off electric power to a 3-way vent valve. Thus, the pressure in the system can be bled off. This is known as venting the grease. Figure 36.11 is a schematic of a traditional single line system using a pneumatically operated pump. As with all three types of systems, the pump can be operated from all three power sources, pneumatically, hydraulically, or electrically. With grease systems, shovel type positive displacement reciprocating pumps are often used. (See Section 36.4 on pumping systems.) Because single line systems can range from small in the sense that the distance between the pump and the farthest injector is short (5 ft) to large in the sense that there are long distances between the pump and the farthest injector (500 ft), the size and power ratio of the pump can vary. When the system is turned on, the 2-way valve is positioned to allow grease to flow to the injectors and thus to the bearings. After the injectors have metered the correct amount of grease to the bearing, the system is shut off by a controller turning off the pump. The 2-way valve is then shifted in a manner that bypasses the pump and redirects the grease back to the container, which is normally under atmospheric pressure only. This allows the line to bleed off the grease pressure or vent, thus allowing the injectors to reset and be ready for the next lube event. The vent valve can be operated pneumatically, hydraulically, or electrically depending on the power source that the pump uses. Figures 36.12–36.15 is a schematic of the operations of injectors with top adjustments for use in medium to large systems. Each injector can be manually adjusted to discharge the precise amount of lubricant each bearing needs. A single injector can be mounted to lubricate one bearing, or grouped in a manifold with feedlines supplying lubricant to multiple bearings. In each case, injectors supplied with lubricant under pump, pressure pump lubricant through a single supply line. Two injector types are available: one with a top adjustment and one with a side adjustment. Both types can be used in the same system; their selection is made on the basis of bearing lubricant requirements and the general distances from pump to the last injector.
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Handbook of Lubrication and Tribology
Injector piston
Slide valve Passage Lubricant supply inlet
FIGURE 36.12
SL-1 or SL-11 type designation.
Indicator stem
Discharge chamber
FIGURE 36.13
SL-1 or SL-11 type designation.
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Centralized Lubrication Systems — Theory and Practice
36-15
Injector piston
Slide valve Passage
FIGURE 36.14 SL-1 or SL-11 type designation.
Measuring chamber
Discharge chamber
FIGURE 36.15 SL-1 or SL-11 type designation.
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Handbook of Lubrication and Tribology
Stage 1 — The injector piston is in its normal, or rest position. The discharge chamber is filled with lubricant from the previous cycle. Under the pressure of incoming lubricant, the slide valve is about to open the passage leading to the piston. Stage 2 — When the slide valve uncovers the passage, lubricant is admitted to the top of the piston, forcing the piston down. The piston forces the lubricant from the discharge chamber through the outlet port to the bearing. Stage 3 — As the piston completes its stroke, it pushes the slide valve past the passage, cutting off further admission of lubricant out the passage. Piston and slide valve remain in the position until lubricant pressure in the supply line is vented (relieved) at the pump. Stage 4 — After pressure is relieved, the compressed spring moves the slide valve to the closed position. The piston opens the port from the measuring chamber and permits the lubricant to be transferred from the top of the piston to the discharge chamber. Typical output for top adjustment injectors range from 0.008 in.3 (0.131 cm3 ) to 0.080 in.3 (1.31 cm3 ) for SL-1 type designation. Larger top adjustment injectors range from 0.050 in.3 (0.82 cm3 ) to 0.500 in.3 (8.2 cm3 ) for SL-11 type designation. Figure 36.16 and Figure 36.17 are schematic of injectors with side adjustment used in small applications. Stage 1 — Under pressure from the supply line, incoming lubricant moves the injector piston forward. The piston forces a precharge of lubricant from the discharge chamber through the outlet check valve to the feedline. Stage 2 — When the system is vented (pressure relieved), the piston returns to the rest position, transferring lubricant from the measuring chamber to the discharge chamber. The venting pressure of 200 psi is typical of the side adjustment injectors. Because of this lower venting pressure, less grease supply line distances can be realized. Typical output of side adjustment injectors’ range from 0.001 in.3 (0.015 cm3 ) to 0.008 in.3 (0.131 cm3 ). Typical injector incorporates seals. Only the bushing and plunger use metal to metal fits. The measuring chamber uses viton seals. The viton seals were chosen because their resistance to attachments by solvents or chemical additives in some greases and because of the wide tolerance in which these seals can operate. 36.3.2.1 The Quick Venting Single Line Injectors What do they do? This is a new technology and is the state of the art in metering valves used in automatic lubrication. It is the most significant development in automatic lubrication since the invention of the traditional single line injector introduced in 1938. The quick venting injectors can reset at higher pressures thus allowing a system to use smaller lines than traditional single line injector systems. With the quick venting injectors, the system can handle heavier grease, which is more desirable. The compromise to use a softer grease for most automatic lubrication systems is greatly reduced. Grease with NLGI rating of 2 can be used in much colder temperatures often down to 20◦ or less. Figure 36.18 is a schematic of the basic operation of a quick venting injector and how they work. Stage 1 — The injector is in its normal position. The discharge chamber is filled with lubricant from the previous cycle. Under pressure of incoming lubricant, lubricant is directed to both sides of the measuring piston through the slide valve. The port of the bearing is closed in this position, which prevents the measuring piston from moving. The indicator stem will be at its innermost position, having pulled away from the stop in the adjusting screw. Stage 2 — Pressure has built up and has moved the slide valve in the position shown. This closes the flow to the upper side of the piston (larger diameter) while simultaneously opens the port to allow lubricant to flow out of the injector to the bearing. Pressure from the supply line continues to apply pressure to the low portion of the measuring piston, which cause a pressure difference across the measuring piston, thus allowing it to move upward.
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Centralized Lubrication Systems — Theory and Practice
Outlet check valve
36-17
Measuring chamber Lock nut
Adjusting nut Indicator stem
Lubricant supply inlet
Injector piston
FIGURE 36.16 Side adjustment injector operation. SL-32 or SL-33 designation.
Discharge chamber
FIGURE 36.17 Side adjustment injector operation. SL-32 or SL-33 designation.
Stage 3 — Movement of the measuring piston shown is caused by the pressure on the lower side of the measuring piston dispensing lubricant to the bearing. The indicator stem will move up against the stop in the adjusting screw when the lubricant has been delivered to the bearing. Stage 4 — As the pressure in the supply line is vented down to 1200 psi, the slide valve moves back to the piston shown. This closes the flow of lubricant to the bearing and simultaneously allows lubricant to flow
© 2006 by Taylor & Francis Group, LLC
36-18
Measuring chamber
Piston
Piston
Piston Piston
Springs
Passage 1
Passage 1
Passage 2
Passage 2 Passage 2 Slide valve
Lubricant supply inlet Stage 1
FIGURE 36.18
SL-V injector operation.
© 2006 by Taylor & Francis Group, LLC
Stage 2
Outlet port
Slide valve
Slide valve Stage 3
Lubricant supply inlet Stage 4
Handbook of Lubrication and Tribology
Outlet port
Centralized Lubrication Systems — Theory and Practice
36-19
Flow characteristics of grease in a centralized lubrication system 3000 Window to measure pressure drop Pressure (psig)
2500
2000
1500 Poor flow Quick venting injectors
1000
500 Standard injectors
Good flow 0 0.8
3.0
4.1
8.0
13.0
18.0
Time (min)
FIGURE 36.19
Flow characteristics of grease.
to the upper (larger diameter) side of the piston. The remaining pressure in the line is directed to both sides of the piston and is equalized both on top and bottom of the piston. Because the piston diameter is larger on the top, a net force results in the down direction causing the piston to move accordingly. The injector is recharged by the residual pressure in the supply line to the upper portion of the measuring chamber. The displacement of the fluid on the lower side of the measuring chamber is also allowed by the slide valve to flow to the upper side of the piston, thus completing the recharge of the injector. The resulting effect by absorbing lubricant from the supply line to recharge the injector is to reduce the pressure in the supply line close to zero. The traditional limitation of a single line system is the necessity of the lubricant to vent back down under system pressure in order for the injector to reset. While this is still required for quick venting injectors, the venting now can be done at a much higher pressure. Put in other words, the pressure in the system only has to be vented to 1000 psi as opposed to 600 psi for traditional injectors. Because of this, systems requiring long runs from the pump to the furthest injectors is no longer a limitation. To illustrate this point, consider the flow properties of grease in a centralized lubrication system. Figure 36.19 is a typical chart on how grease will flow when under pressure. The lubrication supply pump will build up pressure, normally above 2500 psi. At this pressure, the injectors will operate and lubricate the bearing. As the system is then vented, the pressure in the system drops as shown in the figure. At the point above 1000 psig, the grease will flow quickly. Below 1000 psig, the amount of grease that flows and vents back is slowed down considerably. Tables 36.1, 36.2, and 36.3 show the venting performance of injectors and how long lengths that can be achieved with single line systems depending on the type of injector. An important fact with all types of injectors is that they can be individually adjusted. The output of each injector can be adjusted by the position of the adjusting cap. The adjusting cap limits the travel of the piston thus limiting the amount of the lubricant to be delivered to the bearing.
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36-20
Handbook of Lubrication and Tribology 1 2
3
FIGURE 36.20
Top view of SL-1 or SL-11 type top adjustment injectors.
Figure 36.20 is a schematic of a top adjustment injector where (1) indicates the nut for limiting the travel of the injector pin (2). This can be screwed in and out thus increasing or decreasing the output. A lock nut (3) is used to secure the setting after the adjustment is made. 36.3.2.2 Design Considerations in a Single Line System As with all systems, knowing the grease flow characteristics at the lowest temperature the systems is to work under is very important. The grease flow characteristics can be obtained by the Lincoln Ventmeter. Knowing the Lincoln Ventmeter viscosity, the three supply line chart tables provided earlier can be used to determine the size of the supply lines. For single line systems, the work has all been done. Because of the use of seals, the ability to pass solid additives in the grease is greatly enhanced. The sizing of the pump that is appropriate for the type of injectors used in the system is important. With long distances, the amount of lubricant needed to be pumped is often much greater that what will be dispensed out of the system. This is because of line expansion as grease is pumped into the system. For example, for a 28-point injector system using a total of 25 ft of 1/2 in. supply hose, the output of the injectors of the SL-1 type designation (if set at maximum output) would be 2.16 in.3 (0.08 in.3 × 27 = 2.16 in.3 ). Using information from a supply line expansion charts for SAE 100R2 hose, the amount of lubricant that would be absorbed by the hose expansion would be approximately 10 in.3 . For injector systems, the pump will need to supply initially twice the amount of lubricant for each injector for the purpose of dispensing and recharge, which would be 4.32 in.3 plus additional 10 in.3 for supply line expansion, totaling 14.32 in.3 . Therefore, a pump would have to be selected that can produce 14.32 in.3 in the time needed before the next lubrication event. When the system is vented, the expansion grease is returned back to the reservoir. It should be noted that for supply line sizes for steel pipe or tubing the expansion would be much less. For a 1/2 in. schedule 40 pipe, the supply line expansion would be approximately 1 in.3 . A pump could be selected that would require 5 or 6 in.3 . Lincoln industrial design guide for single line systems contains all information for any possible system or system combinations for determining line sizing and pumping requirements. Crossporting is another important feature that is available for single line injector types systems. Crossporting connects two or more injectors together allowing for additional outputs. Figure 36.21 is a schematic on how this is achieved. Output of the injectors can be doubled or tripled depending on how much crossporting can be achieved. Crossporting is achieved by connecting the output of injector 1 to injector 2. The output of injector 2 is sent to the bearing. The output to the bearing in this configuration can be up to twice as much as with one injector.
© 2006 by Taylor & Francis Group, LLC
Centralized Lubrication Systems — Theory and Practice
36-21 Injector 1 Injector 2
Model 81646 connector tube
Feedline to bearing
Supply line
FIGURE 36.21 Quick venting injectors.
