Process Plant Machinery
This Page Intentionally Left Blank
Process Plant Machinery Second Edition
Edited by Heinz P. Bloeh, P.E.
and Claire Soares, P.E.
Boston Oxford Johannesburg Melbourne New Delhi Singapore
Butterworth-Heinemann 225 Wildwood Avenue, Wobum, Massachusetts 01801-204 I Linacre House, Jordan Hill, Oxford OX2 8DP "~
A member of the Reed Elsevier Group
9 1998 by Butterworth-Heinemann All rights reserved. No part of this publication may be reproduced, stored in a retrieval system, or transmitted in any form or by any means, electronic, mechanical, photocopying, recording, or otherwise, without the prior written permission of the publisher. Recognizing the importance of preserving what has been written, Butterworth-Heinemann prints its books on acid-free paper whenever possible. Butterwonh-Heinemann suppons the efforts of American Forests and the Global ReLeaf program in its campaign for the betterment of trees, forests, and our environment.
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Contents
Preface to the Second Edition
ix
Preface to the First Edition
xi
Acknowledgments
Xu
Introduction Chapter 1
Chapter 2
Chapter 3
eee
Xlil
Electric Motors and Controls
Motor Design Motor Controls Wiring System Design Cycling of Motors Standard Induction Motors Large Heavy-Duty Alternating Current Motors Special Industry and Application Designs Major Components of Induction Motors A Motor as Part of a System Adjustable Frequency Motor Considerations
2 5 6 7 7 8 8 11 13 20
Gear Speed Transmission Equipment
27
Types of Gears
27
Gas Turbines
45
Design Selecting a Gas Turbine Appendix 3A: Gas Turbine Cycles Appendix 3B: Life Cycle Usage Appendix 3C: Specific Maintenance Inspections
46 65 103 112 115
Chapter 4
Gas Engines Two-Stroke Gas Engines
121 122
Chapter 5
Steam Turbines
135
Mechanical Drive Steam Turbines Impulse Steam Turbines Reaction Steam Turbines Appendix 5A: Steam Turbines" Some Design Theory Factors Appendix 5B: Selecting and Sizing a Steam Turbine
136 137 155 167 182
vi Process Plant Machinery
Chapter 6
Chapter 7
Chapter 8
Chapter 9
Chapter 10
Chapter 11
Turboexpanders
199
Refrigeration Power Recovery Power Generation Expander Design and Construction Typical Application Operation Maintenance
199 199 199 200 203 204 205
Centrifugal Pumps
207
Conventional Process Pumps Canned Motor and Sealless Magnet-Drive Centrifugal Pumps High-Speed Centrifugal Pumps Appendix 7A: Centrifugal Pump Fundamentals Appendix 7B: Change of Performance Appendix 7C: Reed Frequency Considerations for Vertical Pump Installations Appendix 7D: Vertical Mixed Flow Variable Pitch Vane Pump Appendix 7E: Rotating Case Design
212 237
303
Positive Displacement Pumps
309
Reciprocating Positive Displacement Pumps Rotating Positive Displacement Pumps Appendix 8A: Principles of Operation of Reciprocating Pumps
309 321 329
Vacuum Pumps
337
Single-Stage Liquid Ring Pumps Liquid Jet Vacuum Pumps Air Ejector and/or Booster Liquid Ring Pumps Multistage Combination Units Rotary Oil-Sealed Vacuum Pumps
338 341 341 342 343
Cooling Water Supply Systems
345
Characterization by Air Flow Characterization by Construction Characterization by Shape Characterization by Method of Heat Transfer Mechanical Component Review
347 349 349 350 351
Centrifugal Compressors
363
Overview of Gas Compression Machinery Centrifugal Compressors Selection of a Centrifugal Compressor Appendix llA: Compressor Design
363 364 392 393
239 249 277 285 299
Contents
Appendix Appendix Appendix Appendix
Chapter 12
Chapter 13
Chapter 14
Chapter 15
Chapter 16
Chapter 17
11B: High-Speed Centrifugal Compressors 11C: Barrel Compressors liD: Isotherm Turbocompressors 11E: Gas Seal Design
398 407 430 467
Axial Flow Compressors
475
Field of Application Basic Axial Compressor Performance Capabilities Fundamentals of Axial Compressor Design Operational Limitations Standard Maintenance Considerations Selecting an Axial Compressor
475 477 478 482 484 487
Propeller, Axial, and Centrifugal Fans
517
Propeller Fans Axial Fans Centrifugal Fans Fan Fundamentals Fan Performance and System Effects Capacity Control of Fans
517 517 520 525 526 533
Reciprocating Compressors
539
Ideal Compressor Cycle Classification of Reciprocating Compressors Compressor Arrangement Overview Compressing Difficult Gases Compressors with Dry-Running Piston Rings Labyrinth-Piston Compressors Diaphragm Compressors Appendix 14A: Performance Calculations Appendix 14B: Capacity Control Appendix 14C: Higher Pressures on Dry-Running Compressors
541 542 544 564 566 568 576 579 584 592
Rotating Positive Displacement Compressors
595
Rotary Screw Compressors Rotary Piston Blowers Appendix 15A: Water-Flooded Single Screw Process Compressor Technology
595 607 611
Mixers and Agitators
617
Impeller Fluid Mechanics
617
Separators
633
Introduction Disk Nozzle Centrifuges Washing Applications
633 636 640 647
vii
viii
ProcessPlant Machinery
Chapter 18
Chapter 19
Chapter 20
Chapter 21
Appendix Index
Internal Mixers: Single- and Twin-Screw Extruders
651
Internal Mixers Single- and Twin-Screw Extruders Suggested Reading
651 660 678
Scraped Surface Crystallizers, Dewaxers and Chillers
679
Continuous Crystallizers for Organic Chemicals Scraped Surface Dewaxing Exchangers and Chillers
679 692
Conveyor-Based Processing Systems
695
Belt Grades Cooler/Flakers Heating/Drying Applications Double Belt Press Freezing Conveyor
695 698 703 703 706
Filtration Systems
707
Classification of Filters Filtration with Otherwise Active Media Backwashing Pretreatment Filter Selection Economics
707 711 711 712 712 712
Units Conversion Tables
715 725
Preface to the Second Edition I cut my rotating machinery teeth on the Syncrude tarsands project in northern Alberta, Canada, and was fortunate enough to work alongside some special people, including members of Exxon' s Research and Engineering Department, for about four years. The range of machinery, both in terms of application details and manufacturers that we worked with on that site, was delightfully varied and complex. For those and other process plant times in my life, this book would have been very welcome. One basic rule I followed in this edition was to leave as much of Heinz's original work and information collection, as is still current, strictly alone. His preface to the first edition described his objective in writing this book, in essence: close the gap between the chemical and the mechanical engineer. He succeeded with that objective as only he, in teamwork with all the people who contributed to the first edition, could. I am honoured that he entrusted me with adding additional information to this effort, to make a second edition. My guidelines, besides asking the original equipment manufacturers (OEMs) involved and several others what new products or modifications they had released, were to ask myself what major project work, such as machinery selection for a new plant or plant expansion or retrofits, would have benefited from additional information in this book. Hence, I have added guide sections for selection or design assessment of specific items such as gas turbines, where engineering development has been swifter than with other equipment types, and other major items such as compressors (axial and centrifugal). Some comments in my editing reflect the fact that, particularly in a tight economy, rotating machinery technical issues are hard to separate from the financial viability of the subject facility. Increasingly, arguments in chemical engineers' or rotating machinery engineers' careers are not with each other, but with the proverbial "bean counters". So the selection guide for gas turbines and compressors for instance, while including all pertinent technical facts, are also ones that the process engineer can use in discussions with, and financial justification documents to, nonengineering personnel. The information additions, specifically in the gas turbine chapter, reflect that a better reading of where land-based machinery designs are headed is evident from reading current literature on aircraft engine systems. Much of Rolls-Royce' s aviation RB211 technology ends up in their industrial RB211 models. General Electric's Frame 7s and 9s, are to a large extent, CF6-80C2 aviation technology sitting on a base on land. Last but by no means least, environmental considerations that arise directly from turbomachinery operation have grown hugely in profile since the release of the first edition of this book. In many large corporations chemical engineers are obliged by organizational structure, to leave those considerations to environmental engineers, even though they must be just as aware of them. As the environmental factors also tie in to mechanical engineers' work, I added in references, where possible, along these lines. If I did that to any comprehensive depth, however, we would end up with an impossibly large book. Hence the odd cross-reference in the book to Environmental Engineering & Management: Sustainable Development ix
x
Prefaceto the Second Edition for the Power Generation, Oil & Gas and Process Industries, which was written in response to questions often asked of me, but recently with increasing frequency. So if this book does not satisfy the chemical engineer's curiosity on enx;ironmental issues as they pertain to turbomachinery, to the required depth, know that I had to do that as a separate project. Claire Soares Dallas, Texas claire_ soares @compuserve.com
In the 1990 to 1995 timeframe, I was often asked by technical individuals when they could expect an updated second edition of the 1989 text. I had to decline even the invitation to give a non-binding estimate until late 1996. That's when Claire renewed our acquaintance dating to course presentations and attendance at turbomachinery conferences in our Exxon days. It was then that I explained to Claire that I was just too busy with my various and sundry "post-retirement" jobs of writing a monthly reliability column for Hydocarbon processing Magazine, organizing yearly International Process Plant Reliability Conferences for Gulf Publishing in Houston, teaching machinery reliability improvement courses, playing with three grandchildren, or just plain travelling with my supportive wife June, who had stuck with me for 34 years. To make a long story short, Claire agreed to tackle the job, and as I sit here reviewing the page proofs, I'm even more convinced Claire was the fight person to do the job. My thanks to her for going beyond what's normally needed to compile the next edition. Our readers and course attendees will surely benefit from her commendable and highly relevant efforts. Heinz P. Bloch Montgomery, Texas
Preface to the First Edition When I graduated from the New Jersey Institute of Technology in 1962, I had a degree in Mechanical Engineering and a fairly good knowledge of thermodynamics, fluid mechanics, and machine design. Little did I know, however, that application of these concepts to the real world of process machinery would be an altogether different learning process. My early suspicion that the recent engineering graduate would greatly benefit from a down-to-earth, application-oriented text on process plant machinery was further reinforced when I accepted employment with the Exxon Corporation. There, I dealt with chemical engineers who had an equal and sometimes greater need to understand the basic operating concepts and application criteria of the machines that both of us encountered in modem process plants. It was then that I searched unsuccessfully for such a text and perhaps also thought I might, some day, assemble the material for this book. The opportunity to synthesize the best available information into a cohesive overview text presented itself in 1986 and 1987 when I elected to retire from an interesting career in machinery reliability and maintenance-related engineering work with Exxon. I then asked some of the best companies in the process machinery field for the material to be included in this text, and many of them responded to the challenge. The guidelines were straighforward: We wanted to present the younger engineer or technician with an overview of the machinery categories he or she was likely to encounter in most process plants. Furthermore, even the experienced individual should be able to benefit from a reference text that didn't dwell on theory but went quickly into a thorough explanation of how the equipment was designed and what made it work. Stated differently, this text is an attempt to close the gap between the machinery engineer and the chemical engineer. Together with my many contributors, I express the hope that we have accomplished this goal. Heinz P. Bloch Montgomery, Texas
xi
To Bill Tillman, with thanks for getting me started Heinz Bloch
To John Russell, the late Jim Pugh and Stan Nathanson: for letting me learn as fast as I wanted, even when that was not convenient Claire Soares
Acknowledgments Our thanks are due to the following individuals and companies that helped make the second edition a reality. H. Middleton, D. Lloyd and K. Taylor from Rolls-Royce, Ansty, I&M group; the training staff at Rolls-Royce aircraft engine group, Derby A. Von Rappard from ABB S. Mauriol from Sulzer-Burckhardt, W. Melanson and D. Kitner from Dresser-Rand K. Reich from Demag Delaval C. Coyle from Lightnin, W.C. Chiang and A. Lee from Dorr-Oliver J. Armstrong from Armstrong Engineering Associates It would have been impossible to assemble the first edition text without the help and cooperation of many experienced equipment manufacturers and key individuals at each of the companies who devoted time and effort to the task. When we explained our intention to combine the best available information into a relevant text book, we received support from professionals who shared our goal of providing a practical, up-to-date text on the subject of process plant machinery. The following companies and individuals have earned our gratitude by allowing us to reprint, adapt, or otherwise incorporate their work in this book: Reliance Electric Co., Cleveland, OH; The Marley Cooling Tower Co., Mission, KS (electric motors and control); Prager Inc. Gear and Machine Products, New Orleans, LA; Westech Gear Corporation, Lynnwood, CA; Maag Gear Co., Zurich, Switzerland (gear speed transmission equipment); General Electric Co., Schenectady, NY (gas turbines); Cooper-Bessemer Reciprocating and Mr. J. Helmich, Grove City, PA (gas engines); Elliott Co., Jeannette, PA; Siemens Energy and Automation, Bradentown, FL (steam turbines); Marl-Trench Corp. and Mr. Jens E. Hanson, Santa Maria, CA (turboexpanders); Goulds Pumps, Inc., Seneca Falls, NY; Kontro Co., Inc., Orange, MA; Sundstrand Fluid Handling and Mr. John Alexander, Arvada, CO; Girdlestone Pumps Limited, Woodbridge, England (centrifugal pumps); Industrial Press, Inc., New York, NY; Dresser-Worthington, Harrison, NJ; Foster Pump Works, Inc., Westerley, RI; Milton Roy Co., St. Petersburg, FL; Pulsafeeder Co., Rochester, NY; Wallace & Tieman, Belleville, NJ; The Hydraulic Institute, Cleveland, OH; McGraw-Hill Book Co., New York, NY; National Mastr Pump, Inc., Houston, TX; Netzsch Pumps, Philadelphia, PA; Leistritz Pump Corp., Allendale, NJ; American Lewa, Inc., Holliston, MA (positive displacement pumps); Kinney Vacuum Co., Boston, MA; SIHI Pumps, Inc., Grand Island, NY; Stokes Division of Pennwalt Corp., Philadelphia, PA; Sulzer-Burckhardt, Basel and Zurich, Switzerland (vaccum pumps); The Marley Cooling Tower Co. and Mr. John C. Hensley, Mission, KS (cooling water supply systems); Mannesmann-Demag and Mr. Otto Robert Hoesel, Houston, TX and Duisburg, Germany; Sundstrand Fluid Handling and Mr. Larry Glassbum, Arvada, CO (centrifugal compressors); Dresser-Rand Co. and Ms. Kathie H. Bulson, Phillipsburg, NJ (axial compressors); ooo
XIII
xiv Acknowledgments Garden City Fan Co. and Mr. Eugene Folkers, Niles, MI; Cemax Fans, Boston, MA; ACME Engineering and Manufacturing Corp., Muskogee, OK (propeller, axial, and centrifugal fans); Sulzer-Burckhardt and Messrs. H.R. Klaey and C. Matile, Basel and Zurich, Switzerland; Transamerica-De Laval, Trenton, NJ (reciprocating compressors); Aerzen USA Corp. and Mr. Pierre Noack, Exton, PA (rotating positive displacement compressors); Mixing Equipment Co., Inc. and Dr. James Y. Oldshue, Rochester, NY; Burgmann Seals America, Houston, TX (mixers and agitators); Wemer & Pfleiderer Corp. and Mr. Adam Dreiblatt, Ramsey, NJ (internal mixers, single and twin-screw extruders); Sandvik Process Systems, Inc. and Mr. Jack Kahrs, Totowa, NJ (conveyor-based processing systems). Three well-qualified machinery engineers and long-term associates have assisted me in compiling several chapters of this book. Hurlel Elliott, Gampa Bhat, and Perry Monroe deserve credit for their contributions to the chapters on fans, gas turbines, and electric motors, respectively. Last, but not least, I'm indebted to my editor, Greg Franklin, who was compelled to show more than the usual patience in bringing it all together.
Introduction Modem process plants could not exist without the machinery to transport, modify, mix, compress, or otherwise manipulate the gases, liquids, and solids that move through the plant at any given time. There are literally hundreds of types of plant machines, and it would not be at all unusual to find 20 or 30 different types at a single process plant site. Moreover, these could be further subdivided into numerous modified versions, depending on desired throughput, pressure, processing temperature, product characteristics, design life, and a host of other parameters. In determining the scope of coverage it quickly became evident that we had to be highly selective in our choice of machinery to be included in this text. Dealing with every conceivable machine type would be an encyclopedic task and could result in superficial descriptions. Fortunately for the reader, and also the writer, the many process plant machines can be assigned to a few major classifications, or functional categories. Once the operating principles and typical configurations of important machine components are understood, the reader will find it considerably easier to think through the design variations or derivatives of a given machine. Making this our basic premise, we decided that the majority of chemical engineers working in the process industries will probably encounter machinery that fits into one of the four primary classifications: 9 9 9 9
prime movers and power transmission machinery pumping equipment gas compression machinery mixing, conveying, and separation equipment
We describe the essentials and, if necessary, important details of a number of process machines and drivers that fit into these four primary classifications. Starting with electric motors, the reader will find substantial information on turbines, pumps, and compressors. We describe the essentials and, if necessary, important details of a number of process machines and drivers that fit into these four primary classifications. The book commences with electric motors, and the chapters on turbines, pumps and compressors have been extended considerably. Chapters on centrifuges, scraped surface crystallizers and filtration have been added to the second edition, as well as an appendix on conversion between USCI and SI units. We had to grapple with the question of whether the United Customery System of units (USCI) or International System of units (SI) should be used. From a purely practical view, it was realized t h a t - like it or n o t - the reader will continue to (1) encounter both systems for the foreseeable future, and (2) have to be able to convert from one system to the other. Accordingly, we opted to generally leave the decision to the individual contributor; thus the reader will find himself or herself immersed in both USCI and SI units.
Xu
xvi
Introduction
Note that compressor sealing components may be described in more than one location, depending on component features and applications (for instance barrel compressors and isotherm compresors). Also, since some components are common to more than one genetic type of machine (for example centrifugal compressors and axial compressors), they may not be described again if they are essentially similar. In this case, the reader is encouraged to use the index at the end of this text.
Chapter 1 Electric Motors and Controls* Electric motors are by far the most prevalent "machines" in use in process plants around the world. The engineer, technician, and operator will benefit from an overview-type knowledge of electric motors; accordingly, we have elected first to introduce the reader to this machine category. Basic motor types, their major component parts, selection criteria, and other topics will be reviewed. To assist the reader, we have included a "motor glossary" at the end of this chapter. Squirrel-cage induction motors are the most widely used type in the size range up to 200 horsepower. An induction motor is an alternating current device in which the primary winding on one member (usually the stator) is connected to the power source and the secondary winding or a squirrel-cage secondary winding on the other member (usually the rotor) carries the induced current. There is no physical electrical connection to the secondary winding; its current is induced. Induction motors are simple, rugged, and reliable because they have no rotating windings, slip tings, or commutators. They have good efficiency and high starting torque, but a lagging power factor; the induction motor operates below the synchronous speed that is set by the power cycles and the number of poles in the stator. For example, a two-pole, 60-Hertz (cycle-per-second) motor has a synchronous speed of 3600 RPM, and a four-pole motor operates at 1800 RPM. Normal operating speeds for induction motors are 3550 RPM and 1750 RPM. This deviation from synchronous speed is called slip, and it varies with load. Full load (maximum) slip varies from 1 percent in large motors to as high as 5 percent in small units. Most motors have an average slip of 3 percent. Synchronous motors are used in cases where a fixed speed is required or to correct a lagging power factor in the power distribution system. A synchronous motor operates at constant speed up to full load, with a rotor speed equal to the speed of the rotating magnetic field of the stator, i.e., there is no slip. There are two major types of magnets on the synchronous motor: reluctance and permanent. Because of a relatively low inrush current, the synchronous motor has low starting torque and will trip out on overload when started under a load. Electric motor configurations can be divided into two major categories: horizontal and vertical. The output shafts of horizontal motors are constructed parallel (horizontal) to the ground or mounting base, while the output shaft of a vertical motor will be perpendicular (or at a right angle) to the conventional mounting base. There is a growing interest in the use of vertical motors as pump drivers in the petroleum, petrochemical, and chemical industries, because the entire assembly requires less space than a horizontal pump. Both horizontal and vertical motors are available in a variety of designs and enclosures. They are classified as standard duty, large heavy-duty, special industry designs, and special application designs. Each of these classifications is available in different enclosures to meet a multitude of environmental requirements, t The * Source: Reliance Electric Company, Cleveland, OH. Adapted by permission. t This segment courtesy of the Marley Cooling Tower Corporation. Adapted by permission.
2
ProcessPlant Machinery two basic types of motor enclosures are open and totally enclosed. The open motor circulates external air inside the enclosure for cooling, whereas a totally enclosed motor prevents outside air from entering the enclosure. Both types of motors are available in either fractional or integral horsepowers. Open motors are further classified as drip-proof, splash-proof, guarded, and weather protected, the distinction between them being the degree of protection provided against falling or airborne water gaining access to live and rotating parts. Drip-proof motors, as defined by the National Electrical Manufacturers Association (NEMA), are never used on outside installations. Integral-horsepower open dripproof motors now marketed (sometimes called "protected") usually meet NEMA requirements for weather-protected type I enclosures, except they will not prevent the passage of a 3/4" diameter rod. These enclosures are used on a wide variety of process machinery in outside locations. NEMA-defined weather-protected type II enclosures require oversized housings for special air passages to remove airborne particles. This type of enclosure is not available in the smaller motor sizes. Totally enclosed motors used on such process equipment as cooling towers are classified as non-ventilated (TENV), fan-cooled (TEFC), air-over (TEAO), and explosion-proof. Whether a motor is TENV or TEFC is dependent on the need for an internally mounted fan to keep the operating temperature of the motor within the rating of its insulation. Air-over motors are TEFC motors without the fan and must have an outside cooling source. Totally enclosed motors are recommended and used for locations where fumes, dust, sand, snow, and high humidity conditions are prevalent, and they can provide a high quality installation either in or out of the air stream provided the typical problems of mounting, sealing, and servicing are properly addressed. In all cases, totally enclosed motors should be equipped with drain holes, and explosion-proof motors should be equipped with an approved drain fitting. Explosion-proof motors are manufactured and sold for operation in hazardous atmospheres, as defined by the National Electrical Code. The motor enclosure must withstand an explosion of the specified gas, vapor, or dust within it, and prevent the internal explosion from igniting any gas, vapor, or dust surrounding it. The motors are Underwriters Laboratory (UL) approved and marked to show the class, group, and operating temperature (based on a 40 ~ ambient temperature) for which they are approved. In applying these motors, no external surface of the operating motor can have a temperature greater than 80 percent of the ignition temperature of the gas, vapor, or dust involved. The National Electrical Code defines hazardous locations by class, group, and division. Class I locations contain flammable gases or vapors; class II locations contain combustible dust; and class III locations contain ignitible fibers or flyings. Group defines the specific gas, vapor, dust, fiber, or flying. Division defines whether the explosive atmosphere exists continuously (division 1) or only in case of an accident (division 2). Motors for division 1 applications must be explosion-proof. Standard open or totally enclosed motors that do not have brushes, switching mechanisms, or other arc-producing devices can be used in class I, division 2 applications. In some cases, they can also be used in class II, division 2 and class III, division 2 applications.
MOTOR DESIGN Three-Phase Motors Three-phase squirrel-cage induction motors have become the standard on the overwelming majority of process plant machines. They do not have the switches, brushes, or capacitors of other designs and therefore require somewhat less
Electric Motors and Controls
maintenance. Where three-phase power is not available, single phase capacitor-start motors may be used, usually not exceeding 7.5 horsepower. Concerned process machinery manufacturers will supply motors that are a few steps beyond "off-theshelf" quality. These motors are usually purchased from specifications developed after comprehensive, rigorous testing under simulated operating conditions.
Two-Speed Motors Two-speed motors are of a variable torque design, in which the torque varies directly with the speed, with 1800/900 RPM being the most common speeds. Single-winding design motors enjoy greatest utilization, since they are smaller in size and less expensive than those of a two-winding design. Occasionally, an installation deserves consideration of the use of two-speed motors. Cooling tower fans are a case in point. Whether operated seasonally or year-round, there will be periods when a reduced load and/or a reduced ambient temperature will permit satisfactory cold water temperatures with the fans operating at half-speed. The benefits accrued from this mode of operation will usually offset the additional cost of two-speed motors in a relatively short time. Additionally, since nighttime operation is normally accompanied by a reduced ambient temperature, some operators utilize two-speed motors to preclude a potential noise complaint.
High-Efficiency Motors Several motor manufacturers provide high-efficiency designs that are suitable for use on numerous types of process plant machinery. These motors are in the same frame sizes as standard motors, but they utilize more efficient materials. While the efficiency will vary with the manufacturer and the size of the motor, the efficiency will always be higher than that manufacturer's standard motor. Naturally, there is a price premium for high-efficiency motors, which must be evaluated against their potential for energy savings.
Motor Insulation One of the most important factors contributing to long service life in an electric motor is the quality of the insulation. It must withstand thermal aging, contaminated air, moisture, fumes, expansion and contraction stresses, mechanical vibration and shock, as well as electrical stress. Insulation is categorized by classes, which establish the limit for the maximum operating temperature of the motor. Classes A, B, F, and H are used in the United States, with class A carrying the lowest temperature rating and class H the highest. Standard integral horsepower motors have class B insulation and are designed for a maximum altitude of 3300 feet and a maximum ambient temperature of 40 ~ Class F insulation is used for higher altitudes, as well as higher ambient temperatures, and it is gaining increased use as a means of improving the service factor of a motor of given frame size.
Motor Service Factor The service factor of a motor is an indication of its maximum allowable continuous power output, as compared with its nameplate rating. A 1.0 service factor motor
3
4
Process Plant Machinery
should not be operated beyond its rated horsepower at design ambient conditions, whereas a 1.15 service factor motor will accept a load 15 percent in excess of its nameplate rating. Usually, motor manufacturers will apply the same electrical design to both motors but will use class B insulation on 1.0 service factor motors and class F insulation on 1.15 service factor motors. Class B insulation is rated at a total temperature of 130~ and class F is rated at 155 ~ More important, a 1.15 service factor motor operates at a temperature from 15 ~ to 25 ~ lower (compared with the temperature rating of its insulation) than does a 1.0 service factor motor operating at the same load. This, of course, results in longer insulation life and, therefore, longer service life for the motor. For this reason, many equipment manufacturers will recommend the use of 1.15 service factor motors for loads at or near nominal horsepower ratings. Since increased air density increases the load on air movers, an added attraction for using 1.15 service factor motors is that there is less chance of properly sized over-loads tripping out this equipment category during periods of reduced heat load and low ambient temperatures.
Motor Heaters Although the insulation used in quality electric motors is considered to be nonhygroscopic, it does slowly absorb water and, to the degree that it does, its insulation value is reduced. Also, condensed moisture on insulation surfaces can result in current leakage between pin holes in the insulation varnish. Because of this, it is advisable on installations exposed to high humidity to keep the inside of the motor dry. This can be done by keeping the temperature inside the motor 5 ~ to 10~ higher than the temperature outside the motor. Motors in continuous service will be heated by the losses in the motor, but idle motors require the addition of heat to maintain this desired temperature difference. One recommended method of adding heat is by the use of electric space heaters, sized and installed by the motor manufacturer. Another method is single-phase heating, which is simply the application of reduced voltage (approximately 5 to 7.5 percent of normal) to two leads of the motor winding. Both of these methods require controls to energize the heating system when the motor is idle. If low voltage dynamic braking is used to prevent an inoperative motor from rotating, it will add sufficient heat to the motor windings to prevent condensation. A typical application would be in cooling tower fans.
Motor Torques High starting torque motors are neither required nor recommended for most process machines. Normal torque motors perform satisfactorily for pumps, fans, blowers, etc., and cause far less stress on the driven components. Normal torque motors should be specified for the bulk of single-speed applications, and variable torque in the case of two-speed. There are five points along a motor speed-torque curve that are important to the operation of many machines: (1) locked-rotor torque, (2) pull-up torque (minimum torque during acceleration), (3) breakdown torque (maximum torque during acceleration), (4)full-load torque, and (5)maximum plugging torque (torque applied in reversing an operating motor). Compared with full-load torque, the average percentage values of the other torques are as follows: locked-rotor torque = 200%; pull-up torque = 100%; breakdown torque = 300%; and plugging torque = 250%.
Electric Motors and Controls
MOTOR CONTROLS Control devices and wiring, the responsibility for which usually falls to the purchaser, can also be subjected to demanding service situations. Controls serve to start and stop the motor and to protect it from overload or power supply failure, thereby helping assure continuous reliable equipment operation. They are not routinely supplied as a part of a machinery procurement contract, but because of their importance to the system, the need for adequate consideration in the selection and wiring of these components cannot be overstressed. The various protective devices, controls, and enclosures required by most electrical codes are described in the following paragraphs. In all cases, motors and control boxes must be grounded.
1. Fusible Safety Switch or Circuit Breaker: This device provides the means to disconnect the controller and motor from the power circuit. It also serves to protect the motor-branch-circuit conductors, the motor control apparatus, and the motors against overcurrent due to short circuits or grounds. It must open all ungrounded conductors and be visible (not more than 50 feet distant) from the controller or be designed to lock in the open position. The design must indicate whether the switch is open or closed, and there must be one fuse or circuit breaker in each ungrounded conductor. A disconnect switch must be horsepower rated or must carry 115 percent of full-load current and be capable of interrupting stalled-rotor current. A circuit breaker must also carry 115 percent of full-load current and be capable of interrupting stalled-rotor current. 2. Nonfused Disconnect Switch: This switch is generally only required if the fusible safety switch or circuit breaker either cannot be locked in the open position or cannot be located in sight of the motor. 3. Manual and Magnetic Starters: These controls start and stop the motor. They also protect the motor, motor control apparatus, and the branch-circuit conductors against excessive heating caused by low or unbalanced voltage, overload, stalled rotor, and too frequent cycling. Starter requirements are determined by the basic horsepower and voltage of the motor. Overloads in a starter are sized to trip at not more than 125 percent of full-load current for motors having a 1.15 or higher service factor, or 115 percent of full-load current in the case of 1.0 service factor motors. Single phase starters must have an overload in one ungrounded line. A three-phase starter must have overloads in all lines. If a magnetic controller is used, it may be actuated by devices sensing certain process fluid parameters. Temperature sensors sensing cooling water temperature would be a typical example. 4. Control Enclosures: NEMA has established standard types of enclosures for control equipment. The types most commonly used in conjunction with process plant machinery are as follows: a. NEMA Type 1 - G e n e r a l Purpose: Intended primarily to prevent accidental contact with control apparatus. It is suitable for general purpose applications indoors, under normal atmospheric conditions. Although it serves as a protection against dust, it is not dust-proof. b. NEMA Type 3 - Dust-tight, Rain-tight, and Sleet-resistant: Intended for outdoor use and for protection against wind-blown dust and water. This sheet metal enclosure is usually adequate for use outdoors on a cooling tower. It has a watertight conduit entrance, mounting means external to the box, and provision for locking. Although it is sleet-resistant, it is not sleet-proof. c. NEMA Type 3R: This is similar to type 3, except it also meets UL requirements for being rainproof. When properly installed, rain cannot enter at a level higher than the lowest live part.
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Process Plant Machinery
d. NEMA Type 4 - Watertight and Dust-tight: Enclosure is designed to exclude water. It must pass a hose test for water and a 24-hour salt spray test for corrosion. This enclosure may be used outdoors on a cooling tower. It is usually a gasketed enclosure of cast iron or stainless steel. e. NEMA Type 4X: Similar to type 4, except it must pass a 200-hour salt spray test for corrosion. It is usually a gasketed enclosure of fiber-reinforced polyester. f. NEMA Type 6 - Submersible, Watertight, Dust-tight and Sleet-resistant: Intended for use where occasional submersion may be encountered. Must protect equipment against a static head of water of 6 feet for 30 minutes. g. NEMA Type 1 2 - Dust-tight and Drip-tight: Enclosure intended for indoor use. It provides protection against fibers, flyings, lint, dust, dirt, and light splashing. h. NEMA Type 7 - Hazardous L o c a t i o n s - Class I Air-Break: This enclosure is intended for use indoors in locations defined by the National Electrical Code for class I, division 1, groups A, B, C, or D hazardous locations. i. NEMA Type 9 - Hazardous Locations - Class II Air-Break: Intended for use indoors in areas defined as class II, division 1, groups E, F, or G hazardous locations. WIRING SYSTEM DESIGN
The design of the wiring system for the numerous process machines, fans, compressors, pumps, and controls is the responsibility of the owner's engineer. Although the average installation presents no particular problem, there are some systems that require special consideration if satisfactory operation is to result. Conductors to motors must be sized both for 125 percent of the motor full-load current and for voltage drop. If the voltage drop is excessive at full load, the resultant increased current can cause over-load protection to trip. (Although motors should be operated at nameplate voltage, they can be operated at plus or minus 10 percent of nameplate voltage.) In a normal system with standard components, even the larger machines will often attain operating speed in less than 15 seconds. During this starting cycle, although the motor current is approximately 600 percent of full-load current, the time delays in the overload protective devices prevent them from breaking the circuit. Because of the high starting current, the voltage at the motor terminals is reduced by line losses. Within certain limits, the output torque of a motor varies as the square of the voltage. Thus, under starting conditions, the current increases, the voltage decreases, and the torque decreases, with the result that the starting time is increased. Long conductors that increase voltage drop, low initial voltage, and high-inertia fans can all contribute to increased starting time, which may cause the protective devices to actuate. In extreme cases, the starting voltage may be insufficient to allow acceleration of the fan to full speed regardless of time. The wiring system design must consider pertinent data on the available voltage (its actual value, as well as its stability), length of lines from the power supply to the motor, and the motor horsepower requirements. If this study indicates any question as to the startup time of the motor, the inertia of the load as well as that of the motor should be determined. This is the commonly known "flywheel" effect (WK 2 factor). Once the WK 2 of the load (referred to the motor shaft) is obtained, the acceleration time can be determined using the motor speed-torque and speed-current curves, compared with the speed-torque curves for the fan. If the calculated time and current is greater than allowed by the standard overload protection of the motor,
Electric Motors and Controls
the condition may be corrected by increasing voltage, by increasing conductor size, or by providing special overload relays. Given no solution to the base problem, special motors or low-inertia fans may be necessary. CYCLING OF MOTORS
The high inrush current that occurs at motor startup causes heat to build up in the windings and insulation. For this reason, the number of start-stop or speed-change cycles should be limited in order to allow time for excessive heat to be dissipated. As a general rule, 30 seconds of acceleration time per hour should not be exceeded. A fan-motor system that requires 15 seconds to achieve full speed, therefore, would be limited to two full starts per hour. Smaller or lighter fans, of lesser inertia, permit greater frequency of cycling. STANDARD I N D U C T I O N MOTORS*
Standard induction motors are typically used for industrial applications such as machine tools, material handling equipment, processing lines, pumps, fans, blowers, and countless others. The protected motor shown in Figure 1-1 (sometimes called the drip-proof or open) is typical of general purpose AC motors. This enclosure is suited for most industrial environments when temperatures are 40~ maximum with ambient air relatively clean and dry. The totally enclosed fan-cooled (TEFC) motor (Figure 1-2) is the enclosed motor most often selected for indoor or outdoor industrial environments containing dust,
FIGURE 1-1 Energy-efficient electric motor typical of general purpose AC motors. (Source: Reliance Electric, Cleveland, OH.)
* Source: Reliance Electric Company, Cleveland, OH. Adapted by permission.