36.3.2.3 System Monitoring System monitoring can be done visually or through electronic detection of movement of the indicator pin. By visual inspection during the operation, the injector operation can be verified. For critical bearings, electronic detection can be incorporated by installing transducers to sense the movement of the pin. Should the sensor not detect any movement of the indicator pin, an alarm can be given indicting that the bearing is not getting grease. Single line systems can also be used to detect if the bearing is getting grease. This is done by installing a grease flow sensor at the bearing. The grease flow sensor can detect small amounts of flow to the bearing or a lack of flow. If the sensor does not detect flow during a lube event, an alarm signal can be sent to a controller indicating a system problem. 36.3.2.4 Strengths and Weaknesses — Single Line Systems The weaknesses of single line systems are summarized below: 1. For traditional systems, the venting of the injectors limits the distances than can be achieved from pump to the last injector in the system. Pipe and tubing diameter sizes required are larger than those of progressive or two line systems. For quick venting injectors, there is no practical limitation on distances and the piping sizes required. For long runs, quick venting injectors can outperform single line or dual line systems. 2. Monitoring a complete system is more difficult due to the fact that a separate transducer would have to be placed at each individual injector. 3. The elastomer seals used deteriorate faster with contamination. The strengths of a single line system are summarized below: 1. Each injector services one lubrication point, making it easy to trace the metering valve to a single bearing.
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Handbook of Lubrication and Tribology
2. 3. 4. 5.
Each injector output can easily be set individually to meet the lubrication requirements of a bearing. Only one single line is needed making the installation simple and straightforward. The system is easiest to plan and understand. The system is very flexible as lubrication points can be added or removed after the initial system is installed. To add lubrication points, just add injectors, to remove bearing lube points, just remove injectors. 6. The use of elastomer seals allows contaminants to pass through the system. Close metal to metal fits on the measuring chamber are not used. 7. The injectors can be used and recommended for greases that contain solid lubricants such as molybdenum disulfide or graphite. 8. The grease in the systems does not remain under pressure. This is accomplished by venting of the grease. This reduces the changes of the base oil separating from the thickener.
36.3.3 Dual-Line Systems The name implies that two main lubrication lines are used to set up, install, and operate the system. Correctly designed, the Dual-Line system can handle long lines, relatively high pressure, and more than 1000 lubrication points. A high pressure Dual-Line system is capable of lubricating bearings located at long distance from each other. Major components of the system: • • • • •
Metering valves Reversing, 4-way valve Pressure switch or transducer Lubricant pump Controller/Timer
In the Dual-Line systems, a pump supplies the lubricant to the reversing 4-way valve. From the reversing valve, lubricant is supplied alternately into one of the two main lines. Dual-Line systems can be combined with progressive single line measuring valve as well. The system is suitable for either oil or viscous grease lubricants. 36.3.3.1 Basic Operating Principle The DL lubrication system works in two cycles. The central lubrication pump supplies the lubricant under pressure to main line “A” through the reversing 4-way valve. Main line “B” is connected to the reservoir. The metering valves are connected to the main supply lines “A” and “B.” The lubricant is dispensed under pressure from one side of the metering valves to the point of application. As soon as the lubricant is dispensed from the last metering valve, the first half of the cycle is complete (see Figure 36.22). The lubrication pump will continue to operate, pressurizing the line “A” to the preset pressure. As soon as the preset pressure is reached, the reversing 4-way valve will switch the lubricant supply to the main line “B,” connecting main line “A” to the reservoir. Now the pump supplies the lubricant under pressure to the main line “B.” Line “A” is connected to the reservoir and pressure is relieved. The lubricant will be dispensed from the opposite side of the metering valves to the point of application (see Figure 36.23). Second half cycle is complete as soon as the lubricant is dispensed from the last metering valve. The pump will continue to operate until preset pressure has been reached. At this point, a signal from the end-of-line pressure switch or from the micro switch on the reversing 4-way valve will stop the pump, turning the system off. 36.3.3.2 Metering Valves A dual-line metering valve is a positive displacement metering device with an adjustable stroke piston to dispense measured volumes of oil or grease. Figure 36.24 illustrates a schematic of a metering valve. The valve has two output ports. After adjusting the valve to the desired setting, it will dispense an equal volume of lubricant through each of two outlets. If application requires more lubricant, one outlet port
© 2006 by Taylor & Francis Group, LLC
Centralized Lubrication Systems — Theory and Practice
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Relief line Change over valve
M
R
B
P
A
Central lubrication pump
Lubricant metering device Main line
Pressure line To the lubrication points
FIGURE 36.22 Dual-Line lubrication system — first half cycle.
Relief line Change over valve
M Central lubrication pump
R
B
P
A
Lubricant metering device Main line
To the lubrication points
Pressure line
FIGURE 36.23 Dual-Line lubrication system — second half cycle.
Lock nut
Rotary valve Outlet port
Pilot piston
Inlet port Plug
Metering piston
Indicator pin
Seal
Cover cap
Jam nut
O-ring seal Adjusting sleeve
FIGURE 36.24 Metering valve.
can be closed and plugged and the remaining port receives twice the preset volume of lubricant. The valve is designed to withstand a harsh environment in steel, glass, and mining industries. The valve body and internal parts construction are a carbon steel material. The seal material is a nitrile or fluorocarbon compound rubber. The valve consists of the pilot piston to direct the inlet lubricant flow, output metering piston with indicator pin, adjustment sleeve, two jam nuts, and crossporting rotary valve with lock nut. The sleeve has a transparent cover cap. The cover cap on the adjustment sleeve protects the
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seal of the indicator pin from dust and dirt in harsh environments. The movement of the indicator pin is used for visual confirmation. There are two types of valves manufactured: high pressure valve for up to 5000 psi (340 bar) and medium pressure valve for up to 3500 psi (238 bar). Regardless of the pressure rating the valves operation is identical.
36.3.3.2.1 Operation of the Metering Valves Stage 1 — See Figure 36.25. Pressurized lubricant enters the valve through the pilot lines “A,” forcing pilot piston to the left and opening the right pilot connecting port passage. Small amount of displaced lubricant is relieved or vented through pilot line “B,” which is open to the reservoir. Stage 2 — See Figure 36.26. When the pilot piston uncovers the left connecting port passage, lubricant enters the passage, pressurizing the top of the metering piston. Metering piston moves full stroke to the left, dispensing lubricant out through the outlet passage to the lubrication point. This completes the first half of the lubrication cycle. Stage 3 — See Figure 36.27. With piston at the terminal position at the end of the stroke, lubricant pressure will rise to the preset point. A reversing 4-way valve switches lubricant supply from main line “A” to main line “B.” Now the pressurized lubricant enters the valve through pilot line “B,” forcing the pilot piston to the opposite right position, opening left pilot connecting passage. Again, a small amount of lubricant is displaced through line “A” now open to the reservoir. Stage 4 — See Figure 36.28. Lubricant enters left pilot connecting port passage, pressurizing the bottom of the metering piston. The metering piston moves full stroke to the right, dispensing lubricant out through the outlet passage to the lubrication point. This will complete the second half cycle.
Pilot line
Pilot line “A” Pilot piston
“B”
FIGURE 36.25
Stage 1.
Piston line “A” Outlet passage to main line “B”
Pilot piston Connecting passage
A
Metering piston
FIGURE 36.26
Stage 2.
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Centralized Lubrication Systems — Theory and Practice
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Pilot line “A”
Pilot line “B” Pilot piston B
FIGURE 36.27 Stage 3.
Outlet passage to main line “A”
Pilot line “B”
Pilot piston Connecting passage B
Metering piston
FIGURE 36.28 Stage 4.
36.3.3.2.2 Valve Adjustment The valves are available in 2, 4, 6, or 8 outlet configurations. The output of the lubricant is infinitely adjustable from almost zero to the maximum specification volume of the valve. To adjust the output volume remove the sleeve cover, loosen and remove upper jam nut, and turn the second nut in the desired position, limiting the travel of the indicator pin and metering piston. After desired output has been set replace the upper nut to lock the setting and install the sleeve cover back (see Figure 36.29). 36.3.3.3 Reversing 4-Way Valve There are several reversing 4-way valve constructions. The most commonly used are: • Hydraulic • Electric The hydraulic and electric valves are directional flow valves with pressure sensitive mechanisms to alternate the flow of lubricant from one line to the other at predetermined pressure setting. Reversing action is automatic, actuated by pressure from the lubricant supply pump. The line that is not under pressure is vented back to the reservoir. 36.3.3.3.1 Hydraulic Reversing 4-Way Valve The valve consists of the following basic components: • • • •
Flow control piston with indicator pin Pressure sensing mechanism with adjustment Limit switch Pressure gauges (optional)
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Handbook of Lubrication and Tribology Adjusting sleeve
Cover cap
Jam nut
Indicator pin
FIGURE 36.29
O-ring seal
Output volume adjustment.
Adjusting nut Pressure gauge Tenon Cycle switch
Outlet port D Indicator pin Supply line B Flow reversing piston
Inlet port C Meter valve
Supply line A
FIGURE 36.30
Nonloop stage 1
Nonloop stage 1.
Turning the adjusting nut changes the preset of the pressure to reverse the flow of lubricant, from one main line to another. The adjusting nut will increase or decrease the spring force exerted on the pilot piston. The larger the spring force, the more pressure it takes to overcome it and to reverse the lubricant flow. The pressure sensing mechanism is acting as an overcentering device as well. As soon as the pilot piston crosses the middle position, it will complete the stroke using the full spring force. The optional pressure gauges are for visual monitoring of the switch over pressure only. 36.3.3.3.1.1 Operation of the Hydraulic Reversing 4-Way Valve. Nonreturned system. Stage 1 — The reversing piston directs the lubricant flow from the inlet port to supply line “A.” Supply line “B” is connected to the reservoir (see Figure 36.30).
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Centralized Lubrication Systems — Theory and Practice
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Tenon Pilot piston
Pilot piston
Outlet port D Supply line B Supply line A
Inlet port C
Nonloop stage 2
Meter valve
FIGURE 36.31 Nonloop stage 2.
Stage 2 — Rising pressure in supply line “A” forces the pilot piston to the right, overcoming the spring force of the tenon. The pilot piston shifts, pushing the excess lubricant to the reservoir through the outlet port “C” (see Figure 36.31). Stage 3 — Pressure from supply line moves the flow reversing piston to direct the lubricant flow from main line “A” to main line “B.” When the reversing piston completes the full stroke it trips the cycle switch to stop the pump. When the controller/timer starts the next lubrication cycle, main line “B” will be pressurized and line “A” will be relieved (see Figure 36.32). 36.3.3.3.2 Electric Reversing 4-Way Valve The electrically controlled reversing valves use an electric signal from the pressure switch or the pressure transducer to reverse the lubricant flow from one to the other line. The pilot piston of the electric reversing valve is driven by electric motor with a camshaft or electromagnetic solenoid. When preset pressure has been reached, the signal from the pressure switch starts the electric motor or energizes the solenoid of the reversing valve to move the pilot piston and switches the lubricant flow from one main line “A” to main line “B.” The principal of operation is similar to the hydraulic reversing valve.
36.3.4 Types of Two Line Systems There are three basic layouts of the two line systems that can be installed, depending on application: • End of the line system • Dead end system • Loop system Each of the layouts has advantages and disadvantages in cost, installation, and maintenance.
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Handbook of Lubrication and Tribology Pressure gauge
Adjusting nut
Plug R1
R2
Plug
Cycle switch
Indicator pin
Flow reversing piston Stage 3
FIGURE 36.32
Flow reversing piston stage 3.
Pressure switch Pressure gauge Dual line valve
Dual line Valve
Controller Reversing 4-way valve (electric)
Supply line Air inlet Pump Return line
FIGURE 36.33
Typical end-of-line system.