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Process Plant Machinery
FIGURE 1-2 Totallyenclosed fan-cooled motor. The externally mounted fan pushes cooling air over the fin-equipped frame. (Source: Reliance Electric, Cleveland, OH.)
dirt, water, etc., in modest amounts that are best kept out of the interior of motors. The external fan is used to cool the motor, since there is not a free exchange of air. Explosion-proof motors have a dual purpose: to withstand an explosion from within and to prevent the explosion of gases in the atmosphere. Both requirements place special emphasis on motor design. Explosion-proof motors should meet the rigid requirements of UL for most National Electrical Code class, group, and temperature code restrictions. Motor sizes and dimensions have been standardized by the NEMA. NEMA frames 48, 56, and 140 T encompass single-phase capacitor start and polyphase motors. These motors are designed for continuous duty operations in a 40~ maximum ambient environment- obviously not typical plants. Similarly, NEMA frames 180 T through 449 T include standard AC polyphase induction motors and also certain single-phase motors designed for a maximum environmental temperature of 40 ~ They are, however, available in explosion-proof executions for use in class I, group D, and class II, groups F and G hazardous locations. LARGE HEAVY-DUTY ALTERNATING CURRENT MOTORS
Large heavy-duty AC motors cover the range from 250 to 5000 horsepower and are available in a variety of frames or enclosures. Totally enclosed motors, typically of cast iron frame design, range to roughly 500 horsepower. Weather-protected designs and tube or water-cooled enclosures can be supplied with these motors. They range through 1500 horsepower with cast iron, and 5000 horsepower with fabricated steel construction (Figure 1-3). SPECIAL INDUSTRY AND APPLICATION DESIGNS
Capable motor manufacturers can offer suitably modified motors for specific industries. Typical are motors designed for corrosive atmospheres such as those found in the paper, chemical, petroleum, and metals industries. These motors would be available from fractional to 500 horsepower.
Electric Motors and Controls
FIGURE 1-3 Weather-protected motor with cooling facilities enclosure and fabricated frame design. (Source: Reliance Electric, Cleveland, OH.) Motors for the food and dairy industries must be designed for easy cleaning and hose washdowns to meet the rigid sanitary codes of all government agencies, the Baking Industry Sanitation Standards Committee, and Dairy Standards. Figure 1-4 depicts an important special application design, a brake motor that combines a motor and an integrally mounted disc brake into one unit. These direct action brakes are spring set, electrically released, and designed for stopping and holding a load. Energy-saving motor designs are available and have reduced full load motor losses through the use of optimum electrical designs and increased active material. They are also designed to operate at low noise levels and are available up to 300 horsepower. Multispeed motors are motors with special electrical characteristics for a wide variety of two-speed applications requiting constant or variable torque. Vertical motors (Figure 1-5) are flanged, footless designs used in direct coupled vertical applications. Vertical motors are available with normal and medium thrust capabilities on many different frames. They are widely applied in pumps and mixers, and sizes can exceed 1000 horsepower. A submersible motor is shown in Figure 1-6. These motors are designed for continuous pumping duty submerged in liquids containing a maximum solid content of 10% by weight and 90% liquid. For use in hazardous environments, these motors are UL listed for use in class I, group D, division I hazardous locations in air or submersible in water or sewage.
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Process Plant Machinery
FIGURE 1-4 Brake motor combining a motor and integrally mounted disc brake. (Source: Reliance Electric, Cleveland, OH.)
FIGURE 1-5 Cross section o f NEMA frame, vertical A C motor. 1 - top-mounted thrust bearing; 2 - steel laminations; 3 - rotor fins; 4 - bearing cap; 5 - lower thrust bearing; 6 - grease reservoir; 7 - grease relief ports; 8 - drip cover; 9 - conduit box; 10 - lifting plates; 11 - cover, with grounding provision. (Source: Reliance Electric, Cleveland, OH.)
Electric Motors and Controls
FIGURE 1-6 Submersible motor for use in hazardous areas. (Source: Reliance Electric, Cleveland, OH.)
MAJOR COMPONENTS OF INDUCTION MOTORS The major components of an AC induction motor are illustrated in Figure 1-7. They consist of the following: 9 9 9 9 9 9 9
flame stator or windings rotor bearings end shield, end bell, or end bracket cooling fan conduit box
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Process Plant Machinery
FIGURE 1-7
What to look for in major components of an AC motor. (Source: Reliance Electric, Cleveland,
OH.)
Frame Cast iron is the most popular frame material. The frame must be made of a rigid material that will absorb noise as well as vibration because it provides the structural support for all the other motor components. The stator or windings are mechanically attached to the frame prior to the rotor installation. Tight machining tolerances are required on the mating surfaces of the frame and end bells or end shields. These tight tolerances ensure accurate bearing locations that allow the rotor to run at the proper air gap. All of this helps to increase motor efficiency.
Stator and Rotor The stator is made up of copper wire that is formed into coils that are insulated. Stator coil construction procedures are sometimes complicated and must be entrusted to capable manufacturers. The rotor is the rotating member of an induction motor that is made up of stacked laminations. Thin, high silicone steel stampings are aligned on a keyed mandrel and welded in the axial direction in three or four places to form a rigid cylinder. This cylinder is removed from the mandrel and pressed on the motor shaft. Slots in each lamination of the cylinder are filled with molten aluminum to form the squirrel cage. The cast aluminum bars of the squirrel cage act as conductors for the induced magnetic fields. Some motors are made with copper bars instead of cast aluminum. The shaft portion of the rotor is precision machined on both ends to fit the bearings. Motors in the 5- to 200-horsepower sizes are usually fitted with antifriction (ball) beatings. Roller or journal bearings are used for very large motors.
Bearings and End Shields Bearings are used to reduce friction and wear while supporting the rotating element (rotor). The bearing acts as the connection point between the rotating (rotor) and stationary (end bell or end shield) elements of a motor. There are various types, such as roller, ball, sleeve (journal), and needle beatings. The ball bearing is used in virtually all types and sizes of electric motors. It exhibits low friction loss, is suited for high-speed operation, and is compatible in a wide range of temperatures. There are various types of ball bearings such as
Electric Motors and Controls
open, single-shielded, or sealed. Although not mandatory, some manufacturers offer a special beating design for oil-mist lubrication. End shields or end bells cover the ends of the motor frame. They protect the internal electrical and mechanical parts from moisture and dirt and provide support for the bearings.
Cooling Fan The cooling fan is a small but important part of an electric motor that is often overlooked. It provides the cooling air across the TEFC motor and is made of a non-sparking, corrosion-resistant material. Most motors come with a bidirectional fan so as not to be sensitive to the direction of motor rotation. Some motors have low noise or high-efficiency fan designs and will pump air in only one direction of rotation. This type of fan generally has a direction of rotation arrow cast on the fan hub.
Conduit Box The conduit box is the metal container on the side of the motor that houses the electrical connections. It must provide a waterproof environment where the stator (winding) leads are connected to the incoming power leads. If oil mist is used to lubricate the motor beatings, a gas-tight seal must be provided around the stator leads at the penetration of the conduit box.
A MOTOR AS PART OF A SYSTEM Selecting a Motor Motor selection is a complicated process containing numerous trade-offs. Efficiency is only one of several important considerations. The objective of informed motor selection is to arrive at the best possible installation consistent with minimum cost, horsepower, and frame size for the specified life expectancy, load torque, load inertia, and duty cycle of the specified application. To satisfy the torque, horsepower, and speed requirements of a large variety of motor applications, polyphase AC motors are designed and manufactured in four groups classified design A, B, C, and D by NEMA. Each classification of motors has its own distinctive speed-torque relationship (Figure 1-8) and inherent expectations regarding motor efficiency. Motors intended for loads that are relatively constant and run for long periods of time are of low slip design (less than 5 percent) and are inherently more efficient than design D motors, which are used in applications where heavy loads are suddenly applied, such as hoists, cranes, and heavy metal presses. Design D motors deliver high starting torque and are designed with high slip (more than 5 percent) so that motor speed can drop when fluctuating loads are encountered. Although design D motor efficiency can be less than other NEMA designs, it is not possible to replace a design D motor with a more efficient design B motor, because it would not meet the performance demands of the load. The motor with the highest operating efficiency does not always provide the lowest energy choice. Figure 1-9 compares the watts loss of a NEMA design D and a design B motor, in a duty cycle that accelerates a load inertia of 27 lb. ft. 2 to full speed and runs at full load for 60 seconds. During acceleration, the lower curve
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Process Plant Machinery
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SPEED (PERCENTSYNCHRONOUSSPEED)
FIGURE 1-8 Speed-torque curves for a 5-HP motor, NEMA design A and D, and full-load efficiencies. (Source: Reliance Electric, Cleveland, OH.) represents the performance of the design D motor, while the upper curve reflects the NEMA design B motor. The shaded area between the curves represents the total energy difference during acceleration. In this example, this area is approximately 6.0 watt-hours, the energy saved accelerating this load with a design D motor instead of a design B. During the run portion of this duty cycle, the energy loss differential favors the NEMA design B, because it has a higher operating efficiency. In this example, the energy saved operating this load with a design B motor instead of a design D motor is approximately 2.8 watt-hours. The bar chart shown in Figure 1-10 summarizes acceleration and running loss/cycle on both the NEMA design B and design D. Comparison of the total combined acceleration and running portions of this duty cycle indicates a total energy savings of 3.2 watt-hours favoring the use of the design D motor, even though the design B motor has an improved operating efficiency. The key is the improved ability of the design D motor to accelerate a load inertia at minimum energy cost. Other Components Affecting
Efficiency
Because a motor buyer selects the most efficient motor of a given size and type does not mean that energy savings are being optimized. Every motor is connected to some form of driven equipment: a crane, a machine tool, a pump, etc., and motors are often connected to their loads through gears, belts, or slip couplings. By examining the total system efficiency, the component that offers the greatest potential improvements can be identified and money allocated to the component offering the greatest payback. In the case of new equipment installations, a careful application analysis, including load and duty cycle requirements, might reveal that a 7 89
Electric Motors and Controls
- - - W a t t s Loss F o r Design B M o t o r 83.4% Eff. - 9- W a t t s Loss F o r Design D M o t o r m 8 0 . 5 % E f t .
800C
700C
F~
=
~1
=
Acceleration Energy Loss/Cycle = 6.0 W-Hr R u n n i n g E n e r g y L o s s / C y c l e = 2.8 W - H r T o t a l E n e r g y L o s s / C y c l e = 3.2 W - H r .
600G
For 10 C y c l e s / H r , 8 5 0 0 H r / Y e a r Operation At $.04/KW-Hr E n e r g y Savings O f $ 1 0 . 9 0 Using Design D Vs. Design B /
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.
.
.
.
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FIGURE 1-9 Energy usage on duty cycle application 5-HP, 4-pole, TEFC accelerating 27 lb-fi 2 inertia. (Source: Reliance Electric, Cleveland, OH.)
pump, for example, could be utilized in place of a l O-horsepower pump, thereby reducing motor horsepower requirements by one third. By reducing the mass of the moving parts, the energy required to accelerate the parts is also proportionately reduced. Or in the instance of an air compressor application, the selection as to the size and type of compressor relative to load and duty cycle will affect system efficiency and energy usage. Of course, the most efficient equipment should be selected whenever possible. Reduced system efficiency and increased energy consumption are also possible with existing motor drive systems due to additional friction that can gradually m
t,~ccEl.E.,~l",o,~! I
13.9 W-HR LOSS/CYCLE ,,
ACCELERATION 13.7 W-HR LOSS/CYCLE
DESIGN D
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11.1 W-HR. LOSS/CYCLE
1 2O Energy Loss/Cycle (Watt-Hrs.)
NEMA DESIGN
B
J
30
FIGURE 1-10 Acceleration and running loss per cycle on NEMA D and NEMA B motors. (Source: Reliance Electric, Cleveland, OH.)
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ProcessPlant Machinery
develop within the driven machine. This additional friction could be caused by a build-up of dust on a fan, the wearing of parts causing misalignment of gears or belts, or insufficient lubrication in the driven machine. All of these conditions cause the driven machine to become less efficient, which causes the motor to work harder. Rather than replace the existing motor with a higher efficiency model, replacing either critical machine components or the machine itself may result in greater system efficiency and energy savings.
Choosing the Best Applications Energy-efficient motors may be the most cost-effective answer for certain applications. Simple guidelines are listed below: 9 Choose applications where motor running time exceeds idle time. 9 Review applications involving larger horsepower motors, where energy usage is greatest and the potential for cost savings can be significant. 9 Select applications where loads are fairy constant, and where load operation is at or near the full load point of the motor for the majority of the time. 9 Consider energy-efficient motors in areas where power costs are high. In some areas, power rates can run as much as $. 12 per kilowatt-hour. In these cases, the use of an energy-efficient motor might be justified in spite of long idle times or reduced load operations. Using these simple guidelines, followed by an analysis and cost justification based on various techniques, can yield results that will influence motor choice beyond just-in cost consideration.
Other Determinants of Operating Cost
Voltage Unbalance Although efficiency is a commonly used indicator of energy usage and operating costs, there are several important factors affecting motor operating costs. Rated performance as well as selection and application considerations of polyphase motors requires a balanced power supply at the motor terminals. Unbalanced voltage affects the motor's current, speed, torques, temperature rise, and efficiency. NEMA Standard MG 1-14.34 recommends derating the motor where the voltage unbalance exceeds 1 percent and recommends against motor operation where voltage unbalance exceeds 5 percent. Voltage unbalance is defined as follows: Maximum Voltage Deviation from Average Voltage Voltage Unbalance (%) = 100 x Average Voltage Voltage imbalance is not directly proportional to the increase in motor losses, as a relatively small unbalance in percent will increase motor losses significantly and decrease motor efficiency as Figure 1-11 shows. An effort to reduce losses with the purchase of premium priced, premium efficiency motors that reduce losses by 20 percent can easily be offset by a voltage unbalance of 3.5 percent that increases motor losses by 20 percent. Energy cost can be minimized in many industrial applications by reducing the additional motor watts loss due to voltage unbalance. Uniform application of singlephase loads can assure proper voltage balance in a plant's electrical distribution system used to supply polyphase motors.
Electric Motors and Controls
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FIGURE 1-11 Motor loss percentage as a function of voltage unbalance. (Source: Reliance Electric, Cleveland, OH.)
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FIGURE 1-12 Power factor and efficiency changes as a function of motor load. (Source: Reliance Electric, Cleveland, OH.)
M o t o r Loading
One of the most common sources of motor watts loss is the result of a motor not being properly matched to its load. In general, for standard NEMA frame motors, motor efficiency reaches its maximum at a point below its full-load rating, as indicated in Figure 1-12. This efficiency peaking below full load is a result of the
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Process Plant Machinery
interaction of the fixed and variable motor losses resulting in meeting the design limits of the NEMA standard motor performance values, specifically locked rotor torque and current limits. Power factor is load variable and increases as the motor is loaded, as Figure 1-12 shows. At increased loads, normally in the region beyond full load, this process reverses as the motor's resistance to reactive ratio begins to decrease and power factor begins to decline. In some applications where motors run for an extended period of time at no load, energy could be saved by shutting down the motor and restarting it at the next load period. Maintenance
Proper care of the motor will prolong its life. A basic motor maintenance program requires periodic inspection and, when encountered, the correction of unsatisfactory conditions. Among the items to be checked during inspection are lubrication, ventilation, and presence of dirt or other contaminants; alignment of motor and load; possible changing load conditions; belts, sheaves, and couplings; and tightness of hold-down bolts.
Total Energy Costs There are three basic components of industrial power cost: cost of real power used, power factor penalties, and demand charges. To understand these three charges and how they are determined, a review of the power vector diagram (Figure 1-13) identifies each component of electrical energy and its corresponding energy charge. Real Power
The real power-kilowatt (kw) is the energy consumed by the load. Real power-kw is measured by a watt-hour meter and is billed at a given rate (S/kw-hr). It is the real power component that performs the useful work and is affected by motor efficiency. Power Factor
Power factor is the ratio of real power-kw to total KVA. Total KVA is the vector sum of the real power and reactive KVAR. Although reactive KVAR performs no actual work, an electric utility must maintain an electrical distribution system (i.e., power transformers, transmission lines, etc.) to accommodate this additional electrical energy. To recoup this cost burden, utilities may pass this cost on to industrial customers in the form of a power factor penalty for power factor below a certain value. REAL POWER - KW
~
RV :AAIVAE ~R'I
DEFINITIONS KW KVA KVAR
FIGURE 1-13
-
POWER FACTOR
Kilowatts -
Kilovolt-
Ampere
Reactive Kilovolt -
Ampere
Electrical power vector diagram. (Source: Reliance Electric, Cleveland, OH.)
Electric Motors and Controls
Power factors in industrial plants are usually low due to the inductive or reactive nature of induction motors, transformers, lighting, and certain other industrial process equipment. Low power factor is costly and requires an electric utility to transmit more total KVA than would be required with an improved power factor. Low power factor also reduces the amount of real power that a plant's electrical distribution system can handle, and increased line currents will increase losses in a plant's distribution system. A method to improve power factor, which is typically expensive, is to use a unity or leading power factor synchronous motor or generator in the power system. A less expensive method is to connect properly sized capacitors to the motor supply line. In most cases, the use of capacitors with induction motors provides lower first cost and reduced maintenance expense. Figure 1-14 graphically shows how the total KVA vector approaches the size of the real power vector as reactive KVAR is reduced by corrective capacitors. Because of power factor correction, less power need be generated and distributed to deliver the same amount of useful energy to the motor. Just as the efficiency of an induction motor may be reduced as its load decreases, the same is true for the power factor, only at a faster rate of decline. A typical 10-horsepower, 1800 RPM, three-phase, design B motor with a fullload power factor of about 80 percent decreases to about 65 percent at half-load. Therefore, it is important not to overmotor. Select the fight size motor for the right job. Figure 1-15 shows that the correction of power factor by the addition of capacitors not only improves the overall power factor but also minimizes the fall-off in power factor with reduced load.
Demand Charges The third energy component affecting cost is demand charge, which is based on the peak or maximum power consumed or demanded by an industrial customer during a specific time interval. Because peak power demands may require an electric utility to increase generating equipment capacity, a penalty is assessed when demand exceeds a certain level. This energy demand is measured by a demand meter, and a multiplier is applied to the real power-kw consumed. Industrial plants with varying load requirements may be able to affect demand charges by (1) load cycling, which entails staggering the starting and use of all electrical equipment and discontinuing use during peak power intervals, and (2) using either electrical or mechanical "soft start" hardware, which limits power inrush and permits a gradual increase in power demand.
Power
_.
tactoq
REAL POWER
Fleal Power--KW Total--KVA
- KW
FIGURE 1-14 Effect of corrective capacitance on total KVA vector. (Source: Reliance Electric, Cleveland, OH.)
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Process Plant Machinery 10 fliP, 1780 RPM TEIrr IIIOTOA 100
gXD
CORRECTEO P.F. ~ W/3 KVAC CAPACITOR . . . . . . . EFF. " - - - UNCORRECTE0 P.F.
%
75 50
%
25
%
1
/
f
,f
/
Corrected P.F.
Eft. Uncorrected P.F.
1/4
2/4
3/4 Full L.ol, ElqF.& P3. vs LOAD
5/4
FIGURE 1-15 Effect of capacitors on fall-off in power factor with reduced load. (Source: Reliance Electric, Cleveland, OH.)
ADJUSTABLE FREQUENCY MOTOR CONSIDERATIONS Speed control by way of adjusting power frequency is becoming more and more important for economical throughput or pressure capacity variation of modem process machinery. Several key parameters that must be considered when applying induction motors to adjustable frequency controllers include the load torque requirements, current requirements of the motor and the controller current rating, the effect of the controller wave-shape on the motor temperature rise, and the required speed range for the application. In order to properly size a controller for a given application, it is necessary to define the starting torque requirements, the peak torque requirements, and the full-load torque requirements. These basic application factors require reexamination because the speed-torque characteristics of an induction motor/controller combination are different from the speed-torque characteristics of an induction motor operated on sine-wave power. The motor current requirements should be defined for various load points at various speeds in order to ensure that the controller can provide the current required to drive the load. The current requirements are related to the torque requirements, but there are also additional considerations due to the harmonics of adjustable frequency control power that must be taken into account. Temperature rise and speed range must be considered when applying induction motors to adjustable frequency controllers because this nonsinusoidal power results in additional motor losses, which increase temperature rise and reduce motor insulation life. Before discussing the speed-torque characteristics of a motor/controller combination, it is useful to review the speed-torque characteristics of an induction motor started at full voltage and operated on utility power (Figure 1-16). Here we see the speed-torque curve for a 100-horsepower, 1800 RPM, high-efficiency motor. When this motor is started across the line, the motor develops approximately 150 percent of full-load torque for starting and then accelerates along the speed-torque curve through the pull-up torque point, through the breakdown torque point, and, finally, operates at the full-load torque point, which is determined by the intersection of the load line and the motor speed-torque curve. In this case, we have shown an application such as a conveyor where the load-torque requirement is constant from 0 RPM to approximately 1800 RPM. The difference between the motor speed-torque curve and the load line is the accelerating torque and is indicated by the cross-hatched area.
Electric Motors and Controls 100 HP, XE Motor STARTING TORQUE 30O
:3 E
200
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100
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PULL-UP TORQUE
ACCELERATING TORQUE
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200
600
1000
1400
BREAKDOWN TORQUE
=ULL-LOAD 9TORQUE .LOAD TORQUE REQUIREMENT
1800
RPM
FIGURE 1-16 Speed-torque characteristics of induction motors started at full voltage. (Source: Reliance Electric, Cleveland, OH.)
If the load-torque requirement ever exceeded the maximum torque capability of the induction motor, the motor would not have enough torque to accelerate the load and would stall. For instance, if the load line required more torque than the motor could produce at the pull-up torque point, i.e., 170 percent load torque versus 140 percent pull-up torque, the motor would not increase in speed past the pullup torque speed and would not able to accelerate the load. This would cause the motor to overheat. It is, therefore, important to ensure that the motor has adequate accelerating torque to reach full speed. Normally, the motor accelerates the load and operates at the point of intersection of the load line and the motor speed-torque curve. The motor then always operates between the breakdown torque point and the synchronous speed point that corresponds to the 1800 RPM location on the horizontal axis. If additional load torque is required, the motor slows down and develops more torque by moving up toward the breakdown torque point. Conversely, if less torque is required, the motor would speed up slightly toward the 1800 RPM point. Again, if the breakdown torque requirements were exceeded, the motor would stall. Figure 1-17 depicts the same motor speed-torque curve, but now the motor current has been shown for full voltage starting. Typically, when a NEMA design B induction motor is started across the line, an inrush current of 600 percent to 700 percent occurs corresponding to the starting torque point. As the load is accelerated to the full-load torque point, the current decreases to 100 percent full-load current at 100 percent full-load torque. High currents, however, are drawn during the acceleration time. The amount of time that the motor takes to accelerate the load will depend on the average available accelerating torque, which is the difference between the motor speed-torque curve and the load speed-torque curve, and the load inertia. Figure 1-18 illustrates a blown-up view of the region between the breakdown torque point and the synchronous speed point, which is where the motor would
21
22
Process Plant Machinery
100 HP, XE Full-Voltage Starting 700
500
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100
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FULL-LOAD TORQUE
200 RPM
FIGURE 1-17 Motor current of induction motor started at full voltage. (Source: Reliance Electric, Cleveland, OH.)
100 HP, XE 400
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1720 RPM
1760
1800
FIGURE 1-18 Motor current and torque as full operating speed is approached. (Source: Reliance Electric, Cleveland, OH.) operate. This is of particular interest because the current for various torque requirements can easily be seen. This would directly affect the size of the controller required to produce a given torque because controllers are current-rated. At 100 percent full-load torque, 100 percent full-load nameplate current is required. At 150 percent torque, 150 percent full-load nameplate current is required. Beyond the 150 percent full-load torque point, however, the torque-per-amp ratio is
Electric Motors and Controls
no longer proportional. For this case, 251 percent breakdown torque would require 330 percent current. Adjustable frequency controllers are typically rated for a maximum of 100 percent continous or 150 percent for one minute of the controller full-load current. This would generally provide a maximum of 100 percent or 150 percent of motor full-load torque. This would not, however, provide the same amount of torque as the motor could potentially develop if it were operated from utility power, which could normally provide as much current as the motor required. It would generally be uneconomical to oversize a controller to obtain the same amount of current (torque), since the controller size would actually triple for this example in order to provide 251 percent torque. Two basic concepts that can explain adjustable speed operation of induction motors can be summarized as follows: Speed oc
Frequency Poles
Torque cx Magnetic Flux oc
Volts Hertz
The speed of an induction motor is directly proportional to the applied frequency divided by the number of poles. The number of poles is a function of how the motor is wound. For example, for 60 Hertz power, a two-pole motor would operate at 3600 RPM, a four-pole motor at 1800 RPM, and a six-pole motor at 1200 RPM. The torque developed by the motor is directly proportional to the magnetic flux or magnetic field strength, which is proportional to the applied voltage divided by the applied frequency or Hertz. Thus, in order to change speed, all that must be done is to change the frequency applied to the motor. If the voltage is varied along with the frequency, the available torque would remain constant. It is necessary to vary the voltage with the frequency in order to avoid saturation of the motor, which would result in excessive currents at lower frequencies, and to avoid underexcitation of the motor, which would result in excessive currents, both of which would cause excessive motor heating. In order to vary the speed of an induction motor, an adjustable frequency controller would have an output characteristic as shown in Figure 1-19. The voltage is varied directly with the frequency. For instance, a 460-volt controller would normally be adjusted to provide 460 volts output at 60 Hertz and 230 volts at 30 Hertz. A controller would typically start an induction motor by starting at low voltage and low frequency and increasing the voltage and frequency to the desired operating point. This would contrast with the conventional way of starting induction motors of applying full voltage, 460 volts at 60 Hertz, immediately to the motor. By starting the motor with low voltage and low frequency, the inrush current associated with across-the-line starting is completely eliminated. This results in a soft start for the motor. In addition, the motor operates between the breakdown torque point and synchronous speed point as soon as it is started, as compared with starting across the line, in which case the motor accelerates to a point between the synchronous speed and breakdown torque point.
Summary The maximum torque for an induction motor is limited by the adjustable frequency controller current rating. In order to determine the maximum torque that would
23
24
Process Plant Machinery
400-
o
2~
115
0 0
I lS
|n
I 3o
I 45
Ili0
-
FREQUENCY (Hz)
FIGURE 1-19 Controller output voltage versus frequency relationship for adjustable speed reduction motors. (Source: Reliance Electric, Cleveland, OH.)
be available from an induction motor, it would be necessary to define the motor torque at the controller's maximum current rating. 9 The starting torque equals the maximum torque for a motor/controller combination. 9 The starting torque current is substantially less for an adjustable frequency controller/motor combination than the locked rotor current for an induction motor started across the line. This results in a soft start for the controller/motor combination. 9 The motor load inertia capability for a controller/motor is much higher, since the controller can limit the motor current to 100 percent or less. This would result, however, in longer acceleration times than starting the motor across the line. Harmonics cause additional motor temperature rise over the temperature rise that occurs for sine-wave power operation. As a rule of thumb, for every 10~ rise in temperature, the motor insulation life is cut in half. This explains why it is important to consider the additional temperature rises associated with adjustable frequency control power and to follow the suggested rating curves provided by capable motor manufacturers. 9 NEMA design C and D motors are not recommended for use on adjustable frequency control power because these motors have high watts loss due to higher rotor watts loss over design B motors and resulting high temperature rises when operated on adjustable frequency control power. 9 Key application points must be defined in order to properly apply an induction motor to a solid-state adjustable frequency controller torque, speed range, motor description, and environment. In order to ensure that adequate torque is available to drive the load and adequate current is available to produce the required torque, the starting torque, the peak running torque, and the continuous torque requirements must be defined. The continous torque is usually defined, but the peak and starting torques are more difficult to define. For the case of retrofit applications, the speed-torque curve of the existing motor might be used as a reference to define the starting and peak-load torque. Sizing the controller for these points, however, would frequently result in a larger controller than necessary. 9 The speed range affects the motor thermal rating. The controllers will typically provide a 10 to 1 speed range below 60 Hertz.
Electric Motors and Controls
9 The motor description will permit selection of a controller size for the motor horse-power, voltage, and current rating. The motor insulation class and design type will permit the motor to be rated properly to ensure that its thermal limitations are not exceeded. 9 It is necessary to consider the environment to choose the proper motor enclosure. Explosion-proof motors usually have a UL label certifying that they are suitable for the defined classified area. The UL label, however, is suitable only for 60 Hertz sine-wave power. When a explosion-proof motor is operated on adjustable frequency control power, the 60 Hertz sine-wave UL label is voided. In addition, induction motors are normally rated for 40 ~ (104 ~ ambient temperature. Use in a higher ambient temperature may require additional cooling or overframing.
25
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Chapter 2 Gear Speed Transmission Equipment* In numerous types of process machinery, gears are employed to transmit motion and power from one revolving shaft to another or from a revolving shaft to a reciprocating element. When used between revolving shafts, the shafts can take one of only three positions: they may be parallel, may intersect at an angle, or may cross without intersecting. If the shafts are parallel, the basic friction wheels and the gears developed from them assume the shape of cylinders (Figures 2-1 and 22). When the shafts are not parallel, the shapes of the friction wheels and gears will be different. For example, on intersecting shafts the wheels become cones and the gears developed on these conical surfaces are called bevel gears (Figure 2-3). Where motion is transmitted from a shaft to a reciprocating element, a cylindrical friction wheel may engage a flat surface, and the gears assume a similar form (Figure 2-4). All of these methods of transmitting power between cylinders, cones, and flat surfaces involve rolling motion. When shafts cross (one above the other), the friction wheels may be cylindrical or may be of hyperbolic cross section. The gears developed on these surfaces are helical (Figure 2-5) or hypoid (Figure 2-6). In neither case will there be pure rolling action, because when shafts are crossed there is unavoidable side-slip between the surfaces of the wheels. It is important to appreciate that the type of contact that occurs between the surfaces of friction wheels will also be the type of contact that will occur between the meshing teeth of the corresponding gears. Thus the contact between two cylinders on parallel shafts takes place along a line, and the contact between the teeth developed on those surfaces (Figure 2-2) also occurs along a line. Likewise, the contact between two cones and the teeth developed on those cones (Figure 2-3) will occur along lines. The same condition exists on hypoid gears. However, the contact between friction surfaces is not always a line. Where cylindrical surfaces are on crossed, nonintersecting shafts (Figure 2-5), contact occurs at a point instead of along a line. When this is the case, the teeth developed on these surfaces also make point contact.
TYPES OF GEARS Spur Gears When the shafts are parallel, the teeth of the meshing gears may be cut straight across the faces of the gear blanks (Figure 2-1). Gears of this kind are called spur gears. There are many special kinds of spur gears, some of which are not commonly * Source: Prager Incorporated, New Orleans, LA, unless otherwise noted. Adapted by permission.
27
28
Process Plant Machinery
FIGURE 2-1
Typical external spur gear set. (Source: Prager, Inc., New Orleans, LA.)
LINE CONTACT
FIGURE 2-2 Internal spur gears. The pitch surfaces on friction wheels contact along a line (upper view) to transmit motion between parallel shafts. The teeth are developed on internal and external cylindrical pitch surfaces as shown in the lower view. (Source:Prager, Inc., New Orleans, LA.)
Gear Speed Transmission Equipment
LINE CONTACT
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29
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/
FIGURE 2-3 Bevel gears. These shafts intersect at a right angle, although bevel gears may also be used between shafts that intersect at larger or smaller angles. (Source: Prager, Inc., New Orleans, LA.)
FIGURE 2-4
Rack-and-pinion gear set. (Source: Prager, Inc., New Orleans, LA.)
encountered. Although none of these special forms differ materially in tooth action from the usual spur gears, it may be beneficial to mention some of them that are sometimes found in special applications. A spur gear meshing with a straight element (Figure 2-4) is known as a rack and pinion. The rack may be visualized as a short section of an infinitely large spur gear; a gear so large in diameter that the teeth lie on a straight line. Elliptical gears (Figure 2-6) are used to convert the uniform rotary motion of a driving shaft to a rhythmic, pulsating rotation of the driven shaft. An equalizer gear and eccentric pinion (Figure 2-7) are sometimes used to drive large chain conveyors in order to prevent the changes of conveyor speed that would occur when the long bar links
30
Process Plant Machinery
i
POIN
FIGURE 2-5
Principle of helical gearing. (Source: Prager, Inc., New Orleans, LA.)
FIGURE 2-6 Elliptical gears as occasionally used in special machinery. (Source: Prager, Inc., New Orleans, LA.) pass around the sprockets at the driving end. The eccentric driving pinion revolves at a constant speed, but it imparts an irregular motion to the equalizer gear that is calculated to provide a smooth, unvarying speed to the chains. This type of drive reduces what otherwise might be excessive shock on the chains. Another special type of spur gear is sometimes encountered in what are known as stop motion or intermittent gears (Figure 2-8). In this modification, the driving gear rotates continuously but actuates the driven gear only when the teeth of both gears are engaged. At other times, the driven gear is locked in a fixed position. Ordinary gears are sometimes used in special groupings; for example, the epicyclic or planetary arrangement of gears in Figure 2-9 makes possible a very compact reduction gear set. Such a unit usually consists of a central pinion, several planetary gears mounted on a spider, and an internal gear (ring gear) encircling the entire unit. One of the three elements - pinion, spider, or ring g e a r - must be
Gear Speed Transmission Equipment 31
FIGURE 2-7
Equalizer gear with eccentric pinion. (Source: Prager, Inc., New Orleans, LA.)
FIGURE 2-8
Stop-motion or intermittent gearing. (Source: Prager, Inc., New Orleans, LA.)
ER
:
ST~ RI
FIGURE 2-9
'.-OFF ~.FT ~1
Epicyclic, or planetary gearing. (Source: Prager, Inc., New Orleans, LA.)
32
Process Plant Machinery
held stationary, the other two then being used as power input and output elements. Usually it is the ring gear that is rigidly fixed in the housing of the gear set. The pinion is usually the driving element with the spider the power take off, that is, the driven element. In this arrangement, the driving and driven shafts are in line with each other and rotate in the same direction but at different speeds. Epicyclic gearing is often the ideal solution for transmitting high horsepower at high speeds where a compact, in-line, and lightweight drive unit is required. These epicyclic drives are available in both planetary and star planetary configurations. The free-floating sun design results in balanced, equal load sharing to maximize reliability and life. Whether mounted integrally with a turbine or free standing, epicyclic drives have a low inertia for quick start-up in standby power generation systems. Installation versatility makes these drives convenient for use in fixed, mobile, and marine applications. Capable of handling high ratios, epicyclic drives have an operating efficiency that may reach up to 99 percent. Epicyclic gears are available in different variations and configurations. Figure 2-10 shows the longitudinal section of a planetary gear with rotating planet carder. The sun pinion and the planet carder with the three planets rotate in the bearings. The outer ring with internal teeth (annulus) is fixed firmly to the casing. The permissible speed of the slow-running shaft is limited principally by the centrifugal force of the planets and the resulting beating pressure on the bearing pins. The sense of rotation of input and output shafts is the same. Planetary gears with rotating annuli are shown in Figure 2-11. Here, the stationary planet carder with the three planets is fixed to the gear casing, and the internal tooth annulus connected with the low-speed shaft rotates. Input and output shafts rotate in opposite directions. It should be noted that Figures 2-10 and 2-11 show the gearing with helical teeth, whereas Figure 2-9 illustrates spur gears incorporated in the epicyclic package. The reason for frequently using helical gears is perhaps best explained by reviewing their precursor, stepped gears.
FIGURE 2-10 Longitudinal section of planetary gear with rotating planet carrier. (Source: Maag Gear Company, Zurich, Switzerland.)
Gear Speed Transmission Equipment
FIGURE 2-11 Planetary speed-reducing gear with rotating planet carrier. (Source: Maag Gear Company, Zurich, Switzerland.)
Since only a few teeth of a pair of spur gears are in contact at the same time, meshing of these gears may be accompanied by a slight impact as the load shifts from tooth to tooth. This is because the teeth undergo slight deformation as the load moves over the tooth surface. At low speeds, this is not a serious factor, but as speeds become higher and higher, this deformation makes it more difficult to mesh spur gear teeth without noise and shock. Although accurate machining of spur gear teeth is a major factor in smoothing out any slight irregularities of transmitted motion and torque, these irregularities are minimized only when the gears are designed to distribute the load on several engaging teeth. An early method of accomplishing this result was by the use of stepped gears. In this construction, an assembly of two or more narrow gears are mounted on the shaft in such a way that the teeth are staggered (Figure 2-12). Sudden transfer of load from one tooth to another is minimized as each adjacent stagger tooth absorbs part of the load before the preceding tooth leaves the mesh. This increases the number of teeth in contact at any one time. In rare cases, stepped gears are used to permit operation with zero backlash. When used for this purpose, one of the gears is an ordinary spur gear. The meshing gear consists of two narrow gears bolted together in such a position that all play between the meshing gear teeth is eliminated. In general, the stepped type of gear is seldom used for ordinary transmission of power because of the difficulty of equalizing the load among the various tooth faces. For this use, however, the modem development of the stepped g e a r - the helical g e a r - is widely employed.
Helical Gears, Parallel Shafts If, instead of two or three narrow gears, a stepped gear were composed of innumerable staggered laminations- each lamination so thin that it no longer appeared as an individual unit - the result would be a gear with smoothly twisted teeth. In actual practice, twisted-tooth gears are machined from solid gear blanks with the twist in the same direction. Such a uniform twist is a true helix, and the resulting gears are
33
34
Process Plant Machinery
FIGURE 2-12 LA.)
Stepped gearing predates helical gears. (Source: Prager, Inc., New Orleans,
FIGURE 2-13
Helical gears. (Source: Prager, Inc., New Orleans, LA.)
called helical gears (Figure 2-13). The angle of twist (helix angle) may range from about 20 degrees to 45 degrees. The helix angle is selected so that several teeth will be in mesh at the same time. Even if only two of these helical teeth are in mesh, a very smooth transfer of power results. As the helix angle is increased, the number of teeth in simultaneous contact and the smoothness of tooth engagement are correspondingly increased. Single helical gears are used in speed reducers as well as speed increasers. The speed reducer shown in Figure 2-14 connects a steam turbine to a reciprocating compressor. It accomplishes the speed reduction in two steps, hence the designation double reduction gear. Another single helical gear is shown in Figure 2-15. Used as a speed increaser between a gas turbine operating at 4860 RPM and a compressor operating at 12507 RPM, this speed increaser transmits 23,800 horsepower and has a pitch-line velocity of 540 feet per second (fps).
Gear Speed Transmission Equipment
35
FIGURE 2-14 Double reduction, single helical gear speed reducer for reciprocating compressor drive. (Source: Maag Gear Company, Zurich Switzerland.)
FIGURE 2-15 Single reduction, single helical gear increaser for centrifugal compressor drive. (Source: Maag Gear Company, Zurich, Switzerland.) Due to the angularity of their teeth, the operation of helical gears produces axial thrusts that must be absorbed by thrust bearings. In most cases, properly selected rolling element bearings will take care of this thrust. However, by using two pairs of opposed helical gears (Figure 2-16), the thrust of one set of gears balances that of the other. This practice was developed before adequate thrust bearings were available.
36
Process Plant Machinery
FIGURE 2-16 Two single helical gears mounted mirror-style to equalize axial thrust generation. (Source: Prager, Inc., New Orleans, LA.)
FIGURE 2-17 LA.)
Double-helical or herringbone gear (Source: Prager, Inc., New Orleans,
Later developments in the economical cutting of gear teeth made it possible to machine two opposed helical gears on a single gear blank. Such a gear is commonly known as a double-helical or herringbone gear (Figure 2-17). High-speed herringbone gears often have a continuous groove machined between the sets of teeth to assist the escape of oil as the gears pass through mesh. Alternatively, the two sets of teeth may be staggered for the same purpose. Either double- or single-helical gear units are widely used in demanding applications where reliability and low maintenance are a must. For example, in the hydrocarbon processing industry, they are specified for process compressors, pump pipeline compressors, fan drives, generator drives, and blower drives. Figure 2-18 illustrates a typically demanding application.
Gear Speed Transmission Equipment
37
FIGURE 2-18 Single helical speed increaser set driving high-speed compressor wheels. (Source: Bayerische Huettenwerke Sonthofen, Germany.)
The centrally located input gear drives two pinions that are fitted with compressor impellers at each end.
Helical Gears, Nonparallel Shafts When shafts are not parallel but cross one another, the provision of slanted helical teeth will allow a limited amount of power to be transmitted irrespective of the angle between the shafts. The gears are true helical gears (Figure 2-19) but are sometimes called spiral gears.
Worm Gears Where the driving gear of a helical right angle drive is much smaller in diameter than the driven gear, the combination could be called a nonthroated worm gear set (Figure 2-20), the smaller gear being the worm. Worm length, as compared with
FIGURE 2-19
Spiral gearing used in certain right-angle drives. (Source: Prager, Inc., New Orleans, LA.)
38
Process Plant Machinery
@
(" 0
FIGURE 2-20 Nonthroated worm gear set. The teeth are straight and do not envelop the worm. (Source: Prager, Inc., New Orleans, 1_.4.) diameter, permits the helical teeth to encircle the shaft more than once, thus giving the teeth the appearance of threads and giving the worm the appearance of a screw. When the worm has only one thread (tooth), it is commonly called a single-thread worm. If there is more than one thread, it is known as a double-thread, triple-thread, etc., worm. The relative number of teeth on the worm and wheel determines the ratio of speed reduction. When the teeth of both worm and worm gear are of true helical form, the contacts concentrate on a series of points. This limits the power that can be transmitted by such gears. Although the gears transmit motion very smoothly, excessive wear occurs if much power is involved. For this reason, nonthroated worm gears and helical gears on crossed shafts are not very extensively used. Since commercial worm gear sets must transmit considerable power, it is usual to machine the worm gear so that a considerably increased area of tooth surface will make contact (Figure 2-21). This is done by changing the shape of the teeth of the driven gear so that these teeth partly encircle the worm. Such a type is called a throated, or single-enveloping, worm gear and is the type most commonly used in worm gear sets. Another type of worm gear is the double-throated, or double-enveloping, gear set, employing a throated worm and a throated gar (Figure 2-22). Not only does the gear partly envelop the worm, but the worm also partly envelops the gear, thus further increasing the area of the contacting surfaces. When properly designed and manufactured, these gears are able to carry very heavy loads, although not generally at high speeds.
@ FIGURE 2-21 Single-throated worm and gear. This is the type of gear used most commonly in industrial worm gear sets. The teeth of the gear are throated (curved) to partially envelop the worm. (Source: Prager, Inc., New Orleans, 1..4.)
Gear Speed Transmission Equipment
39
FIGURE 2-22 Double-throated worm and gear. Both the worm and gear are throated. As a result, they partially envelop each other. The increased contact area permits heavy loads to be carried. (Source: Prager, Inc., New Orleans, LA.)
DRIVING SHAFT
DRIVING PINION PLANETARY ,~, e-P, .'~IL||/'~hS
RING GEAR
DIFFERENTIAL CASE
TAKE-I ~,FT
SHAFT
SHAFT
SI DE GEAR PLANETARY IDLER PINION
FIGURE 2-23 Differential gear set. Turning of the idler pinions causes an increase in the speed of the less heavily loaded shaft. (Source: Prager, Inc., New Orleans, LA.)
Bevel Gears When shafts intersect, the teeth of meshing gears may be cut straight across the faces of conical gear blanks. Such gears are called bevel gears (Figure 2-3). Bevel gears are widely used where a fight angle change in direction of shafting is required, although occasionally the shafts may intersect at acute or obtuse angles. When of equal size and mounted on shafts at fight angles, they are sometimes referred to as miter gears. Usually, however, the driving gear is smaller than the driven gear, because in the majority of cases gear sets are employed to obtain a reduction of operating speed. Bevel gears may be assembled in a special grouping known as a differential gear set (Figure 2-23) such as is used in automotive vehicles. This arrangement of gears is intended to divide power between two variable speed shafts, e.g., to permit the wheels of motor vehicles to rotate at different speeds when the vehicle is turning comers. Observing the left side of Figure 2-23, we note a large ring gear that is rigidly attached to the differential gear case. The planetary idler pinions, meshing
40
Process Plant Machinery
with the side gears (fight view) are pivoted on and revolve with the case, thereby turning the takeoff shafts. When the load is equal on the takeoff shafts, the entire unit revolves as a solid block. Unbalanced load, however, causes the more heavily loaded shaft and its side gear to slow down. Since the ring gear and case are driven at a constant speed, the planetary idler pinions are forced to turn on their pins as they revolve around the slower side gear. This turning of the idler pinions causes an increase in the speed of the less heavily loaded shaft. Bevel gears may also be arranged to form a strong and compact planetary reduction gear set (Figure 2-24) similar in principle to that shown in Figure 2-9. The drive pinion engages several idler (planetary) gears, causing them to rotate. These gears then roll around the stationary gear, dragging the spider with them and causing the takeoff shaft to rotate. Speed reduction depends on the ratio of the number of teeth on the drive pinion to the number of teeth on the stationary gear.
Spiral Bevel Gears In the same way that the teeth of a spur gear can be twisted to make a helical gear, the teeth of an ordinary bevel gear can be twisted to form a spiral bevel gear (Figure 2-25). Because the teeth of a bevel gear are developed on the surface of a cone, these twisted teeth will take the form of a spiral; thus the gears are called spiral bevel gears. The angle of the spiral is selected so that one end of each tooth enters mesh before the other end of the preceding tooth has disengaged. As with helical gears, this results in very smooth transfer of power.
Hypoid Gears Hypoid gears (Figure 2-26) are used where shafts cross, one below the other, and the design of the machine precludes the use of a worm and gear. This may result where space limitations require that one of the shafts be moved aside, where several pinions on a single shaft drive several cross shafts, where a small pinion must transmit high power, or where rigidity requires a supporting beating on each side of each gear. Where any of these conditions exist, hypoid gears provide a strong, smooth, and quiet drive. The shafts of practically all hypoid gear sets cross at right angles. Although the ordinary hypoid gear is similar in appearance to a spiral bevel gear, it is not developed on the same type of pitch surfaces. Two conical pitch SPIDER DRIVE
TAKE-OFF SHAFT
IDLER (PLANETARY) GEARS
STATIONARY GEAR
FIGURE 2-24 Bevel gears arranged to form a strong and compact planetary reduction gear set. (Source: Prager, Inc., New Orleans, LA.)
Gear Speed Transmission Equipment
FIGURE 2-25
41
Spiral bevel gears. (Source: Prager, Inc., New Orleans, LA.)
LINE
CONTACT
FIGURE 2-26 Hypoid gear and pinion. These gears transmit motion between nonintersecting shafts crossing at a right angle. The pitch surfaces are hyperbolic in form. (Source: Prager, Inc., New Orleans, LA.) surfaces on intersecting shafts roll on each other with line contact and without sideslip, but if the shafts do not intersect, i.e., if they cross one below the other, the cones do not make contact along a line and do not roll without sliding. Instead, they meet at a point and roll with more or less side-slip, depending on the positions of the shafts. Gear teeth developed on such surfaces on crossed shafts would also make point contact and would give poor service. To increase contact area and improve gear service, it is necessary to replace the cones with surfaces that will bear on each other along a line of contact. This line contact is obtained by employing curved pitch surfaces of hyperbolic contour. The teeth developed on these hyperbolic pitch surfaces also meet in line contact, thus distributing the load over considerable tooth surface. Unit loading on the metal is reduced, and the ability to transmit power is increased. The working surfaces of the meshing teeth, however, are always subject to side-slip and consequent friction. Such gears are called hypoid gears. Although most hypoid gears look like spiral bevel gears, this is not always the case. In rare cases, extreme offset of the shafts and high ratios of reduction will require gears such as shown in Figure 2-27. In this case, the pinion has only one tooth, which causes it to resemble a worm.
42
Process Plant Machinery
Special Gear Sets for Process Plant Applications Process plant gear sets often incorporate one or more of the gear types, configurations, and arrangements that we have discussed here. Figures 2-28 through 2-31 show high-power-density parallel shaft double-reduction gears (Figure 2-28), special or customized process machine drives (Figure 2-29), cooling tower fan drives with double-reduction and right-angle output (Figure 2-30), and completely packaged drive systems that include driver, speed change, driven machine, and support systems such as lube oil supplies (Figure 2-31).
FIGURE 2-27 Hypoid gear with extreme offset will require pinion to resemble a worm. (Source: Prager, Inc., New Orleans, LA.)
FIGURE 2-28 This double-reduction helical gear set employs the gear geometries shown earlier in Figures 2-16 and 2-17. (Source: WesTech Gear corporation, Los Angeles, CA.)
FIGURE 2-29 Special process machine drives can be built in almost limitless variations of geometry and arrangements. (Source: WesTech Gear Corporation, Los Angeles, CA.)
Gear Speed Transmission Equipment
FIGURE 2-30 Right-angle drives such as this large cooling tower fan gear often use a combination of gear types to accomplish both speed change and changes in drive direction. (Source: WesTech Gear Corporation, Los Angeles,
CA.)
43
FIGURE 2-31 Complete process systems often include speed-up gears for motor-driven blowers and high-speed pumps. This entire unit is being applied in a solvents recovery process. (Source: WesTech Gear Corporation, Los Angeles, CA.)
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Chapter 3 Gas Turbines In process plants, gas turbines are used in mechanical drive or power generation applications. They are more efficient than steam turbines and their simple cycle thermal efficiency can be greatly increased by waste heat recovery (combined cycle) and other cycle adaptions discussed later in this chapter. In the gas turbine field, information flow starts with the highest budgets and most demanding technological requirements in research and development and moves towards industry. Which means that technology flow goes from NASA to aircraft engines to land-based gas turbines. The 1950s provided industry with a large shot in the gas turbine's technological arm in terms of efficiency, good design, operational features, environmental correctness and every other commercial facet. Nowadays, a number of industrial engines made by manufacturers who also make aircraft engines, such as Rolls-Royce and General Electric, are essentially similar to their aircraft engine sibling but are mounted on a skid. Support systems however, both operational and condition monitoring, will not have the same complexity and degrees of redundancy as the aircraft engine counterpart. The factors which can influence selection of the gas turbine versus the steam turbine in power generation are as varied as the potential users for them. High efficiency is often quoted as the reason for selection of the gas turbine versus the steam turbine. While this is true, in most of the high-growth areas of the world, the choice is made mainly on the basis of readily available, inexpensive fuel. Good illustrations of this fact are China with its high-sulphur coal and Thailand with its lignite. Both fuels occur naturally, are used in steam turbine plants and have been responsible for bad pollution scenarios that both countries are playing "catch up" with. Both these countries' interest in the gas turbine in power generation service is as a peaking unit, as gaseous fuel is an expensive commodity for countries where it doesn't occur naturally. In SE Asia, Japan is the only country that pays the premium for buy liquid natural gas (LNG). They do this in order to maintain the toughest environmental standards in the region, is Japan. The gas turbine and systems associated with it are often more complex in terms of operational systems and repairability than the steam turbine. In many instances, the main reasons for selection of the gas turbine is its versatility with respect to size, at the small horsepower end. In oil and gas production for instance, varying mixed field volumes require that gas compression occur in smaller packages that can be started or shut down with little or no lead time, as opposed to one larger unit. There is also the question of fuel availability. Gas pipeline transport compressors are driven by turbines that run on the fuel they contain. Pumps that transport water away from mixed field separation are generally driven by gas turbines, although in some instances electric motor drivers are used if power is cheap enough. In these and many other applications, mature and reliable gas turbine models are available. Also in power generation, new technology has put some of the latest 45
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ProcessPlant Machinery models of gas turbines in a size category that was otherwise the exclusive domain of steam turbines. Consider also that coal gasification for running gas turbines has very recently reached commercial standards in some applications in the US and Europe. With major current improvements in gas turbine system optimization, metallurgical and design development and component repairability, there will be no stopping the gas turbine when this technology gains a foothold globally. Once the choice is made for a gas turbine, factors that may influence the particular unit selected, include its track record on reliability and availability. Cost per fired hour, spare parts usage and parts repairability are governed by design features. In the case of remotely located or operated plants, such as high Arctic or unmanned stations, reliability and availability feature highly enough that users may favor oversized machinery (by as much as 100%) versus a newer model that fits the kilowatt requirement. Efficiency and fuel flexibility are also major factors. Environmental factors, such as NOx emissions ratings for burner designs, lower temperature designs (that affect NOx production rates), water injection capabilities, ability to burn lower grade fuels (such as residual fuel), play an increasingly large part in selection. This is a large enough topic that it will be covered in depth in a separate book: Environmental Engineering and Management: Sustainable Development in the Power Generation, Oil & Gas and Process Industries (ButterworthHeinemann, 1998). This chapter however, does include a detailed look at one low NOx burner design in commercial operation. Despite all these technical considerations, by far the biggest factor in gas turbine sales in fast developing countries is the financing or monetary incentives that the OEM (original equipment manufacturer) can provide. In one instance, a customer turned down a model he preferred technically, in favor of another where the OEM paid him the difference in fuel efficiency (a sizable 4%) over the estimated life of the engines - which essentially meant no capital outlay with initial purchase of the units.
DESIGN Basically, the gas turbine consists of a compressor that draws in atmospheric air and compresses it to a certain power ratio; a combustion section at the start of which fuel is let in and ignited; and a turbine section through which the combustion gases expand. In power generation applications, the power turbine drives a generator. In mechanical drive applications, the power turbine powers the item being driven (generally a compressor or pump). Sometimes a gearbox is used to alter the speed of the power turbine. A discussion of gas turbine cycles forms one of the appendices of this chapter. The major components and processes of the gas turbine, as well as guidelines in selecting one, will now be discussed.*
The Compressor In the gas turbine engine, compression of the air before expansion through the turbine is effected by one of two basic types of compressor, one giving centrifugal flow and the other axial flow. Both types are driven by the engine turbine and are usually coupled direct to the turbine shaft. * Source: Rolls-RoyceLimited, UK. Adapted with permission.
Gas Turbines The centrifugal flow compressor (Figure 3-1) is a single-or two-stage unit employing an impeller to accelerate the air and a diffuser to produce the required pressure rise. The axial flow compressor is a multistage unit employing alternate rows of rotating (rotor) blades and stationary (stator) vanes, to accelerate and diffuse the air until the required pressure rise is obtained. In some cases, particularly on small engines, an axial compressor is used to boost the inlet pressure to the centrifugal. With regard to the advantages and disadvantages of the two types, the centrifugal compressor is usually more robust than the axial compressor and is also easier to develop and manufacture. The axial compressor, however, ingests far more air than a centrifugal compressor of the same frontal area and can be designed to attain much higher pressure ratios. Since the airflow is an important factor in determining the amount of thrust, this means the axial compressor engine will also give more thrust for the same frontal area. This, plus the ability to increase the pressure ratio by addition of extra stages, has led to the adoption of axial compressors in most engine designs. However, the centrifugal compressor is still favored for smaller engines where its simplicity and ruggedness outweigh any other disadvantages.
FIGURE 3-1
A typical centrifugal flow compressor. (Source: Rolls-Royce, UK)
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ProcessPlant Machinery Z
O_2.00 ~ :03D150
~
z
\
o 1.00 o D tt
0.50
.
.
.
" ~ ~ .
.
.
----
o tt tll
0
03
FIGURE 3-2
10
20
PRESSURE RATIO
30
40
Specific fuel consumption and pressure ratio. (Source: Rolls-Royce, UK)
The trend to high pressure ratios which has favored the adoption of axial compressors is because of the improved efficiency that results, which in turn leads to improved specific fuel consumption for a given thrust (Figure 3-2). Combustion Chambers
The combustion chamber (Figure 3-3) has the difficult task of buming large quantities of fuel, supplied through the fuel spray nozzles, with extensive volumes of air, supplied by the compressor and releasing the heat in such a manner that the air is expanded and accelerated to give a smooth stream of uniformly heated gas at all conditions required by the turbine. This task must be accomplished with the minimum loss in pressure and with the maximum heat release for the limited space available. The amount of fuel added to the air will depend upon the temperature rise required. However, the maximum temperature is limited to within the range of 850 to 1700 ~ by the materials from which the turbine blades and nozzles are made. The air has already been heated to between 200 and 550~ by the work done during compression, giving a temperature rise requirement of 650 to 1150 ~ from the combustion process. Since the gas temperature required at the turbine varies with engine thrust, and in the case of the turbo-propeller engine, upon the power required, the combustion chamber must also be capable of maintaining stable and efficient combustion over a wide range of engine operating conditions. Efficient combustion has become increasingly important because of the consequent increase in atmospheric pollution.
The Combustion Process Air from the engine compressor enters the combustion chamber at a velocity up to 500 feet per second, but because at this velocity the air speed is far too high for combustion, the first thing that the chamber must do is to diffuse it, i.e. decelerate it and raise its static pressure. Since the speed of burning kerosine at normal mixture ratios is only a few feet per second, any fuel lit even in the diffused air stream, which now has a velocity of about 80 feet per second, would be blown away. A region of low axial velocity has therefore to be created in the chamber, so that the flame will remain alight throughout the range of engine operating conditions.
Gas Turbines
FIGURE 3-3
A combustion chamber. (Source: Rolls-Royce, UK)
In normal operation, the overall air/fuel ratio of a combustion chamber can vary between 45:1 and 130:1. However, kerosine will only burn efficiently at, or close to, a ratio of 15:1, so the fuel must be burned with only part of the air entering the chamber, in what is called a primary combustion zone. This is achieved by means of a flame tube (combustion liner) that has various devices for metering the airflow distribution along the chamber. Approximately 20 percent of the air mass flow is taken in by the snout or entry section (Figure 3-4). Immediately downstream of the snout are swirl vanes and a perforated flare, through which air passes into the primary combustion zone. The swirling air induces a flow upstream of the center of the flame tube and promotes the desired recirculation. The air not picked up by the snout flows into the annular space between the flame tube and the air casing. Through the wall of the flame tube body, adjacent to the combustion zone, are a selected number of secondary holes through which a further 20 percent of the main flow of air passes into the primary zone. The air from the swirl vanes and that from the secondary air holes interacts and creates a region of low velocity recirculation. This takes the form of a toroidal vortex, similar to a smoke ring, which has the effect of stabilizing and anchoring the flame (Figure 3-5). The recirculating gases hasten the burning of freshly injected fuel droplets by rapidly bringing them to ignition temperature. It is arranged that the conical fuel spray from the nozzle intersects the recirculation vortex at its center. This action, together with the general turbulence in the primary zone, greatly assists in breaking up the fuel and mixing it with the incoming air. The temperature of the gases released by combustion is about 1,800 to 2,000 ~ which is far too hot for entry to the nozzle guide vanes of the turbine. The air
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Process Plant Machinery
FIGURE 3-4 Apportioning the airflow. (Source: Rolls-Royce, UK)
FIGURE 3-5
Flame stabilizing and general airflow pattern. (Source: Rolls-Royce, UK)
not used for combustion, which amounts to about 60 percent of the total airflow, is therefore introduced progressively into the flame tube. Approximately a third of this is used to lower the gas temperature in the dilution zone before it enters the turbine and the remainder is used for cooling the walls of the flame tube. This is achieved by a film of cooling air flowing along the inside surface of the flame tubewall, insulating it from the hot combustion gases (Figure 3-6). A recent development allows cooling air to enter a network of passages within the flame tube wall before exiting to form an insulating film of air, this can reduce the required wall cooling airflow by up to 50 percent. Combustion should be completed before the dilution air enters the flame tube, otherwise the incoming air will cool the flame and incomplete combustion will result. An electric spark from an igniter plug initiates combustion and the flame is then self-sustained. The design of a combustion chamber and the method of adding the fuel may vary considerably, but the airflow distribution used to effect and maintain combustion is always very similar to that described.
Gas Turbines
CORRUGATED STRIP
~
FLAME TUBE
FLAME TUBE
COOLING AIR MACHINED COOLING RING
CORRUGATED STRIP COOLING LAMINATED FLAME TUBE WALL
COOLING AIR IN
OF
COOLING AIR
~r INTERNAL COOLING
SPLASH COOLING STRIP
FILM OF COOLING AIR OUT
TRANSPIRATION COOLING
FIGURE 3-6
Flame tube cooling methods. (Source: Rolls-Royce, UK)
Fuel Supply Fuel supply can be gaseous, liquid or both (dual). Liquid fuel is supplied to the airstream by one of two distinct methods. The most common is the injection of a fine atomized spray into the recirculating airstream through spray nozzles. The second method is based on the pre-vaporization of the fuel before it enters the combustion zone. In the vaporizing method the fuel is sprayed from feed tubes into vaporizing tubes which are positioned inside the flame tube. These tubes turn the fuel through 180 degrees and, as they are heated by combustion, the fuel vaporizes before passing into the flame tube. The primary airflow passes down the vaporizing tubes with the fuel and also through holes in the flame tube entry section which provide 'fans' of air to sweep the flame rearwards. Cooling and dilution air is metered into the flame tube in a manner similar to the atomizer flame tube.
Types of Combustion Chamber There are two main types of combustion chamber in use for gas turbine engines. These are the multiple chamber and the annular chamber.
Multiple Combustion Chamber This type of combustion chamber is used on centrifugal compressor engines and the earlier types of axial flow compressor engines. The chambers are disposed around the engine Figure 3-7 and compressor delivery air is directed by ducts to pass into the individual chambers. Each chamber
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FIGURE 3-7
Multiple combustion chambers. (Source: Rolls-Royce, UK)
has an inner flame tube around which there is an air casing. The air passes through the flame tube snout and also between the tube and the outer casing as already described. The separate flame tubes are all interconnected. This allows each tube to operate at the same pressure and also allows combustion to propagate around the flame tubes during engine starting.
Annular Combustion Chamber This type of combustion chamber consists of a single flame tube, completely annular in form, which is contained in an inner and outer casing (Figure 3-8). The airflow through the flame tube is similar to that already described, the chamber being open at the front to the compressor and at the rear to the turbine nozzles. The main advantage of the annual chamber is that, for the same power output, the length of the chamber is only 75 percent of that of a tubo-annular system of the same diameter, resulting in considerable saving of weight and production cost. Another advantage is the elimination of combustion propagation problems from chamber to chamber. The introduction of the air spray type fuel spray nozzle to this type of combustion chamber also greatly improves the preparation of fuel for combustion by
Gas Turbines
FIGURE 3-8
Annular combustion chamber. (Source: Rolls-Royce, UK)
aerating the over-rich pockets of fuel vapors close to the spray nozzle; this results in a large reduction in initial carbon formation. A combustion chamber must be capable of allowing fuel to burn efficiently over a wide range of operating conditions without incurring a large pressure loss. In addition, if flame extinction occurs, then it must be possible to relight. In performing these functions, the flame tube and spray nozzle atomizer components must be mechanically reliable. The gas turbine engine operates on a constant pressure cycle, therefore any loss of pressure during the process of combustion must be kept to a minimum. In providing adequate turbulence and mixing, a total pressure loss varying from about 3 to 8 percent of the air pressure at entry to the chamber is incurred. The heat released by a combustion chamber or any other heat generating unit is dependent on the volume of the combustion area. Thus, to obtain the required high power output, a comparatively small and compact gas turbine combustion chamber must release heat at exceptionally high rates.
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Process Plant Machinery
The Turbine The turbine has the task of providing the power to drive the compressor and accessories and, in the case of engines which do not make use solely of a jet for propulsion, of providing shaft power for a propeller or rotor. It does this by extracting energy from the hot gases released from the combustion system and expanding them to a lower pressure and temperature. High stresses are involved in this process, and for efficient operation, the turbine blade tips may rotate at speeds over 1500 feet per second. The continuous flow of gas to which the turbine is exposed may have an entry temperature between 850 and 1700 ~ and may reach a velocity of over 2500 feet per second in parts of the turbine. To produce the driving torque, the turbine may consist of several stages each employing one row of stationary nozzle guide vanes and one row of moving blades (Figure 3-9). The number of stages depends upon the relationship between the power required from the gas flow, the rotational speed at which it must be produced and the diameter of turbine permitted.
FIGURE 3-9 A triple-stage turbine with single shaft system. (Source: Rolls-Royce, UK)
Gas Turbines
The number of shafts, and therefore turbines, varies with the type of engine; high compression ratio engines usually have two shafts, driving high and low pressure compressors. On mechanical drive engines, driving torque is derived from a free power turbine (Figure 3-10). This method allows the turbine to run at its optimum speed because it is mechanically independent of other turbine and compressor shafts. In normal land-based applications, component weight is not a major consideration. For offshore platform applications, however, so as to arrive at a common model for all applications, most manufactures attempt to conserve weight. Also many of today's land-based gas turbines today are derived from aircraft engines, where weight is a major consideration. The mean blade speed of a turbine has a considerable effect on the maximum efficiency possible for a given stage output. For a given output the gas velocities, deflections, and hence losses, are reduced in proportion to the square of higher mean blade speeds. Stress in the turbine disk increases as the square of the speed, therefore to maintain the same stress level at higher speed the sectional thickness, hence the weight, must be increased disproportionately. For this reason, the final design is a compromise between efficiency and weight. Engines operating at higher turbine inlet temperatures are thermally more efficient and have an improved power to weight ratio. The design of the nozzle guide vane and turbine blade passages is based broadly on aerodynamic considerations, and to obtain optimum efficiency, compatible with
FIGURE 3-10
A typical free power turbine. (Source: Rolls-Royce, UK)
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Process Plant Machinery
compressor and combustion design, the nozzle guide vanes and turbine blades are of a basic aerofoil shape. There are three types of turbine; impulse, reaction and a combination of the two known as impulse-reaction. In the impulse type the total pressure drop across each stage occurs in the fixed nozzle guide vanes which, because of their convergent shape, increase the gas velocity whilst reducing the pressure. The gas is directed onto the turbine blades which experience an impulse force caused by the impact of the gas on the blades. In the reaction type the fixed nozzle guide vanes are designed to alter the gas flow direction without changing the pressure. The converging blade passages experience a reaction force resulting from the expansion and acceleration of the gas. Normally gas turbine engines do not use pure impulse or pure reaction turbine blades but the impulse-reaction combination Figure 3-11. The proportion of each principle incorporated in the design of a turbine is largely dependent on the type of engine in which the turbine is to operate, but in general it is about 50 percent impulse and 50 percent reaction. Impulse-type turbines are used for cartridge and air starters. From the previous description, it will be seen that the turbine depends for its operation on the transfer of energy between the combustion gases and the turbine. This transfer is never 100 percent because of thermodynamic and mechanical losses. When the gas is expanded by the combustion process, it forces its way into the discharge nozzles of the turbine where, because of their convergent shape, it is accelerated to about the speed of sound which, at the gas temperature, is about 2500 feet per second. At the same time the gas flow is given a 'spin' or 'whirl' in the direction of rotation of the turbine blades by the nozzle guide vanes. On impact with the blades and during the subsequent reaction through the blades, energy is absorbed, causing the turbine to rotate at high speed and so provide the power for driving the turbine shaft and compressor.
.__..•I I
NOZZLE
TURBINE
I
NOZZLE
II
TURBINE
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% Turbine driven by the impulse of the gas flow only
19 Turbine driven by the impulse of the gas flow and its subsequent reaction as it accelerates through the converging blade passage
FIGURE 3-11 Comparison Between a pure impulse turbine and an impulse~reaction turbine. (Source: Rolls-Royce, UK)
Gas Turbines
The torque or turning power applied to the turbine is governed by the rate of gas flow and the energy change of the gas between the inlet and the outlet of the turbine blades. The design of the turbine is such that the whirl will be removed from the gas stream so that the flow at exit from the turbine will be substantially "straightened out" to give an axial flow into the exhaust system. Excessive residual whirl reduces the efficiency of the exhaust system and also tends to produce jet pipe vibration which has a detrimental effect on the exhaust cone supports and struts. It will be seen that the nozzle guide vanes and blades of the turbine are "twisted", the blades having a stagger angle that is greater at the tip than at the root, Figure 3-12. The reason for the twist is to make the gas flow from the combustion system do equal work at all positions along the length of the blade and to ensure that the flow enters the exhaust system with a uniform axial velocity. This results in certain changes in velocity, pressure and temperature occurring through the turbine, as shown diagrammatically in Figure 3-13. The 'degree of reaction' varies from root to tip, being least at the root and highest at the tip, with the mean section having the chosen value of about 50 percent. The losses which prevent the turbine from being 100 percent efficient are due to a number of reasons. A typical uncooled three-stage turbine would suffer a 3.5 percent loss because of aerodynamic losses in the turbine blades. A further 4.5 percent loss would be incurred by aerodynamic losses in the nozzle guide vanes, gas leakage over the turbine blade tips and exhaust system losses; these losses are of approximately equal proportions. The total losses result in an overall efficiency of approximately 92 percent.
FIGURE 3-12
A typical turbine blade showing twisted contour. (Source: Rolls-Royce, UK)
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~
/
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/
VELOCITY DECREASES
PRESSURE INCREASES (From root to tip
across nozzles )
q
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Velocity . . . . Static pressure
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?, NOZZLE
FIGURE 3-13
Pressure and velocity uniform
on entering exhaust system
I BLADE
Gas flow pattern through nozzle and blade. (Source: Rolls-Royce, UK)
Construction The basic components of the turbine are the combustion discharge nozzles, the nozzle guide vanes, the turbine disks and the turbine blades. The rotating assembly is carried on bearings mounted in the turbine casing and the turbine shaft may be common to the compressor shaft or connected to it by a self-aligning coupling.
Nozzle Guide Vanes. The nozzle guide vanes are of an aerofoil shape with the passage between adjacent vanes forming a convergent duct. The vanes are located in the turbine casing in a manner that allows for expansion (Figure 3-14). The nozzle guide vanes are usually of hollow form and may be cooled by passing compressor delivery air through them to reduce the effects of high thermal stresses and gas loads. Turbine Disks. Turbine disks are usually manufactured from a machined forging with an integral shaft or with a flange onto which the shaft may be bolted. The disk also has, around its perimeter, provision for the attachment of the turbine blades. To limit the effect of heat conduction from the turbine blades to the disk a flow of cooling air is passed across both sides of each disk. Turbine Blades. The turbine blades are of an aerofoil shape designed to provide passages between adjacent blades that give a steady acceleration of the flow up to the "throat", where the area is smallest and the velocity reaches that required at exit to produce the required degree of reaction. The actual area of each blade cross-section is fixed by the permitted stress in the material used and by the size of any holes which may be required for cooling purposes. High efficiency demands thin trailing edges to the sections, but a compromise has to be made so as to prevent the blades from cracking due to the temperature changes during engine operation. The method of attaching the blades to the turbine disk is of considerable importance, since the stress in the disk around the fixing or in the blade root has an important bearing on the limiting rim speed. Various methods of blade attachment
Gas Turbines
FIGURE 3-14
Typical nozzle guide vanes showing their shape and location. (Source: Rolls-Royce, UK)
FIGURE 3-15
Variousmethods of attaching blades to turbine disks. (Source: Rolls-Royce, UK)
are shown in Figure 3-15; however, the B.M.W. hollow blade and the de Laval bulb root types are not now generally used on gas turbine engines. A gap exists between the blade tips and casing, which varies in size due to the different rates of expansion and contraction. To reduce the loss of efficiency through gas leakage across the blade tips, a shroud is often fitted. This is made up by a small segment at the tip of each blade which forms a peripheral ring around the blade tips. An abradable lining in the casing may also be used to reduce gas leakage.
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ProcessPlant Machinery Contra-Rotating Turbine. Figure 3-16 shows a twelve-stage contra-rotating free power turbine driving a contra-rotating rear fan. This design has only one row of static nozzle guide vanes. The remaining nozzle guide vanes are, in effect, turbine blades attached to a rotating casing which revolves in the opposite direction to a rotating drum. Since all but one aerofoil row extracts energy from the gas stream, contra-rotating turbines are capable of operating at much higher stage loadings than conventional turbines, making them attractive for direct drive applications. Dual Alloy Disks. Very high stresses are imposed on the blade root fixing of high work rate turbines, which make conventional methods of blade attachment impractical. A dual alloy disk, or 'blisk' as shown in Figure 3-17, has a ring of cast turbine blades bonded to the disk. Materials Nozzle Guide Vanes. Due to their static condition, the nozzle guide vanes do not endure the same rotational stresses as the turbine blades. Therefore, heat resistance is the property most required. Nickel alloys are used, although cooling is required to prevent melting. Ceramic coatings can enhance the heat resisting properties and, for the same set of conditions, reduce the amount of cooling air required, thus improving engine efficiency.