36.3.4.1 End of the Line System This system is preferred whenever the lubrication points are spread over long distances. Use of the electrical reversing 4-way valve is recommended. Figure 36.33 is a schematic of a typical end of the line system. The pump station and electrical reversing valve are positioned in the middle to minimize the pressure drop and to equalize the pressure in both branches of the lubrication line. End-of-line pressure switches or pressure transducers are installed to monitor the line pressure and give the signal to the electric valve to switch the flow at the preset pressure. The controller or timer starts the lubrication cycle and set the lubrication frequency of the system. Upgraded systems are using flow and pressure monitoring accessories to alert the failures.
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Supply line Controller 4-way reversing valve Main line “A”
Air inlet
Main line “B” Dual line valve
Pump Return line Bearing
Bearing
FIGURE 36.34 Typical dead-end (nonreturn) system.
The cost of the system will depend on the selected pump and electrical reversing valve. In addition to the lubrication lines, electrical wires are needed to connect the pressure switches or transducers to the controller or timer, adding to the cost of material and installation. The advantage of this type of the system is that it does not depend on the temperature fluctuation and lubricant variations in viscosity. 36.3.4.2 Dead-End System This system is preferred whenever lubrication points are located in a long line in close proximity to each other or in clusters/zones. The system can use either hydraulic or electrical reversing valve. Figure 36.34 is a schematic of a typical dead-end system. The pump station and the reversing valve are installed at one end of the system. The controller or timer starts the lubrication cycle and set the on and off time of the system. With a hydraulic reversing valve, this is the most economical system and commonly used in the steel and cement industries. The switch over pressure of the hydraulic reversing valve has to be preset considering the pressure fluctuations due to the temperature and lubricant viscosity variations. The disadvantage is that it requires more maintenance and service. Seasonal adjustments of the hydraulic reversing valve switch over pressure are recommended. Adjustments are recommended if the lubricant brand or lubricant formulation changes as well. The system can be upgraded with additional accessories to monitor the lubricant flow and pressure. 36.3.4.3 Loop System This type of the system is preferred with the use of hydraulic reversing valve and whenever better control of the line pressure is necessary. The main lubrication lines are connected to the reversing valve in a complete closed loop. Figure 36.35 is a schematic of a typical loop system. See the description of the valve for line connection in a loop system. The better control of the line is achieved by connecting each line to the opposite ends of the pilot piston and sensing the pressure at the end of the line. The system requires additional line connections increasing the installation cost. The size and the length of the feedlines to the metering valves have to be selected properly to maintain appropriate pressure drop for the valves to operate correctly. Properly designed, the advantage of the loop system is in more stable operation during temperature and lubricant property changes, without using costly monitoring accessories. 36.3.4.4
Design Considerations in a Two Line System
Proper selection of the grease and components for the lubrication system is very important. Dual-Line systems have certain limitations in selecting main supply lines. Total pressure drop between the lubrication point and hydraulic reversing valve should not exceed 2250 psig (153 bar). Pressure available at each
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Bearing Supply line Controller
Bearing 4-way reversing valve
Air inlet
Main line “A”
Main line “B”
Pump Return line Bearing
FIGURE 36.35
Typical loop system.
metering valve should be 500 psig (34 bar) or greater. Pressure drop in the main supply line should not exceed 1500 psig (102 bar) for dead-end (nonreturn) system and 2000 psig (136 bar) for the loop systems. Pressure drop of the branch lines should be less than half of the pressure drop in the main line. Selecting the pump consideration should be given to the pump output. Grease in long lines can compress up to 20% in volume. To better monitor the metering blocks, install the monitoring blocks parallel to the main line. Connect the block inlets and outlets to the same respective lines, so, after completion of one cycle or half-cycle the indicator pins will be in the same position. 36.3.4.5 Advantages and Disadvantages of the Dual-Line Systems The advantage of the Dual-Line system is summarized below: 1. 2. 3. 4. 5.
Long distances can be achieved Additional lubricant points can be added without changing the main lubrication lines Long history in industry Can be used in a system with a large number of lubrication points Uses metal to metal fits in the metering valves. This can be an advantage because there are no elastomer seals to wear
The disadvantages of the Dual-Line system is summarized below: 1. Requires two lubrication lines and double the amount of fittings and mounting hardware. 2. Old technology. 3. Most two line valves use metal to metal fits in the metering valve. Use of lubricants with solid additives may be prohibited. With metal to metal fits, the system is more susceptible to failure when contamination is present.
36.4 Pumping of Grease and Viscous Materials In automatic grease lubrication systems, one common component is the pump. The pump must be able to remain primed (avoid cavitation), deliver the correct amount of volume under back pressure. Because greases can have high apparent viscosity, back pressure is created to overcome the yield pressure of the grease and the friction loss that is produced when grease is flowing in the supply line. Typically, for NLGI
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number 2 grease, with a Lincoln Ventmeter viscosity of 400 psi, the apparent viscosity can be in the range of 70,000 to 100,000 cP or even higher. The pump must have the capability to remain primed. That is to say, the amount of grease entering the pump must be sufficient to charge the pumping chamber. Almost all grease pumps operate with positive displacement reciprocating action. If the grease is too stiff or has too high of an apparent viscosity, grease may not flow into the pump. When this happens, the pump cavitates. Cavitation should be avoided as it can cause premature pump failure. Moreover, a pump that is cavitating cannot pump grease into the system, thus the bearing will not be getting lubrication. For pumps mounted in refinery drums such as a 55-gal (400 lb) or 120-lb drums, the pump must be able to pick up the grease under operating temperatures. A positive displacement double acting shovel action pump is most used for pumping viscous material. The pumps are described as double acting because they output grease when pump is in both the up- and downstroke. There are two pumping chambers, one for the upstroke and one for the downstroke. Positive displacement pumps create suction in their action. The pump operates in a piston/cylinder arrangement. The piston displacement creates a vacuum, this vacuum is used to create a pressure differential causing grease to flow. Because there is only 14.7 psi vacuum pressure possible, this may not be enough to produce flow from the grease reservoir to the pumping chamber. To overcome this situation, a mechanical shovel is used to mechanically push the grease into the pumping chamber. See Figure 36.36 and Figure 36.37 for a description on how they work.
Grease out Second pump chamber 1/2 volume of 1st pump chamber Outlet check
Lower inlet check
Mechanical shovel
FIGURE 36.36 Shovel pump in downstroke.
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Grease out
Lower inlet check
Outlet check
First pump chamber twice volume of chamber 2
Mechanical shovel
Grease in (upstroke only)
FIGURE 36.37
Shovel pump in upstroke.
The operation of a double acting shovel pumps produces the same output when the piston is in the up- or downstroke. After the pump is inserted in a grease reservoir, the pump is first primed by turning on the pump and removing any air out of the pump. When the pump is turned on, the pump in the up stroke uses the mechanical shovel to force the grease into the pumping chamber. All volume of grease entering the pump occurs on the upstroke only. The pump does not accept grease on the downstroke. There are two pumping chambers with double acting shovel pumps. The grease that is entering the pump tube first enters the first pumping chamber. The inlet check opens during the upstroke cycle. Simultaneously, the second pump chamber volume is compressed forcing lubricant out of the pump. During the downstroke, the outlet check closes and the pump piston fills the second pumping chamber while dispensing lubricant out of the pump. Because the displaced volume of the first pump chamber is twice that of the second, the grease fills the second pump chamber as it dispenses. It is important that a pump produces the same amount of pressure and flow on the downstroke as the upstroke. If the pump ratio is, say 50 : 1, the pump should be able to generate the same pressure on either the up- or downstroke at that ratio.
36.4.1 Sealing Methods for Pumping Viscous Materials The weakest point of any reciprocating pump is the gland area. This is the seal area at the top of the pump. There are two major factors affecting the gland of the reciprocation pump. The first factor is
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Plunger
Scrapper ring
Seal Labyrinth bushing
FIGURE 36.38 High pressure pump seal configuration.
the environment of the exposed pump piston. The second is the fluctuation of internal pressure produced when the pump changes from an upstroke to a downstroke and vice versa. The exposed portion of the pump plunger can be contaminated and gradually destroys the gland seal. The internal pressure fluctuation, developed by pump during up- and downstrokes, deteriorates the gland seal contributing to possible premature failures. To improve a pump’s ability to maintain a good seal, there are several designs available. Some pumps use a staked seal design. The features of this seal arrangement allows for retightening when grease begins to leak out the top of the pump. This retightening reforms the seal and is often effective in stopping the leak. This is only good to a point and retightening cannot be performed infinitely. The most effective design is the special gland protection package incorporated into the pump. The gland protection consists of a spring type scraper ring and a labyrinth bushing. The scraper sharp edge is in constant circumference contact with the plunger and cleans/scrapes the plunger surface from any particles that can damage the gland seal. The labyrinth bushing is protecting the gland seal from the pressure fluctuations due to the labyrinth path of the material before it reaches the gland seal (see Figure 36.38). This new technology is successfully proven and tested in many toughest environments. Techniques are available to keep the pump primed when the grease being used is very stiff or has a high apparent viscosity. Often, the ambient temperature drops resulting in a drastic increase in the grease stiffness. The solution is to change to a lighter grade grease or to improve the priming capability of the pump. Those techniques are described below. 36.4.1.1 Create Positive Head Pressure by Using a Follower Plate A follower plate can be used to create an additional pressure head, which will prevent a void from forming around the pump inlet. If grease is too stiff, and the pump draws in grease, a void could be created. This void will cause cavitation. The principal behind this technique is using differential pressure produced by the pumps ability to produce a vacuum. When the pump displacement causes a vacuum, some grease will flow into the pump chamber. Simultaneously, this will create a pressure differential across the follower plate. The pressure differential across the follower plate may be small, but the net force produced becomes
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Handbook of Lubrication and Tribology Shovel pump
Drum cover 55-gal refinery drum
Follower plate Provides downward force to aid in pump priming
FIGURE 36.39
Greases
Grease drum follower plate.
large, creating positive head that will prevent voids or pockets. The simple relationship can be illustrated as follows. A typical follower plate for a 55-gal refinery drum may be 24 in. in diameter. This results in an area of 452 in.2 . If the pump can produce just a small vacuum of 2 psi, the net force acting on the follower plate would be F = Pressure × Area or 2 psi × 452 = 904 lbs. This net downward force will cause any void in the grease to collapse. Figure 36.39 is an example of a pumping arrangement using a follower plate. 36.4.1.2 Create Positive Pressure by Using a Pressurized Reservoir The grease reservoir can be pressurized within the structural limits of the reservoir. A contained grease reservoir can be pressurized to, say, 10 psig. This 10 psig plus ambient pressure can be added to force the grease into the pumping chamber. Good positive displacement pumps should be able to produce 12 psi of vacuum pressure. This would result in a differential pressure between the grease and the pumping chamber of 12 psi vacuum plus 10 psig to develop 22 psig of differential pressure. 36.4.1.3 Use Pressure Primer to Force the Grease into the Pumping Chamber For more severe conditions, a pressure primer can be used. A pressure primer uses mechanical actuators pressurized by air to apply a downward force onto the follower plate. This action forces the grease into the pumping chamber. Pressure primers are often used on applications where the grease is stiff due to the requirement of the application combined with cold temperature environments. Figure 36.40 is an illustration of a pump set up in a pressure primer system. The illustration shows a pressure primer with the actuating cylinder in the retracted position. When the actuation cylinder is extended, a 55-gal refinery drum containing the viscous material is placed under the follower plate. The air control valve is switched to the retraction position thereby producing a continuous downward force on the follower plate; thus providing positive head pressure to keep the pump primed. When the 55-gal refinery drum is empty, the actuating cylinder is extended pushing the follower plate out of the drum. The barrel supports keep the drum in place while the follower plate is being removed. Grease pumps are usually specified by the amount of pressure that can be developed. For pneumatically and hydraulically operated pumps, the pump ratio is the amount of grease pressure the pump can develop over the amount of air/hydraulic pressure supplied to the motor. For example, a 50 : 1 pressure ratio means
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Centralized Lubrication Systems — Theory and Practice
Shovel pump
36-35
Actuating cylinder
55-gal barrel support
Follower plate
FIGURE 36.40 High pressure primer configuration. Pump chart 75:1 Ratio 100 psi
100 psi
90 80
6000
70 5000
70 psi
70 psi
4000
60 50 40
3000
40 psi
40 psi 2000
20
1000 0 0.0
30
0.1
0.2
0.3
0.4 Output (gpm)
Mean effective pressure Air consumption 0.5 0.6 0.7
Air consumption (scfm)
Mean effective pressure (psi)
7000
10 0 0.8
FIGURE 36.41 Mean effective pressure vs. output chart.
that if 80 psi of air pressure is supplied to the motor, the pump will be able to produce 4000 psi. The pump would stall out at this pressure. When specifying pumps, manufactures produce flow and output curves based on the pressure needed to be developed. Figure 36.41 is an example of typical pump curves for positive displacement pneumatically driven pumps. For electric driven grease pumps, the ratio is not as straightforward. For DC and AC operated grease pumps, the manufacturer will specify the maximum pressure the pump will achieve, depending on the motor horsepower and gear transmission design of the pump. Because grease pumps can produce high pressure, safety must be a concern when designing a system. The designer of the system must select the hardware and piping to handle the amount of pressure the system will develop. Relief valves must be installed in the system at the correct locations to vent out any excessive pressure before damage or harm could occur.