FIGURE 3-16
Freepower contra-rotating turbine. (Source: Rolls-Royce, UK)
Gas Turbines
CAST BLADE RING DIFFUSION BOND
POWDER DISC
FIGURE 3-17
Section through a dual alloy disk. (Source: Rolls-Royce, UK)
Turbine Disks. A turbine disk has to rotate at high speed in a relatively cool environment and is subjected to large rotational stresses. The limiting factor which affects the useful disk life is its resistance to fatigue cracking. In the past, turbine disks have been made in ferritic and austenitic steels but nickel-based alloys are currently used. Increasing the alloying elements in nickel extend the life limits of a disk by increasing fatigue resistance. Alternatively, expensive powder metallurgy disks, which offer an additional 10% in strength, allow faster rotational speeds to be achieved. Turbine Blades. A brief mention of some of the points to be considered in connection with turbine blade design will give an idea of the importance of the correct choice of blade material. The blades, while glowing red-hot, must be strong enough to carry the centrifugal loads due to rotation at high speed. A small turbine blade weighing only two ounces may exert a load of over two tons at top speed and it must withstand the high bending loads applied by the gas to produce the many thousands of turbine horse-power necessary to drive the compressor. Turbine blades must also be resistant to fatigue and thermal shock, so that they will not fail under the influence of high-frequency fluctuations in the gas conditions, and they must also be resistant to corrosion and oxidization. In spite of all these demands, the blades must be made in a material that can be accurately formed and machined by current manufacturing methods. From the foregoing, it follows that for a particular blade material and an acceptable safe life there is an associated maximum permissible turbine entry temperature and a corresponding maximum engine power. It is not surprising, therefore, that metallurgists and designers are constantly searching for better turbine blade materials and improved methods of blade cooling. Over a period of operational time the turbine blades slowly grow in length. This phenomenon is known as 'creep' and there is a finite useful life limit before failure occurs. The early materials used were high-temperature steel forgings, but these were rapidly replaced by cast nickel base alloys which give better creep and fatigue properties. Close examination of a conventional turbine blade reveals a myriad of crystals that lie in all directions (equi-axed). Improved service life can be obtained by aligning the crystals to form columns along the blade length, produced by a method known as "Directional Solidification". A further advance of this technique is to make the blade out of a single crystal. Examples of these structures are shown in Figure 3-18. Each method extends the useful creep life of the blade (Figure 3-19)
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Process Plant Machinery
FIGURE 3-18
Various turbine blade crystal structures. (Source: Rolls-Royce, UK)
and in the case of the single crystal blade, the operating temperature can be substantially increased.
Balancing The balancing of a turbine is an extremely important operation in its assembly. In view of the high rotational speeds and the mass of materials, any unbalance could seriously affect the rotating assembly bearings and engine operation. Balancing is effected on a special balancing machine.
Gas Turbines .,, ......... ,,
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9 FRACTURE SINGLE CRYSTAL BLADES
Z
DIRECTIONALLY SOLIDIFIED BLADES
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FIGURE 3-19
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Comparison of turbine blade life properties. (Source: Rolls-Royce, UK)
Internal Air System A typical system is discussed below (as used in the Rolls-Royce Avon). To disperse the heat conducted from the main hot gas stream to the turbine, bearings and the internal casings, a flow of cooling air is directed over these components, and this same air supply is used to pressurize various oil and gas seals. Low pressure air from stage 3 of the compressor is used for cooling the main bearings and for pressurizing the oil seals, and is then dispersed to atmosphere. High pressure air from stage 15 cools the turbine and pressurizes the gas seals before dispersing.
Low Pressure Air. Air from stage 3 of the compressor is directed into the hollow compressor shaft, and holes at the front of the shaft direct the flow to the rear of the front bearing to pressurize the oil seals and prevent oil leakage into the compressor. Holes at the rear of the compressor shaft allow some LP air into the area around the rear bearing housing. This air adequately reduces the transfer of heat from the combustion chambers to the turbine shaft and rear bearing, and prevents oil leakage by pressurizing both the center bearing oil seals and the rear beating front oil seal. The air then flows to the location between the combustion chambers and the exhaust unit, i.e., cooling air manifold, over the nozzle box and turbine casings to the upper part of the manifold and then to atmosphere. The remainder of the LP cooling air flows through the turbine shaft and outwards to cool the rear beating inner track, then onwards to the cooling air manifold.
High PressureAir. Some of the compressor delivery air is directed via the combustion chamber inner heat shield to the front face of the HP turbine disk, after which the flow divides into two main streams. One air stream enters the annulus formed by the outer air seal carrier and the HP nozzle guide vane inner locating ring. The air passes into the labyrinth seal for
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FIGURE 3-20
Main intend air system flows. (Source: Rolls-Royce, UK)
FIGURE 3-21
Air system flow in the turbine. (Source: Rolls-Royce, UK)
pressurizing the gap between the roots of the HP turbine blades and the HP nozzle guide vanes. This air is also used to balance the pressures across the front of the nozzle box and to provide a flow of cooling air for the HP turbine blades (which is a feature of the higher powered Avon gas generators). The air enters the hollow blades via holes in the aerofoil sections, passes out through the blade tips, then flows into the main gas stream through the space at the rear of the rotor blade tip shrouds. The second air stream flows inwards to pressurize the inner air seal before passing to the area between the turbine HP and IP disks. From there the flow
Gas Turbines
divides, cools the IP and LP disks and exhaust deflection plate, and also pressurizes the gap at the turbine blade root before passing into the main gas stream. The HP and IP nozzle guide vanes are cooled by high pressure air which is passed through the vanes along the leading and trailing edges, and exhausted via trailing edge drillings. HP air also cools the HP and IP shroud tings, the same supply preventing hot gases passing behind the turbine shrouds.
Burners
The following burner designs have generic features typical for the fuel applications featured. The sketches depict Rolls-Royce Avon configurations.
Liquid Fuel Burners (Figure 3-22) Liquid fuel is injected at high pressure into the combustion chambers using duplex burners (or fuel injection nozzles). A duplex burner is fitted in each of the eight combustion chambers. Each burner incorporates a primary filter and primary nozzle co-axial with a main nozzle. Additionally an air shroud directs air across the nozzle faces to minimize carbon deposition, and the amount of air flow is controlled between limits to optimize weak burning stability and ignition performance. The atomized fuels spray from the two nozzles takes the form of a hollow cone, the spray cone angle of the primary being less than that of the main.
Gas Fuel Burners (Figure 3-23) The gas burners have been designed to simulate as far as possible the conical injection pattern of the liquid fuelled burners, and each burner incorporates a deflector valve and shroud to produce the conical shape. An alternative p e p p e r p o t burner is available for use with heavier or wet hydrocarbon gases.
Dual Fuel Burners (Figure 3-24) The dual fuel burner nozzle was developed to enable liquid fuels and gaseous fuels to be used alternatively or at the same time. While a pressure jet atomizer is used for start purposes, atomization at power conditions is achieved by the airspray principle - the subject of a Rolls-Royce patent in 1952. This method of atomization makes use of the air pressure drop across the flametube which is usually about 5% of the compressor outlet pressure. The pressure drop results in an air velocity in the air atomizing duct of around 400 ft/s (122 m/s), and this remains fairly constant over the power range. Atomization is achieved by the fuel being presented in a thin film to the air, which accelerates the fuel, and in so doing produces drag forces which, acting against surface tension, tear the film into droplets. In the airspray principle, air velocity is most important- and this does not change greatly over a wide range of power settings. SELECTING A GAS TURBINE
At some point in their working life, process engineers may be asked to select or be part of a team for, selecting a gas turbine. The main features to be considered
65
66
Process Plant Machinery
~
FEEDARM
MARYFILTER
PRIMARYNOZZLE-
~
MAINNOZZLE
SPRINGLOCKINGRING
SHROUD
FIGURE 3-22
Liquid fuel burner. (Source: Rolls-Royce, UK)
in this process are described below. Items on this list are also a good check list during normal operation, packaging and shipping when major repairs off-site are required, and when modifications are proposed by the manufacturer (to assess the gross effect of those changes). Reference literature used for this subsection was edited from Rolls-Royce literature on their Avon and RB211 models.
Gas Turbines
: U E ~ CONIX : C T , O N
...,/
FEFD A R M - - -
,. ,.~,
"~
"': ~.t,
~
L~_~
.:
SH ROL,I r]
DEFLEU IOR V A L V E . " f "
FIGURE 3-23
Gas fuel burner. (Source: Rolls-Royce, UK)
Required Nominal Power Output See Figure 3-25. Also see subsections on growth potential and performance characteristics.
Thermal Efficiency and Cogeneration Potential Consider if simple cycle application is what is required, or if regeneration, waste heat recovery or combined cycle would benefit overall process profitability (see the appendix on cycles in this chapter). Optimized cogeneration is responsible operation from both economic and environmental perspectives. Cogeneration is the utilization of normally wasted gas generator exhaust heat to produce additional energy via a heat recovery system. (Figure 3-26). This conversion of a fuel can supply all the energy requirements of a given building, factory, or process utility at their point of utilization, satisfying the demands of heating, cooling and electrical generation from a single energy source. The type of fuel used in the gas turbine may be liquid or gaseous, and it is the factors of cost, availability and transportation which determine the selection of fuel type. The gas turbine is a basic energy converter, and is therefore considered to be the heart of the system. The power turbine drives an a.c. generator or an item of driven equipment such as a compressor. Exhaust gas from the turbine is directed into an exhaust gas heat exchanger to produce steam or hot water to power a steam turbine or service district heating requirements. Additionally, these products can then be passed through an absorption cycle water chilling package to satisfy air conditioning and other refrigeration applications.
Operational Track Record Consider this item in terms of running hours for the engine being assessed, and availability and reliability figures for the specific application in question (Figure 3-27).
67
68
Process Plant Machinery
FIGURE 3-24
Dual fuel burner. (Source: Rolls-Royce, UK)
In pipeline service the Avon has achieved a recorded: 9 Availability 99.8% 9 Reliability 99.7%
Growth Potential This is more of a concern for aeroengines. There are, however, genetic industrial gas turbines, usually aeroengine derived that have a design that has growth potential.
Gas Turbines
FIGURE 3-25
69
Nominal power output for mechanical drive and power generation. (Source: Rolls-Royce, UK)
FIGURE 3-26(a)
Cogeneration schematic. (Source: Rolls-Royce, UK)
This usually means that design development work being conducted will produce options including new components that will fit within the same core diameter as the original engine, which can then be run at a higher power rating. (See Tables 3-1, 3-2 and 3-3. Also see subsection on performance characteristic.
Start Up and Cool Down Full power production in a few minutes from cold start is frequently a requirement in a mechanical drive machine that is not base or continually loaded. Long warm up and cool down periods help reduce the effective operating hours used up by the turbine. Proper warm up and cool down are safeguards against shaft bowing, which occurs in older designs.
Engine Material and Design Features Consider overall engine features in terms of corrosion resistance features such as:
70
Process Plant Machinery
FIGURE 3-26(b) High efficiency cogeneration plant in Rotterdam, The Netherlands. Up to 90% of the total fuel energy can be used in this type of plant. (Source: Rolls-Royce, UK)
TABLE 3.1 Industrial Avon: Entry Into Service and Comparative Performance ISO C o n d i t i o n s - No Losses - Base L o a d - Gaseous Fuel
Gas generator types (variants)
Year of entry into service (approx.)
Nominal power output
Mk 1533 -56, -75, -76, -80
EGHP
Nominal thermal efficiency % (combined gas generator and matched power turbine)
1964
19580
27.6
Mk 1534 101, - 102
1975
21860
27.6
-
Mk 1535-120 121, - 122, - 130
1978
23800
27.6
-
Mk 1535-160 161, - 162, - 170
1990
24000
29.4
-
(Source: Rolls-Royce Industrial & Marine Gas Turbines Limited)
Gas Turbines
FIGURE 3-27 TABLE 3.2
Data for the Rolls-Royce Avon. (Source: Rolls-Royce, UK)
Industrial Avon: Gas Generator Variants
The Avon has been developed progressively over many years giving incremental power output increases as material and cooling changes allowed. Five Marks of Avon have existed (including the short lived Mk.1535E) and those remaining in service are the Mk.1533, 1534, 1535-120 and the Mk.1535-160. The gas generator is further identified by a sub-mark number, this being a suffix which identifies whether the lubricating oil, hydraulic oil and fuel system components are mounted on or off the engine itself, i.e. Phase 1 or Phase 2. Phase 2 are served by floor-mounted ancillaries. The chart below gives a useful reference as a means of identifying gas generator types by the suffix number.
Type
Phase 1
Phase 2
Phase 1 (Close coupled)
Phase 2 (Close coupled)
Mk.1533" Mk.1534 Mk.1535-120 1535-160
75,77,78 102 122 162
76 101 121 161
X X X 171
80 X 130 170
*very early versions include -51, 56, 60 A close-coupled unit is one which is directly attached to the power turbine, in this case the power turbine is an RT56. Power output and efficiency improvements are realized with a close-coupled configuration as inlet ducting losses are reclaimed. (Source: Rolls-Royce Industrial & Marine Gas Turbines Limited) compressor materials such as titanium alloy selections, ceramic coatings rotor dynamics, such as blade shrouding turbine materials such as alloys for corrosion resistance and creep life flexibility afforded by bleed valve system combustor design fuel flexibility (gas, liquid, dual) internal cooling on turbine blades and nozzle guide vanes
71
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Process Plant Machinery
TABLE 3.3 Industrial Avon Nominal P e r f o r m a n c e - M e c h a n i c a l Imperial U n i t s - ISO Conditions - No Losses Performance of combined gas generator and power turbine Gas generator type
Mk. Mk. Mk. Mk. Mk. Mk. Mk. Mk. Mk. Mk.
Power t u r b i n e Fuel type type
1533 1533 1534 1534 1535-120 1535-120 1535-160 1535-160 1535-160 1535-160
RT 48 RT 48 RT 48 RT 48 RT 48 RT 48 RT 48S RT 48S RT 48S RT 48S
Gas Gas Gas Gas Gas Gas Gas Gas Liquid Liquid
Rating
Max. cont. Base Max. cont. Base Max. cont. Base Max. cont. Base Max. cont. Base
Power B.P.H.
20806 16154 22658 18016 23582 19568 23734 20363 23202 19855
Specific heat rate BTU/BHPhr
8722 9213 8816 9219 8909 9216 8498 8663 8434 8630
Drive
Applications,
Gas generator data
Thermal efficiency Speed
Exhaust gas horsepower
%
29.2 27.6 28.9 27.6 28.6 27.6 29.9 29.4 30.2 29.5
7900 7500 7900 7500 7900 7500 7820 7540 7770 7500
25280 19580 27890 21860 29100 23800 28400 24000 27700 23355
(Source: Rolls-Royce Industrial & Marine Gas Turbines Limited A g e n e r a l specification for a gas turbine will h i g h l i g h t these features, as the outline description for R o l l s - R o y c e R B 2 1 1 does in F i g u r e 3-28. a. S e v e n - s t a g e IP c o m p r e s s o r R o t o r blades ~ T i t a n i u m Stator blades ~ C o r r o s i o n
alloy resistant alloy steel
b. Six-stage HP c o m p r e s s o r B l a d e materials are c h o s e n for m a x i m u m c o r r o s i o n r e s i s t a n c e and cyclic fatigue life at the h i g h e r o p e r a t i n g t e m p e r a t u r e for this a s s e m b l y .
Rotor blades" Stages 1 to 3 Stage 4 Stages 5 and 6
FIGURE 3-28
T i t a n i u m alloy C o r r o s i o n resistant alloy steel C o r r o s i o n and heat resistant nickel base alloy
Schematic o f gas turbine. (Source: Rolls-Royce, UK)
Gas Turbines
Stator blades: Stages 1 to 5 Corrosion resistant alloy steel Stage 6 Corrosion and heat resistant nickel base alloy The steel stator blades on both compressors are ceramic coated to give maximum resistance to corrosion and oxidization. c. Single-stage lip turbine Nozzle guide vanes and rotor blades are air cooled. Rotor blades are shrouded to ensure that sealing remains good throughout the operating life of the unit and that power and thermal efficiency are maintained at a high level close to the initial pass-off test. d. Single-stage IP turbine Cooled nozzle guide vanes and shrouded turbine blades. Blades in both turbines Corrosion resistant nickel base alloys (giving maximum creep life) e. Rotating assemblies Co-axial drive shafts connect the turbines to the compressors. f. Inlet guide vanes A single stage of variable incidence inlet guide vanes, in conjunction with bleed valves on both compressors, maintains optimum performance over the whole speed range with the minimum mechanical complexity. g. Bleed air Bleed valves on the compressors assist low power starting and give good handling characteristics. Bleed ports on the casings provide IP and HP compressor air for power turbine cooling and other services. h. Combustion Annular combustion chamber with 18 burners. The burners are designed to function with gaseous fuel, distillate liquid fuel, or a mixture of the two, with or without water injection. The annular design of the chamber provides a clean aerodynamic extension to the HP compressor outlet, minimizes the area to be cooled, gives good temperature distribution, and avoids interconnector problems.
Modular Construction This feature is of growing importance in cutting down on maintenance times, reducing how much of the process system around the gas turbine needs to be dismantled for component removal and ease of transport of major modules, especially in remote and offshore applications The gas generator shown in Figure 3-29 (a Rolls-Royce RB 211) is constructed from five m o d u l e s - all fully interchangeable with standard replacements. Modular construction offers the following advantages: 9 Modules may be changed without removal of the gas generator from site. 9 The modules, of low weight and size, are easily transported to remote sites where access may be difficult. All modules, and the complete gas generator, are portable by helicopter. 9 Reduced need for holdings of spare gas generators and range of spare parts. 9 Reduced down-time and maintenance costs. Lower down-time improves plant availability.
73
74
Process Plant Machinery
-/
-03 .....
J 04
05
FIGURE 3-29 Modules in gas turbine. (Source: Rolls-Royce, UK) 9 Life capability of all modules can be maximized with flexibility of inspection intervals.
Bearings The number of bearings, type of beatings and proximity of beatings to hot section zones, needs to be considered (see Figure 3-30). Beatings are positioned away from the hot combustion zone. Main roller bearings are hydraulically damped by a squeeze film (Figure 3-31). The thin film of oil between bearing outer track and bearing housing provides a high degree of damping which: 9 Reduces vibration 9 Increases component life 9 Ensures correct operation of rolling elements of beatings at all loads and shaft speeds
IP COMPRE$SO
UST BEARING
HP TURBINE BEARING
P TURBINE BEARINGS<:
FIGURE 3-30 Bearing positions in a gas turbine. (Source: Rolls-Royce, UK)
Gas Turbines OIL
:~ SQUEEZE
Ftt.M
FEED
, ; /
TO BEARING LUBRICATION BEARING OUTE R RACE
:\ \
FIGURE 3-31
Squeeze film damped bearing. (Source: Rolls-Royce, UK)
Auxiliary Systems and Packaging Options Some of the service problems experienced with gas turbines originate with the auxiliary systems on the gas turbine, which might have been made by a subvendor to the original equipment manufacturer (OEM). The packager of the overall system may be the OEM or one of the subvendors. Operational problems that occur with the gas turbine system could originate with one or more of the components indicated in the following line diagrams.
Starting System (See Figure 3-32) The starting system for the RB211 comprises a gas generator-mounted air/gas-driven starter motor (comp|ete with a speed sensing unit), a skid-mounted shut-off valve and pressure regulator, and a control panel.
RB211
AI R/GAS STARTER MOTOR & SPEED SENSOR
VALVE
SI'~ I_D 5 :rG'~'AL--"I
I I I
E X H - A U S T DI.)CTE D~r IO A IMOSPHERI ~
SPEED SENSING & C ON T R OL UNIT l
A t F~:'G AS SUPPLY
., l - I --1
41,,.- ,....,. - ..,..7
VALVE ,
,
SHUT-OFF VALVE
.....
FIGURE 3-32
I I I
t
NrJ R M A L - ]-STARTER MOTOR ] CUT-OFF SIGNAL ,~,. . . . . . .
I
..,.i
1
[ L~TARTE R OVERSPEED SIGNA...
1} CONTROL PANEL__;
I 1 _
SIGNAL TO I N I T I A l E T OT A L 5;HUTI3OWN OF SFT.,~
Starting System Schematic. (Source: Rolls-Royce, UK)
75
76
Process Plant Machinery
An engine-mounted hydraulic starter can be fitted as an option. Hydraulic power is supplied by the Packager. The output drive from the starter motor is by a drive-shaft and bevel gear to the HP compressor rotor assembly. Light-up is achieved by energizing two igniter/burners while the HP spool is being turned. The starter drive shaft is automatically disengaged when the gas generator has reached self-sustaining speed. Speed sensing components are incorporated in the control system to provide safeguards against malfunction. The ability of the starting system to operate successfully in extremely low temperatures is being checked during development testing.
Gas Generator Options In cases like that of the Rolls-Royce Avon, a Phase 1 model with all accessories mounted on the gas generator is available. A Phase II model with these mounted off the gas generator on a separate skid is also available. Both these models are dual fuelled (i.e. they can switch to gas or liquid or a mixture of both these fuels without shutdown).
Phase I. The Phase I Avon gas generator is equipped with gas generator mounted accessories. For example, the oil system, including a combined sump and tank, pressure and scavenge pumps and filter, are all contained in a wheelcase (Figure 3-33). This wheelcase, although having the facility of being removable, forms an integral part of the Phase I gas generator. Similarly, on the liquid fuelled Phase I gas generator, the fuel control system is gas generator mounted, together with such items
/ ,'
i#
C O N \ E E l If~t,I I O :_;~TIE~ D,A_ ~'~LIrPL" T&NK
Z ) IFT " O L P.JPIP GrA.'I:,
-LL
=UMP E-P.IVr Cil TO C r,.)r,,,)LE ~t r,);L J ~ ~,., rRot-I :_OOLER A
OII PRES~UP, E C C N NECXLOIN
- - r Cr.J r, rCTIOI'.4 ~Gl,, t X l ~I',NAL S U I ~ . Y T A N K ""
F,I#.T~RIN C I - ~ 0 " 1 FUEl. P..HP I I.'),.51N:.,
FIGURE 3-33
f E E ~ ,~1. T O
ILE"..RN C I rK~,,I rl~O,N- n r A ! I',IG NOUN-ING rA~ r
Phase I oil sump and wheelcase. (Source: Rolls-Royce, UK)
Gas Turbines
77
as the high pressure (HP) fuel pump, HP shut-off cock, pressurizing valve, fuel manifold, burners and piping - all being supplied by Rolls-Royce, with a minimum of equipment being supplied by the purchaser. The compressor airflow control system can be gas generator mounted and hydraulically powered by the HP fuel pump. Phase II. The Phase II Avon gas generator has the entire hydraulic and lubricating oil pressure supply and scavenge systems, together with the fuel supply and control system mounted off the gas generator. All these systems will generally be supplied and equipped predominantly with floor/pedestal mounted accessories provided by the purchaser (Figure 3-34 and 3-35). Normally the purchaser is responsible for supplying the entire gas generator fuel system up to the gas generator liquid HP fuel and gas fuel inlets. All Phase I and Phase II Avon gas generators have electromagnetic pick-ups for speed monitoring purposes. Exhaust gas thermocouples are also supplied. In the case of gaseous and dual fuelled Phase I gas generators, it is generally the purchaser who supplies the fuel control and forwarding system, with only the fuel manifolds, burners and engine piping being supplied by Rolls-Royce. The compressor air flow control system remains gas generator mounted and is normally powered by a Rolls-Royce supplied hydraulic pump which is fitted in place of the gas-generator-mounted HP fuel pump. Dual fuelled gas generators are also supplied with the liquid fuel pressurizing valve, HP fuel inlet filter and where appropriate the gas fuel manifold purge valve. Those parts of the systems which will be supplied on the Phase II gas generator by Rolls-Royce are: Hydraulic supply system:
9 Compressor air flow control system, disconnect block, and interconnecting pipes.
FRONT BEARING HOUSING AND ANTI-ICING AIR MANIFOLD
STAGE 7 BLEED VALVE ~
ANTI-ICING AIR VALVE
COMBUSTION AIR CASING \
NOZZ LE BOX AN D TURBINE HOUSING EXHAUST UNIT
COOLING AIR OUTLET
COMPRESSOR
FILTER AND FUEL PRESSURISING VALVE
THERMOCOUPLE
COMPRESSOR OUTLET CASING
HARNESS CONNECTION
FIGURE 3-34(a) Phase H liquid fuelled (left-hand view). (Source: Rolls-Royce, UK)
78
Process Plant Machinery
REAR MOUNTI NG TRUNNION / THERMOCOUPLES
/
FRONT SUSPENSION
/
.,..~
ELECTRO M A G N E T I C
/
/
LUBRICATING OIL CONNECTION BLOCK
FIGURE 3-34(b)
COMPRESSOR A I R F L O W C O N T R O L MECHANISM
Phase H liquid fuelled (right-hand view). (Source: Rolls-Royce, UK)
FRONT BEARING HOUSING AND ANTI-ICING AIR M A N I F O L D
STAGE 7 BLEED V A L V E
~_._, ,.,
ANTI-ICING AIR VALVE
\
'
~ - ,
COMBUSTION AIR CASING
'\,~, ~
~ l [ ~ i l ~ l ; ~ - ' - ~ ~ _
"
STAGE 4 BLEED V A L V E
.
,,[(,,.~,~7~
_ , , ~ ~ =
~_~
NOZZLE BOX AND TURBINE HOUSING
\ ~
,
- ~,
/
i
EXHAUST UNIT
? / .-.dill;
,-p>Tt 9 =-~;! , . s iV.l v/.4/i
COMPRESSOR
,-. - I
I
.
.i~
:~
o
~.,
.,-.-;- ~ . ~ e ~ - I ' ~
FUEL M A N I F O L D INLET
~ -
- ""
i
,:7";
,;:.~' z.'5..".,j" /'" .....
" "
COMPRESSOR O U T L E T CASI NG
"
.'
THERMOCOUPLE
t n~-:. .
m
i=t.~
:x-Dual fuelled
HP FUEL INLET F I L T E R LIQUID F U E L PRESSURISING V A L V E
M A N I F O L D PURGE V A L V E L I Q U I D FUEL M A N I F O L D S D U A L FUEL BURNERS
FIGURE 3-35(a)
1'~) J ~.I"
)
,.~L 4:,'-~ ......
-
-
~j
Phase H gaseous fuelled (left-hand view). (Source: Rolls-Royce, UK)
Gas Turbines
FIGURE 3-35(b)
Phase II gaseous fuelled (right-hand view). (Source: Rolls-Royce, UK)
Lubricating oil system: 9 Lubricating oil (pressure and scavenge) connection block, including magnetic chip detectors in each beating scavenge line, and oil inlet pressure tapping. 9 Associated pipework for bearing feed, scavenge and venting. Fuel system (liquid fuelled): 9 Fuel pressuring valve 9 Fuel manifold 9 Burners Fuel system (gaseous fuelled): 9 Fuel manifold 9 Burners Fuel system (dual fuelled): 9 Liquid fuel pressurizing valve 9 Liquid fuel manifold 9 Gas fuel manifold 9 Gas manifold purge valve 9 Dual fuel burners
Fuel Supply and Control Systems Liquid Fuelled Phase I Gas Generators The gas generator fuel system (see Figure 3-36) provides a means of forwarding the fuel from the plant fuel system installation to the gas generator.
79
Process Plant Machinery
80
D
GAS GENERATOR
GAS GENERATOR EXHAUST GAS TEMPER ATU RE (CONTROL! THERMOCOUPLES
POWER TURBINE
T,-IERIVAL
HFLIFF
SYS~'E~:
,z LC t.c cc
-,P. t'UEL
t
<.
m iiii
L.P. FIJEL PRESSUHI: REGULATING VALVE
1
FCT CC ~ITROI I FR
PURCHASER FUEL CONTHOL SYSTFM
rLOW%tETER
L.P. FUEL. ISOLATING VALVC
FUEL DELIVERY LINE FI LTErl
LIQUID FUFL SUnPLY
A
9 L.P. FUEL FILTER, H.P. FUEL PUMP A N D H.P. FUEL SHUT-OFF COCK
B
9 FUEL C O N T R O L SYSTEM INCLUOtNG C O M B I N E D THROTTLE/H.P. FUEL SHUT-OFF COCK
C
9 PRESSURIZING V A L V E
D
9 I N T A K E GUIDE V A N E A N D BLEED V A L V E RAMS
FIGURE 3-36
Liquid fuelled phase I gas generator schematic. (Source: Rolls-Royce, UK)
Gas Turbines
The gas generator fuel control system is designed to satisfy the following basic requirements: 9 Full control of the gas generator at any power level while protecting it from damage should a malfunction occur 9 Control of fuel delivery according to engine demand 9 Shaft speed control 9 Control of fuel flow during rapid acceleration 9 Prevention of flame failure during rapid deceleration 9 Protection against sensor failure 9 Efficient burning of fuel 9 Control of fuel flow during a start 9 Maximum power limitation 9 Control mode indication 9 Stopping the gas generator
Dual Fuelled Gas Generators The gas generator fuel system (see Figure 3-37) provides a means of forwarding the fuel from the plant fuel system installation to the gas generator. The gas generator fuel control system is again designed to satisfy the following basic requirements: 9 Full control of the gas generator at any power level while protecting it from damage should a malfunction occur 9 Control of fuel delivery according to engine demand 9 Shaft speed control 9 Maximum gas temperature limitation 9 Control of fuel flow during rapid acceleration LIQUID SUPPLY
TO BURNERS
1
J PR~SSUR,SiN'r, VALVE
TO BURNERS POWER TURBINE SPEED
J
(N31
IJ
COMPRESSOR DELIVERY AIR IP2I EXHAUST GAS TEMPERATURE (ECT)
LP COMPRESSOR SPEED IN1)
AMBIENT
i
J.LIQUID F LOW CONTROL UNIT
I
LIQUID
I
J METERING ~ I VALVE I
i-
.
.
.
.
I ACTUATOR ; I =
ELECTRONIC CONTR 0 L UNIT
TEMPERATURE
(T1]
GAS F LOW CONTROL UNIT ~ , I I I GAS J ACTUATOR ~ METERING I. I I VALVE
POWER INPUT
PRESSURISING RELIEr VALVE
GAS SHUT-OFF COCK
HYDRAULIC POWER PACK
HP F I LTER
SPILL
T
VENT VALVE TO ATMOSPHER E GAS REGULATING VALVE
POWE R I NP UT
l FIGURE 3-37
TO ATMOSPHERE
SENSOR
3 WAY SOLENOID VALVE
Dual fuelled gas generator schematic. (Source: Rolls-Royce, UK)
GASSUPPLY
81
82
ProcessPlant Machinery 9 9 9 9 9 9 9
Prevention of flame failure during rapid deceleration Protection against sensor failure Efficient burning of fuel Control of fuel flow during a start Maximum power limitation Control mode indication Stopping the gas generator
Lubrication and Hydraulic Oil Supply Liquid fuelled Phase I gas generators As shown in Figure 3-38, the lubricating oil system, incorporates a combined sump and tank, the pressure and scavenge pumps and filters. It is situated within the wheelcase to form an integral part of the gas generator. Oil is pumped from the sump to an externally mounted cooler, from where it is returned to a pressure filter integral with the sump and wheelcase. From there it is piped to the main bearings and returned via the scavenge pumps and filters to the sump. A gas generator mounted fuel cooled oil cooler is available for liquid fuelled engines, if required, to replace the contractor-supplied external cooler. Each main shaft beating is protected by a thread-type filter, and each bearing housing incorporates air pressurized seals to retain the oil in the system. The wheelcase and bearings are vented to atmosphere through a breather with a centrifugal separator to minimize oil vapor loss. For continuous industrial duties, and on gas generators intended to operate by remote control, provision is made for a contractor-supplied auxiliary oil tank to be fitted in addition to the engine mounted tank. This tank is vented to the engine tank and must be provided with its own level checking arrangements. The Phase I liquid fuelled gas generator takes its hydraulic supply from the gas-generator-mounted HP fuel pump. A high pressure tapping is taken from the pump outlet to supply hydraulic power to actuate the compressor airflow control system. Gaseous and Dual Fuelled Phase I Gas Generators As with the oil system of the liquid fuelled Phase I gas generators, the lubricating oil system for the above generators is self-contained on the gas generator - but with the addition of a supplementary oil tank, an external oil cooler, temperature control valve and pressure filter (Figure 3-39). Depending upon the type of fuel control system selected, a separate hydraulic oil supply is required to operate the compressor airflow control system. The compressor airflow control system remains gas generator mounted and is normally powered by a Rolls-Royce supplied hydraulic pump which is fitted in place of the gas-generator-mounted high pressure fuel pump. All items not in the Rolls-Royce extent of supply must be designed to meet the requirements outlined in the Rolls-Royce Installation Manual. Phase II Gas Generators For Phase II gas generators (Figure 3-40) the scavenge and lubricating oil supply system components (with the exception of the lubricating oil connection block, including magnetic chip detectors in each beating scavenge line, and oil inlet pressure tapping, together with associated pipework for bearing feed scavenge and venting) are removed from the gas generator. The main contractor takes
Gas Turbines
83
GAS GENE RATOR
/
VALVES
:;
::o%~
CENTRE BEARING
IGV RAMV 9 HYDRAULIC SUPPLY FROM H.P. FUEL PUMP
AI
I 1
r o ATMOSPHERE
A2
v
.r---~J...L:-~. DEAERATOR
AUXILIARY OI L TANK
An alternative to the L8 connection Is a R o l l s - R o y c e supplied gas generator mounted pressure switch
COOLER FILTER OR STRAINER INTERFACE It
FLEXIBLE PIPE RELIEF VALVE PRESSURE ~ T C H PO~TIVE OISPLACEMENT PUMP ~GHT GLASS
FIGURE 3-38 Liquid fuelled Phase I gas generator schematic for lubrication and hydraulic oil schematic. (Source: Rolls-Royce, UK) the responsibility of supplying the remote mounted pumps, tanks, cooler and filters. All Phase II gas generators require a complete off-gas generator hydraulic supply system, including a high pressure hydraulic pump and drive motor. Again it is the responsibility of the main contractor to supply (apart from the compressor air flow control system, disconnect block, and interconnecting pipes), the high pressure hydraulic pump and drive motor, storage tank, cooler, heater (if required), filter, pressure relief valve, and interconnecting pipework and wiring.
84
Process Plant Machinery L9 ~ . , ~ j
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~ ,a,.t3G-=IkERATUI'
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i -~ I[ ..... "'-'~~:J--~ -] i ...................... /
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'
,, 9 \'o";"1 II
Lfl
!
I *~
ElrY
.___
.-.
.I
/
_
/'1
" ", "~'.. " . " ~ ' _ ~
I~
.
.
~
~----.'-,-.',---~
.
demalJve to the 1.8 connection is a Rolls.Royce BUpplIm:I gas ~
.
.
Y : ' " ~:.-)It'VEN '
'
I
ml~uw111dI=reuure switch
COOLER
r~ ~,
FILTEROR STRAINER INTERFACE
FLEXIBLE PiPE I
INSTRUMENTATION PiPS ] REMOTELY CONTROLLED TEMPERATURE V A L V E RELIEF VALVE
PRESSURESWITCH TEMPERATURE SWITCH O
OSITIVE DISPLACEMENT PUMP SIGHT GLASS
FIGURE 3 - 3 9 gaseous and dual fuelled Phase I gas generator lubrication and hydraulic schematic. (Source: Rolls-Royce, UK)
All items not in the Rolls-Royce extent of supply must be designed to meet the requirements outlined in the Rolls-Royce Installation Manual.
Performance
Characteristics
This item becomes particularly important in cases where power output at specific inlet temperature ranges is critical. Note that corrections for altitude affect developed horsepower considerably. (See Figure 3-41 and subsection on growth potential.)
Gas Turbines
GAS G :_NERATOP.
.,,/I
.,^~.~
.."1
B .E~D
"~
{l~ ~-:--~-.="-'-I
~.~,; ~M !1 I, ~ ...............
A
I
I
j.
STE,~
,
"
~
:,
~,,,~-~.~I
r , ~ . o.~T.A,NE.
v
FL "
~
~! .AG.~r,c ~,,,~ DrrEcroR
l
l
lJ PRIM NG FEED
I
INSTRUMENTATION PIPE
II
~ESrRICTOR
w
AV.EN.SE "
COOLER
*, .! ~t ~=.....~,~ v
:l ":
~'~
~
POSITIVE DISPLACEVENT PUMP
:
9
FIGURE 3-40
-
'
Phase H gas generator, lubrication and hydraulic schematic. (Source: Rolls-Royce, UK)
TABLE 3.4 The Effective Swallowing Capacities for Power Turbine/Gas Generator Matches Mk.1533 Mk.1534 Mk.1535-120 Mk.1535-160
352 344 338 332 332
sq sq sq sq sq
in. (A size) in. (B size) in. in. in.
Gas Generator and Power Turbine Matching
The operation of the gas generator is intrinsically linked with the power turbine it is driving. The gas generator working conditions are determined by the power turbine effective swallowing capacity. The power turbine capacity directly affects the power output available. Consequently, to achieve the outputs and s.f.c's quoted, it is essential that the gas generator is operated with a power turbine of the correct specific capacity at the design conditions. For retrofit situations, any gas generator may be operated into any power turbine by customer choice, and performance details for these combinations may be obtained from the manufacturer.
85
86
Process Plant Machinery
24" 22" Nominal shaft
output power x 1000 (shp)
~"'"----.---.......~~90,~.
~Oo~
20" 18" 16" 14"
1210
-:/o -:;o -io
b lb
~;o 3'0 do
~o
Air intake temperature ~ FIGURE 3-41 Performance characteristics of the industrial Avon. (a) Coberra 2348 Avon 1533/RT48 Variation of nominal shaft power with air intake temperature. No losses, gaseous fuel, sea level operation. (Source: Rolls-Royce, UK)
262422-
Nominal shaft
20-
output power x 1000 18(shp) 16141210
-~o -:;o -io
b
lb
:;o ~o do ~o
Air intake temperature ~ FIGURE 3-41 (b) Coberra 2448 Avon 1534/RT48 Variation of nominal shaft power with air intake temperature. No losses, gaseous fuel, sea level operation. (Source: Rolls-Royce, UK)
Gas Turbines
87
2826-
24-
~/ot
_
o
"~~
~
22-
Nominal shaft output power x 1000 20(shp)
181614-
12
-30
-20
-10
'0
1'0
20
30
40
50
Air intake temperature ~ FIGURE 3-41(c) Coberra 2558 Avon 1535-120/RT48 Variation of nominal shaft power with air intake temperature. No losses, gaseous fuel, sea level operation. (Source: Rolls-Royce, UK)
282624-
~_~
":L%%oo,
22-
Nominal shaft output power x 1000 20(shp)
1816-
1412
-~o -ko -i o
b
lb
~o ~o
;~o ~o
Air intake temperature ~ FIGURE 3-41 (d) Coberra 2648 Avon 1535-160/RT48S Variation of nominal shaft power with air intake temperature. No losses, gaseous fuel, sea level operation. (Source: Rolls-Royce, UK)
88
Process Plant Machinery
3230" Nominal overall thermal efficiency
28"
26242220-
1'0 12 1'4 16 1'8 Shaft power x 1000
6
2'0
22
FIGURE 3-41(e) Coberra 2348 Avon 1533/RT48 Variation of nominal overall thermal efficiency with shaft horsepower. (Source: Rolls-Royce, UK)
32"
30" Nominal overall thermal efficiency
28-
I
26 24-
2220
6
8'
1'0 1~
1'4 1~
1'8 ~
~2
Shaft power x 1000 FIGURE 3-41(f) Coberra 2448 Avon 1534/RT48 Variation of nominal overall thermal efficiency with shaft horsepower. (Source: Rolls-Royce, UK)
Condition Monitoring and M a i n t e n a n c e Features "On condition" is the preferred operating mode for gas turbines today. Modular design helps the operator take advantage of this feature. The fact that aircraft engine technology as developed by manufactures like Rolls-Royce, General Electric and
Gas Turbines
89
32" 30" 28"
Nominal overall thermal efficiency
26 2422-
20
6
13
1'0 12
1'4 16
Shaft power x 1000
1'8
2'0
'~2
FIGURE 3-41 (g) Coberra 2548 Avon 1535-120/RT48 Variation of nominal overall thermal efficiency with shaft horsepower. (Source: Rolls-Royce, UK)
32" 30"
Nominal overall thermal efficiency
28" 2624220
6
|
8
!
!
!
|
10 12 14 16 18 Shaft power x 1000
!
20
!
22
FIGURE 3-41(h) Coberra 2648 Avon 1535-160/RT48S Variation of nominal overall thermal efficiency with shaft horsepower. (Source: Rolls-Royce, UK) Pratt & Whitney is transferred directly into industrial models, accentuate the trends towards modularized construction and on condition maintenance. The object of condition monitoring is to achieve maximum economic installed life by the identification of evidence of deterioration of the gas generator, or one of its modules, at an early stage.
90
ProcessPlant Machinery By the recording and analyzing of the trend of changes in the values of the gas generator running parameters, evidence of deterioration can be identified sufficiently early to allow for planned maintenance. This procedure thus reduces unplanned non-availability of plant, minimizes inconvenience and offers cost savings.
Mechanical Monitoring This consists of monitoring: 9 Vibration levels (Figure 3-42) 9 Hydraulic and lubricating oil pressures 9 Metal wear (via magnetic chip detectors, Figure 3-42) 9 Oil consumption and acidity.
Performance Monitoring This consists of monitoring: 9 Variable inlet guide vane angle 9 Exhaust gas temperature 9 Exhaust gas pressure 9 IP spool speed 9 HP spool speed.
Borescope Inspection Borescope inspections (Figure 3-43) can be carried out when trend analysis (as the result of performance monitoring), external examination, or running symptoms, suggest the possibility of internal damage or deterioration. A number of borescope access ports are located on the gas turbine to facilitate examination of all rotating blades in the compressor and turbine assemblies. Static compressor blades. HP turbine nozzle guide vanes, the combustion chamber liners, and the fuel burners can also be inspected. The time required to conduct an examination is determined by the nature of the inspection. For example, inspection of the IP compressor inlet and outlet blades for suspected damage from a foreign object, could be expected to take 30 minutes.
Routine Servicing Gas turbines that run reliably on minimum routine servicing requirements have an advantage over those with more time consuming requirements, particularly in applications where continuous running is critical. Gas generators require to periodic inspection, maintenance and reconditioning to maintain maximum performance. The length of time before some form of reconditioning is required depends on such factors as: 9 Type and cleanliness of fuel 9 Operating environment 9 Efficiency of air filtration 9 Power output 9 Number and frequency of starts It is important that a servicing routine is established appropriate to the operation and duty to which the unit is applied. Servicing intervals may be varied as experience
Gas Turbines
'\
Vibration transducer /
/~'~~
/
/~\..~;-- ~--~-
~x~~
/
~
O L S IC A V E N G E B L O C K I
INLET
UNION
Magnetic chip detectors
FIGURE3-42 Typical positions of vibration transducer and magnetic chip detectors. (Source: Rolls-Royce, UK)
91
92
Process Plant Machinery
FIGURE 3.43(a) UK)
.
.
.
Borescope inspection ports on combustion chamber. (Source: Rolls-Royce,
.
.
.
.
.
.
.
.
.
1
Boreseopo
access
ports
r
.
.
.
.
.
.
.
.
.
.
.
.
.
1
I
i !
,PCOM~-s~O~ ROTOR
a
: :
HPCOM~RESSO~ HOTOR
L,~tR
TU~S,NES
FIGURE 3-43(b) Borescope access ports on gas turbine. (Source: Rolls-Royce, UK)
and unit accessibility allow, and should be increased in frequency if any adverse operating condition manifests itself. Occasional checks of the integrity of the unit should be made coveting security of equipment, evidence of leaks, and deterioration or wear of components. Initially a weekly schedule may be recommended, but the frequency can be reduced as satisfactory operation proceeds. Lubricating oil levels, and fuel oil filters should be checked at appropriate intervals depending on the particular installation and its mode of operation. Inspection of magnetic chip detectors, which may be carried out with the unit in operation, forms an important part of the condition monitoring routine for the gas generator. Chip detectors should be inspected frequently during commissioning and at any time when there is evidence of unit deterioration. During normal operation, an infrequent check, for example monthly, should suffice as long as there is evidence of satisfactory operation of the units. Weekly Checks
Examination of external pipework, conduit and electrical leads Check lubricating oil tank level
15 min 5 min
Monthly Inspections Examine air intake flare, VIGVs and visible compressor blades Inspect magnetic chip detectors
20 min 5 min
Gas Turbines
Six-Monthly/Annually Inspections Inspect filters Borescope inspection of internal components
15 min 7 hours
Compressor Cleaning Gas generators will often be operated in an atmosphere of fog, industrial pollution or high salt concentrations. Contamination of the compressor in these conditions, if allowed to go unchecked, will result in a loss of compressor efficiency leading to a reduction in gas generator performance. The effects of this contamination can be offset by compressor cleaning at regular intervals or when a deterioration in performance becomes apparent. The compressor can be cleaned by either of the following methods: 9 Crank/soak washing This method is used when operating conditions permit the gas generator to be shut down for cleaning 9 Fired fluid washing For occasions when it is inconvenient to shut down the gas generator
S P R A Y JETS
C R A N K WASH S P R A Y R I N G
9
a
FIGURE 3-44
Location of compressor wash system. (Source: Rolls-Royce, UK)
93
94
ProcessPlant Machinery
Environmental Considerations This facet of gas turbine design, operation and maintenance has grown immensely since the release of the first edition of this book. From vague regulatory restrictions that some "green person" on staff came around to check on, it is fast becoming part of the process and the rotating machinery engineer's daily bread. Those who ignore this trend risk major unnecessary and additional expenditure further down their plants' operating lives. What follows is a description of low NOx burners. There is considerably more material on this aspect, including details of other low NOx burner designs, as well as NOx reduction techniques such as water injection, low temperature gas turbine designs (that affect NOx production rates), ability to burn lower grade fuels (such as residual fuel), in Environmental Engineering and Management: Sustainable Development in the Power Generation, Oil & Gas and Process Plant Industries (Butterworth-Heinemann, 1998) Low NOx Burners*
Present nitrogen oxide (NOx) emission limits for gas turbines in Japan and the United States (notably California) can hardly be satisfied with commonly used NOx abatement techniques. Although the 75 ppm, 15% oxygen EPA bench-mark value is the prevailing US Federal standard, new projects now call for Best Available Control Technology (BACT) or specify an overall limit of < 25 ppm. Before investing in new plant equipment, operators of power plants must seriously consider the public's interest in having ecologically-clean as well as economically-sound power plants. Similar trends are now emerging in Europe and with the increasing environmental awareness in the industrial nations it may not be long before a < 40 ppm limit is imposed throughout Western Europe. Reducing harmful pollutants in the process of burning fossil fuel is a compromise: reducing carbon monoxide (CO) normally results in increased output of harmful NOx and vice versa. Whereas CO is very poisonous, NOx can cause acid rain and, in combination with sunlight, increase ground-level ozone to unhealthy values. Carbon dioxide (CO2) is another unwelcome pollutant; as some scientists believe, it can be a contributor to global warming. In addition, limited fossil fuel resources should be used as efficiently as possible - and this calls for an economical technology. In the past, steam or water injection was the most widely used method for NOx reduction. The drawback of this method lies in its efficiency loss (when injecting water) and in its increased costs, arising from the injection water treatment. Selective Catalytic Reduction (SCR), a further NOx reduction method, leads to even higher operating costs and lower efficiency. ABB now presents an ecological as well as economical solution to the problem, by introducing a new dual fuel dry low-NOx burner which is designed to meet NOx emission limits down to 9 ppm on natural g a s - without steam injection or SCR. A pragmatic approach combining theory, experimentation, and intuition resulted in a remarkably simple and efficient device: the double cone burner, also called EV-burner (Figure 3-45). The EV-burner- the result of research started in 1987 - is the latest step in ABB's development program. Rather than just concentrating on ever lower NOx levels, ABB has chosen a total solution that limits pollutants and at the same time increases energy efficiency. The "environmental" (EV) burner delivers NOx levels lower than 15 ppm, a value considered unrealistic in dry combustion only * Source: ABB Limited. Adapted with permission.
Gas Turbines
FIGURE 3-45
Dual fuel double cone burner. (Source: ABB, Sweden)
a short time ago. When gas is burned, ABB can guarantee 25 ppm dry, while 42 ppm can be achieved burning oil, with water injection. Years of research into new combustor technologies and extensive test programs have led to a system that also yields particularly low emission values for carbon monoxide (CO) and unburnt hydrocarbons (UHC). ABB's new EV-burner is cleaner and more efficient than the company's conventional burners, while being reliable and safe due to its simple design.
Operating Principle. The EV-burner is a dual fuel burner, this means that it can be operated on gaseous fuel, on liquid fuel, or in dual fuel operation. The burner is shaped like two half-cones (400 mm long, 150 mm in diameter) slightly offset sideways to form two inlet slots of constant width running the full length, (Figure 3-46). Combustion air enters the cone through these slots and gas is injected through a series of fine holes in their edges. With this arrangement fuel and air are intimately mixed (Figure 3-47).
FIGURE 3-46
Two half-cones, forming two inlet slots. (Source: ABB, Sweden)
95
96
Process Plant Machinery
FIGURE 3-47
Fuel and air are mixed at the inlet slots. (Source: ABB, Sweden)
FIGURE 3.48
EV-silo combustor. (Source: ABB, Sweden)
Gas Turbines
FIGURE 3-49
FIGURE 3-50
Annular combustor of the GT13E2. (Source: ABB, Sweden)
Developmentof the dry low-NOx burner. (Source: ABB, Sweden)
97
98
ProcessPlant Machinery
FIGURE 3-51
FIGURE 3-52
First-generation premix burner. (Source: ABB, Sweden)
Assembly of the EV-burner. (Source: ABB, Sweden)
Gas Turbines
FIGURE 3-53 The EV-burner, designed to provide low NOx emission and high fuel efficiency. (Source: ABB, Sweden)
FIGURE 3-54
Emissions measured for a given fuel to air ratio. (Source: ABB, Sweden)
The liquid fuels are sprayed into the cone through an atomizing nozzle. The combustion air enters the cone in the usual way, where it is mixed with the vaporized fuel. Water injection can also be applied. The design of the EV-burner is based on the vortex breakdown principle: A lean mixture leaves the cone and enters the flame. At the exit of the burner the vortex breaks down, forming a recirculation zone which stabilizes the flame in free space, keeping combustion temperatures and emissions low.
99
100
ProcessPlant Machinery As there is no flameholder body exposed to ignitable mixtures, and no danger of any flashback damaging the burner, it is a very reliable and safe design. Excess air is a feature of the EV-burner design, resulting in a flame temperature around 500 ~ lower than in a conventional burner and in a very low NOx level.
Applications. In the combustor there is no other burning mode than premix combustion. A group of burners act as a pilot system for all bumers. The EV-bumer is suited to both annular combustion chambers and to standard silo combustors (Figure 3-48) and it can be fitted to all of the company's gas turbine models. Existing units working with ABB's conventional burners can be retrofitted with EV-bumers. EV silo combustors are equipped with 19-54 bumers (number of burners depending on gas turbine type) and achieve NOx emission values below 25 ppm with natural gas. The EV-bumer can be fitted to the annular combustion system of the 24 MW GT10 turbine (Figure 3-49) Another application is with ABB's latest heavy-duty gas turbine of the proven GT13 family; the combustion chamber on the GT13E2 is of the single annular type, with 72 bumers arranged symmetrically around the annulus in four rows. With natural gas, the NOx value is well below 25 ppm.
Development History As the formation of nitrogen oxides in the process depends both on the firing temperature and the residence time of the fuel/air mixture in the combustion zone, conventional burners produce high amounts of NOx. By injecting water directly into the flame, it is possible to lower the temperature and thus reduce NOx emission to levels of 25-75 ppm. This traditional method of lowering NOx produced in gas turbines is used widely throughout the world. But this "wet" combustion process also considerably lowers energy efficiency, increases the output of CO due to incomplete combustion, uses large amounts of water, and causes higher maintenance costs.
Development of the Dry Low-NOx Emission Techniques. ABB's approach to this problem was the development of the dry low-NOx burner. The design and development history (Figure 3-50) leads from ABB's conventional diffusion burner to the advanced EV-bumer, covering various evolution p h a s e s - starting with the first-generation lean premix burner (Figure 3-51). Its concept is based on the simple principle of premixing air and fuel, with the maximum amount of excess air, before combustion. One burner consists of a premixing section, a separate combustion zone and a solid flame holder. The temperature of the flame is determined by the homogeneity of the air/fuel mixture and the amount of excess air in the combustion. About twice the theoretical amount of air required for combustion is u s e d - thus giving the method its name "lean premix". The flame temperature is at least 500 ~ lower than in the company's regular diffusion burner design and therefore NOx emission is reduced to 4 0 - 6 0 ppm. First commercial operation of these dry low-NOx burners occurred in 1984 on a Type 13 gas turbine, when the turbine's diffusion burner was replaced by a bundle of lean premix burners.
Gas Turbines
The first-generation lean premix bumers need careful monitoring and often use diffusion or pilot flames against combustion instabilities. Assembly of the EV-bumer is illustrated in Figure 3-52. A photograph of the EV-burner is illustrated in Figure 3-53. Figure 3-54 shows emissions measured for a given fuel to air ratio.
101
This Page Intentionally Left Blank
APPENDIX 3A
Gas Turbine Cycles*
From a thermodynamic point of view, a gas t u r b i n e - or gas turbine engine - is a machine that accepts and rejects heat at different energy levels and, in the process, produces work. While this work is converted to pressure and velocity energy in the aircraft jet engine, the commercial or industrial gas turbine is arranged to convert this work into shaft rotation or, more correctly, torque. The gas turbine (Figure 3A-l) consists of an air compressor and gas combustion, gas expansion, and exhaust sections. The gas turbines cycle is composed of four energy exchange processes: an adiabatic compressor, a constant-pressure heat addition, an adiabatic expansion, and a constant-pressure heat rejection. The four thermodynamic processes can be accomplished either in an open-cycle or a closedcycle system. The open-cycle gas turbine takes ambient air into the compressor as the working substance that, after compression, is passed through a combustion chamber where the temperature is raised to a suitable level by the combustion of fuel. It is then expanded inside the turbine and exhausted back to the atmosphere. Most industrial-type gas turbines work on this principle, and Figures 3A-2 and 3A-3 illustrate simple, open-cycle gas turbines. The use of two or more hot gas expansion stages makes it possible to produce the two-shaft turbine of Figure 3A-3. This configuration has greater speed flexibility than single-shaft machines. The closed-cycle gas turbine uses any gas as the working substance. The gas passes through the compressor, then through a heat exchanger where energy is added from a source, then expanded through the turbine and finally back to the compressor through a precooler where some energy may be rejected from the cycle. Perhaps the most important reasons why process plants use gas turbines are summarized as high system reliability and high combined energy system and process efficiency. Where the forced outages of a single driver can shut down an entire complex, highest reliability is a must. For projects involving process system modifications of a new process design, choosing the most reliable turbine or energy system rather than maintaining an already existing process design can result in significantly higher reliability and reduced financial loss due to excessive process shutdowns. With regard to the second point, high efficiency, the potential user may be confronted with an apparent mismatch between project needs and available machine sizes. In that case, it may save considerable money initially and over the life of the plant to revise the process design to match the best equipment and energy system available.
Simple-Cycle Gas Turbines Most gas turbines in the process industries are operating in base load, or continuous, service. Fuel costs are an important consideration in determining the type of prime * Source: General Electric Company, Schenectady, NY. Adapted with permission. 103
104
Process Plant Machinery
FIGURE 3A-1 Typicalindustrial gas turbine. (Source: General Electric Company, Schenectady, NY.)
FIGURE 3A-2 Single-shaft, simple open-cycle industrial gas turbine. (Source: General Electric Company, Schenectady, NY.)
FIGURE 3A-3 Two-shaft, simple open-cycle industrial gas turbine. (Source: General Electric Company, Schenectady, NY.) mover in these applications. However, there are hundreds of simple-cycle gas turbines installed in many areas of the world where fuel is relatively low cost, in underdeveloped areas, or in remote or harsh environments. Examples of simplecycle gas turbine installations are in Indonesia, the North Sea, the Sahara Desert, and the Alaskan North Slope. The advantages of simple-cycle gas turbines include:
Appendix 3A: Gas Turbine Cycles
105
Low capital cost Minimum installation cost 9 No external power or cooling water required 9 Minimum operating labor 9 Low maintenance costs 9 High reliability 9 High availability 9 9
The disadvantage of the simple-cycle gas turbine is its relatively low system efficiency and higher fuel costs, compared with gas turbine systems with exhaust heat recovery. Heat Recovery Cycles Usually the economics of gas turbines in the process industries depend on effective use of the gas turbine exhaust energy. The most common use of this energy is for steam generation in heat recovery steam generators (HRSG), unfired, as well as fired designs. However, the gas turbine exhaust gases can also be used as a source of energy for unfired and fired process fluid heaters and direct drying applications, as well as for combustion air for power boilers, reformers, or other process equipment. One of the more common gas turbine/heat recovery cycles is one where the exhaust energy is used to generate steam at conditions suitable for the process steam header (Figure 3A-4). The HRSG may be unfired or have supplementary firing to increase steam output. Power generation capability for these gas turbine-HRSG cycles, per unit of heat delivered to process, ranges from approximately 150 to 250 kW per million British thermal units (Btu)-net heat to process (NHP). Generation of steam at higher initial steam conditions than those required for process heat will allow use of a steam turbine in the cycle in addition to the gas turbine, as shown in Figure 3A-5. This combined cycle will result in a higher power generation-to-process heat ratio than the gas turbine-HRSG cycle shown in Figure 3A-4, with power generation in the range of 200 to 400 kW per million Btu NHP. A typical upper limit for steam conditions of unfired HRSGs is 1315 pounds per square inch gauge (psig)/950~ The HRSG steam temperature is usually 75 TYPICAL INDUSTRIAL GAS TURBINE CYCLES
AIR
FUEL
KW (OR HP)
PROCESS STEAM
0 b~ V
HRSG FIGURE 3A-4 Typical industrial gas turbine cycle employing heat recovery steam generator (HRSG). (Source: General Electric Company, Schenectady, NY.)
106
Process Plant Machinery
TYPICAL INDUSTRIAL GAS TURBINE CYCLES
TURBINE INLET STEAM i
i
O FUEL
HRSG
KW
PROCESS STEAM
FIGURE 3A-5 Combined cycles employ both gas and steam turbines. (Source: General Electric Company, Schenectady, NY.) to 100 ~ or more below the gas turbine exhaust gas temperature. Fired HRSGs have been applied with steam generation pressure and temperature as high as 1525 psig/955 ~ A multiple-pressure-level HRSG combined-cycle system is shown in Figure 3A-6. This arrangement is common for unfired and moderately fired (up to approximately 1200 ~ HRSGs. The multipressure-level HRSG results in increased recovery of the gas turbine exhaust energy compared with an unfired, singlepressure-level HRSG system, thus increasing the cycle thermal efficiency. The steam turbine in a combined cycle may be a noncondensing or a condensing design, depending primarily on process heat requirements. The steam turbine design shown schematically in Figure 3A-6 provides considerable cycle flexibility in industrial process application. The condenser provides a heat sink for HRSG TYPICAL INDUSTRIAL GAS TURBINE CYCLES
M E T 8TEAM
FUlL
{W KW
8TEAM
FIGURE 3A-6 Cycle flexibility is provided through utilization of a two-level heat recovery steam generator and admission-extraction steam turbine. (Source: General Electric Company, Schenectady, NY.)
Appendix 3A: Gas Turbine Cycles
steam-generating capability in excess of that extracted from the turbine for process use. Furthermore, the optional admission capability permits the introduction of lower pressure process steam into the turbine for expansion to the condenser during times of excess low-pressure steam. Even though gas turbines are not available in an infinite number of ratings, the application of a helper steam turbine may permit utilization of the capability of standard, proven gas turbine units in certain mechanical drive or generator drive applications. In this cycle, shown schematically in Figure 3A-7, the helper steam turbine can argument gas turbine power generation as load requirements vary. In most instances, depending on horsepower and steam conditions, the helper steam turbine can be mounted on the gas turbine base and shipped as an integral driver unit to minimize installation costs. For instance, base-mounted, noncondensing helper steam turbines for a medium-to-large gas turbine may range from small, singlestage turbines to multistage, multivalue turbines rated up to 8000 horsepower (HP). Control of the steam turbine helper and the gas turbine is integrated into a single governing system. In the regenerative-cycle gas turbine, exhaust heat is recovered by heating the turbine's combustion air after compression, but before it enters the combustion chambers to reduce gas turbine fuel. The Figure 3A-8 schematic diagram shows a regenerative-cycle gas turbine followed by a low-pressure process HRSG. One of the consequences of the low fuel consumption of the regenerative-cycle gas turbine is a reduction of the gas turbine exhaust gas temperature to approximately 600 ~ This cycle arrangement can be an option when a relatively small amount of process steam is required. In petroleum refineries, power generator drives predominate, as illustrated in Figure 3A-9. Petroleum production facilities, both onshore and offshore, make use of gas turbines for gas reinjection compressor, waterflood pumping, and power generator drive applications. In modem ethylene production facilities, three and sometimes four major compressor strings require variable speed drivers. The charge gas compressor driver is typically an extraction-condensing steam turbine that supplies 600 psig steam to
TURBINE INLET STEAM
I
S T E A M TURBINE PROCESS STEAM HEADER
FIGURE 3A-7 Starter~helper steam turbine arrangement in a combined gas turbine cycle. (Source: General Electric Company, Schenectady, NY.)
107
108
Process Plant Machinery ..- P R O C E S S STEAM
§
v
HRSG
REGENERATOR
FIGURE 3A-8 Regenerative-cycle gas turbine followed by a low-pressure process HRSG. (Source: General Electric Company, Schenectady, NY.) 8 5 0 TO 1 2 5 0 PSIG
Bo,LE.s ~_.w
~Kw
W
6 0 0 PSIG
I L
Kw
)
I
,U,
150 PSIG
5 0 PSIG
.w
ux
ox
r"
TO PROCESS PROCESS HEAT J RECOVERY BOILERS
FIGURE 3A-9 Typical refinery power plant schematic shows power generator drives predominating. (Source: General Electric Company, Schenectady, NY.) the cracking process. Additional process steam can be furnished by the extractioncondensing propylene compressor driver or sometimes by a noncondensing ethylene driver. However, typical energy balances require large quantities of power boiler fuel and large blocks of condensing power. To reduce overall fuel consumption and cooling water requirements, gas turbines have been selected as the propylene and ethylene compressor drivers for a number of large plants. A typical layout is shown in Figure 3A-10.
Appendix 3A: Gas Turbine Cycles
109
1250 TO 1500PSIG
PROCESS HEAT RECOVERY
G
POWER BOILERS
z_
600 PSIG J
150 PSIG FIGURE 3A-10 NY.)
~
AUX
TOPROCES~
Ethylene plant with combined cycle drives. (Source: General Electric Company, Schenectady,
PROCESSSTEAM ~I~ RETURNOIL COOLER ,
..
9
I EXHAUST
_
o
--
SUPPLEMENTARY
~
FIRING
-
OIL HEATED
~k
x,
1, '
~ ~
GASTURBINE
KW
FIGURE 3A-11 Special process requirements may employ gas turbines in a variety of arrangements. (Source: General Electric Company, Schenectady, NY.) One major plastics plant uses the gas turbine exhaust to first heat the oil used in a high-temperature distillation process. The exhaust gases then pass to a supplementary fired heat recovery steam generator where plant process steam is produced. Figure 3A-11 depicts this sequence. Heavy-duty industrial gas turbines are available as both simple cycle and regenerative cycle machines. These can be single- or two-shaft configurations for both
110
Process Plant Machinery TABLE 3A.1 Data
Performance Specifications for Major Mechanical Drive Gas Turbines: 1988
Model a
LM5OOO-PC
LM1600
LM2500-PH STIG b
Yeaff ISO rating continuous (HP) d Heat rate (Btu/HP-hr) e Air flow (Ib/sec) f Turbine speed (rpm)g Pressure ratio h Turbine inlet temp. (~ i Exhaust temp. (~ i Dry weight (--rib) Length x width x height (,~ft)
1986 46,210 7,040 265 3,600 30.0 834 89,000 58 x 11 x 12
1988 16,500 7,120 96 7,000-9,000 21.5 1,344J 880 7,000 17 x 6 x 6.5
1986 36,000 6,321 151 3,600 20.0 1,475J 952 10,500 21 x 8 x 7
All ratings without losses and zero humidity, except for STIG units. aSeries designation bAll ratings on gas fuel, 4"/10" H20 inlet/exhaust losses CFirst year unit was available d6,000 hours per year or more eLower heating value fAt base load gOutput shaft speed hOverall compressor iAt continuous rating JPower turbine inlet temperature
TABLE 3A.2
Performance Specifications for Major Electric Power Generator Drive Gas Turbines: 1988 Data
Model a
Yearb ISO base rating (kw) c Heat rate (Btu/kw-hr) d ISO peak rating (kw) e Heat rate (Btu/kw-hr) d Pressure ratio f Air flow (lb/sec) g Turbine speed (rpm) h Turbine inlet temp. (~ Exhaust temp. (~ Dry weight (-~lb) Length x width x height(~ft)
PG5371(PA)
PG6541(B)
PG7111(EA)
PG9161(E)
PGLM5OOO-PC
1987 26,300 11,820 28,150 11,730 10.2 270 5,100 1,755 901 570,000
1978 38,340 10,860 41,400 10,780 11.8 301 5,100 2,020 1,003 700,000
1976 81,700 10,610 89,200 10,580 12.4 641 3,600 2,020 992 1,070,000
1987 116,900 10,310 126,100 10,280 12.1 889 3,000 2,020 985 1,900,000
1986 33,090 9,860 30.0 278 3,000 834 314,200
132x71x31
115x77x39
115 x 19 x 3 4
123x24x34
All units are package power plants; all ratings are on gas fuel. aSeries designation bFirst year unit was available c6,000 hours per year or more dLower heating value eGood for up to 2,000 hours per year fOverall compressor gAt base load hOutput shaft speed iAt base rating
119x20x31
Appendix 3A: Gas Turbine Cycles
mechanical and generator drive applications. Mechanical drive gas turbines cover a range from 14,000 HP to 45,000 HP at International Standards Organization (ISO) conditions burning gaseous fuels. One leading U.S. manufacturer has five different models that can be used as mechanical- or power generation-type turbines. Tables 3A.1 and 3A.2 indicate cycle, fuel used, output, heat rate, and shaft speed of both mechanical drive and generator drive gas turbines available from this manufacturer. Of course, other manufacturers may produce machines in different or overlapping size categories.
111
APPEN DIX 3 B
Life Cycle Usage Estimated Hot-Gas Path Parts Lives (Peaking Duty) The application, cyclic or continuous duty, starting frequency and time, internal temperatures as a result of loading duty, and type of fuel used all determine parts life and maintenance cost. A peaking plant has many thermal cycles, resulting in a requirement to inspect the unit on a shorter fired-hour basis than is required with a continuous-duty unit. The normal variance of peaking units will be one start per one fired hour to one start per six fired hours. Within this range, the planned hot-gas path inspection should take place each 6,000 to 10,000 fired hours of operation, depending on the evaluations made at combustion inspections, factoring in the effects of fuel and metal temperature. This is approximately one third of the fired hours expected to be attained on a continuous-duty unit before an inspection will be routinely scheduled. At this inspection, the affected parts may be replaced for minimum downtime or may be repaired and reinstalled in the unit with a longer outage. The parts under consideration for this inspection interval are transition pieces, first-stage nozzles, and second-stage nozzles. The combustion system parts will be repaired, using these criteria at approximately 1000- to 1500-hour intervals, or once per year. These parts include the combustion caps, liners, and cross-fire tubes. Turbine buckets should require little repair except for foreign object damage caused by ingesting external material or for restoration of bucket tip clearances for continued efficiency. The inspection interval hours stated previously for the hot-gas path parts will be maximized by optimization of the combustion system. It is important that the maintenance program be used to maintain proper control settings and that the combustion parts be kept in proper working order. The fuel nozzles, for example, will have a direct effect on the liners, transition pieces, and nozzles. Balanced firing temperatures will maintain minimum temperature differentials and assure that one combustion chamber and nozzle segment will not experience excessive temperatures. This occurs because the transition pieces and first-stage nozzles are exposed to the direct discharge from the combustion process. First- and second-stage turbine nozzles can be repaired several times with a resultant extension of the total life. The economic determination of repair versus replacement will govern the feasibility and number of times the nozzles are repaired.
Operating Factors Affecting Component Lives The factors having the greatest influence on the life of parts for any given machine are type of fuel, starting frequency, load duty, environment, and maintenance practices. 112
Appendix 3B: Life Cycle Usage
Fuel The effect of the type of fuel on parts life is associated with the radiant energy in the combustion process and the ability to atomize various liquid fuels. Therefore, natural gas, which does not require atomization, has the lowest level of radiant energy and will produce the longest life of parts. Diesel fuels will produce the next highest life, and the crude oils and residual oils, with higher radiant energy and more difficult atomization, will produce shorter life of parts, as shown in Figure 3B-1. Contaminants in the fuel also affect the maintenance interval. This is particularly true for liquid fuels where dirt results in accelerated replacement of pumps, metering elements, and fuel nozzles. Contaminants in fuel gas systems can erode or corrode control valves and fuel nozzles. Filters must be observed and changed when practical to assure against the carrying of these contaminants through the fuel system. Clean fuels will invariably result in reduced maintenance and extended parts life.
Starting Frequency and Time Each start, stop, and load change of a combustion gas turbine subjects its hot-gas path parts to thermal cycling. Control systems are designed, programmed, and adjusted to apply temperatures that are compatible with material properties to minimize required maintenance from this cycling effect. However, a unit in a peaking application will demonstrate parts lives (Figures 3-19 and 3-20) that are shorter than a similar unit in base-load continuous duty service, as with any equipment subject to cycling conditions. The normal programmed starting time for a peaking unit is designed to minimize transient thermal stresses and maximize parts life. Fast start/load programs are available that compromise these objectives and are therefore used primarily in emergencies or to periodically demonstrate fast starting capability. These effects are shown in Figure 3B-2 as a function of starting frequency. The maintenance penalty for fast start/load occurs mainly from the load application, since a fast start from standstill to rated speed occurs in approximately 2.25 minutes, with a temperature change of 800 ~ maximum; but a fast load application is accomplished in 30 seconds, with a resulting temperature change of 1000 ~ The differences in rate of temperature change are obvious and explain the increased maintenance cost.
'F 3
MAINTENANCE FACTOR
RESIDUAL ,
,
-CRUDE
2 DISTILLATE ,
1
,
NATURAL GAS
0 FUEL TYPE
FIGURE 3B-1 Effect of fuel on gas turbine maintenance. (Source: General Electric
Company, Schenectady, NY. )
113
114
Process Plant Machinery
MAINTENANCE FACTOR
j
FAST START, NORMAL LOAD OR NOR
B
3
NORMAL START, NORMAL LOAD
I
I
1/1000
I
1
1/100 1/10 STARTS / FIRED HOUR
1/1
FIGURE 3B-2 Effect of number of starts on gas turbine maintenance. (Source: General Electric Company, Schenectady, NY.)
4
'-"
BASE ~ LOAD
3 -MAINTENANCE COST FACTOR
w I ' PEAK I LOAD i
,
,
2 --
Ii~ ~
i
I
"-
/,
I I
1 "-
9 I
l
80
IO0
I I I
10S
I
I
I
110
I
,
I
11S
120
% OF BASE LOAD RATING AVERAGE PEAK LOAD 108%
FIGURE 3B-3 Effect of load duty on gas turbine maintenance. (Source: General Electric Company, Schenectady, NY.)