© 2006 by Taylor & Francis Group, LLC
37 Used Oil Recycling and Environmental Considerations 37.1 37.2 37.3 37.4
Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Terminology . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Quantifying the Resource . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Common Contaminants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
37-1 37-2 37-3 37-4
Contaminants of Special Interest
37.5 Typical Uses . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 37.6 Technologies. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Dennis W. Brinkman Indiana Wesleyan University
Barbara J. Parry Newalta Corporation
37-5 37-6
Re-Refining Process
37.7 Environmental Regulations . . . . . . . . . . . . . . . . . . . . . . . . . . . 37.8 Pollution Prevention/Lifecycle Assessment . . . . . . . . . . 37.9 Conclusion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
37-10 37-11 37-11 37-13
37.1 Introduction Lubricants are unique among all petroleum products in that often they are not consumed during use. Solid lubricants, such as greases are not considered here as they are seldom in a form to be easily recovered at the end of their useful life. Most fluid lubricants, such as crankcase oil and transmission fluid are collected and replaced as part of normal equipment service. These “used oils” are now available to become either a useful resource if handled properly or a potentially hazardous waste if ignored or discarded. Used oil reclamation is not a new idea. As early as World War I, there was recognition that used oils represented a resource that could be accumulated and utilized instead of thrown away. Starting with World War II, there was a concerted effort to recycle used oils, such that by 1960 there were 150 separate businesses throughout the United States producing several hundred million gallons of crankcase oils and other petroleum products each year. Similar infrastructure was in place throughout the world [1]. “Today, there are around 400 known re-refining facilities [worldwide] with an overall capacity of 1800 kt/yr” [2]. Tremendous changes in technology and environmental concerns have significantly impacted the used oil recycling industry in the past 50 years. Lubricants have become much more complex in composition and function. This makes the challenge of processing these fluids after use increasingly difficult. Additionally, 37-1
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the increased use of bio-based oils and fuels is a challenge to the re-refining industry, since these are incompatible with most current technologies. Used oils have been shown to have the potential for serious environmental damage as well as health and safety risks if not handled properly. This has led to a significantly higher intensity of government regulations, which is usually more than a small business can handle. Thus, the industry has gone through several decades of consolidation and enhancements that have resulted in collection of used oil in larger volumes to the betterment of the environment. Many studies have been done on used oil characteristics, recycling technologies, and related topics [3,4]. The most recent wave of research into used oil recycling technology peaked in the 1970s and 1980s, during the period of oil embargoes. Once the OPEC crude oil production rates had stabilized at a rate acceptable to the developed countries, government funding for this research was eliminated and the economic driving force for private investment faded. More recently, used oil studies have focused on environmental stewardship and life cycle analysis. Interest in finding ways to increase collected used oil volumes and improve the quality of recycled products is again on the rise. This renewed interest is primarily due to pollution control and prevention advocacy initiatives such as the Kyoto Protocol, risk averse management practices, and recognition that used oil is, arguably, the largest single source of hazardous recyclable material. When one does an internet search using the term “used oil recycling,” thousands of entries are found. With a few exceptions, each entry represents a government agency or program that is intended to inform its constituents about the potential harm caused by uncontrolled dumping of used oil and filters, new and proposed legislation regarding recycling and stewardship programs and suggestions for recycling. The websites often contain links to other information sources as well as listings of locations that accept used oil. The number of sites underscores the importance that environmental agencies have placed on maximizing the proportion of used oil generated that is collected for recycling.
37.2 Terminology After collection, used oil recycling consists of a wide variety of activities yielding a range of products. Thus, it became necessary to appropriately define terms used to describe various recycling approaches. While regulatory agencies have often crafted independent definitions, consensus terminology was created and published by American Society for Testing Materials (ASTM) International in the 1980s. These are found in ASTM Standard D4175 and are reproduced in Table 37.1 [5]. From a practical standpoint, the following are typical descriptive terms in common use: 1. Used oil — oil that has, through use or contamination, become unsuitable for continued use in its current application, but which is likely to have some use in another application or as a feedstock for a process that generates a useful product.
TABLE 37.1
Consensus Terminology
Reclaiming
The use of cleaning methods during recycling primarily to remove insoluble contaminants, thus making the oil suitable for further use. The methods may include settling, heating, dehydration, filtration, and centrifuging. The processing of oil that has become unsuitable for its intended use, in order to regain useful material. The use of refining processes during recycling to produce base stock for lubricants or other petroleum products from used oil. In petroleum product recycling, oil whose characteristics have changed since being originally manufactured, and which is suitable for recycling (see also waste oil). In petroleum technology, oil having characteristics making it unsuitable either for further use or for economic recycling.
Recycling Re-refining Used oil Waste oil
Source: ASTM Annual Book of Standards, Vol. 5.02, D4175, Standard Terminology Relating to Petroleum, Petroleum Products, and Lubricants.
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2. Waste oil — oil that has become so degraded or contaminated that it is impractical to recover anything useful from it other than its heat content (if that can be done in an environmentally sound manner). With the increased understanding of lubricant chemistry, a focus on pollution prevention, and the enhanced sophistication of recycling facilities, very little lubricant needs to end up here. 3. Recycling — This is usually the umbrella term that covers all aspects of used oil collection, processing, and reuse. Thus, reprocessing, reclaiming, and re-refining are all subsets of recycling. 4. Reprocessing — used oil recycling where the primary objective is producing fuels, whether for burning in small space heaters or large industrial boilers. This may include simple settling and filtration techniques to remove bulk water and solids. 5. Reclaiming — used oil recycling where the primary objective is rejuvenating a lubricant so that it can be reintroduced into the original application. This is especially useful for hydraulic fluids and other industrial lubricants that have relatively simple compositions and less demanding applications. This involves simple dehydration, settling, and filtration techniques followed by additive replenishment. 6. Re-refining — used oil recycling when the primary objective is a clean base oil equivalent to virgin base oil from which any and all petroleum-based lubricants can be blended. This involves sophisticated processing and testing.
37.3 Quantifying the Resource Worldwide production of lubricants in 2002 was estimated to be 37 million tonnes (roughly 11.2 billion gal). Table 37.2 shows the approximate breakdown by region [6]. Regional volumes have changed over time but the total volume has not changed significantly in the last 10 years. While exact numbers are difficult to break out, one can assume these volumes are split roughly equally between automotive and industrial lubricants. In round numbers, the production of petroleum-based lubricants in the United States is about 2.5 billion gal/yr [7]. However, not all used oil is recoverable. Many oils are either consumed during processing (e.g., quench oils) or are lost during use (e.g., chain saw and 2-cycle oils) but of the estimated 65 to 70% that is recoverable [8], most information agencies agree that approximately 40% of the original product is collected; roughly equivalent to 15 million tonnes (4.5 billion gal). The balance disappears into the environment potentially creating nonpoint source pollution. TABLE 37.2 Production
Estimated Regional Lubricant
Region North America Central and South America Western Europe Central/Eastern Europe Near/Middle East Africa Asia Pacific
Estimated lubricant production (2003) billion gal/million tonnes 2.7/8.9 1.0/3.2 1.5/5.1 1.4/4.9 0.6/2.0 0.6/1.8 3.4/11.2
Source: Adapted from Europalub publication extracts “Consumption by Countries 1996–1999” July, 2004, Singh, H., “Lubricant Technology Today” Science in Africa on line Science Magazine Nov, 2002 and Cheuveux Germany Annual Report for FuchsPetrolube Q2,2004.
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In addition to used lubricants themselves, used oil filters and containers usually hold residual, recoverable hydrocarbon fluids and can also be recycled. Most jurisdictions that have legislation pertaining to used oil also have provisions for recycling used oil filters provided they have been thoroughly drained. Recycling plastic bottles with lube oil residues is more problematic. Capacity to recycle plastic containers into such items as new plastic containers, flower pots, pipe, fencing, and patio furniture is slowly growing in most jurisdictions. Draining the bottles prior to disposal may be the optimum scenario for the present.
37.4 Common Contaminants Used oils most often come to the end of their usefulness not because the hydrocarbons in the base oil have broken down, but because of a combination of additive depletion and contaminant accumulation. The additives (described in detail in Chapter 20) provide many (if not most) of the desired lubrication properties. The combination of additive depletion and introduction of contaminants creates the potential for damaging the hardware being lubricated. Contaminants include solids (sludge, varnish, rust, and wear debris), additive degradation by-products (oxidized additive molecules, and sheared viscosity improver polymers), water (free, emulsified, and dissolved), fuels, and process chemicals. Once the used lubricant is isolated from a crankcase, hydraulic cylinder, or other application — and is awaiting collection and recycling — further contamination often occurs as other wastes are added to the container. This secondary adulteration is often more problematic than the original, since the sources are less predictable and often are unrelated to the petroleum-based lubricant. While no two truckloads of used oil are exactly alike, typical parameters for used oil arriving at a collection tank farm and processing plant might include: • • • • • •
Water content Flash point Acid number Chlorine Sulfur Nondistillables
10 to 30% >100◦ C 2 to 3 500 to 3000 ppm 1000 to 4000 ppm 5 to 15%
A major breakthrough for used oil recycling came during the 1970s and 1980s when researchers at the U.S. Department of Energy at its Bartlesville (OK) Energy Technology Center demonstrated that hydrocarbons in the base oils of lubricants were not being broken down or oxidized during use. Evidence was clear that the lubricants were just becoming dirty and, therefore, could be cleaned up to their original condition [9]. This was significant because it implied that high-quality base oil could be recovered with standard refinery techniques. This research was foundational to the development of the modern rerefineries and acceptance of re-refined base oils.