Load Duty Utility units are usually supplied with a designated peak and peak reserve rating higher than the normal base rating. These ratings will affect the life of hot-gas path parts due to the higher firing temperatures that exist in the unit (Figure 3B-3). These ratings are used to allow the operator flexibility in the use of this equipment for the system needs. Usual daily peaking applications justify loading the units only to the base rating, with peak and peak reserve capabilities assigned for additional flexibility in emergency conditions. Maintenance requirements are affected by the assignment of loading temperatures, and the economics of use must be balanced to arrive at overall use factors.
APPENDIX 3C
Specific Maintenance Inspections The combustion gas turbine, as does any rotating power equipment, requires a program of planned periodic inspection, with repair and replacement of parts to achieve optimum availability and reliability. The major structural components of the heavy-duty combustion gas turbine are designed according to long-established standards derived from steam turbine design and manufacture. Major differences occur between the steam turbine and the combustion gas turbine due to the fact that the combustion gas turbine is a complete, self-contained, prime mover. This combustion process to develop energy does not require a boiler with its associated limitations; therefore, the cycle temperatures are considerably higher. The parts that are unique to the gas turbine because of this feature are combustion caps, liners, and transition pieces. These, along with the turbine nozzles and buckets, are referred to as the "hot-gas path" parts. The inspection and repair requirements of the gas turbine lend themselves to establishing a pattern of inspections, starting with very minor work and increasing in magnitude to a major overhaul, and then repeating the cycle. These inspections can be optimized to reduce unit outages and maintenance cost for the user's specific mode of operation, while maintaining maximum availability and reliability. Inspections can be classified as operational or shut down. The operational inspections are used as indicators of the general condition of the equipment and as guides for planning the disassembly maintenance program. The entire scope of inspections can be described as standby, running, combustion, hot-gas path, and major.
Standby Inspection Standby inspections pertain particularly to gas turbines used in intermittent service, such as peaking and emergency duty. Starting reliability is of prime concern, as a delay in starting usually means that the demand for the unit has passed. This includes routine servicing of the battery system, lubrication, changing of filters, checking oil and water levels, cleaning relays, checking device calibrations, and other general preventive maintenance. This servicing can be performed in off-peak hours without interrupting the availability of the turbine. A periodic test run is an essential part of the standby inspection.
Running Inspection Running inspections consist of the observations made while a unit is in service. The turbine should be observed on a programmed schedule, which should be established as part of the unit maintenance program consistent with the operator's requirements. Operating data should be recorded to permit an evaluation of equipment performance and maintenance requirements. Typical running inspections (Table 3C.1) 115
116
ProcessPlant Machinery
Typical Running Inspections Recommended for Gas Turbines
TABLE 3C.1
Load versus exhaust temperature Vibration Fuel flow and pressure Exhaust temperature control Exhaust temperature variation Start-up time
include load versus exhaust temperature; vibration; fuel flow and pressure; exhaust temperature control and variation; and startup time. The general relationship between load and exhaust temperature should be observed and compared with previous data. Ambient temperature and barometric pressure will have some effect on the absolute temperature level. High exhaust temperature can be an indicator of deterioration of internal parts, excessive leaks, axial-flow compressor fouling, or improper control settings. Initial startup data should be used as the reference point for checking. Power loss resulting from deteriorated parts or leaks may require disassembly of the turbine to restore power. This can be done with on-site labor and equipment. Loss due to dirt fouling of the axial flow compressor can usually be restored by cleaning the compressor while in service. This is accomplished by injecting 10 to 20 pounds of mild abrasives such as hard rice or screened crushed nut shells into the compressor inlet. A successful cleaning will reduce the exhaust temperature for a given load and will increase the compressor discharge pressure. If the need to clean the compressor is frequent, the causes of the fouling condition should be determined and corrected. The vibration level of the unit should be observed and recorded. Minor changes will occur with changes in operating conditions. However, major changes, or a continuous trend to increase, indicate that corrective action is required. The fuel system should be observed for general fuel flow versus load relationship. Fuel pressures through the system should be observed. Changes in fuel pressure can indicate that fuel nozzle passages are plugged or fuel metering elements are damaged or out of calibration. Probably the most important control function to be observed is the exhaust temperature-fuel override system and its backup overtemperature trip system. Routine verification of the operation and calibration of these devices will minimize wear on the hot-gas path parts. The variation in turbine exhaust temperature should be measured. An increase in temperature spread indicates combustion deterioration or fuel distribution problems. If not corrected, reduced life of downstream parts can be expected. Startup time (when the gas turbine is new) is an excellent reference against which subsequent operating parameters can be compared and evaluated. A curve of starting parameters of speed, fuel signal, exhaust temperature, and critical sequence benchmarks versus time from the initial start signal will give a good indication of the condition of the control system. Deviations from normal conditions help pinpoint impending trouble, changes in calibration, or damaged components.
Combustion Inspection This is a shutdown inspection to inspect combustion liners and fuel nozzles; these are recognized as the first parts requiring replacement and repair for a good maintenance
Appendix 3C: Specific Maintenance Inspections
117
program. Proper attention to these items will optimize the life cycle of downstream parts, such as turbine nozzles and buckets. Figure 3C-1 illustrates the section of a typical unit that is disassembled for this inspection. The combustion liners and fuel nozzles should be removed and replaced with new or repaired liners and new or clean fuel nozzles. This method of inspecting allows for minimum unit downtime and maximum utilization of manpower. A visual inspection of the transition pieces and nozzles (first-stage) at this time optimizes the scheduling of the hot-gas path inspection. This visual inspection is accomplished by an optical instrument called a "borescope," which is inserted through the combustion liner area to allow examination of the transition pieces and first-stage nozzle. The typical intervals for combustion inspections are shown in Tables 3C.2 and 3C.3.
FIGURE 3C-1 Major areas of gas turbine inspection. (Source: General Electric Company, Schenectady, NY.)
TABLE 3C.2
Typical Inspection Intervals for Gas Turbines in Continuous Duty*
Combustion
Hot Gas
Major
A. Gas Distillate
10,000-14,000
-
30,000-40,000 20,000- 28,000
B. Gas Distillate
8,000-10,000 5,000- 7,000
16,000- 20,000 10,000-14,000
30,000-40,000 20,000- 28,000
*One start per 1,000 fired hours. A. Continuous process- Gas turbine outage results in process shutdowns. Loss in production exceeds savings from optimum maintenance. B. lnterruptable process- Scheduled outages coincide with other equipment inspections (i.e., underwriter requirement). Maintenance costs optimized.
118
Process Plant Machinery
TABLE 3C.3 Typical Inspection Intervals Recommended for Gas Turbines in Peaking Service Starts per Fired Hour
1-3 1-5 1-10
Base Temperature Control
Peak Temperature Control
2,100 3,000 3,450
700 1,000 1,150
Fuel: Natural gas/light distillate TABLE 3C.4 Time
Recommended Inspection Intervals Linked to Fired Hours, Starts, and Elapsed i
Whichever Comes First
Combustion inspection Natural gas Distillate Hot-gas path Natural gas Distillate Major inspection Natural gas Distillate
TABLE 3C.5
Fired Hours
Starts
Time
8,000-10,000 5,000- 8,000
300-400 300 -400
Annually
20,000-24,000 20,000- 24,000
600-800 600- 800
-
42,000-48,000 42,000-48,000
1,600-2,400
6 Years
Recommended Work Scope for Major Inspections* Action
Inspection For
Bearings, seals
Clean
Blading
Clean manually, loose parts check
Buckets
Remove from rotor-grit blast- loose parts check Clean, loose parts check in dovetail area
Wear, fouling, leaks, wiping, scoring, deterioration of babbitt Foreign object damage, erosion, corrosion, cracks, fouling Foreign object damage, cracks, erosion, corrosion
Part
Turbine wheel Journals and seal fits Inlet system and exhaust system
Inspect, repair, paint
Cracks in dovetail area Wear, scoring, wear on seal fits Corrosion, cracks, loose parts
*Step 1: Same as for combustion and hot-gas-path inspections; Step 2: Remove remaining upper half casings and beating covers; Step 3: Remove rotors.
Hot-Gas Path Inspection This inspection includes the work necessary for a combustion inspection plus the removal of the upper half turbine shell, and on applicable turbines, removal of the upper half combustion chamber wrapper.
Appendix 3C: Specific Maintenance Inspections TABLE 3C.6 Estimated Time to Perform Recommended Gas Turbine Inspections on Popular GE Sizes
Inspection
Model
Hours
Work Shifts
Combustion
5001 6001 9001 5001 6001 9001 5001 6001 9001
160 240 480 480 672 1,120 1,280 1,600 2,560
5 6 10 10 12 20 20 25 40
Hot-gas path
Major
The inspection involves all hot-gas path parts. These include the turbine buckets, shrouds, nozzles, transition pieces, and exhaust hood turning vanes. The typical recommended intervals are given in Table 3C.4.
Major Inspection A major inspection includes the work items outlined for combustion inspections and hot-gas path inspections; it also includes "laying the turbine on the half shell," and completely inspecting the axial flow compressor stator and rotor parts, turbine buckets and shrouds, bearings, and seals. Table 3C.5 gives the recommended work scope for major inspections and Table 3C.6 gives estimated man-hours and work shifts for popular gas turbines.
119
This Page Intentionally Left Blank
Chapter 4 Gas Engines Gas engines are internal combustion engines and incorporate many of the operating principles of modem automobile engines. They are reciprocating machines and come in a variety of sizes and configurations. Gas engines, as the term implies, operate on gaseous fuel that is ignited by an electric spark. Smaller gas engines, typically in the 200- to 400-kilowatt (kw) range, are frequently used to drive emergency generators or fire water pumps. Integral gas engines, with compressor cylinders mounted on the engine crankcase (Figure 4-1), are often found in the larger size ranges, up to several thousand kilowatts of power output. Again analogous to automotive engines, reciprocating gas engines are manufactured either as two-stroke-per-cycle or four-stroke-per-cycle machines. With twice as many power strokes per revolution, the two-stroke-per-cycle engines tend to be smaller than four-stroke-per-cycle versions built for the source power output. Moreover, the two-stroke machine is generally less complex because it can dispense with valves and their associated mechanisms. Reciprocating gas engines display typical efficiencies in the 28 to 42 percent range. The upper portion of this range belongs to turbocharged engines, whereas the lower range is populated by naturally aspirated machines. Turbocharged equipment uses exhaust gas to drive a blower that forces combustion air through a suitable heat exchanger into the intake manifold. This cooled, compressed air is used for combustion and, in certain engine types, "scavenging." Scavenging air purges exhaust gases from engine cylinders before combustion air is admitted. Reciprocating gas engines suffer from a few disadvantages that must be considered when selecting process plant equipment: 9 They require a fair amount of competent surveillance and routine maintenance. Minor overhauls are typically needed after about 2500 operating hours. Major overhauls should be anticipated every three to five years. 9 Their low speed, typically in the 180 to 900 RPM range, requires step-up gears for such process duties as pump drives. 9 Fluctuations in the heating value of the fuel may require constant adjustment of spark timing and could also require derating of the power output capability. 9 The sulfur content of gas engine fuels may adversely influence the extent and frequency of maintenance required. Gas engines typically consume between 6500 and 8000 British thermal units (Btu)/brake horsepower (BHP)-hour. * Source: Cooper-BessemerReciprocating, Grove City, PA. Adapted by permission. 121
122
Process Plant Machinery
FIGURE 4-1 Large gas engine (ten cylinders) with built-on reciprocating compressor cylinders. (Source: Cooper-Bessemer Reciprocating, Grove City, PA) TWO-STROKE GAS ENGINES
It could be stated that gas engines are basically blown-up versions of the conventional automotive engine with just a few important modifications" they are slow-running, they use a gaseous fuel/air mixture instead of the liquid fuel/air mixture typically found in most automotive engines, and they are often integrally arranged (or combined) with the process gas compressor cylinders that they are driving. Since the combustion process takes place inside the cylinder, gas engines belong to the family of internal combustion engines. Like their cousins the automotive gasoline and automotive diesel engines, gas engines are either of the two-stroke or of the four-stroke per cycle variety. Two-stroke engines have one power stroke for every full revolution of the crankshaft, whereas in four-stroke engines, only every second revolution is accompanied by a power stroke. Four-stroke engines have inlet and exhaust valves; two-stroke engines have inlet and exhaust ports. Each type of engine has a spark plug; the two-stroke engine also incorporates a fuel admission valve. With two-stroke engines considerably more prevalent in the process industries, we will confine our considerations to this type of engine. Modem units are typically configured as shown in Figure 4-2. This cutaway view illustrates a Cooper-Bessemer model GMVH, essentially a combination V-type
Gas Engines
123
gas engine and horizontal compressor built into one compact unit. The typical twostroke engine is built in units of six, eight, ten, and twelve cylinders; the number of compressor cylinders varies according to requirements and arrangements yield any combination of volume and pressure within the rating of the engine. The GMVH is a two-stroke-cycle loop-scavenged V-type engine, designed to use natural gas as a fuel. Scavenging air is supplied by an exhaust-driven turbocharger and a highly developed control system maintains optimum combustion under varying conditions of load and speed. Perhaps the best way to study this engine is to review its principal components identified in Figure 4-3. The engine base (Figure 4-3, item 17) is a complex iron casting that forms the backbone of the entire structure. The main journal beatings are vertically split, leaving one side of the base open for easy access to the bearings and the crankshaft. Once the alignment and fit of the main bearings have been established, the caps and outer shells are removed and the crankshaft (Figure 4-4, item 24) is installed. The bearing halves and caps are then installed and properly tightened once more. The crankshaft is a high-quality steel forging, machined with great precision and having a very fine finish on all bearing surfaces and fillets. Oil holes are drilled from the main journals to the adjacent crank pins to transmit large quantities of lubricating oil from the pressure-fed main beatings for lubrication of the connecting rod beatings and cooling of the power pistons. The remaining connecting rod assembly depicted in Figure 4-4 is installed next. The main connecting rod, item 7, is used to drive the compressor piston. The power connecting rods, item 11, are articulated to the main connecting rod in a manner similar to that used in radial aircraft engines. In an engine of this type, the load on the power connecting rod is always compressive.
FIGURE 4-2
Cutaway view of a modern gas engine reciprocating compressor combination. (Source: Cooper-Bessemer Reciprocating, Grove City, PA)
124
Process Plant Machinery
FIGURE 4-3 Cross-section of a Cooper-Bessemer-type GMVH gas engine compressor. 1 - p l a t f o r m and railing; 2 - j e t cell igniter; 3 - gas injection valve; 4 - load balancing valve; 5 - j a c k e t water outlet header; 6 - insulated exhaust manifold; 7 - j a c k e t water inlet header; 8 - power cylinder head; 9 - power piston; l O - p o w e r cylinder; 1 1 - a i r inlet manifold and intercooler; 12-layshaft; 13-articulated power rods; 14 - crankcase relief valve; 15 - lube oil pressure regulator; 1 6 - lube oil suction header; 1 7 - engine base (crankcase); 1 8 - m a s t e r rod; 19-crosshead and shoe; 20-crosshead guide housing and support; 21-crosshead balance weights; 22-crosshead diaphragm and packing; 23-compressor cylinder rod packing; 24 - valve cap; 25 - plug-type suction valve unloader; 26 - unloader volume bottle; 27 - compressor cylinder head; 2 8 - compressor cylinder body; 29-compressor piston; 3 0 - p i s t o n rod and nut. (Source: Cooper-Bessemer Reciprocating, Grove City, PA) The power cylinders (Figure 4-3, item 10) are next mounted on top of the base. The cylinder is a high-strength iron casting of some complexity, and Figure 4-3, item 10 shows clearly the air induction ports as well as the higher exhaust ports and the passages leading to each. The self-contained jacket, providing a flow of cooling water, is also shown. The bore of the power cylinder is chrome plated and then honed to a high degree of precision. It is common for such cylinders to be in continuous operation for several years without significant wear of the cylinder bore. The power piston (Figure 4-3), item 9) is an oil-cooled trunk-type piston using four compression tings and two oil control rings. Figure 4-4 shows these ring grooves as items 9 and 8, respectively. The piston pin housing, Figure 4-4, item 13, is a separate casting, bolted into the piston, containing the bronze bushing for the pin at the upper end of the power connecting rod. The space between the pin housing and the piston crown receives a continuous flow of lubricating oil for cooling. This oil comes through a longitudinal drilled hole in one flange of the power connecting rod and up through one of the vertical tubes in the pin housing. Oil is continually
Gas Engines
125
FIGURE 4-4 Gas engine reciprocating compressor crankshaft assembly showing power pistons and compressor crosshead. (Source: Cooper-Bessemer Reciprocating, Grove City, PA) drained through the other tube and down through the other side of the connecting rod. This cooling prevents excessive thermal strains in the piston even though the engine is operating at a high level of output. The cylinder head, (Figure 4-3, item 8) is an open-style iron casting with a cover plate to form a complete water jacket. The fuel gas injection valve is located in the center of the head, and there are two spark plugs, one on either side of the gas valve. The crosshead guide, (Figure 4-3 item 19), is now installed. This serves both as a mount for a compressor cylinder and as a stationary slide for the crosshead at the outer end of the main connecting rod. The crosshead, Figure 4-4, item 6, has separate top and bottom shoes made either of aluminum or of cast iron with a babbitt overlay. These shoes are adjustable to fit with the proper clearance within the bore of the crosshead guide. The crosshead pin, item 5, is used to connect the crosshead to the eye in the outer end of the main connecting rod. The opposite side of the crosshead receives the end of the piston rod that drives the compressor piston. The piston rod packing will be discussed in more detail later. The last major component to be added to the basic mechanical structure is the flywheel. The flywheel is bolted and doweled to a flange on the end of the crankshaft. Being of generous size, the flywheel serves to maintain the engine speed essentially constant, in spite of the variable turning effort of the power and compressor cylinders.
Gas Engine Compressor Support Systems Pressure lubrication is supplied to practically all lubrication points of the engine except the power and compressor cylinders, which are lubricated by a force-feed lubricator system. The external lubricating oil system will vary according to installation requirements. Figure 4-5 is a schematic diagram of a typical system.
126
Process Plant Machinery
The wet-sump-type engine base serves as a reservoir for the lubricating oil. A sight gauge, located in the forward end cover, indicates the oil level at all times. The pump suction header, in the bottom of the base, is perforated to form a strainer that prevents foreign matter from entering the pump. The oil pump (21) discharges the oil through an oil cooler (16), full flow filter (12), and strainer (11) and delivers it to the main oil header (32) in the engine base. A thermostatic-operated three-way valve (15) is located in the system upstream of the cooler to direct oil through the cooler or around it to maintain the proper operating temperature. Thermometers should be installed in the line ahead of and after the oil cooler to give a constant reading of oil temperature. From the lube oil distribution header, connections supply oil to all main bearings. From the main beatings, oil is delivered through drilled passages in
p
,
,.
i
FIGURE 4-5 Schematic representation o f a gas engine pressure lube system. 1 - turbocharger; 2 - engine outline; 3 - turbo oil variable pressure regulator; 4 - turbo oil filter; 5 - check valve; 6 - engine oil pressure relief valve; 7 - oil in temperature indicator; 8 - f i l t e r and strainer vent line; 9 - needle valve; 10 - check valve; 11 - oil strainer; 12 - oil filter; 13 - differential pressure gauge; 14 - drain valve; 15 - three-way thermostatic valve; 16 - oil cooler; 1 7 - p u m p relief valve; 18 - engine prepost-lube pump; 19 - oil strainer; 20 - engine oil sight glass; 21 - engine main oil pump; 22 - f i r e safe shut-off valve; 23 - oil level indicator; 24 - engine low oil level alarm; 25 - oil supply to engine; 2 6 - f l o w meter (optional); 2 7 - oil level regulator; 28 - engine high oil temperature shutdown; 29 - p u m p relief valve; 30 - turbo prepost-lube pump; 31 - p u m p relief valve; 32 - oil to engine bearings; 3 3 - engine low oil pressure shutdown; 3 4 - turbo low oil pressure shutdown. (Source: Cooper-Bessemer Reciprocating, Grove City, PA)
Gas Engines 127 the crankshaft to the master connecting rod bearings. The crankshaft (Figure 4-4, item 24) is drilled so that each master connecting rod receives oil from the two adjacent main bearings. From the master connecting rod, oil flows to the piston pin through a drilled passage in the articulated connecting rod. The crown of the power piston is jacketed, and oil from the piston pin is circulated through the jacket to cool the piston. Oil returns from the piston jacket through a second drilled passage in the articulated connecting rod, which connects with passages in the master connecting rod and is discharged to the base through holes in the master connecting rod cap. Oil also flows through a drilled passage in the master connecting rod to lubricate the crosshead pin and guide. Vertical oil lines at the flywheel end carry oil from the main header to all the chain sprockets and bearings. Oil is returned to the base by gravity. At the other end, lines carry oil under pressure to all auxiliary drive shaft beatings, gears, chains, etc. A pressure relief valve, located at the flywheel end of the oil header inside the base, protects the system against excessive pressure. A pressure gauge on the control panel indicates the oil pressure in the system. To maintain the required oil level in the base, an automatic control valve is sometimes installed in an oil makeup line. The filter consists of a housing or shell containing replaceable, yarn-wrapped elements operating in parallel to give the desired capacity. The frequency of replacing the elements will vary with operating conditions. As the elements become contaminated, the pressure drop across the filter will increase. Modem machines have the main lube-oil header installed in the engine base. High-pressure flexible lines supply oil from the header to the main bearing caps, from where it is carried to the rest of the running gear as previously described. The engine is equipped with either multiple pumps or a block distribution-type force-feed lubrication system. All force-feed lubricator systems are divided into two separate sections. One section supplies lubricant for the power cylinders while the other section supplies lubricant to the compressor cylinders. This arrangement permits the use of different oils to lubricate the compressor cylinders when required. For certain types of compressor service, this is unnecessary and the same oil may be used for both power and compressor cylinders.
Fuel System The fuel piping system (Figure 4-6) consists of a variable fuel gas pressure regulator, gas receiver, manual gas cock, safety shut-off and vent valve, gas accelerating valve, governor-operated gas regulating valve, gas injection valves, and isolating valves. The variable fuel gas pressure regulator regulates the gas supply pressure according to the governor speed signal. The receiver (located as close to the engine as possible) absorbs pulsations in the gas flow and ensures a more uniform gas pressure at the engine. The safety shut-off and vent valve will shut off the gas supply and vent the line to the engine if an abnormal operating condition occurs. The gas accelerating valve controls the amount of gas supplied to the engine by the governor-operated gas valve during starting. The variable fuel gas pressure regulator is installed in the main fuel gas supply line upstream of the receiver. The gas pressure required at the engine will vary with the number of power cylinders and the heat content of the gas. In every case, the pressure should only be high enough to enable the engine to carry about 10 percent overload. The gas regulating valve (Figure 4-7) is located in the gas inlet on the operating end. It is controlled by the speed-regulating governor to regulate the amount of gas according to load requirements. Gas enters the valve body from the supply line connected to the bottom of the body, passes through the valve port, and enters the
128
Process Plant Machinery
/
.
,,;
/
/
FIGURE 4-6 Fuel piping system for two-stroke gas engine. 1 - engine gas header; 2 - load balance valve; 3 - j e t cell igniter; 4 - gas injection valve; 5 - cylinder head; 6 - pilot gas header to igniters; 1 0 - r e g u l a t o r feedback line; 1 1 - gas pressure regulator; 1 2 - starting pressure adjustment; 13 - speed signal inlet pressure and gauge; 14 - gas inlet, psi maximum; 1 5 - m a n u a l gas shut-off valve; 1 6 - r e g u l a t o r pilot filter; 1 7 - f u e l gas command signal; 18 - gas pressure gauge tap; 19 - igniterpilot gas filter and pressure gauges; 20 - gas supply to pilot gas filter; 2 1 - governor-operated gas regulating valve. (Source: Cooper-Bessemer Reciprocating, Grove City, PA)
Gas Engines VALVE S L O T
129
B A L A N C E PISTON
FIGURE 4-7 Gas-regulating valve for modem gas engine. (Source: Cooper-Bessemer Reciprocating, Grove City, PA)
inlet header of the engine. The valve is of the ported type and is designed to give a very fine regulation of flow for minimum travel of the valve and governor. A balancing piston on the valve stem will equalize the gas pressure in both directions of travel. These design features give close regulation of gas flow and ensure steady operation and close regulation of engine speed at all loads. Figure 4-8 illustrates the gas injection valve in cross section. Gas from the header is admitted to the injection valve through a cylinder-isolating plug valve. This valve is normally wide open and is used to restrict the flow to the injection valve to obtain load balance for equal distribution of load to the power cylinders. The injection valve has a conical surface that seats on the valve seat insert. The valve is opened mechanically by a push rod and rocker arm operated by a cam attached to the crankshaft and is closed by the spring in the injection valve. Packing at the upper end of the gas valve stem prevents leakage of gas at this point. The gas valve operating mechanism (Figure 4-9) consists of a rocker arm assembly, cam follower, and push rod with a hydraulic valve lifter. The rocker arm assembly is mounted on the cylinder head. An adjustable tappet is provided in the end of each rocker arm to adjust the hydraulic valve lifter. The cam follower is located in the engine base and is held in place by the push rod and crosshead bracket. Each follower consists of a crosshead and hardened steel roller that rides on a gas cam attached to the crankshaft. The push rod and hydraulic valve lifter assembly connects the rocker arm and cam follower. This assembly consists of a two-section push rod, the lower section being a tubular steel rod and adapter, with the hydraulic valve lifter installed in the lower end. A push rod guide supports the upper end of the lower section of the push rod at the point where it protrudes through the base. The upper section of the push rod is also tubular steel and is connected to the lower section by a ball pivot.
Cooling System Cooling the engine is accomplished by two separate systems: the jacket water system, which circulates through the engine jackets and heat exchangers, and the aftercooler water system, which circulates through the heat exchanger, aftercoolers, and lube oil cooler.
130
Process Plant Machinery OIL H E R E ~ A D J U S T I N G
N
" ,~
E:/
SCREW
~::~:~.....~
II STARRETT
~ 1 9 6 DI.AL INDICA TOR 196F
------;.L ~, "r ~ ~ / /,, // ,"
~
~J
/ / / /
MAGNETIC BASE
PUSH ROD
DI /
VE
I ~
l
GUIDE BA LL ASSEMBLY r
,i
DRIVE ROD ~NAP RING ~,DAPTER HYDRAULIC VALVE LIFTER CROSSHEAD BA LL END CRO~SH EAD ROLLER
FIGURE 4-8 Gas injection valve cross section. (Source: Cooper-Bessemer Reciprocating, Grove
City, PA.) GAS CA!
FIGURE 4-9 Gas valve operating mechanism. (Source: Cooper-Bessemer Reciprocating, Grove City, PA.) One of the most important factors involved in the design of the cooling system is an adequate supply of clean water, free from sediment and scale forming ingredients, since even a very thin layer of scale or dirt on any heat transfer surface will act as an insulator, which may cause overheating and breakage. It is preferable to circulate a large volume of water accompanied by a small temperature rise than to
Gas Engines
131
circulate a small volume of water accompanied by a large temperature rise. A large volume results in higher velocity through the system and retards the formation of scale and deposits of sediment in the jackets. Likewise, a low temperature rise means more uniform temperature at all points and less possibility of casting strains from this source. For these and other reasons, a closed cooling system is recommended. In such a system, a minimum of makeup water is required. Therefore, treated water that removes scaleforming ingredients is not expensive. There are numerous piping arrangements that can be used, and these will vary according to the number of engines installed, cooling equipment used, and other individual requirements. Figure 4-10 is a typical diagram of the cooling system for an engine with a built-in jacket water pump and a motor-driven aftercooler water pump. Tracing the flow of the jacket water system starting with the water pump, it is directed to the cooling equipment. This may be a cooling tower, radiator, or any other type of suitable equipment. A three-way thermostatically operated proportioning valve is located ahead of the cooling equipment to maintain the proper jacket water inlet temperature to the engine by bypassing a portion or all of the water around the cooling equipment. The water flows from the cooler, or bypass, to the engine inlet header. Cooling water enters the engine through the inlet header located in the vee of the engine between the two cylinder banks. From the header, water enters the bottom of the cylinder jacket and passes upward around the ports and enters the cylinder head jacket through outside jumper connections. Outlets from the cylinder heads connect with the water discharge headers. A jumper connects the two outlet headers at the flywheel end to give a common outlet connection. Turbocharger cooling water
___..__..._------~
1
@ FIGURE 4-10 Typical jacket water cooling system f o r a gas engine. 1 - water outlet temperature; 2 - outline of engine; 3 - standpipe fill line; 4 - vent; 5 - standpipe overflow; 6 - sightglass; 7 - standpipe; 8 - drain; 9 - j a c k e t water cooler; 1 0 - thermostatic valve; 11 - p l u g valve; 1 2 - to intercooler water system (balance line); 13 - gate valve; 14 - pump flexible pipe connections; 15 - j a c k e t water pump; 16 - to engine water inlet header; 1 7 - turbo-charger; 1 8 - engine water outlet header. (Source: Cooper-Bessemer Reciprocating, Grove City, PA)
132
Process Plant Machinery
is supplied from the inlet header at the flywheel end of the engine, circulates through the turbocharger, and discharges into the engine outlet header. From the outlet header, the jacket water flows to the standpipe where makeup water is added when necessary. From the standpipe, water returns to the suction of the jacket water pump where it is again recirculated. The standpipe should be high enough to maintain a positive head at the suction of the pump. Its diameter should be large enough to limit the downward flow velocity to 0.5 feet per second, thus allowing any entrained air bubbles to rise to the surface. All engines are equipped with two fin-tube-type aftercoolers. In some localities, additional cooling of the air may be required to enable the engine to carry rated load. Precooling is then recommended whereby the air is cooled before entering the turbocharger by passing it through an aquatower. A separate cooling water system is required for the water circulated through the aftercoolers. Engine jacket water is not suitable, as the water inlet temperature should not exceed 120 ~ for proper cooling. Higher water inlet temperatures to the aftercoolers will not cool the air sufficiently, and the engine will fail to carry rated load. Each aftercooler must receive ample cooling water for efficient operation. Tracing the flow of the aftercooler water system (Figure 4-11), starting with the pump, water is discharged from the pump through the heat exchanger, the aftercoolers, the lube oil cooler, and then back to the pump where it is again recirculated.
The Power Train Engines equipped with motor starting are cranked by air- or gas-driven reduction gear motors attached to a Bendix starter drive. Four- and six-cylinder engines have
Q-
t
J
t|
I
FIGURE 4-11 Typical intercooler water system diagram f o r a gas engine. I - outline o f engine; 2 - intercooler water outlet temperature; 3 - engine air inlet manifold; 4 - air intercoolers; 5 - intercooler water pump; 6 - pump flexible pipe connections; 7 - balance line (supply) from jacket water system; 8 - intercooler water cooler; 9 - t e m p e r a t u r e control valve; I 0 - c o o l e r bypass; 1 1 - p l u g valve; 1 2 - oil cooler; 1 3 - temperature controller; 14 - temperature control signal; 15 - intercooler vent; 16 - globe valve; 17 - intercooler water drain. (Source: Cooper-Bessemer Reciprocating, Grove City, PA)
Gas Engines
133
one starting motor; eight-, ten-, and twelve-cylinder engines have two. The pinion on the Bendix drive then engages the ring gear on the flywheel to crank the engine. After the engine "fires," the starting valve is closed, pressure to the motors is shut off, and the Bendix drive pinion gear disengages the flywheel ring gear. Pressure to the starting motors is filtered and then lubricated by an oil-fog-type lubricator. The layshaft is located in the vee between the power cylinders. It transmits power from the crankshaft to the auxiliary drive and is chain driven by the crankshaft at the flywheel end. The shaft is constructed in two sections and supported at both ends and in the middle by bronze beatings (Figure 4-12, items 1, 22, 32). It is enclosed throughout its length by a tubular housing that is oil- and dust-tight. The layshaft chain tightener is located in the engine base at the flywheel end. The main speed control governor, lubricators, and ignition timer are mounted on the auxiliary end drive cover and are driven by the layshaft. The governor is driven directly off the end of the layshaft and the timer is chain driven from the layshaft. The lubricators are chain driven from the timer drive. The positioning of these components is seen in Figure 4-1, foreground.
The Turbocharger A turbocharger consists of a centrifugal blower and turbine mounted on a common shaft surrounded by five major castings. It is mounted on a diesel or gas engine and is driven by the engine exhaust gases. The exhaust gases, due to their elevated temperatures and high velocity, transmit enough energy to the turbine and blower to force 50 to 100 percent more air into the engine for scavenging and combustion. This additional air "supercharges" the engine, making it possible to burn more fuel to produce additional power. The scavenging air flow provides cooling for the cylinder head, cylinder walls, and piston. For this reason, a greater amount of fuel can be burned without harmful effects to the engine and turbocharger due to excessive heat. The turbocharger output is proportional to engine load. It automatically slows down if the engine load is decreased and speeds up and delivers more air if engine load is increased or if barometric pressure drops. It maintains engine operation at or near optimum air-fuel ratio, resulting in high engine efficiency over a very O I L LINE CONNECTION
~__!~_ N~ :'v. t ~
1______1~!
!__.-._ _o__17-7 ~
~o ~ - o i ~
FIGURE 4-12 Power train layshafi for a two-stroke gas engine. (Source: Cooper-Bessemer Reciprocating, Grove City, PA)
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Process Plant Machinery
wide operating range. The turbocharger operates in one direction only, regardless of engine rotation. Turbochargers used on "pure turbocharged" engines are equipped with air-assist nozzles on the blower diffuser. Air discharged through these nozzles, at the time of engine starting, assists in purging the engine and delivers sufficient energy to the turbocharger to maintain the required air-fuel ratio until exhaust energy becomes sufficient to drive the turbocharger.
Engine Control System The control system of the engine has two purposes: first, to ensure safety, and second, to provide optimum operation. The pneumatically operated safety shutdown controls will stop the engine and indicate (on the control panel) the system in which the unsafe condition occurred. Any one of the safety shutdown devices (some are optional) will vent control air pressure from the fuel gas shutoff and vent valve to stop the engine should any of the following unusual malfunctions occur: 1. 2. 3. 4. 5. 6. 7. 8. 9. 0. 1. 2. 3.
High jacket water temperature. High air manifold temperature. Force-feed lubricator, power cylinder, or compressor cylinder failure. Low lubricating oil pressure. High lubricating oil temperature. Engine overspeeding. High main bearing temperature. High connecting rod bearing temperature. Excessive turbocharger vibration. Excessive engine vibration. Low crankcase oil level. Low aftercooler water pressure. Low jacket water pressure.
Any number of these sensing devices as required by the installation can be used in the system to stop the engine via fuel gas shutoff and vent valves. Engine speed is controlled by the main governor. The governor most commonly used is a Woodward hydraulic relay type. However, other types may be used according to service requirements and user preference. The governor is mounted on the auxiliary drive housing and driven through a set of bevel gears. The bevel gears are lubricated from the engine lube oil system. The governor is a self-contained unit with its own lubricating system. Governor signals are transmitted to the fuel gas accelerating valve to control the flow of gas to the headers in accordance with engine requirements. An increase in load causes a decrease in engine RPM, which in turn causes the governor to further open the fuel gas-regulating valve to admit more fuel gas to the engine. A decrease in load has the opposite effect. The result is that constant speed is maintained regardless of load and speed conditions. Fuel gas header pressure and inlet manifold pressure change with load and speed. The manifold pressure regulator balances fuel gas pressure against air manifold pressure. Any change in sensed pressures will correspondingly move a pilot valve that directs oil under pressure to an actuator. This actuator repositions the exhaust bypass butterfly valve, thus maintaining the correct air manifold pressure in accordance with engine requirements.