37.4.1 Contaminants of Special Interest Water and sludge are contaminants that affect product quality, but do not contribute to any health concerns. There are, however, other contaminants present at much lower levels, which provide significant motivation for the controlled collection and disposition of used oils [10,11]. Polychlorinated biphenyl compounds (PCBs) were used extensively as the fill fluids in large electrical transformers, as heat transfer fluids, and as hydraulic fluids. While PCBs are attractive due to their resistance to oxidation, the entire family of compounds (all 209 congeners) are listed as probable human carcinogens [12]. Because PCBs tend to look and behave like oils and are used in the same locations as oils, past practice has been to combine waste PCBs with used oils. This custom was halted in the 1970s as the health concerns of PCBs became more widely publicized and regulations were put in place that led to the banning of their production in 1977. Because PCBs are specifically regulated in the United States under the Toxic Substances Control Act (TSCA), the used oil recycling industry has been forced to incorporate specific testing and
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control measures to make sure the used oil collected and processed contains no more than 50 ppm of total PCBs. Similar legislation exists in other regions of the world. Other halogenated compounds are also of interest. The most prominent are halogenated solvents that are often used as cleaning agents, such as perchloroethylene (tetrachloroethane), methylene chloride, trichloroethylene, and 1,1,1-trichloroethane. Since these compounds are never components of lubricants, they are indicators of adulteration of the used oil. Further, they have toxicity issues of their own, although concentrations found are almost always below 100 ppm, and often below detection limits for modern analytical instruments [6]. Ethylene glycol is widely used as antifreeze in automotive cooling systems. As such, it is another waste stream generated in the automotive servicing industry and is often found blended into used oils for disposal. Some modern used oil rerefineries isolate and market the ethylene glycol contained in their feed stream. While antifreeze is typically not a regulated waste, ethylene glycol is toxic and contributes to the overall toxicity of used oil. Lead and benzene contamination in used oil originally came from gasoline that found its way into the oil as blowby in engines (mist pushed past the piston rings). In the mid-1980s, environmental legislation in North America forced the elimination of lead in gasoline. A direct result of this decision was the decreased lead content of used oils. The primary source now is aftermarket additives and wear metal. Benzene, a known carcinogen, is present at levels of 0.5 to 3% in gasoline. Thus, any gasoline contamination incorporated into used oils contributes a volatile carcinogen, albeit one that is handled safely everyday by millions of people. Benzene is the simplest example of a family of hydrocarbons known as aromatics. These compounds of hydrogen and carbon exist as rings with alternating single and double bonds connecting the carbons. If more than one of these rings are connected together, polycyclic aromatic hydrocarbons (PAHs) are formed. These are found in high-boiling fractions of crude oil (such as asphalt) and are often mutagenic and carcinogenic. These are closely related to polynuclear aromatic hydrocarbons (PNAs), which tend to be mutagenic and carcinogenic [13] and also contain oxygen, sulfur, nitrogen, and other elements. Toxic metals can be introduced during use or from contamination after use. The four of primary interest in used oil are arsenic, cadmium, chromium, and lead. Of these, only chromium and lead are found at above 1 ppm with any frequency, and both tend to be below 50 ppm. While specific contaminants found in used oils are of some interest, the primary concern is the overall health effect on those exposed to used oils. Further, the primary pathway for transmission into humans can be assumed to be through dermal exposure. Relatively simple tests are now available for testing used oils to determine their relative carcinogenicity [14]. Testing has clearly shown that more advanced used oil processing techniques can remove the carcinogenic properties of used oils as well as the heavy metals.
37.5 Typical Uses After collection, used lubricating oil is used in or sold as feed for a variety of uses. Prior to 1988, the most prevalent use for used oil was fuel for energy recovery and dust control, such as road oiling. Since that time, environmental legislation has ensured that road oiling with used oils has almost completely disappeared in developed countries. Today, the primary use worldwide is as a fuel. It is estimated that at least two-thirds of all collected used oil in North America is eventually burned for energy recovery either directly in space heaters and industrial boilers, as blended fuels used in commercial applications or from reprocessing facilities, which produce fuels for various applications. Worldwide this number may be closer to 90%. Space heaters are small furnaces that provide heat in automotive garages and manufacturing locations. Often the user also provides the fuel in the form of used oil generated from on-site operations. There are two basic types: aspirating (fuel aspirated with air into a flame) and vaporizing (fuel converted into a vapor, which is then burned). The former has caused some concerns due to air emissions of metals contained in the used oil.
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Industrial burners and commercial boilers operate in the same manner but on a much larger scale. In North America, these units are usually required to comply with local air emission standards and in some regions must register the fuels they are burning with regulatory agencies. In order to encourage a positive experience with the use of fuels blended with used oils in industrial and commercial boilers, ASTM International issued two separate fuel specifications [15,16]. These specifications provide guidance on what properties should be monitored when purchasing fuels made with used oils to ensure good performance. Road oiling with used oils has almost completely disappeared in developed countries. The most famous event involving used oil applied to roads occurred in 1982 in Times Beach, MO, when the Environmental Protection Agency (EPA) purchased all properties, closed the town of 2000 people, and demolished all the homes due to extensive road oiling with used oil contaminated with dioxin and PCBs [17]. 265,000 t of soils were incinerated as part of the cleanup effort. In the years that followed, road oiling with used oil has been banned in almost every state in the United States and most developed countries throughout the world. Another widespread use for used oil is in hot-mix asphalt. While incorporating used oil into asphalt might seem related to road oiling, it is not. Some asphalt plants buy used oil for use as a substitute fuel with the residues incorporated into the asphalt. Asphalt manufacturers also purchase the vacuum distillation bottoms from rerefiners for direct blending into their product as an “asphalt extender” or “asphalt flux.” It has been found that these bottoms contribute positively to the properties of the resulting asphalt. Further, the asphalt binds trace metal contaminants to prevent leaching into the environment as they would with road oiling and volatile organics are consumed in the manufacturing process. Used oil is mixed with crude oil as refinery catalytic cracking feedstock. It also provides feedstock for reprocessing facilities, in Europe particularly, that produce distillate fuels for such applications as industrial fuel, home heating oil, and diesel fuel additives. Of all the potential uses, re-refining (recycling used oil to produce lubricant base oil) is often viewed as the optimum pathway for used oil. By making a product that can be used over and over, re-refining saves a valuable natural resource while diverting a potentially hazardous waste from loss into the environment. The modern rerefinery now uses vacuum distillation with a finishing stage, such as catalytic hydrotreating, just like a modern crude oil refinery. It is therefore not surprising that the quality of the products made from re-refined oils is equivalent to that made from virgin oils [18]. This has been demonstrated multiple times with the passage of engine sequence tests and the certification of specific formulations by such organizations as the American Petroleum Institute (API) and the European Automobile Manufacturers Association (Association des Constructeurs Européens d’ Automobiles; ACEA).
37.6 Technologies In-plant recycling of used oil is a specific form of reclaiming in which the life of the oil is extended, often through the simple steps of filtration to remove solids, heating to remove water, and refortification of the depleted additives. If the performance requirements for the original application are too stringent for efficient reclaiming (too much processing and testing required), alternative uses can often be found for the oil, for example, as a metalworking fluid. The fact that no outside party is involved and no transportation of the used oil is required can often mean elimination of any regulatory paperwork (such as manifests) and reduction in process costs. Reclaiming using outside resources can still result in the lubricant returning to the original producer. More likely, the used oil generator has spent material removed and recycled product delivered continuously such that there is no identification of the source of the material in use at the time. The processes used by the commercial reclaimer are similar to those that would be used in an in-house operation, but the cost benefits of larger-scale operations can make a third-party, central operation attractive. Reprocessing normally involves chemical or physical treatment of the used oil to produce a fuel oil that meets customer requirements. While this also means removal of solids and water as in reclaiming,
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the standards may well be less stringent. The large volumes of fuel oil purchased by industrial users also lends itself to the blending of used oils into the virgin fuel oil stream at low proportions to minimize any handling difficulties or change in emissions characteristics. Alternatively, many fuel producers are using variations of re-refining techniques without the final finishing step used in base oil production. These methods include catalytic cracking, distillation, thin-film evaporation, thermal de-asphalting, and propane or direct hydrogen extraction. These processes are more costly to build and operate, but the fuels produced are of consistently higher quality and generally provide a higher rate of return. An emerging trend in fuel production is processing used bio-oils (lubricant fluids made from renewable plant materials) that have been segregated from used mineral oils into bio-fuels for transportation and energy recovery applications. These fuels are generally blended into virgin diesel at 5 to 20%. In the used oil recycling hierarchy, re-refining to make high quality base oil is the highest tier and requires the use of more sophisticated technology. When looking at the distillation characteristics of used oil, as shown in Figure 37.1, it becomes obvious that many of the contaminants can be removed by distillation. The base oil is the fraction boiling between roughly 315◦ C (600◦ F) and 540◦ C (1000◦ F). Many approaches to isolating commercially attractive lubricant products have been attempted over the years. These include [19]: • • • • •
Blending used oil into crude oil refinery streams Sulfuric acid/clay treatment Vacuum distillation/solvent treatment/extraction Vacuum distillation/clay treatment Vacuum distillation/hydrotreatment
While there are a number of specific proprietary processes currently in place around the world, they all tend to have a few similar basic steps [18–20]. A sequential distillation process, as shown in Figure 37.2, allows the rerefiner to first remove the water and solvents, then to remove the distillate fuels, and finally to volatilize the base oil away from the distillation bottoms, a very high viscosity material will be blended into asphalt. The base oil is then passed over a precious metal catalyst along with high-pressure hydrogen to produce a clean, stable hydrocarbon base oil. Additional fractionation after hydrotreating allows the rerefinery more control over the viscosity and volatility of the final products. A more detailed description is offered in the sidebar in Section 37.9.
Dehydration
0
200
Fuel stripping
400
Lube oil
600
800
Bottoms
1000
Boiling point °F
FIGURE 37.1 Simulated distillation curve showing typical cut points used by rerefiners.
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1200
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Handbook of Lubrication and Tribology Dehydration fuel Pretreatment drum
Dehydration drum
VFS fuel Vacuum fuel stripper
Used oil
Light lube tower
Thin film evaporator #1
Thin film evaporator #2
Light vacuum oil to HT
Medium vacuum oil to HT Asphalt Light vacuum oil to HT
Recycle hydrogen
Make-up hydrogen feed
Lt., Med., or heavy Vac. oil feed
HT FUEL Gas/liquid separation Hydrofinishing reactors
HT spindle oil — 60 SUS Stripping tower HT lube oil — 100, 240 SUS
FIGURE 37.2
Used oil rerefinery schematic showing stepwise distillation followed by catalytic hydrofinishing.
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A benefit of this modern approach is that re-refining no longer produces hazardous by-products, such as the sludges resulting from treatment with sulfuric acid and clay that was so prevalent in the middle of the last century and still is used in less developed countries where lubricant volumes or collection networks are too small to support a high technology plant. Light hydrocarbon by-products are used as fuels, water is treated and released, and even the catalyst can be recycled for reuse in the process.
37.6.1 Re-Refining Process The re-refining process illustrated in Figure 37.2 may not reflect any particular facility, but it does represent the common steps found at any modern rerefinery. In sequence, these are: 1. Pretreatment — Incoming used oils often require some chemical treatment to protect the plant against corrosive attack and the hydrocarbons in the oil against chemical degradation during the process. Most often, this pretreatment involves a mild caustic wash, but it could include demulsifier, antioxidant, or other chemical agents. (Note: one advantage the much larger rerefineries of today have over their predecessors of 50 years ago is sheer size. The large volume throughput requires massive storage tanks, which are continuously being filled and drained. This provides a natural averaging of the composition of the feed, such that the process operates in relative equilibrium.) 2. Dehydration — Water is typically a major constituent of the used oil stream. It comes from metalworking fluids, antifreeze, and contamination of the used oil as it resides in drums and tanks waiting for collection. An atmospheric distillation is usually sufficient to remove most water and light hydrocarbons (e.g., gasoline, solvents, and glycols). The water is treated for plant use or discharge while the light hydrocarbons are often used as fuel within the plant due to their high odor level. The by-product produced may represent 10 to 30% of the incoming volume depending on the level of contamination. 3. Stripper Tower — The next stage involves distillation at a moderate vacuum to remove distillate fuels (e.g., fuel oils, jet fuel, and diesel fuel). These fuels can be sold on the industrial market to blend into boiler fuel, but again are often used within the re-refining process. This stream is usually less than 5% of the original input. 4. Vacuum Distillation — The distinctive of most rerefineries is the manner in which they perform the distillation step(s) that volatilize the base oil present in the used oil. A maximum vacuum (typically 10 to 50 mm Hg) is desired to keep the oil temperatures below that resulting in thermal cracking (225◦ C). Many installations use some variation of a thin-film evaporator that combines short residence time with efficient heat transfer. The by-product of this stage is the nondistillable bottoms, which is sold to the asphalt industry as an asphalt extender or flux. The bottoms can be as much as 10% of the input to the plant. 5. Finishing Step — The vacuum distillate contains too many unsaturated and polar molecules to be stable in typical lubricant applications. Thus, some type of final treatment is required to convert or remove these reactive compounds. In the past, this has been accomplished with activated clay (attapulgus clay with bonding sites converted to the acid form). However, this generates a large quantity of solid waste and has some limits on the quality of the final product. Thus, most large rerefineries now utilize catalytic hydrotreating (passing the oil distillates over a fixed bed of precious metal catalyst at temperatures of 350◦ C or more and pressures of 600–1000 psi). The final base stocks usually add up to be 60–70% of the volume of the original used oil feed. Fractional distillation or multiple fractionation towers are used to generate the series of distillate cuts required to produce a variety of base oil viscosities. Some oil components that are required for final lubricant blends, such as the nondistillable bright stock, are purchased from crude oil refineries and the additive packages are purchased from special chemical suppliers. The base oils from these modern rerefineries are sufficiently rejuvenated and stabilized so that they respond in an identical manner to the additives as with virgin base oils.
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As with any manufacturing process, the economics of this process depends on efficient operation and minimal downtime. However, somewhat unique to used oil re-refining, the operating component that often causes financial troubles is the feedstock. Reliably obtaining enough good quality used oil to keep the feed tanks reasonably full (to benefit from blend averaging) and the plant operating at full capacity can be more difficult than new entrepreneurs might suspect. If the scale of the plant and the volume of available used oil are well-matched, the economics of re-refining can be attractive [28].
37.7 Environmental Regulations With the advent of enhanced environmental consciousness in the 1960s and 1970s, disposal and recycling practices for waste streams such as used oil came under a new level of scrutiny. Given the volumes involved and levels of contamination described above, it is not surprising that used oil was specifically targeted for pollution control. As already mentioned, an excellent summary of the toxicology of used oils was released in 1997 by the U.S. Department of Health and Human Resources [11]. A broader discussion of the backdrop to environmental regulations in Europe is provided in a report from CONCAWE in 1996 [4]. Most developed countries have enacted federal, regional, and local environmental protection legislation. This legislation varies considerably in the treatment of used oil from generic pollution prevention and spill response strategies to detailed collection and treatment requirements to mandated use of recycled products. In the United States, the direct regulation of the disposition of used oils is usually considered to have begun with language in the Energy Policy and Conservation Act (EPCA) of 1975, followed immediately by the Resource Conservation and Recovery Act (RCRA) in 1976. These established guidelines determine which wastes would be considered hazardous and require special handling. Used oil was then addressed specifically in the Used Oil Recycling Act of 1980. All of these tried to make sure that used oil was considered a resource for further use and not disposed of in landfills. These three legislative actions resulted in a sizeable body of regulation issued by the U.S. EPA and contained within Part 279 of Volume 40 of the U.S. Code of Federal Regulations (40CFR279). The fact that an entire section of regulations was devoted exclusively to used lubricants illustrates the significance given to this commodity. A primary focus of the regulations is used oil burned for energy recovery. The used oil specifications currently in place for fuels are provided in Table 37.3. If there are more than 1000 ppm total halogens, but less than 4000 ppm, an analysis showing that the contaminants are not halogenated solvents listed by EPA as hazardous will allow the material to be used as a fuel. TABLE 37.3 Used Oil Fuel Regulatory Limits — U.S. EPA Constituent
Allowable limit
Flash point Lead Chromium Arsenic Cadmium PCBs Total halogens
60◦ C (min.) 100 ppm (max.) 10 ppm (max.) 5 ppm (max.) 2 ppm (max.) 50 ppm (max.) 1000 ppm (max.)a
a 4000 ppm if nontoxic source (see 40CFR279 for details).
Source: United States Code of Federal Regulations, Vol. 40, Part 279, Section 11 (40CFR279.11).
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On a positive note, Executive Orders 12873 (1993) and 13101 (1998) issued by President Clinton encourages the use of recycled products, including re-refined oil, which has demonstrated compliance with purchase specifications [21,22]. The direct pressure on government agencies to procure re-refined lubricants has provided a market for these products, as well as many others, the most successful of which has been paper. In Europe, governments have taken a more activist role. Countries such as Germany and Italy subsidize the collection and re-refining of used oils [23]. This direct encouragement has led to an output of approximately 100,000 t of base oil in an Italian market of around 600,000 t. In Germany about 35,000 t of base oils are produced by rerefiners in a market of 1,100,000 t. Those numbers would suggest that even government subsidies are not enough to divert used oils from the less difficult pathway of use as substitute fuels. A report released in 2001, considers the fairly radical step of prohibiting the burning of used oils in the United Kingdom. As stated in the report, this would have to be implemented very carefully, since the re-refining capacity to absorb that volume does not exist [24]. Another example of comprehensive environmental legislation can be found in Australia, where the federal and state governments have published and enforced extensive Codes of Practice for the management of used oil, filters, and oil containers. Federally funded public education programs and tiered tax incentives for processors have created a positive environment for the used oil recycling industry. Canada regulates used oil at the provincial level. Most provinces have environmental legislation, which includes sections on used oil and currently four of the provinces have incentive-based stewardship programs for the collection of used oil and filters managed by nonprofit organizations. Environmental handling fees are charged on oil sales. The fees are then used to pay registered collectors as a return incentive. These programs are supported by provincial recycling and used oil management legislation and are being considered for implementation across all provinces. Japan, having no onshore petroleum resources, currently is studying the most effective legislative method specifically to encourage used oil re-refining.
37.8 Pollution Prevention/Lifecycle Assessment Environmental protection and pollution prevention legislation exits in various forms in most developed countries. Recognition that used oil management is crucial to pollution prevention and resource conservation while also having a positive economic impact has been a major turning point in both government regulation and private sector interest. As has been previously shown, the environmental hazards, health, and safety aspects of used oil have been well defined. Current studies are now focusing on two main areas, (1) defining and quantifying the economics and marketplace drivers for used oil recovery and (2) the life cycle and environmental fate of used oil [25]. Life cycle analysis is a complex project, which must take into consideration such things as original production and sale, detailed use analysis (including loss mechanisms), recoverable volumes after use, available recycling options, transportation impacts (including emissions, from transport vehicles), energy use and emissions, and residues from the recycling process whether from burning or re-refining. Figure 37.3 is a very simple schematic to illustrate the life cycle and environmental fate analysis process for lubricant fluids.
37.9 Conclusion Regardless of the motivation mechanism, government and industry now share to some extent the twin goals of reducing the amount of waste going to landfill while preserving nonrenewable resources through recycling and removal of hazardous materials from air, land, and water. The evidence is found in regulatory documents and in environmental policy statements in corporate literature across the world.
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Loss in use, unrecoverable product
Lubricant use
Transport End of life – waste oil for disposal
Emissions
Fuel for space heaters
Used oil: (Removed from use/ drained from equipment)
Energy* Transport Feed for refinery cat crackers
Transportation
(Refinery life cycle)
Emissions
On-site reclaiming
Off-site reclaiming
Reprocessing
water, solids, catalyst**
Re-refining emissions
Energy*
Energy*
Industrial fuel
Water, solids, residues**
Asphalt flux
Transport Burning for Energy Recovery (end of life)
Chemicals
Water, solids, residues**
Transport
Industrial fuel
Additives Lubricant for re-use
Additives
Transport
Energy* Transport
Burning for energy recovery (end-of-life)
Emissions
Energy* Packaging Chemicals Lubricant for re-use
(*energy can be mechanical, thermal, electrical, etc.) (**water, solids, residues, spent catalyst, etc., may be treated for re-use and/or nonhazardous disposal)
Life cycle and environmental fate analysis for lubricant fluids.
© 2006 by Taylor & Francis Group, LLC
(asphalt life cycle)
Blended lubricant
Base oil
Emissions
FIGURE 37.3
Asphalt manufacture & use
emissions
Transport (Packaging life cycle)
Handbook of Lubrication and Tribology
Emissions
Used Oil Recycling and Environmental Considerations
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The plethora of suggested methods to achieve these goals indicates that a universal solution has yet to be identified. Indeed, based on waste diversity and quantity, technological advancement and local economies, there may be many routes chosen to achieve the best possible recycling solutions. Ultimately, business interest in maximizing profits and environmental interest in minimizing the generation of wastes should converge to optimize the use and recycling of lubricants. Looking at more than just the initial price of lubricants is a first step. Starting with high quality re-refined lubricants, maximizing the lifetime of these lubricants through good system maintenance, and collecting the used oils in such a way as to preserve them for optimum recycling, all lead to decreased costs and increased natural resource conservation. The steadily increasing lubricant prices, feedstock shortages, and the cost of legal disposal make conservation in use and recycling of used oil a necessity. By assessing the entire life cycle of a lubricant, one can minimize the overall costs incurred. By contracting with a responsible recycler, one can be assured of not contributing to pollution once the lubricant is no longer useful.
References [1] Brinkman and Dennis, W., Used oil: Resource or pollutant?, Technology Review, 88(5), 1985, p. 47. [2] European Commission, Integrated Pollution Prevention and Control, Draft Reference Document on the Best Available Techniques for the Waste Treatment Industries, Draft — February 2003. [3] Cotton, F.O., Waste Lubricating Oil: An Annotated Review, U.S. Department of Energy, DOE/BETC/IC-82/4, 1982, 84 pages. [4] Pedenaud, M., A. Bruce, M. Cayla, B. Descotis, G. Fisicaro, M.R.S. Manton, R. Prince, F.J. Sheppard, and J. Smith, Collection and Disposal of Used Lubricating Oil, CONCAWE Report No. 5/96, November 1996, 107 pages. [5] ASTM Annual Book of Standards, Vol. 5.02, D4175, Standard Terminology Relating to Petroleum, Petroleum Products, and Lubricants. [6] Fitzsimons, D., P. Lee, and N.J. Moreley, Oakdene Hollins, Ltd., WASTE OILS — Policy Options in the Light of German and Italian Experience, Final Report to UK Dept. for Environment, Food, and Rural Affairs, March, 2003, p. 20. [7] Sullivan and Tim, U.S. Lubes Dipped in 2002, Lube Report, LNG Publishing Co., November 2003. [8] American Petroleum Institute, National Used Oil Collection Study, May 1996. [9] Cotton, F.O., M.L. Whisman, J.W. Goetzinger, and J.W. Reynolds, Analysis of 30 Used Motor Oils, Hydrocarbon Processing, 56(9), 1977, p. 131. [10] Brinkman, D.W. and J.R. Dickson, Contaminants in used lubricating oils and their fate during Distillation/Hydrotreatment Re-refining, Environmental Science and Technology, 29, 1995, p. 81. [11] U.S. Department of Health and Human Services, Toxicological Profile for Used Mineral-Based Crankcase Oil, September 1997, 175 pages (available at www.atsdr.cdc.gov/toxprofiles/tp102.pdf). [12] Sax, N.I., Dangerous Properties of Industrial Materials, 7th ed., Van Nostrand Reinhold, NY, 1989, p. 2815. [13] International Agency for Research on Cancer (IARC). IARC Monographs on the Evaluation of Carcinogenic Risk to Humans; Polynuclear Aromatic Compounds; Part 1: Chemical, Environmental and Experimental Data, Vol. 32, 1983, pp. 57–62. [14] Dickson, J.R., D.W. Brinkman, and G.R. Blackburn, Evaluation of the Dermal Carcinogenic Potential of Re-Refined Base Stocks Using the Modified Ames Assay, PAC Analysis, and the 32 P-Postlabeling Assay for DNA Adduct Induction, Journal of Applied Toxicology, 17(2), 1997, p. 113. [15] ASTM D6823, Specification for Commercial Boiler Fuels with Used Lubricating Oils. [16] ASTM D6448, Specification for Industrial Boiler Fuels from Used Lubricating Oils. [17] U.S. EPA, http://www.epa.gov/superfund/programs/recycle/success/1-pagers/timesbch.htm. [18] Brinkman, D.W., Large grassroots lube rerefinery in operation, Oil and Gas Journal, 89(33), 1991, p. 60. [19] Pyziak, T. and D.W. Brinkman, Recycling and re-refining used lubricating oils, Lubrication Engineering, 49(5), 1993, p. 339.