Chapter 5 Steam Turbines Steam turbines occur mainly in mechanical drive applications in the process engineer's world. When the plant also has small power generation requirements, the power plant selected is generally a gas turbine. Such is the case at Esso's Sriracha refinery in Thailand which owns two ABB GT35s. Larger facilities often find it economical to make most of their own power. The Syncrude Tar Sands facility in northern Alberta, Canada which refines 170,000 barrels of oil a day is such a facility. Syncrude uses a combination of gas and steam turbines. There is also a current emerging global trend that may grow: for oil and gas and petrochemical facilities to make their own power and sell the excess to the local national power grid. However, due to the relative proportion of mechanical drive versus power generation applications that a process engineer is likely to see, this chapter is essentially devoted to mechanical drive turbines. In steam turbines, apart from the load factor (close to, or at base or steady load in most power generation applications), the subdivision between mechanical and power generation applications is size. The approximate cut-off for mechanical drive applications is about 100 MW. Sizes above 100 MW are not normally used in compressor drive services. Although less versatile than the gas turbine (a steam plant has to provide design inlet pressure steam, before the steam turbine can run), steam turbines are reliable and easy to operate. The global distribution of natural gas still favors the utilization of steam turbines. At recent tally, the world has about 70 years of natural gas supply left. This figure increases as gas exploration becomes more ambitious. However, as the world has 200 to 300 years worth of coal, a great deal of that in the newly developing global regions, it's easy to see which fuel comes out ahead. Political factors also enter the selection issue. Note also that although steam turbine plant systems have been touted as having a slower response time than gas turbines, that gap is diminishing with new advances in steam turbine control systems. Environmental considerations caused by steam turbine plant boiler emissions need not be an issue for new plants, if current technology is used. This technology is described in Environmental Engineering and Management: Sustainable Development for the Power Generation, Oil and Gas and Process Industries in far more detail than the scope of this chapter allows. That text would be useful for any process engineer who has to face new equipment selections, plant expansion design, changes in environmental legislation or major plant refurbishments today.
135
136
Process Plant Machinery
MECHANICAL DRIVE STEAM TURBINES*
Steam turbines for mechanical drive applications were among the first real machines to usher in the Age of Industrialization. In 1629, an Italian inventor, Giovanni de Branca, envisioned a boiling pot whose nozzle opening was aimed against the paddle wheels of a roasting spit. Clearly, then, the use of expanding steam to turn a shaft is anything but a new idea. By now, hundreds of thousands of steam turbines are installed and working in every conceivable type of process plant the world over. As shown in Table 5.1, inlet steam pressures range from a few pounds per square inch (psi) to over 2000 psi (140 bar) and power output covers the field from a single kilowatt (kw) in the emergency lube oil pump driver to almost 100,000 kw in large compressor drive applications at modem petrochemical plants. This should be no surprise, since the ability to efficiently convert large amounts of heat energy into mechanical work makes the steam turbine the logical choice for many industrial drive applications. Its reliability, smoothness of operation, and versatility also contribute to its popularity. Before discussing turbine selection, let's review how a steam turbine converts the heat energy of steam into useful work. The nozzles and diaphragms in a turbine are designed to direct the steam flow into well-formed, high-speed jets as the steam expands from inlet to exhaust pressure. These jets strike moving rows of blades mounted on the rotor. The blades convert the kinetic energy of the steam into rotation energy of the shaft. In a reaction turbine, the steam expands in both the stationary and moving blades. The moving blades are designed to utilize the steam jet energy of the stationary blades and to act as nozzles themselves. Because they are moving nozzles, a reaction f o r c e - produced by the pressure drop across t h e m - supplements the steam jet force of the stationary blades. These combined forces cause rotation. To operate efficiently, the reaction turbine must be designed to minimize leakage around the moving blades. This is done by making most internal clearances quite small. The reaction turbine also usually requires a thrust balance piston (similar to large centrifugal compressors) due to the large thrust loads generated. Because of these considerations, the reaction turbine is less often used for mechanical drive in the United States, despite its higher initial efficiency. However, reaction turbines are quite often used in other parts of the world. This text will explain them later. Moreover, since impulse and reaction turbines share many construction details and nomenclature, the reader is encouraged to become familiar with both. TABLE 5.1 Unit
Small Medium Large Very large
Typical Steam and Power Conditions for Process Plant Turbines Steam
150-400 psig; (10-27 bar; 400-600 psig; (27-41 bar; 600-900 psig; (41-62 bar; 900-2,000 psig; (62-140 bar;
500-750~ 260-400 ~ 750-825 ~ 400-440 ~ 750-900~ 400-482 ~ 825-1000 ~ 440-538 ~
Power
1-1,000 HP (0.75-750 kw) 1,000-5,000 HP (750-3,750 kw) 5,000-60,000 HP (3,750-45,000 kw) 15,000-120,000 HP (11,200-90,000 kw)
Source: The Elliott Company, Jeannette, PA. * Sources: The Elliott Company, Jeanette, PA (impulse turbines), and Siemens Energy and Automation, Inc., Bradenton, FL (reaction turbines). Adapted by permission.
Steam Turbines
IMPULSE STEAM TURBINES The impulse turbine has little or no pressure drop across its moving blades. Steam energy is transferred to the rotor entirely by the steam jets striking the moving blades (Figure 5-1). Since there is theoretically no pressure drop across the moving blades (and thus no reaction), internal clearances are large and no balance piston is needed. These features make the impulse turbine a rugged and durable machine that can withstand the heavy-duty service of today's mechanical drive applications. Steam turbine materials are tabulated in Table 5.2.
Velocity-Compounded (Curtis) Staging A Curtis stage (Figure 5-2) consists of two rows of moving blades. Stationary nozzles direct the steam against the first row; reversing blades (not nozzles) then redirect it to the second row. The large pressure drop through the nozzle produces a high-speed steam jet. This high velocity is absorbed in a series of constant pressure steps (see below). The two rotating rows of blades make effective use of the high-speed jet, resulting in small wheel diameters and tip speeds, fewer stages, and a shorter, more rugged turbine for a given rating.
Pressure-Compounded (Rateau) Staging Again referring to Figure 5-2, we observe how in Rateau staging the steam path is slightly different. Here the heat energy of the steam is converted into work by stationary nozzles (diaphragms) directing the steam against a single row of moving blades. As in a Curtis stage, pressure drops occur almost entirely across the stationary nozzles.
Turbine Configuration Overview Single-flow condensing units (Figure 5-3) require less steam for a given horsepower than other types. They expand steam from inlet pressure to a pressure less than atmospheric. The exhaust pressure is maintained by a condenser, providing for recovery of the spent steam. A condensing unit thus minimizes the need for makeup water. Because of these advantages, the straight condensing turbine is much in demand, as evidenced by the literally thousands of units installed world-wide. Double-flow condensing units (Figure 5-4) are very similar to single-flow units except that the last-stage flow is divided between two rows of blades. This enables a double-flow turbine to operate at higher horsepowers and speeds than single-flow units of similar steam conditions. Automatic extraction units are schematically represented in Figure 5-5. Extraction turbines are used when there is a need for process steam at a pressure between turbine inlet and exhaust pressures. They are designed to simultaneously maintain the desired extraction steam pressure and the speed of the driven machine. They can do this even though the demand for extraction steam and the horsepower requirements of the driven unit may vary over a wide range. Noncondensing steam turbines (Figure 5-6) exhaust steam at greater than atmospheric pressure and while it still contains a great deal of energy. It can therefore be
137
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ProcessPlant Machinery
Vj (Jet v e l o c i t y ] .
.
V b (blade v e l o c i t y ] = 0 .
.
.
If the t u r b i n e rotor is locked, the steam jet e x e r t s maxim u m forr on the blades, but no w o r k is d o n e since the blade d o e s n ' t move.
9
= Vj / 4
If the blade is m o v i n g at ~ of the jet v e l o c i t y , the f o r c e on the blade is r e d u c e d , but s o m e w o r k is d o n e by m o v i n g the blades.
--•
Vj
,=
V b --
Vj/2
M a x i m u m w o r k is d o n e w h e n the blades are m o v i n g at let speed. Relative v e l o c i t y of steam leaving blades is zero.
/
0.5
// 0
0.5
1.0
Vb/vj
FIGURE 5-1
The impulse principle. (Source: The Elliott Company, Jeannette, PA.)
Steam Turbines TABLE 5.2
139
Typical Standards Materials of Construction for Mechanical Drive Steam Turbines
Material Steam chest and casing 600 p s i - 750 ~ b a r - 399 ~ 600 psi - 825 ~ bar - 440 ~ 900 psi - 900 ~ bar - 482 ~ 2,000 psi - 950 ~ bar - 510 ~ Exhaust casing Condensing and non-condensing (cast) Non-condensing (cast) Fabricated Nozzles Diaphragm centers Fabricated Cast Disks Forged Cross-rolled plate Integral with shaft Blades Shroud bands Damping wire Shaft Built-up Integral Bearing shells Bearing liners Labyrinth seals Shaft sleeves up to 750 ~176 751-875 ~ ~ Stationary baffles Governor valves Governor valve stems and seals Governor valve seats Bar lift rods and bushings Steam strainer screen Bearing housings
Commercial Specifications
Cast carbon steel Carbon-molybdenum steel Chromium-molybdenum steel Chromium-molybdenum steel
ASTM ASTM ASTM ASTM
High-strength cast iron Cast steel Steel 12% Chromium-stainless steel
ASTM A-278 Class 40 ASTM A-216 Grade WCB ASTM SA 285 Grade C AISI-405
Steel High-strength cast iron
ASTM SA 285 Grade C ASTM A-278 Class 40
Chromium-nickel-molybdenum steel Constructional alloy steel Chromium-nickel-molybdenumvanadium steel 12% Chromium-stainless steel 12% Chromium-stainless steel 15% Chromium steel
AISI 4340 USS T- 1 ASTM A-470 Class 4, 7, or 8
A-216 A-217 A-217 A-217
Grade Grade Grade Grade
WCB WC 1 WC6 WC9
AISI Type 403 AISI Type 410 Inconel X750
Chromium-molybdenum steel Chromium-nickel-molybdenumvanadium steel Steel Bonded tin-base babbitt
AISI 4140 ASTM A-470 Class 4, 7, or 8
Carbon steel Nickel-chromium-molybdenum-steel Chromium-molybdenum-steel forgings 12% Chromium-stainless steel 12% Chromium-stainless steel, nitrided 12% Chromium-stainless steel 12% Chromium-stainless steel, nitrided Stainless steel Ductile iron
ASTM A- 179 ASTM 4340 AISI 4140
ASTM SA 285 Grade C ASTM B-23 Alloy #7
AISI Type 410 AISI Type 416 AISI Type 416 AISI Type 416 AISI Type 321 ASTM A-536 Grade 60-45-12
Source: The Elliott Company, Jeannette, PA.
used for other purposes in a given plant. Units of this type are much more compact than condensing turbines of the same horsepower due to the smaller volume of steam handled at the exhaust end. Figure 5-7 depicts three typical rotor categories: single-flow condensing, double-flow condensing, and noncondensing.
140
Process Plant Machinery
FIGURE 5-2 Steam flow through turbine stages. S = Stationary. (Source: The Elliott Company, Jeannette, PA.)
Turbine Components
Casing Overview. A good casing should be just thick enough to contain the steam pressure for which it is designed. Overly thick walls act as heat sinks and can restrain expansion and contraction and lead to premature cracking. Thin, contoured walls prevent large gradients between inside and outside "skin" temperatures. A turbine casing designed this way conducts heat rapidly. This protects against cracking of the nozzle partitions (critical area) and assures long life for casings subjected to load changes and start-stop operation.
Steam Turbines
FIGURE 5-3
Single-flow condensing turbine. (Source: The Elliott Company, Jeannette, PA.)
1
FIGURE 5-4
Double-flow condensing turbine. (Source: The Elliott Company, Jeannette, PA.)
|
FIGURE 5-5
r
Automatic extraction turbine. (Source: The Elliott Company, Jeannette, PA.)
I FIGURE 5-6
Noncondensing steam turbine. (Source: The Elliott Company, Jeannette, PA.)
141
142
Process Plant Machinery
FIGURE 5-7 Principal rotor categories: Single-flow condensing (A, B), double-flow condensing (C, D), noncondensing (E, F). (Source: The Elliott Company, Jeannette, PA.)
Steam Turbines 143 This casing philosophy is carried a step further on turbines designed for 2000 psi and 950~ (140 bar and 510~ The nozzle chambers and partitions are free of both the front and side walls. The result, in effect, is a double-casing that relieves the internal stress in an area where there can be large inside-outside temperature differences.
Casing Construction. A typical steam turbine casing consists of a steam chest, intermediate barrel section, and separate exhaust casing. To prevent leakage at the high-pressure end, the steam chest and barrel section are often cast as one piece. This permits the vertical casing joint to be made at the low-pressure end. The high-pressure end of the turbine is supported by the steam end bearing housing. This housing is flexibly supported to permit axial expansion caused by temperature changes. The exhaust casing is centerline-supported on pedestals that maintain alignment with the driven machine while allowing for lateral expansion. Figures 5-8 and 5-9 illustrate important casing details. Steam Turbine Rotors. Steam turbine rotors are either of the solid or built-up type. Because each has its advantages, the user should not be tied to any one method of rotor construction. Operating conditions often call for a solid rotor, but sometimes the user may choose between the two or accept a recommendation from capable manufacturers. With solid rotor construction, the shaft and disks are one piece. These rotors are generally used in units operating at high temperatures and/or high speeds. Built-up rotors (Figure 5-10) cost less to make, less to buy, and are easier to repair if damaged. They are generally used in units operating at lower temperatures and speeds. Speeds are generally limited by the shrink fit required to overcome centrifugal growth of the disk. With either solid or built-up construction, shaft and disk materials must be selected to match the user's speed and temperature conditions. The shaft must be accurately proportioned to make sure that critical speeds are safely outside the operating speed range. Disk profiles should be designed to minimize centrifugal stresses, thermal gradients, and blade loading at the disk rims. Combination solid and built-up rotors are also available. Figure 5-11 shows a seventeen-stage, 28,000 HP (21 MW) rotor weighing 17,000 lbs. (7700 kg) and embodying the combination principle. Blades. The size and configuration of turbine blades depend on the operating conditions imposed on them. Machinery engineers have been designing blades to match horsepower, speed, and all types of steam conditions for many years. This continuing accumulation of blade knowledge has provided a wide selection of tested and proved blade designs. The blade performance record of reputable manufacturers usually reflects substantial design and manufacturing experience. Figure 5-12 illustrates first-stage blades that have to withstand punishment of steam at its highest pressure and temperature. The blade at left is used for 2000 psi (140 bar) service; others are for lower pressures. A chromium stainless steel alloy, selected for its strength, erosion/corrosion resistance, and damping qualities is used on all blades. Last-stage blading is configured to efficiently expand large volumes of steam. Consequently, the blades are often considerably larger than those located in the higher pressure regions of a steam turbine.
Nozzle Rings and Diaphragms. A stage in multistage turbines consists of both rotating and stationary blades. The stationary blades can be part of a nozzle ring
144
ProcessPlant Machinery
FIGURE 5-8
Steam turbine casing components. (Source: The Elliott Company, Jeannette, PA.)
or a diaphragm. In either case, their function is to direct steam onto the rotating blades, turning the rotor and producing mechanical work. The first-stage nozzle ring is made by milling steam passages into stainless steel blocks, which are then welded together (Figure 5-13). Other manufacturing methods are, of course, available.
Steam Turbines
145
FIGURE 5-9 Steam end (2000 psi; 140 bar) showing steam space between nozzle chambers and front and side walls. This allows nozzle chambers to expand and contract freely in response to load and temperature changes. (Source: The Elliott Company, Jeannette, PA.)
FIGURE 5-10 Typical built-up single-flow rotor with 14-inch (360-mm) last-stage blades. Every disk is shrunk and keyed to the shaft. (Source: The Elliott Company, Jeannette, PA.)
The nozzles in the intermediate pressure stages are formed from stainless steel nozzle sections and inner and outer bands. These are then welded to a round center section and an outer ring, as shown in Figure 5-14. Low-pressure diaphragms of condensing turbines are often made by casting the stainless nozzle sections directly into high-strength cast iron. This design
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Process Plant Machinery
FIGURE 5-11 Combination solid and built-up rotor for 28,000-HP (21-mW) steam turbine. (Source: The Elliott Company, Jeannette, PA.)
FIGURE 5-12
First-stage blades for steam turbines. (Source: The Elliott Company, Jean-
nette, PA.)
(Figure 5-15) includes a moisture catcher to trap condensed droplets of water and keep them from re-entering the steam path. These diaphragms can also be completely fabricated, if desired, at extra cost. Figure 5-16 illustrates steam turbine diaphragms as ultimately installed. The figure shows how the steam goes through the nozzles and strikes the rotor blades, causing rotation. Each diaphragm is located in its own groove in the casing. The keyway assures alignment of the diaphragm halves and helps minimize steam leakage across the split line.
Steam Turbines
FIGURE 5-13
First-stage nozzle ring. (Source: The Elliott Company, Jeannette, PA.)
Thrust Bearings. Thrust bearings in modern steam turbines must withstand greater loads and higher speeds. The double-acting, self-leveling, fully equalizing Kingsbury type best meets these demands. It is long-wearing, requires little maintenance, and is capable of handling variable speeds. As the wedge-shaped oil film is formed between the thrust collar and the surface of each shoe (Figure 5-17), the upper and lower leveling plates assure that the thrust load is distributed evenly among the shoes. The base ring then transmits this thrust load to the beating housing. Stops prevent the base ring from rotating with the shaft. Thrust beatings are generally sized to provide thrust capacity well in excess of that required for normal operating conditions.
Journal Bearings. The turbine rotor runs in positively aligned, pressure-lubricated journal bearings that are horizontally split and lined with babbitt of the highest quality. Pressure-dam or tilting shoe bearings are used when required to assure bearing stability throughout the speed range. Pressure-dam journal beatings (Figure 5-18) are generally used in high-speed units. This type of bearing is produced from a standard liner in which "dams" have
147
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Process Plant Machinery
FIGURE 5-14 nette, PA.)
Intermediate pressure diaphragm. (Source: The Elliott Company, Jean-
been machined to create a "pressure pad" of oil. This pad forces the shaft to remain in position in the lower beating liner, thus maintaining proper shaft attitude angle. Tilting shoe journal bearings, shown in Figure 5-19 with five shoes (or pads), help ensure rotor stability. Labyrinth Seals. Experience has proved the labyrinth seal to be most reliable for steam turbines. It forms an effective seal without contact between its component parts, which means turbines will rarely if ever be shut down for seal maintenance. The function of the seal, simply put, is to keep steam in and air out of the turbine where the rotor passes through the casing. It restricts steam leakage along the shaft by using the sealing strips on the shaft sleeve to form a "labyrinth" that hinders the flow of steam (Figure 5-20). In noncondensing or back-pressure turbines, the labyrinth seal must prevent steam leakage out of the casing at both inlet and exhaust because these pressures are greater than atmospheric. In condensing turbines, the pressure at the exhaust
Steam Turbines
FIGURE 5-15
Low-pressure diaphragm. (Source: The Elliott Company, Jeannette, PA.)
FIGURE 5-16 Passage of steam through nozzles in diaphragm. (Source: The Elliott Company, Jeannette, PA.) end is less than atmospheric. In this case, the labyrinth seal and sealing steam are used to prevent the flow of air into the casing.
Extraction Control The use of extraction steam has gained wide acceptance in industry today. An extraction turbine can supply steam at some constant pressure between turbine inlet and exhaust pressures. This flexibility can be a great asset when designing a plant steam balance. To automatica|]y supply extraction steam, a second steam chest and a pressure control system are added to the speed control (governing) system of the turbine. This creates, in effect, two turbines on one shaft.
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Process Plant Machinery
FIGURE 5-17 Self-leveling thrust bearing (Kingsbury-type) installed in modem steam turbine. (Source: The Elliott Company, Jeannette, PA.) Extraction steam pressure and turbine speed are controlled by the extraction regulator as follows: Constant Extraction Demand. Let's assume that extraction steam demand is constant, but the load on the turbine is reduced slightly and it begins to increase speed. The governor calls for reduced steam flow to the turbine. The regulator is designed to reduce the steam flow proportionally through both the steam inlet valves and extraction valves. Extraction flow, therefore, remains unchanged, while both inlet and exhaust flows are reduced. This diminished steam flow through the turbine decreases the power developed, and the turbine slows to its set speed. (Figure 5-21A).
Steam Turbines 151
FIGURE 5-18
Pressure-damjournal bearing. (Source: The Elliott Company, Jeannette, PA.)
FIGURE 5-19
Tilting-shoejournal bearing. (Source: The Elliott Company, Jeannette, PA.)
Constant Load. Let's assume that the load on the turbine is constant, but there is a change in the process and the demand for extraction steam decreases. The extraction pressure increases, thus compressing the bellows in the regulator and rotating the three-arm linkage. This simultaneously closes the inlet valves and opens the extraction valves. The inlet and extraction flows are thus decreased, while the exhaust flow is increased. Speed holds at setpoint because the three-arm linkage is proportioned to maintain a constant total horsepower developed by the rotor during these changes in extraction flow (Figure 5-21B). In actual operation, both speed and extraction steam demands are constantly changing, and the sequences described are occurring simultaneously. When excess steam is available, it can be "induced" into the turbine and expanded to exhaust pressure. Major manufacturers of steam turbines can supply automatic extraction units, automatic induction units, or a combination of the two. It should also be stressed that fully electronic controls are sometimes an attractive feature that could be considered instead of the more traditional, time-tested mechanical controls. Many of the 1980-vintage steam turbines are fitted with highly reliable electronic controls.
152
Process Plant Machinery
FIGURE 5-20 Labyrinth seal details. (A) Labyrinth seal showing how high and low sealing strips combine with the stationary baffle to hinder the flow of steam. The spring allows the stationary baffle to move away from the shaft if a rub occurs. The heat generated is absorbed by the stationary baffle. This protects the shaft and minimizes rotor damage. (B) Leakage flow and sealing steam arrangement for steam end of condensing turbine. Conditions that follow are typical values: (1) First-stage pressure equals 300 psia (21 bar) at full load. (2) Leakoff to turbine stage at approximately 125 psia (8.7 bar). (3) Leakoff to turbine stage at approximately 40 psia (2.8 bar). (4) Sealing steam at start-up is approximately 18 psia (1.25 bar). (5) Steam and air drawn to gland condenser at approximately 13.5 psia (0.9 bar). (6) Small amount of atmospheric air drawn in. (Source: The Elliott Company, Jeannette, PA.)
Steam Turbines
;........ 1 ~ BellOqSl~~ Bell s~ ~
j
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f
sor,o
i
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Governor
~.~
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l
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ExtractionPressure-/ SensingL~ne
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H_: _- _.U_L_______
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FIGURE 5-21
Constant extraction demand (A) versus constant load output (B). (Source: The Elliott Company, Jeannette, PA.)
An electronic extraction control system for a turbine driving a compressor is shown in Figure 5-22. Speed and pressure signals are fed to the control and electrically compared with the speed and pressure reference set points. Corrective signals are generated and sent to the inlet and extraction servo-motors. The optional operator control station is just one of a wide variety of optional features available with this type of control. Trip Devices. Steam turbines must be protected against excessive speed by automatic overspeed trip devices. These devices are either of the mechanical or electronic type. Mechanical overspeed devices are based on the centrifugal force principles as shown in Figures 5-23 and 5-24. Upon actuation of the trip lever, hydraulic oil is dumped from a pressurized cylinder that normally holds inlet steam or trip valves open. When depressurization occurs, these valves close instantly.
153
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Process Plant Machinery
ELECTRONIC EXTRACTION CONTROL
SPEED
~,
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r
'
Electronic extraction control for steam turbine. (Source: The Elliott Company, Jeannette, PA.)
FIGURE 5 - 2 3
Eccentric overspeed trip device. (Source: The Elliott Company, Jeannette, PA.)
Steam Turbines 155
FIGURE 5-24
Disk-type overspeed device. (Source: The Elliott Company, Jeannette, PA.)
Electronic overspeed devices operate on a somewhat similar principle. Their actuation results in the opening of a solenoid dump valve that depressures the hydraulic cylinder that normally keeps the inlet and trip valves open. REACTION STEAM TURBINES
Our introductory comments on steam turbines made the point that steam turbines extract heat energy from the steam and convert it into mechanical work. Heat and mechanical work are both forms of energy, and, therefore, can be converted from one to the other. First, heat energy is converted into velocity (kinetic) energy. In this first step, steam expands in a nozzle discharging at high velocity. The total heat (enthalpy) of the steam is converted to kinetic energy (velocity). Second, a steam turbine does mechanical work by virtue of the steam velocity that strikes moving blades. There are two types of stages to convert this velocity energy into usable work, namely, impulse and reaction. With an impulse design, the pressure drop is taken across the nozzles only, whereas with reaction design the drop is evenly distributed across the nozzle and blades. If the nozzle is fixed and the force of the jet is directed against a crescent shape (the blade), the jet's impulse force pushes the blade in the direction of the jet. Further, the leaving velocity is less than the initial jet velocity. Such an example would be a water hose jetting water against a wall. If the nozzle is free to move, the reaction pushes against the nozzle and it will travel in a direction opposite to the jet. The leaving velocity is greater than the jet velocity. An example would be a revolving lawn sprinkler. The basic characteristics of impulse-type turbines and reaction-type turbines can be summarized as follows. In impulse turbines we find the following:
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Process Plant Machinery
9 Large bucket clearances permit quicker loading with minimal danger from thermal stress. 9 Since most of the pressure drop is in stationary nozzles, efficiency does not depend on blade clearances. 9 Low thrust generation. In reaction turbines, we note the following: 9 Expansion takes place in stationary nozzles and moving blades. 9 More bucket sealing (sealing strips) is required at the end of blades. 9 Efficiency is theoretically greater but depends on close clearances. 9 The so-called greater efficiency of a reaction machine over an impulse machine in practice becomes a function of the staging (number of stages, efficiency of stages available, etc.). A reaction machine will have more stages because to keep losses to a minimum, the pressure drop per stage must be kept low. 9 Reaction turbines require larger thrust bearings and/or a balancing piston. Reaction staging is similar in concept to Rateau staging except that pressure drop occurs through both the stationary nozzle and matching blades. Figure 5-25 shows reaction staging and can readily be compared with Rateau staging. A cutaway view of a typical reaction-type turbine is shown in Figure 5-26. It highlights the principal components as follows.
Foxed
P
Foxed
I
x____,\
~
teom pressure
Fixed I Moving
V
A Velocity-compounded (Curtis) stage and reaction stages
B (Rateau)
Pressure-compounded stage
C Reaction stage
FIGURE 5-25 Different types of stages employed in modem steam turbines. (Source: The Elliott Company, Jeannette, PA.)
Steam Turbines
157
FIGURE 5-26 Cutaway view of a typical reaction-type steam turbine. (Source: Siemens Energy and Automation, Inc., Medium Power Generation Division, Bradenton, FL.)
Casing and Guide Blade Carriers To achieve satisfactory performance with the high steam pressures involved and to simplify erection and overhauls, a double-shell casing is used. The outer shell is split horizontally at the machine axis, the top half incorporating the integrally cast admission chest. This is in the form of a transverse tube with an opening at each end for assembly. The emergency stop valve body is welded to one end of the tube. The two halves of the casing are flanged and bolted together. The casing flange bolts are heated to a predetermined temperature when being fitted. The cooling of the bolts prestresses them exactly to a predetermined value. The guide blades are fitted in carriers. The front carrier is mounted in the onepiece inner shell, while the real carrier is mounted directly in the outer shell. The longer machines incorporate an additional carrier. The unsplit inner shell allows the
158
Process Plant Machinery
useful feature of a split outer casing to be retained while at the same time offering the advantages of the "barrel" type of construction. The inner shell is shown in Figure 5-27. The front guide blade carder is split axially and clamped together by a conical ring. It is located centrally in the inner casing by a bayonet coupling that is pinned to prevent it from turning.
FIGURE 5-27 One-piece high-pressure inner casing for reaction turbine. (Source: Siemens Energy and Automation, Inc., Medium Power Generation Division, Bradenton, FL.)
FIGURE 5-28 Rear guide blade carrier for reaction turbine. (Source: Siemens Energy and Automation, Inc., Medium Power Generation Division, Bradenton, FL.)
Steam Turbines
Figure 5-28 shows the completely assembled rear blade carder, which is kinematically supported in the outer casing. The guide blade carders are aligned in the bottom half of the casing before final assembly. The steam pressure in the annular space between the inner and outer shells is approximately half the initial steam pressure. This means that both the inner and outer shells are subjected to relatively low stresses. Since there is only a small difference in the temperature of the steam inside and outside the inner shell when the turbine is running normally, thermal stresses are also minimal. The largely symmetrical shape ensures that thermal expansion is uniform at all sections. The fact that the front part of the outer casing is subjected to only half the initial steam pressure and the rear part of the casing to the exhaust pressure means that flanges and bolts can be of a size consistent with their reduced stress. The inner shell rests on its integrally cast brackets in the bottom half of the outer casing. Eccentric pins (Figure 5-29) are incorporated to adjust the position of the inner shell in relation to the outer casing and to align the rear guide blade carrier in the outer casing. Self-sealing, L-section rings free to expand in any direction are used for the seal between the admission chest and the inner shell. The initial steam entering the various nozzle boxes of the admission chest gives rise to a downward thrust on the inner casing. This thrust is partially compensated by allowing steam from the two outer nozzle boxes to pass through holes in the bottom of the inner casing and build up an opposing force on the underside of the casing in two small chambers of appropriate cross-sectional area sealed from the outer casing by L-section rings (Figure 5-30). The exhaust, gland, and drain connections are made to the bottom half of the outer casing with the result that only the initial steam line must first be removed to lift the top half of the casing. Since the bottom half remains in its aligned position on the pedestals, the amount of work involved in overhauls and inspections is considerably reduced. Once the top half of the casing has been lifted, the rotor, complete with the inner casing and front guide blade carrier, can be taken out (Figure 5-31). The
FIGURE 5-29 Fitted eccentric pin for internal component alignment of reaction turbine. (Source: Siemens Energy and Automation, Inc., Medium Power Generation Division, Bradenton, FL. )
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Process Plant Machinery
FIGURE 5-30 Vertical thrust balancing chambers with L-ring sealing to casing bottom half (Source: Siemens Energy and Automation, Inc., Medium Power Generation Division, Bradenton, FL.)
FIGURE 5-31 Turbine rotor complete with inner casing and front guide blade carrier. (Source: Siemens Energy and Automation, Inc., Medium Power Generation Division, Bradenton, FL.)
Steam Turbines
combined weight of these parts is only approximately one eighth of the total weight of the turbine.
Rotor and Blading The solid rotor shaft and control stage wheel are forged of one piece. The individual blades, which are machined from the solid, have integral shrouding and a T-root that fits into a groove in the rotor shaft. The drum blading is approximately 50 percent reaction.
Blade Tip Sealing The rotor blades are milled from the solid together with the shrouding. After they have been fitted in the rotor groove, the shrouding forms a complete ring, which is then skimmed. A circumferential step is formed in the surface of the shrouding, which, together with the sealing strips caulked into the casing opposite the shrouding, produces an efficient sealing effect. Riveted shrouding is used for the guide blades as they are not subjected to centrifugal force. A labyrinth sealing effect is again produced by caulking sealing strips in the rotor opposite the shrouding. Figure 5-32 shows some shrouded blading and also a diagrammatic section of the tip sealing employed. The radial clearance may be kept extremely small to reduce losses without adversely affecting operational reliability. Should maloperation cause contact to occur, the sealing strips will be rubbed away without producing a dangerous rise in temperature, and they can be replaced at the next overhaul without disturbing the blading in any way.
Shaft Glands The points where the shaft passes through the casing are sealed by means of labyrinth glands, which are composed of sealing strips in the stationary part projecting into grooves in the turbine rotor shaft (Figure 5-33). Any steam leaking through the gland is drawn off at an intermediate stage to a region of lower pressure in the turbine or to a special condenser in order to limit the amount of steam discharged into the atmosphere.
FIGURE 5-32 Tip sealing of blade by integral shrouding. (Source: Siemens Energy and Automation, Inc., Medium Power Generation Division, Bradenton, FL.)
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Process Plant Machinery
FIGURE 5-33 Labyrinth shaft gland. (Source: Siemens Energy and Automation, Inc., Medium Power Generation Division, Bradenton, FL.)
FIGURE 5-34 Front and rear bearing pedestals for reaction turbine. (Source: Siemens Energy and Automation, Inc., Medium Power Generation Division, Bradenton, FL.)
Casing Supports, Iournal and Thrust Bearings The top half of the casing is supported on four brackets, two resting on the front beating pedestal and two on the rear (Figure 5-34). Vertical alignment devices are placed between the feet and the pedestal surfaces to adjust the height accurately. The rear bearing pedestal is bolted rigidly to the foundation. When the casing expands as the temperature rises, it slides on the front bearing pedestal. In order to locate the casing to the beating pedestals in the longitudinal and transverse directions, both pedestals have guide rails that engage with guide forks on the casing bottom half fitted with horizontal alignment devices.
Steam Turbines
FIGURE 5-35 Double-acting segmented thrust bearing. (Source: Siemens Energy and Automation, Inc., Medium Power Generation Division, Bradenton, FL.)
The journal bearings are of the multiwedge type. They are white-metal lined. The oil-film wedges uniformly spaced around the circumference hold the rotor in a stable position. The position of the bearing housings in the pedestals can be adjusted with alignment devices to align the rotor accurately. This alignment ensures that the correct radial clearance is obtained in the blading and labyrinth seals. The axial position of the rotor within the casing is determined by the thrust bearing in the front beating housing. It is of the double-acting segmented type (Figure 5-35). The bearing segments are also lined with white metal. As its temperature rises, the rotor expands in the direction of the exhaust hood. Since the casing is fixed at the rear bearing housing, the front beating housing carrying the thrust bearing expands in the opposite direction. Hence, only the very much smaller differential expansion between rotor and casing is temporarily effective. The thrust bearing accepts any residual thrust that has not been eliminated by the balance piston.
Valves and Governing System The initial steam is admitted to the admission chest through the emergency stop valve (Figure 5-36). During normal operation, this valve is held open against spring
FIGURE 5-36 Emergency stop valve incorporating a steam strainer. (Source: Siemens Energy and Automation, Inc., Medium Power Generation Division, Bradenton, FL.)
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Process Plant Machinery
load by hydraulic pressure. In the event of a defect, the trip oil circuit is vented and the springs close the valve in a fraction of a second. A steam strainer is incorporated in the valve body. It is comprised of corrugated steel strip wound spirally on edge on a former. In comparison with the perforated type of strainer, the wound type is considerably stronger, has a greater flow area and a smaller aperture size. The nozzle control system of standard turbines makes extensive use of lever mechanisms. The five nozzle control valves are suspended from a beam
FIGURE 5-37 Steam turbine control valves. (Source: Siemens Energy and Automation, Inc., Medium Power Generation Division, Bradenton, FL.)
FIGURE 5-38 Valve diffusors and L-section sealing rings in the top half casing. (Source: Siemens Energy and Automation, Inc., Medium Power Generation Division, Bradenton, FL.)
Steam Turbines
(Figure 5-37). The beam is raised and lowered on two spindles through a system of levers by a front-mounted hydraulic servomotor. This arrangement means that there are only two spindle glands in the admission chest. Both spindles move whenever there is any control action so that the possibility of seizure is virtually eliminated. The distance between the top of the adjusting bushing in the beam and the underside of the backing nut at the top of the valve spindle varies from valve to valve. Thus when the beam is lifted, the valves open in sequence, allowing steam to reach individual nozzle segments progressively. The valve diffusors in the top half of the casing are shown in Figure 5-38.
165
This Page Intentionally Left Blank
APPENDIX 5A
Steam Turbines: Some Design Theory Factors
There are many factors affecting overall blade performance. They should all be carefully weighed before a design is released for production. The discussion here will be limited to just two of these factors: frequency and stress analysis. We shall see how they can affect the final design of a blade.
ENGINEERING Stress Analysis The stress levels in each blade must often be given strong consideration. As explained later, stress is best evaluated by a modified Goodman diagram, which provides a ratio of allowable cyclic stress to the maximum steam bending stress. This ratio should have certain minimum values for each stage under consideration. There are many factors affecting the mechanical integrity of a blade. The damping of the stage, caused by the inherent damping qualities of the material, together with the method of attaching the blades to the disk, is one. Disk configuration and the type of shrouding are other factors. A Campbell diagram, described below, is merely another one of the tools for designing blades. It cannot alone determine if a blade will operate reliably.