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[20] Brinkman, D.W., Technologies for re-refining used oil, Lubrication Engineering, 43(5), 1987, p. 324. [21] Executive Order 12873, 10/20/1993, Federal Acquisition, Recycling, and Waste Prevention, Sec. 506. [22] Executive Order 13101, 9/14/1998, Greening the Government Through Waste Prevention, Recycling, and Federal Acquisition, Sec. 507. [23] Fitzsimons, D., P. Lee, and N. J. Morley, Oakdene Hollins, Ltd., http://www.oakdenehollins.co.uk/ pdf/wasteoils2.doc. [24] Fitzsimons, D., N. Morley, and P. Lee, UK Waste Oils Market 2001, Oakdene Hollins, Ltd., http://www.oakdenehollins.co.uk/pdf/wasteoilsreport.doc. [25] Boughton, Bob and Arpad Horvath, Environmental assessment of used oil management methods, Environmental Science and Technology, 38(2), 2004, p. 353. [26] ICIS LOR Base Oils Conference 2003 (based on estimated 2001 regional differences) [see also www.europalube.org]. [27] United States Code of Federal Regulations, Vol. 40, Part 279, Section 11 (40CFR279.11). [28] McKeagan, D.J., Economics of rerefining used lubricants, Lubrication Engineering, 48(5), 1992, p. 418.
© 2006 by Taylor & Francis Group, LLC
Appendices
© 2006 by Taylor & Francis Group, LLC
Appendix 1
Symbols for Lubrication Problems α g a α, β, Y , . . . ω A C, c f α ∂ D, d e
m J M n H p R, r N,n τ, ζ S c
mass mechanical equivalent of heat moment of force, torque number power pressure, normal force per unit area radius revolutions per unit time shear rate shear stress Sommerfeld number, (R/C)2 (µN/P) specific heat
T
temperature, absolute
η
acceleration, angular acceleration, gravitational acceleration, linear angles angular velocity area clearance, radial coefficient of friction coefficient of linear expansion density, mass per unit volume diameter eccentricity, bearing-journal center distance eccentricity ratio, eccentricity/ radial clearance efficiency
t
δ E
elongation, deformation, displacement energy
t k
h W F q
film thickness, thickness force, load friction force heat rate, heat flow rate, heat per unit time heat transfer coefficient, heat flow rate per unit area per degree length (perpendicular to direction of motion)
t M u, U , v, V µ
temperature difference, temperature rise temperature, ordinary thermal conductivity, heat flow rate per unit area per unit length per degree time torque, moment of force velocity viscosity, absolute (η is also used)
v
viscosity, kinematic
V
volume
ε
λ L, I
A1-3
© 2006 by Taylor & Francis Group, LLC
A1-4
Handbook of Lubrication and Tribology
W
load, force
Q
P
load per unit projected area
γ B, b Wk
volume rate, flow rate, volume per unit time weight per unit volume width (parallel to direction of motion) work
Source: Reprinted by permission of the American National Standard Institute, New York, Copyright 1962 (Rev. 1973), Z211.116.
© 2006 by Taylor & Francis Group, LLC
Appendix 2 SI Units and Conversion Factors Compiled by Douglas Godfrey
Introduction The American Society of Lubrication Engineers (ASLE) adopted the International System of Units (SI) on January 1, 1976 because of its logic and convenience of calculation and for national and international uniformity. For publication, conventional units may be added in parenthesis. The following units and conversions were taken primarily from Reference 1 and also from Reference 2. These publications give rules on the use of SI. Table 1 gives the basic and supplementary SI units. Tables 2 and 3 give the derived units divided into those with special names and other common derived units. Tables 4 and 5 give the prefixes and units used with the SI system. Table 6 gives conversion factors from common to SI units.
TABLE A2.1
International System of Units (SI)
Quantity
Unit
SI symbol
Base units Length Mass Time Electric current Thermodynamic temperature Amount of substance Luminous intensity
metre kilogram second ampere kelvin mole candela
m kg s A K mol cd
Supplementary Plane angle Solid angle
radian steradian
rad Sr
Note: All quantities must be given in the above units or their derivatives.
A2-5
© 2006 by Taylor & Francis Group, LLC
A2-6
Handbook of Lubrication and Tribology TABLE A2.2
Derived SI Units with Special Names
Quantity Frequency (of a periodic phenomenon) Force Pressure, stress Energy, work, quantity of heat Power, radiant flux Quantity of electricity, electric charge Electric potential, potential difference, electromotive force Electric capacitance Electric resistance Conductance Magnetic flux Magnetic flux density Inductance Celsius temperature Luminous flux Illuminance Activity (of a radionuclide) Absorbed dose Dose equivalent
Unit
Symbol
Formula
hertz newton pascal joule watt coulomb volt
Hz N Pa J W C V
1/s kg· m/s2 N/m2 N·m J/s A·s W/A
farad ohm siemens weber tesla henry degree Celsius lumen lux becquerel gray sievert
F S Wb T H ◦C lm lx Bq Gy Sv
C/V V/A A/V V·s Wb/m2 Wb/A K − 273.15 cd·sr lm/m2 l/s J/kg J/kg
Source: Reprinted with permission from the Annual Book of ASTM Standards, ASTM E 380–79, Copyright 1979, American Society for Testing and Materials, 1916 Race St., Philadelphia, PA, 19103.
© 2006 by Taylor & Francis Group, LLC
Appendix 2 SI Units and Conversion Factors TABLE A2.3
A2-7
Some Common Derived Units of SI
Quantity Absorbed dose rate Acceleration Angular acceleration Angular velocity Area Concentration (of amount of substance) Current density Density, mass Electric charge density Electric field strength Electric flux density Energy density Entropy Exposure (X- and gamma rays) Heat capacity Heat flux density irradiance Luminance Magnetic field strength Molar energy Molar entropy Molar heat capacity Moment of force Permeability (magnetic) Permittivity Power density Radiance Radiant intensity Specific heat capacity Specific energy Specific entropy Specific volume Surface tension Thermal conductivity Velocity Viscosity, dynamic Viscosity, kinematic Volume Wave number
Unit
Symbol
gray per second metre per second squared radian per second squared radian per second square metre mole per cubic metre ampere per square metre kilogram per cubic metre coulomb per cubic metre volt per metre coulomb per square metre joule per cubic metre joule per kelvin coulomb per kilogram joule per kelvin watt per square metre
Gy/s m/s2 rad/s2 rad/s m2 mol/m3 A/m2 kg/m3 C/m3 V/m C/m2 J/m3 J/K C/kg J/K W/m2
candela per square ampere per metre joule per mole joule per mole kelvin joule per mole kelvin newton metre henry per metre farad per metre watt per square metre watt per square metre steradian watt per steradian joule per kilogram kelvin joule per kilogram joule per kilogram kelvin cubic meter per kilogram newton per metre watt per metre kelvin metre per second pascal second square metre per second cubic metre 1 per metre
od/m2 A/m J/mol J/(mol · K) J/(mol · K) N·m H/m F/m W/m2 W/(m2 -sr) W/sr J/(kg · K) J/kg J/(kg · K) m3 /kg N/m W/(m · K) m/s Pa · s m2 /s m3 1/m
Source: Reprinted with permission from the Annual Book of ASTM Standards, ASTM E 380–79, Copyright 1979, American Society for Testing and Materials, 1916 Race St., Philadelphia, PA, 19103.
© 2006 by Taylor & Francis Group, LLC
A2-8
Handbook of Lubrication and Tribology TABLE A2.4
SI Prefixes
Multiplication factors 1,000,000,000,000,000,000 1,000,000,000,000,000 1,000,000,000,000 1,000,000,000 1,000,000 1,000 100 10 0.1 0.01 0.001 0.000,001 0.000,000,001 0.000,000,000,001 0.000,000,000,000,001 0.000,000,000,000,000,001
= = = = = = = = = = = = = = = =
1018 1015 1012 109 106 103 102 101 10−1 10−2 10−3 10−6 10−9 10−12 10−15 10−18
Prefix
SI Symbol
exa peta tera giga mega kilo hectoa dekaa decia centi milli micro nano pico femto atto
E P T G M k h da d c m µ n p f a
a To be avoided where possible.
Note: The above prefixes are to be used to indicate orders of magnitude. Source: Reprinted with permission from the Annual Book of ASTM Standards, ASTM E 380-79, Copyright 1979, American Society for Testing and Materials, 1916 Race Street, Philadelphia, PA, 19103.
TABLE A2.5 Quantity Time
Plane angle
Volume Mass Area
© 2006 by Taylor & Francis Group, LLC
Units in Use with SI Unit
Symbol
Definition
minute hour day week, month, etc. degree minute
min h d
1 min = 60 s 1 h = 60 min = 3600 s 1 d = 24 h = 86 400 s
◦
1◦ = (π/180) rad 1 = (1/60)◦ = (π/10, 800) rad 1 = (1/60) = (π/648, 000) rad 1 = 1 dm3 = 10−3 m3 1 t = 103 kg 1 ha = 1 hm2 = 104 m2
second
litre metric ton hectare
t ha
Appendix 2 SI Units and Conversion Factors TABLE A2.6
A2-9
Conversion Factors [1,2] Listing by Physical Quantity
To convert from
To
Multiply by
Acceleration foot/second2 free fall, standard gal inch/second2
metre/second2 metre/second2 metre/second2 metre/second2
−01 +00 −02 −02
3.048∗ 9.806 65∗ 1.00∗ 2.54∗
Area acre circular mil foot2 hectare inch2 mile2 (U.S. statute) section township yard2
metre2 metre2 metre2 metre2 metre2 metre2 metre2 metre2 metre2
+03 −10 −02 +04 −04 +06 +06 +07 −01
4.046 856 422 4∗ 5.067 074 8∗ 9.290 304∗ 1.00∗ 6.4516∗ 2.589 988 110 336∗ 2.589 988 110 336∗ 9.323 957 2 8.361 273 6∗
Bending Moment or Torque dyne centimeter kilogram (force) metre ounce (force) inch pound (force) inch pound (force) foot
newton metre newton metre newton metre newton metre newton metre
−07 +00 −03 −01 +00
1.00∗ 4.706 650∗ 7.061 552 1.129 848 1.355 818
Bending Moment or Torque per Unit Length pound (force) foot/inch pound (force) inch/inch
newton metre per metre newton metre per metre
+01 +00
5.337 866 4.448 222
Density gram/centimeter3 Ibm/inch3 Ibm/foot3 slug/foot3
kitogram/metre3 kilogram/metre3 kilogram/metre3 kilogram/metre3
+03 +04 +01 +02
1.00∗ 2.767 990 5 1.601 846 3 5.153 79
Energy British thermal unit (IST before 1956) (IST after 1956) British thermal unit (mean) British thermal unit (thermochemical) British thermal unit (39 F) British thermal unit (60 F) calorie (International Steam Table) calorie (mean) calorie (thermochemical) calorie(15 C) calorie (20 C) calorie kilogram, International Steam Table
joule joule joule joule joule joule joule joule joule joule joule jouie
+03 +03 +03 +03 +03 +03 +00 +00 +00 +00 +00 +03
1.055 04 1.055 056 1.055 87 1.054 350 1.059 67 1.054 68 4.1868 4.190 02 4.184∗ 4.185 80 4.181 90 4.1868
Energy calorie (kilogram, mean) calorie (kilogram, thermochemical) electron volt erg foot lbf foot poundal
joule joule joule joule joule joule
+03 +03 −19 −07 +00 −02
4.19002 4.184∗ 1.602 191 7 1.00∗ 1.355 817 9 4.214 011 0 (Continued)
© 2006 by Taylor & Francis Group, LLC
A2-10
Handbook of Lubrication and Tribology
TABLE A2.6 Continued To convert from
To
Multiply by
joule (international of 1948) kilocalorie (International Steam Table) kilocalorie (mean) kilocalorie (thermochemical) kilowatt hour kilowatt hour (international of 1948) therm ton (nuclear equivalent of TNT) watt hour watt second
joule joule joule joule joule joule joule joule joule joule
+00 +03 +03 +03 +06 +06 +08 +09 +03 +00
1.000 165 4.1868 4.190 02 4.184∗ 3.60∗ 3.600 59 1.055 056 4.20 3.60∗ 1.00
Energy/Area Time Btu (thermochemical)/foot2 second Btu (thermochemical)/foot2 minute Btu (thermochemical)/foot2 hour Btu (thermochemical)/inch2 second calorie (thermochemical)/cm2 minute erg/centimetre2 second watt/centimetre2 watt/inch2
watt/metre2 watt/metre2 watt/metre2 watt/metre2 watt/metre2 watt/metre2 watt/metre2 watt/metre2
+04 +02 +00 +06 +02 −03 +04 +03
1.134 893 1 1.891 488 5 3.152 480 8 1.634 246 2 6.973 333 3 1.00∗ 1.00∗ 1.550 003
Force dyne kilogram force (kgf) kilopond force lbf (pound force, avoirdupois) ounce force (avoirdupois) pound force, Ibf (avoirdupois) poundal
newton newton newton newton newton newton newton
−05 +00 +00 +00 −01 +00 −01
1.00∗ 9.806 65∗ 9.806 65∗ 4.448 221615 260 5∗ 2.780 138 5 4.448 221 615 260 5∗ 1.382 549 543 76∗
Length angstrom astronomical unit (IAU) astronomical unit (radio) cable chain (surveyor or gunter) chain (engineer or ramden) cubit fathom foot foot (U.S. survey) furlong hand inch league (U.K. nautical) league (international nautical) league (statute) light year link (engineer or ramden) link (surveyor or gunter) metre micron (micrometer) mil mile (U.S. statute) mile (U.K. nautical) mile (international nautical) mile (U.S. nautical) nautical mile (U.K.)
metre metre metre metre metre metre metre metre metre metre metre metre metre metre metre metre metre metre metre wavelengths Kr 86 metre metre metre metre metre metre metre
−10 +11 +11 +02 +01 +01 −01 +00 −01 −01 +02 −01 −02 +03 +03 +03 +15 −01 −01 +06 −06 −05 +03 +03 +03 +03 +03
1.00∗ 1.496 00 1.495 978 9 2.194 56∗ 2.011 68∗ 3.048∗ 4.572∗ 1.8288∗ 3.048∗ 3.048 006 096 2.011 68∗ 1.016∗ 2.54∗ 5.559 552∗ 5.556∗ 4.828 032∗ 9.460 55 3.048∗ 2.011 68∗ 1.650 763 73∗ 1.00∗ 2.54∗ 1.609 344∗ 1.853 184∗ 1.852∗ 1.852∗ 1.853 184∗
© 2006 by Taylor & Francis Group, LLC
Appendix 2 SI Units and Conversion Factors TABLE A2.6
A2-11
Continued
To convert from
To
Multiply by
nautical mile (international) nautical mile (U.S.) pace rod statute mile (U.S.) yard
metre metre metre metre metre metre
+03 +03 −01 +00 +03 −01
1.852∗ 1.852∗ 7.62∗ 5.0292∗ 1.609 344∗ 9.144∗
Mass carat (metric) grain gram hundredweight (long) hundredweight (short) kgf second2 metre (mass) kilogram mass lbm (pound mass, avoirdupois) ounce mass (avoirdupois) ounce mass (troy or apothecary). pennyweight pound mass, lbm (avoirdupois) pound mass (troy or apothecary) scruple (apothecary). slug ton (assay) ton (long) ton (metric) ton (short, 2000 pound) tonne
kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram
−04 −05 −03 +01 +01 +00 +00 −01 −02 −02 −03 −01 −01 −03 +01 −02 +03 +03 +02 +03
2.00∗ 6.479 891∗ 1.00∗ 5.080 234 544∗ 4.535 923 7∗ 9.806 65∗ 1.00∗ 4.535 923 7∗ 2.834 952 312 5∗ 3.110 347 68∗ 1.555 173 84∗ 4.535 923 7∗ 3.732 417 216∗ 1.295 978 2∗ 1.459 390 29 2.916 666 6 1.016 046 908 8∗ 1.00∗ 9.071 847 4∗ 1.00∗
Power Btu (thermochemical)/second Btu (thermochemical)/minute calorie (thermochemical)/second calorie (thermochemical)/minute erg/second foot Ibf/hour foot lbf/minute footlbf/second horsepower (550 foot Ibf/second) horsepower (boiler) horsepower (electric) horsepower (metric) horsepower (U.K.) horsepower (water) kilocalorie (thermochemical)/minute kilocalorie (thermochemical)/second watt (international of 1948)
watt watt watt watt watt watt watt watt watt watt watt watt watt watt watt watt watt
+03 +01 +00 −02 −07 −04 −02 +00 +02 +03 +02 +02 +02 +02 +01 +03 +00
1.054 350 264 488 1.757 250 4 4.184∗ 6.973 333 3 1.00∗ 3.766 161 0 2.259 696 6 1.355 817 9 7.456 998 7 9.809 50 7.46∗ 7.354 99 7.457 7.460 43 6.973 333 3 4.184∗ 1.000 165
Pressure or Stress atmosphere bar barye centimetre of mercury (0◦ C) centimetre of water (4◦ C) dyne/centimetre2 foot of water (39.2◦ F)
pascal pascal pascat pascal pascal pascal pascal
+05 +05 +01 +03 +01 −01 +03
1.013 25∗ 1.00∗ 1.00∗ 1.333 22 9.806 38 1.00∗ 2.988 98 (Continued)
© 2006 by Taylor & Francis Group, LLC
A2-12
Handbook of Lubrication and Tribology
TABLE A2.6 Continued To convert from mercury (32◦ F) mercury (60◦ F) water (39.2◦ F) water (60◦ F)
To
Multiply by
inch of inch of inch of inch of kgf/centimetre2 kgf/metre2 Ibf/foot2 Ibf/inch2 (psi) millibar millimeter of mercury (0◦ C) torr (0◦ C)
pascal pascal pascal pascal pascal pascal pascal pascal pascal pascal pascal
+03 +03 +02 +02 +04 +00 +01 +03 +02 +02 +02
3.386 389 3.376 85 2.490 82 2.488 4 9.806 65∗ 9.806 65∗ 4.788 025 8 6.894 757 2 1.00∗ 1.333 224 1.333 22
Speed or Velocity foot/hour foot/minute foot/second inch/second kilometre/hour knot (international) mile/hour (U.S. statute) mile/minute (U.S. statute) mile/second (U.S. statute)
metre/second metre/second metre/second metre/second metre/second metre/second metre/second metre/second metre/second
−05 −03 −01 −02 −01 −01 −01 +01 +03
8.466 666 6 5.08∗ 3.048∗ 2.54∗ 2.777 777 8 5.144 444 444 4.4704∗ 2.682 24∗ 1.609 344∗
Temperature Celsius Fahrenheit Fahrenheit Kelvin Rankine
kelvin kelvin Celsius Celsius kelvin
tK tK tC ◦ tC tK
= tc +273.15 = (5/9)(tF +459.67) = (5/9)(tF − 32) = TK − 273.15 = (5/9)tK
Time day (mean solar) hour (mean solar) minute (mean solar) Month (mean calendar) Year (calendar)
second second second second second (mean solar)
+04 +03 +01 +06 +07
8.64∗ 3.60∗ 600∗ 2.628∗ 3.1536∗
Viscocity centipoisea centistokea foot2 /second poise poundal second/foot2 Ib/foot second Ibf second/foot2 Ibf second/inch2 rhe stoke
pascal second metre2 /second metre2 /second pascal second pascal second pascal second pascal second pascal second 1 per pascal second metre2 /ssecond
−03 −06 −02 −01 +00 +00 +01 +03 +01 −04
1.00∗ 1.00∗ 9.290 304 1.00∗ 1.487 164 1.488 164 4.788 026 6.894 757 1.00∗ 1.00∗
Volume acre foot barrel (petroleum 42 gallons) board foot bushel (U.S.) cord cup fluid ounce (U.S.) foot3 gallon (U.K. liquid)
metre3 metre3 metre3 metre3 metre3 metre3 metre3 metre3 metre3
+03 −01 −03 −02 +00 −04 −05 −02 −03
1.233 481 837 547 52∗ 1.589 873 2.359 737 216∗ 3.523 907 016 688∗ 3.624 556 3 2.365 882 365∗ 2.957 352 956 25∗ 2.831 684 659 2∗ 4.546 087
© 2006 by Taylor & Francis Group, LLC
Appendix 2 SI Units and Conversion Factors TABLE A2.6
A2-13
Continued
To convert from
To
Multiply by
gallon (U.S. dry) gallon (U.S. liquid) inch3 litre ounce (U.S. fluid) peck (U.S.) pint (U.S. dry) pint (U.S. liquid) quart (U.S. dry) quart (U.S. liquid) tablespoon teaspoon ton (register) yard3
metre3 metre3 metre3 metre3 metre3 metre3 metre3 metre3 metre3 metre3 metre3 metre3 metre3 metre3
−03 −03 −05 −03 −05 −03 −04 −04 −03 −04 −05 −06 +00 −01
4.404 883 770 86∗ 3.785 411 784∗ 1.638 706 4∗ 1.00∗ 2.957 352 956 25∗ 8.809 767 541 72∗ 5.506 104 713 575∗ 4.731 764 73∗ 1.101 220 942 715∗ 9.463 529 5 1.478 676 478 125∗ 4.928 921 593 75∗ 2.831 684 659 2∗ 7.645 548 579 84∗
Volume per Unit Time (Includes Flow) foot3 /minute foot3 /second gallon (U.S.)/horse power hour inch3 /minute yard3 /minute gallon (U.S.)/day gallon (U.S.)/minute
metre3 /second metre3 /second metre3 /joule metre3 /second metre3 /second metre3 /second metre3 /second
−04 −02 −09 −07 −02 −08 −05
4.719 474 2.831 685 1.410 089 2.731 177 1.274 258 4.381 264 6.309 020
Note: The above tables express the definitions of miscellaneous units of measure as exact numerical multiples of coherent SI units, and provide multiplying factors for converting numbers and miscellaneous units to correspondingly new numbers and SI units. The first two digits of each numerical entry represent a power of 10. An asterisk follows each number which expresses an exact definition. For example, the entry “−022.54∗ ” epxresses the fact that 1 inch = 2.54 × 10−2 metre, exactly, by definition. Numbers not followed by an asterisk are only approximate representations of definitions, or are the result of physical measurements. a Allowable non-SI unit.
References [1] Mechtly, E. A., “International System of Units, Physical Constants and Conversion Factors,” 2nd revision, NASA SP-7012, U.S. Government Printing Office, Washington, D.C., 1973. [2] “Standard for Metric Practice,” ASTM E 380–79, Annual Book of ASTM Standards, American Society for Testing and Materials, Philadelphia, 1979.
© 2006 by Taylor & Francis Group, LLC