Blade Frequency The simplest method for blade frequency analysis has been to use the natural frequencies of individual blades. This can be misleading because blades vibrate differently in assembled packets. Individual blade frequencies are easy to obtain. However, static packet testing will yield more complete data. Dynamic packet testing is clearly the most accurate method, but it has been and always will be extremely expensive. The most widely-used frequency consideration is the Campbell diagram. Mr. Campbell experimented with disk frequencies back in the early days of steam turbines. There were numerous disk failures then, caused primarily by disk resonance at wheel critical speeds. The industry has since coined the name "Campbell diagram" to mean "interference diagram." These are useful in showing when natural blade frequencies coincide, or interfere, with various operating harmonics such as multiples of running speed or nozzle passing frequencies (NPF). Many engineers feel that if any interference exists, the blade will f a i l - but is this true? Experience indicates it is almost impossible to design a multistage turbine with no interference. The turbine must therefore be designed to operate reliably in certain modes of resonance. 167
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Process Plant Machinery
Example To prove the point that frequency is not the only factor to consider in the design of a turbine blade, let's review a blade in terms of frequency and stress. The example is an 0.847-inch (21.5-mm) blade on an 18-inch (460-mm) diameter. The first Campbell diagram, Figure 5A-l, shows no interference of the fundamental tangential mode with NPF. Keep in mind that these data are based on individual blades. Realize, also, that the various "modes" simply describe vibration behavior in different directions or at multiples of a fundamental vibratory frequency. The second interference diagram, Figure 5A-2, is on the basis of static packet data and the situation worsens; the fundamental axial mode intersects the NPF. (Packets describe several blades joined by a shroud band or similar means.) Centrifugal force will tend to increase the natural frequencies, but this particular blade will still be operating in resonance. The conclusion might be that this blade will Speed R a n g e ~ Fundamental Axial
12000 11000 10000
90001 8000
7000
Fundamental Tangential
.~
6000 w
5000 400O
"
3000
.~"
Harmomc of Running
2000 1000 0
0
1
2
3
4
5
6
7
Rotor Speed--r/ram
8 g 10 11 12 X 1000
FIGURE 5 A - 1 Campbell diagram for an individual steam turbine blade. (Source: The Elliott Company, Jeannette, PA.)
12000~
Speed Range " = "
11000 r
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10000[
9000
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~. 8ooc / cU
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2
3
4
5
6
7
Rotor Speed-r/ram
8 9 10 11 X 1000
12
FIGURE 5A-2 Campbelldiagram for an eightblade "packet" (Source: The Elliott Company, Jeannette, PA.)
Appendix 5A: Steam Turbines: Some Design Theory Factors fail. This is not necessarily so. The stress in the blade has not been considered, and it is a fact that failure occurs only when the operating stress exceeds the endurance limit, regardless of what causes the stress.
Modified Goodman Diagrams Many experienced manufacturers rely on a modified Goodman diagram for stress evaluation. For the ultimate stress, the strength of the blade material at the design operating temperature is used. The endurance limit used is based on the corrected ultimate stress and other considerations such as the fatigue notch factor. A straight line is then constructed between the ultimate stress and the corrected endurance limit. Theory states that failure will occur when the total stress in the material is to the fight of the failure line. The steady state stress is the sum of the centrifugal stress and steam bending stress. This total steady state stress is plotted on the abscissa. A straight horizontal line is then drawn over to the failure line, a vertical line is constructed, and this becomes the allowable alternating stress. This is shown in Figure 5A-3. The allowable alternating stress, divided by the steam bending stress, is the Goodman factor. In theory, the steady state stress can equal the ultimate stress, without failure, when the alternating stress is zero. Further, the alternating stress can equal the corrected endurance limit, without failure, when the steady state stress is zero. Blade reliability is thus seen to be a combination of steady state stresses and alternating stresses. Past operating history has established safe, minimum Goodman factors. The blade in question has a 25.8 Goodman factor, as indicated in Figure 5A-4. This is well above the minimum requirements for this stage. Many manufacturers would feel safe in recommending it for this application even though a Campbell diagram shows frequency interference. They can support their recommendation by showing that this stage has been in actual operation, with both the fundamental tangential and axial modes at nozzle passing frequency, for many years. The example adds emphasis to what was said at the beginning: many factors affect reliability, and it behooves the designer to look at all of them.
I
SULT I~ ~'~ ~'~ ~,~
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GOODMAN DIAGRAM 2.25" B l a d e S ULT -- U l t i m a t e Stress S E I -- U n c o r r e c t e d E n d u r a n c e Limit ( S u L T X .50) SE 2 -- C o r r e c t e d E n d u r a n c e L i m i t (SULT X . 5 0 / 1 . 5 ) 1.50 is Fatigue Notch Factor S G B -- Gas B e n d i n g Stress SC -- C e n t r i f u g a l Stress s..
3~
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Alternating Stress-- PSI
FIGURE 5A-3 Modified Goodman diagram used for blade stress evaluation. (Source: The Elliott Company, Jeannette, PA.)
169
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Process Plant Machinery
GOOOM,,. O..G.,,M 22 " B..O.
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Alternating S t r e s s - P S I X 1000
Data used in modified Goodman diagram. (Source: The Elliott Company,
FIGURE 5A-4
Jeannette, PA.) Inlet S t e a m C o n d i t I o n s
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~ "-- 77%
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1 5 0 0 0 TO 3 0 0 0 0
8 2 % .J
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84%
t
HP
> 50000
Approximate efficiency chart for steam turbines. (Source: The Elliott Company, Jeannette, PA.)
FIGURE 5A-5
Appendix 5A: Steam Turbines: Some Design Theory Factors
171
Steam Consumption (Approximate Steam Rates) F i g u r e 5A-5 and Table 5A.1 can be used to d e t e r m i n e an a p p r o x i m a t e turbine efficiency w h e n h o r s e p o w e r , speed, and s t e a m c o n d i t i o n s are k n o w n . In m a n y instances, an a p p r o x i m a t e efficiency m a y well serve y o u r purpose, since the p a r a m eters m e n t i o n e d m a y change. As can be seen in the e x a m p l e , a 25,000 HP, 5 0 0 0 r/min turbine using s t e a m at 600 p s i / 7 5 0 ~ H g abs. has an a p p r o x i m a t e efficiency of 77 percent. A p p l y i n g this a p p r o x i m a t e efficiency to the theoretical s t e a m rate ( T S R ) results in a s t e a m rate and steam flow as follows: T S R 600 psi/750~
H g abs. = 7.64 l b / k w h r
Approximate efficiencyA p p r o x i m a t e steam rate ( A S R ) -
0a-
77%
TSR/r/a-
(7.64 l b / k w h r ) ( 0 . 7 4 6 k w / H P ) / 0 . 7 7
= 7.40 l b / H P - h r A p p r o x i m a t e steam f l o w -
A S R • H P = (7.40 l b / H P - h r ) ( 2 5 , 0 0 0 H P )
= 185,000 lb/hr A M o l l i e r d i a g r a m can be used to d e t e r m i n e the T S R for s t e a m conditions not tabulated. Figure 5A-6 can be used to find a p p r o x i m a t e s t e a m rates for turbines operating at part-load and speed. F o r e x a m p l e , find the a p p r o x i m a t e steam rate w h e n the 25,000 HP, 5000 r/min turbine w e ' v e discussed is o p e r a t e d at 20,000 H P and 4 5 0 0 r/min. % HP-
20,000 H P / 2 5 , 0 0 0 H P -
80%
% r/min -- 4 5 0 0 r / m i n / 5 0 0 0 r/min -- 9 0 %
TABLE 5A.1 Theoretical Steam Rates (lbs/kwhr)*
2" Hg abs. 4" Hg abs. 0 psig 50 psig 100 psig 200 psig 300 psig 400 psig 600 psig
250 psig 550~
400 psig 750 ~
600 psig 750~
600 psig 850 ~F
900 psig 900 ~F
1,500 psig 900 oF
2000 psig 950 ~F
8.78 9.67 14.57 26.75 42.40 . .
7.36 7.98 11.19 17.56 23.86 43.51 . .
7.08 7.64 10.40 15.36 19.43 29.00 43.72
6.66 7.17 9.64 13.98 17.64 26.33 39.70
6.26 6.69 8.74 12.06 14.50 19.45 25.37 33.22 63.40
6.08 6.48 8.26 10.94 12.75 15.84 18.94 22.32 30.75
5.84 6.20 7.78 10.07 11.55 13.96 16.19 18.49 23.63
. .
. .
Note: Interpolate, where necessary, for approximate values. *From Theoretical Steam Rate Tables, copyright 1969 by A.S.M.E. The TSR for steam conditions not tabulated can be found by using the equation: At 175 psig (190 psia) and an S of 1583 the isentropic TSR = 3413 Btu/kwhr/AHi Btu/lb exhaust enthalpy, H exhaust, would be 1227 Btu/lb. where AHi = isentropic enthalpy drop from AHi would therefore be: inlet conditions to exhaust pressure. AHi = H inlet - H exhaust For example, find TSR for steam conditions of 1250 psig/ = 1439- 1227 900 ~ psig using the Mollier diagram. = 212 Btu/lb At 1250 psig (1265 psia) and 900 ~ The TSR would then be" H inlet = 1439 Btu/lb TSR = 3413 Btu/kwhr/212 Btu/lb S = 1583 Btu/lb~ = 16.1 lb/kwhr Source: The Elliott Company, Jeannette, PA.
172
Process Plant Machinery
1 . 1 5 !1.14 r 1.131--
1.12~1.11 1.10 .1.09 1.08 1.07 1.06 HP
1.05 1 . 0 4 ,1.03 I .O2 1.01
[;,,,
110
105
100
95
90
85
80
75
Per Cent H P O r Speed
FIGURE 5A-6 PA.)
Part-load~speedcorrection curves. (Source: The Elliott Company, Jeannette,
From the curve, the HP correction is 1.04 and the r/min correction 1.05. Total correction is 1.04 x 1.05 - 1.09. The part-load steam rate is therefore 7.40 x 1.09 = 8.06 lb/HP-hr.
TURBINE SELECTION: DESIGN FACTORS Staging, Pressures, and Temperatures The following examples illustrate the performance of various combinations of impulse staging. It should be understood, of course, that stage selection and overall turbine efficiency are affected by many important considerations other than stage efficiency. Speed limitations, mechanical stresses, leakage and throttling losses, windage, bearing friction, and r e h e a t - all these must be factored into the ultimate turbine design. That's the job of the factory specialist. The approximate efficiencies of a Curtis stage and a Rateau stage are shown in Figure 5A-7.
Basic Formulae (Refer to Table 5A.2 for English to metric conversions.) Vb = where
zrDN 720
Vj = 224v/h~ - h 2 = 224~/AH
Vb = Pitch line (blade) velocity, ft/s D = Pitch diameter of wheel, inches (base diameter plus height of blade) N = Rotative speed, r/min Vj = Steam jet velocity, ft/s hi = Inlet steam enthalpy, Btu/lb h2 -- Isentropic exhaust steam enthalpy, Btu/lb h2e "- Stage exit steam enthalpy, Btu/lb AH = Isentropic heat drop, Btu/lb (h] - h 2 ) Vb/Vj = Velocity ratio, dimensionless
Appendix 5A: Steam Turbines: Some Design Theory Factors
"O[ .70
.
g
I
I
.
.
7T' .
I
.-',/
|
FIGURE 5A-7
o
i
dl_gl~
I#/I 'or / 3
i
1
I~
i
i :
:
' i I o ~ .,o .,o.,~.~o , o .,~.~o ~.,oo,,, , . , , o [ v o / ~ j
Velocityratio versus efficiency. (Source: The Elliott Company, Jeannette, PA.)
Example 1" Curtis Stage Performance Conditions" 1500 psig (1515 psia); 950 ~ 5000 r/min, 25-in wheel diameter Assume 1-in blade height Find isentropic heat drop and end point (see Figure 5A-9). V b --
7rDN 720
--
(3.14)(25 4- 1)(5000) 720
= 568 ft/s
From Figure 5A-8, velocity ratio for optimum Curtis stage efficiency = 0.23 Vb/Vj - - 0 . 2 3 ,
Vb 568 Vj = 0.23 = 0.23 = 2470 ft/s
Vj -- 2 2 4 ~ / A H -
2470 ft/s
AH = 121.5 Btu/lb hi (from Mollier chart or steam t a b l e ) - 1459.9 Btu/lb he - 1459.9-121.5 -- 1338.4 Btu/lb Exhaust pressure (from Mollier chart) = 590 psia
TABLE 5A.2
English-Metric Conversion Table
To Obtain kg/cm 2 Atmospheres kg/cm 2 mm of mercury kw
~ cm kg
Multiply
By
psig psig Inches of mercury CHg abs.) Inches of mercury CHg abs.) HP ~ -32 Inches Pounds
0.0703 0.06804 0.03453 25.4 0.746 0.556 2.54 0.454
Source: The Elliott Company, Jeannette, PA.
173
174
ProcessPlant Machinery Assuming a stage efficiency of 70 percent, the stage exit conditions are: exhaust pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . 590 psia h2e = 1 4 5 9 . 9 - (0.7)(121.5) . . . . . . . . . . . . . . . 1374.8 Btu/lb
Example 2: Rateau Stage Performance Conditions: 400 psig (415 psia); 600~ 5000 r/min, 35-in wheel diameter Assume 1-in blade height Find isentropic heat drop and end point (see Figure 5A-9). Vb =
7rDN 720
(3.14)(35 + ])(5000) 720 = 785 ft/s
From Figure 5A-7, velocity ratio for optimum Rateau stage efficiency = 46% V b / V j = 0.46,
Vb 785 Vj = 0.46 -- 0.46 = 1705 ft/s
Vj = 224~/AH = 1705 ft/s AH = (Vj/224) 2 = (1705/224) 1 = 58.0 Btu/lb hi (from Mollier chart or steam tables) = 1305.7 Btu/lb h2 = 1 3 0 5 . 7 - 58.0 = 1247.7 Btu/lb Exhaust pressure (from Mollier chart) = 230 psia Assuming a stage efficiency of 80 percent, the stage exit conditions are: exhaust pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . 230 psia h2e = 1 3 0 5 . 7 - (0.8)(58.0) . . . . . . . . . . . . . . . . 1259.3 Btu/lb
Example 3: Straight Rateau Staging Conditions: 400 psig (415 psia); 600 ~ Exhaust pressure 100 psig (115 psia)
1459.9
~
~ .'~950~
~~
~7
1374.8
i I
/ /
1338.4
Entropy
FIGURE 5A-8
Jeannette, PA.)
Entropyversus enthalpy chartfor Example 1. (Source: The Elliott Company,
Appendix 5A: Steam Turbines: Some Design Theory Factors
1305.7
t
/
, ~/ /
-~
/
--
//
=
/
__ m _ _ ~
600 ~ F
I1130 ~,
/ ~ / v.
I'/i
_~ I /
/<~ /
=7 V
"' ~259.3/[-- . . . . .
~t / ~
/
I
-~176
/ ~ 4 7 0 ~
1247 7
EntroPy
FIGURE 5A-9 Entropy versus enthalpy chart for Example 2. (Source: The Elliott Company, Jeannette, PA.)
1305~ 7
~ . . . .
eOO~
m ..J Ira >, 1250 Q
i
.IE: .,.., UJ
i t/
1208.5 1200 1184.2
\~
I __ D m
I
//
/
1150
Entropy
FIGURE 5A-10 Jeannette, PA.)
Entropy versus enthalpy chart for Example 3. (Source: The Elliott Company,
Find the number of 35-in diameter Rateau stages required, assuming optimum stage efficiency (see Figure 5A-l 0). From Mollier chart, the isentropic heat available is: 1305.7-
1184.2-
121.5 Btu/lb
Alternatively, using a TSR table or Polar Mollier chart: T S R - 28.08 lb/kwhr
175
176
Process Plant Machinery Using the basic definition of TSR: 3413 = -- 121.5 Btu/lb TSR 28.08 From Example 2, the optimum isentropic heat drop per Rateau stage = 58.0 Btu/lb Approximate stages required = 121.5/58.0 = 2.1 (thermodynamically) The turbine will require 2 Rateau stages. Assuming 80% stage efficiency" 121.5 x 0.80 = 97.2 Btu/lb 1305.7 - 97.2 = 1208.5 Btu/lb AH = hi - h2 --
3413
Example 4: Curtis and Rateau Staging Conditions: 1500 psig (1515 psia); 950~ Exhaust pressure 150 psig (165 psia) Find the number of 35-in-diameter Rateau stages required when using one 25-indiameter Curtis stage, assuming optimum stage efficiencies (see Figure 5A-11). From Example 1, we found that a Curtis stage removes 121.5 Btu/lb 1 4 5 9 . 9 - 121.5 = 1338.4 Btu/lb pressure = 590 psia Assuming 70% stage efficiency: 121.5 • 0.70 = 85.05 Btu/lb 1459.9 - 85.05 = 1374.8 Btu/lb From this end point to 165 psia, the isentropic heat drop for the Rateau stages is 1374.8 - 1239.0 = 135.8 Btu/lb From Example 2, the optimum isentropic heat drop per Rateau stage = 58.0 Btu/lb Approximate stages required = 135.8/58.0 = 2.34 (thermodynamically) The turbine will therefore require one 25-in Curtis stage and two 35-in Rateau stages. Assuming an overall Rateau efficiency of 81%, the end point will be 1374.8 - 0.81 (137.3) = 1264.8 Btu/lb
1459.6
I
------~~,
e/
m ..J I-~,~ 1374.5 (3. .r
~ 1338.0 m
~v~ A~"
7
___
I
/
"
1262.8 i-- . . . . / . . . . 1236.6 i-" 1211.0
950~
I/ i
'lL,.,--." 676~ ~o
i i
i l;='=l
i Ii
I _ _.1 _ , , V ~ ,-'T~ - J ' ~
------
I
/
o<
482~
435~
390~
Entropy
FIGURE 5A-11
Jeannette, PA.)
Entropy versus enthalpy chart for Example 4. (Source: The Elliott Company,
Appendix 5A: Steam Turbines: Some Design Theory Factors
EXTRACTION TURBINE PERFORMANCE Today's fuel costs demand that the maximum amount of energy be squeezed from each pound of steam generated. To help in this effort, when both process steam and shaft power are required, many plant designers are turning to extraction turbines. The extraction turbine can substantially reduce the energy charged to the driven machine if process steam would otherwise be supplied through pressure-reducing valves. Even though back-pressure turbines can be used to supply process steam, they are rather inflexible, since the shaft power and process steam requirements must be closely matched. An extraction turbine, however, can cope with changes in these variables and satisfy the requirements of each over a broad range. The diagram in Figure 5A-12 shows the performance map for a typical extraction turbine. Determining the shape of this diagram is a problem that often 34(
Z6C
~40
;;2~ ooI-_
,
,~-x~"
,8o -
A
.~ G~'~
80 -
o,,O~ ~f~ ,~ ~"
o90 o~'~
b4o
,
, .oo
20
O0
~ ~~
~
80
~
0
40 20
Le 0
1
5
1
10 Shaft
1
15 HP/1000
I
20
I
25
FIGURE 5A-12 Performance map for typical extraction turbine. (Source: The Elliott
Company, Jeannette, PA.)
177
178
Process Plant Machinery
arises. Here is an example that demonstrates the procedure to follow in drawing an approximate extraction diagram. Assume: Shaft HP and speed: 25,000 HP at 4500 r/min Steam conditions: 600 psig/750 ~ in Hg abs. Extraction requirements" 150,000 lb/hr @ 250 psig First tabulate the TSRs. TSR (inlet to extraction) 600 psig/750 ~
psig = 35.4 lb/kwhr
TSR (inlet to exhaust) 600 psig/750 ~
in Hg abs. = 7.64 lb/kwhr
Now assume that the efficiency of the entire turbine (inlet to exhaust) is 75 percent and the efficiency of the inlet-to-extraction section is 70 percent. Therefore: approximate steam rate (SR) (inlet to extraction) will be (35.4 lb/kwhr x 0.746 kw/HP) - 0.70 = 37.8 lb/HP-hr approximate SR (inlet to exhaust) will be (7.64 lb/kwhr x 0.746 k w / H P ) - 0.75 = 7.60 lb/HP-hr Point A is the first point to be located on the diagram by multiplying Total HP x approximate SR (inlet to exhaust) = 25,000 HP x 7.60 lb/HP-hr = 190,000 lb/hr Locate Point A at 25,000 HP and 190,000 lb/hr throttle flow. Point B is located at zero HP and a throttle flow of 5 percent of A or 0.05 x 190,000 lb/hr = 9500 lb/hr The zero extraction line results from connecting Points A and B. Point C is located by dividing extraction flow requirement by approximate steam rate (inlet to extraction)" (150,000 l b / h r ) - (37.8 lb/HP-hr)= 3970 HP Locate Point C at 3970 HP and 150,000 lb/hr throttle flow. Now draw line C-D parallel to A-B and another line C to B. Label A-B "zero lb/hr extraction at 250 psig" and C-D "150,000 lb/hr extraction at 250 psig." Notice that the general shape of the diagram and the slopes of the lines are determined mainly by the steam conditions used. The turbine investigated here could be built, but one must remember that we have dealt only with the thermodynamic aspects of this application. The mechanical aspects, such as blade stresses, nozzle flow limits, cooling steam, and other factors must also be checked. This example can also be carded further to determine the number of stages in each section elsewhere. To do so requires finding the energy available in each portion of the turbine. The energy available (AHi) to the inlet-to-extraction section is as follows: AHi -- 3413 Btu/kwhr -- TSR (inlet to extraction) = 3413 Btu/kwhr - 35.4 lb/kwhr = 96.4 Btu/lb Wb =
zr(25 + 1)(4500)/720 = 511 ft/s
Appendix 5A: Steam Turbines: Some Design Theory Factors Vj for an ideal Curtis stage would therefore be: Vb 511 = 2220 ft/s 0.23 -- 0.23
Vj-
Vj for an ideal Rateau stage would be" Vb 511 Vj = 0.46 = 0.46 = 1110 ft/s With a AHi of 96.5 Btu/lb, Vj for the inlet-to-extraction section with one stage would be: Wj -~- 224~/AH = 224~/96.5 = 2200 ft/s This is seen to be very close to the 2220 ft/s for an ideal Curtis stage. We will therefore assume that the inlet-to-extraction section will contain one 25-in Curtis stage. Now for the extraction-to-exhaust section. To find the energy available to this section we need the temperature of the steam entering this portion of the turbine (extraction steam temperature). The enthalpy of this steam will be as follows: Inlet steam e n t h a l p y - AHi(inlet to extraction) x r/(inlet to extraction) = 1378 B t u / l b - 96.4 Btu/lb • 0.70 = 1378 B t u / l b - 67.5 B t u / l b - 1310.5 Btu/lb From a good Mollier diagram, at 250 psig and 1310.5 Btu/lb the extraction steam temperature is found to be close to 590 ~ (say 600 ~ The extraction-to-exhaust portion of this turbine therefore operates on steam conditions of 250 psig/600 ~ in Hg abs. TSR (extraction to exhaust) is 9.35 lb/kwhr. The energy available to the extraction-to-exhaust section is therefore: AHi -- (3413 Btu/kwhr) -- (9.35 lb/kwhr) -- 365 Btu/lb The blade velocity for 35-in nominal diameter staging with a l-in blade height will be" Vb = rr(35 + 1)(4500)/720 = 706 ft/s If all staging is of the Rateau type in this portion of the turbine" Vj-
706/0.46 = 1535 ft/s
AHi per stage is therefore" AH~ =
~
--
224
= (6.85) 2 = 47.0 Btu/lb
The number of Rateau stages in this section would therefore be: Total energy available Energy removed per stage
=
365 47.0
= 7.77 (say 8)
This turbine will, therefore, contain one 25-in diameter Curtis stage followed by eight 35-in diameter Rateau stages with the extraction opening after the Curtis stage.
STEAM BALANCE CONSIDERATIONS The steam balance of a process plant can be quite complicated due to the multiple steam pressure levels often required.
179
180
Process Plant Machinery
Selecting a turbine to complement a particular steam balance is made easier, however, by the wide variety of turbines available. Condensing, back-pressure, or extraction/induction turbines can be used, as required, in designing both new plants and additions to existing plants. Steam for process use, for example, can be supplied from the exhaust of a back-pressure turbine or from an extraction turbine. The choice would depend on the number of pressure levels involved, the design of the remainder of the plant, the number of turbines required, etc. This versatility simplifies the job of optimizing a steam balance. The steam balance diagrams shown in Figures 5A-13 through 5A-15 illustrate how various types of turbines have been used to supply both shaft horsepower and steam for other uses. TYPICAL STEAM BALANCE
J
1500 P S I G / 9 O O ~
|
~
Pressu re Control
HzO
|
103.4 bar/482~ 25000 HP/18 650 kW--4500 r/rain
20000 HP/14 920 kW--4800 r/min
135 mbar Temperature Control
~1,
I~
L . ~ 5000 HP/3730 kW--10500 r/min ~ . I " ~ . ' I'f'l ~" I
Flow A 266000 LB/HR 121 000 kg/h Flow B 256000 LB/HR 116 000 kg/h Flow C 50000 LB/HR 23 000 kglh
,o,,
i
-
v
~ 4" HGA Extraction To ( ~ ) 135 tuber Process II 150000 LB/HR 68 000 kg/h at 255 PSlG at 17 6 bar
Extraction To Process I
140000 LB/HR at 400 PSIG
135 mbar
84 000 kg/h at 27.6 bar
FIGURE 5A-13 Steam balance, Example 1. Steam is extracted from turbine A at 400 psig, and steam is extracted from turbine B at 225 psig. (Source: The Elliott Company, Jeannette, PA.) TYPICAL STEAM BALANCE 900 PSIG/850~ I(~)
H20
~
Temperature Control
1w
Flow A Flow B Flow C Total Boiler Flow
62 bar/455~ 41000 HP/30 600 kW--4100 r/rain
l|
35000 HP/26 100 kW-3800 r/min
~ 4 - H G
Jr ~
135 mbar
340000 LB/HR 154 000 kg/h 320000 LB/HR 145 000 kg/h 153000 LB/HR 69 000 kg/h 813000 LB/HR 368 000 kg/h
J
~
Extraction To Process I
220000 LB/HR at 410 PSIG
i
7000 HP/5220 kW-8000 r/rain
A
135 mbar
100 000 kg/h at 28.3 bar
Exhaust To Process II
153000LB/HR at 190 PSIG
69 000 kg/h at 13.1 bar
FIGURE 5A-14 Steam balance, Example 2. Three turbines all use steam from the 900psig/850~ (62 bar/4550C) boiler. Steam is extracted from both larger units at 410 psig (28.3 bar)for process I, and the remainder is condensed at 4 in Hg abs. (135 mbar). Smaller back-pressure turbine exhausts steam at 190 psig (13.1 bar)for use in process II. (Source: The Elliott Company, Jeannette, PA.)
Appendix 5A: Steam Turbines: Some Design Theory Factors
181
TYPICAL STEAM BALANCE 850 PSIG/9OO~
|
58.6 bar/482~
7000 HP/5220 kW--8500 r/min
H20 Temperature Control
1
Exhaust To Process
i l l
64000 LB/HR 29 000 kg/h
Flow A Flow B Flow C Total Boiler Flow
155 PSIG/6 (~) 10.7 bar/315~
l
6000 HP/4475 kW--8000 r/rain
145000 LB/HR 66 000 kg/h 65000 LB/HR 30 000 kg/h 16000LB/HR 7 000 kg/h 145000 LB/HR 66 000 kg/h
~
~
3H120 5" mbGaAr
15~00 HP/1120 kW-10800 r/man
+"
FIGURE 5A-15 Steam balance, Example 3. Topping turbine concept was used in this plant, where exhaust from a high-pressure turbine as well as the low-pressure turbines is used to supply process demands. Exhaust steam from these low-pressure units is then condensed at 3.5 in Hg abs. (120 mbar). (Source: The Elliott .Company, Jeannette, PA. ) We note from Figure 5A-13 that two 1500 psig/950~ turbines drive the large compressors in this application. Two different extraction pressures were used (400 psig and 255 psig), with lower pressure steam being supplied to a process and to the smaller turbine. Exhaust steam from all three turbines is then condensed at 4 in Hg abs.
APPEN DIX 5B Selecting and Sizing a Steam Turbine*
There are three main Steam Turbine model types (generic model letters and numbers in this appendix refer to the contributing source's model designation). (1) High-pressure backpressure turbine type KR after KanisR6der; (2) Backpressure turbine type G (3) Condensing turbine type V. The turbines of these model ranges are not only used as single-casing machines for driving generators and driven machines but, by combining several turbines, can also be employed as dual- and triple-casing steam turbosets. Our standard design includes the possibility of coupling other machinery on either end of the turbines. The turbosets can be equipped with reheating systems, multiple feedwater heaters and various extraction points. Secondary steam admissions are also possible by standard.
BACKPRESSURE TURBINE TYPE G (TABLE 5B.2 AND FIGURE 5B-1) The backpressure turbine type range G, comprising eight model sizes, covers an output range from 1 to over 100 MW. With the exception of the largest type G80, all the maximum type speeds of the standard program are higher than those of the two-pole alternators and the smallest type attains a speed of 16,000 RPM. Since it is quite normal nowadays to use gear units to transmit outputs of up to approx. 40 MW, a direct alternator drive is also available only from type G40 onwards. The maximum live steam conditions of the standard model comply with the values normally used in industry today. The maximum casing backpressure for continuous operation lies at 16 bar. At backpressures of 6 bar and above, leakage steam is extracted from the shaft labyrinth glands. The type designation relates to the nominal diameter of the blading in cm. The product of blading diameter and maximum type speeds is fairly constant for all sizes, and all types are similar. The maximum inlet volume via the nozzles is limited by the possible nozzle cross-sectional area at maximum type speed. The maximum inlet volume via the bypass is dictated by the cross-section available in the valve chest. The maximum outlet volume is determined by the permissible steam velocity in the exhaust nozzle, the velocity being lower with higher pressures.
182
volumes at the turbine inlet, and higher speeds or outputs, it would be necessary to consult the design engineers.
Description and Dimensions, Type G (Figures 5B-2, 5B-3 and Table 5B.3) The backpressure turbines of the type range G are designed as multistage reaction turbines with approx. 50 percent reaction. A single-row or, with direct alternator drivers, also a double-row flow-controlled impulse control wheel is provided upstream of the reaction section that is always fitted with shrouded blading. The guide blades are mounted in thermoelastic guide-blade carriers which are suspended centrally in the casing. The nozzle chests for steam admission to the control wheel are assembled separately in the upper and, when necessary, in the lower part of the casing and welded in position. They are connected to the valve block, arranged above the main casing in most cases, by means of welding nozzles and piping. The horizontally divided turbine casing is of simple and fully symmetrical construction. It serves only to guide the steam pressure and to accommodate the guide-blade carrier and the labyrinths. It is symmetrically supported on the two bearing housings. Radial bolts serve to avoid a displacement of the casing center line with regard to the bearing center and thus to the rotor axis under all operating conditions. The turbine rotor, manufactured from one single forging, is guided axially in the front bearing housing. The rear bearing housing is usually designed as the fixed point, so that the exhaust-side shaft end is subjected only to minimal axial movement at startup.
CONDENSING TURBINES TYPE V (TABLE 5B.4 AND FIGURE 5B-4) The condensing turbine type range V consists of eight model sizes and covers an output range from 1 to over 100 MW. With the exception of the largest type V90, the maximum type speed is higher than that of the two-pole alternators and the smallest turbine attains a speed of 14,000 RPM. Since it is quite normal nowadays to use gear units to transmit outputs of up to 40 MW, direct alternator drive is only provided for from type V50 upwards. The maximum live steam conditions are comparable with those of the backpressure turbines of similar size. The exhaust casings are designed for backpressures of less than 1 bar, but will withstand short-time atmospheric exhaust operation in case of a failure of the evacuating equipment, for example.
In cases where the operating conditions differ from the standard design parameters, such as higher backpressures, larger
The type designation corresponds to the nominal diameter of the high-pressure blading up to type V32, and from type V40 upwards to the mean value of the HP and LP blading diameters. All types are of similar thermodynamic design.
* Source: ABB Turbinen GmbH, N0mberg, Germany
The maximum inlet volumes via the nozzles are limited by the nozzle cross-sectional areas available at maximum type
Appendix 5B: Selecting and Sizing a Steam Turbine TABLE 5B.1
183
Multlstage Steam Turblnes (Source: ABB Turblnen GmbH, NOrnberg, Germany)
Model range KR Normal operating range
Max. inlet steam condition*) bar/~ Max. backpressure*) bar Output range MW Type speed range RPM
190/580 50 0.5-90 3.000-18,000
[
7
Design possibilities Controlled extraction Bleed steam Type of construction No. of stages Flow control Throttle control
X reaction multistage X X
J
Model range G Normal operating range
Max. inlet steam condition*) bar/~ Max. backpressure* bar Output range MW Type speed range RPM
Design possibilities
Controlled extraction Bleed steam Type of construction No. of stages Flow control Throttle control
140/540 16 0.5-150 3,000-16,000 X X reaction multistage X X
r
[-
0
~,.
Model range V Normal operating range
Max. inlet steam condition*) bar/~ Max. backpressure*) bar Output range MW Type speed range RPM
Design possibilities
Controlled extraction Bleed steam Type of construction No. of stages Flow control Throttle control
140/540 1 0.5-125 3,000-14,000 X X reaction multistage X X
*~ Higher ratings with specially built turbines X Design possible - Design not possible
speed, whereas the bypass inlet volumes are determined by the maximum steam velocity in the valve chest.
Description and Dimensions, Type V (Figures 5B-5, 5B-6 and Table 5B.5)
In cases where the operating conditions differ from the standard design parameters, such as higher live steam pressures and temperatures, larger inlet volumes, or higher outputs and speeds, it would be necessary to consult the design engineers.
The condensing turbines, type range V, are designed as multistage reaction turbines with approx. 50 percent reaction. A single-row or double-row flow-controlled impulse wheel is incorporated ahead of the shrouded HP blading. The LP blading, not equipped with shrouding on account of the
184 Process Plant Machinery TABLE 5B.2 Standard Backpressure Turbine Range G Type Max. type speed Turbine speed with generator 50 Hz Turbine speed with generator 60 Hz
G 16
G 20
G 25
G 32
G 40
G 50
G 63
G 80
RPM RPM
16000 12000
12000 12000
10000 10000
8000 8000
6300 3000
5000 3000
4000 3000
3600 3000
RPM
16000
12000
10000
8000
3600
3600
3600
3600
Nominal type output*)
MW
6
10
16
25
40
63
100
150
Max. intel steam conditions*) (permissible deviation at cont. operation 5%/8.3 ~ Max. backpressure*) with/without shaft labyrinth steam leakoff Max. inlet flow via nozzles via bypass Max. inlet steam condition with bypass Max. exhaust volume at 3 bar
bar/~
133/532
133/532
133/532
133/532
133/532
133/532
133/532
133/532
bar/bar
6/16
6/16
6/16
6/16
6/16
6/16
6/16
m3/s m3/s bar/~C
0.92 1.20 63/500
1.44 2.20 85/500
2.15 3.50 851500
3.24 5.00 85/500
7.95 10.00 85/500
12.70 17.50 85/500
5.0
8.3
13.1
21.2
33.5
53.0
85.0
160
200
320
400
m3/s
Nominal blading diameter
mm
Inlet flangesmin./max, Exhaust flanges, size 1 size 2
mm mm
1001150 100/200 300 300 400
250 150/250 500
5.18 6.80 851500
2001350 250/2x350 600 700
500
630
250/2x350 900 1000
300/2x400 1200 1600
Further technical details upon request
,
*) Higher ratings possible with specially built turbines.
Fig. 5B-1 Selection diagram for backpressure turbines type G (approximate values). (Source: ABB Turblnen GmbH, N0rnberg, Germany) centrifugal forces, is usually fitted with damping wire. All guide-blade rows are mounted in carriers which are centrally suspended in the turbine casing. The nozzle chests for steam admission to the control wheel are welded into the upper and, where necessary, also the lower part of the casing. They are connected to the valve block above the main casing with welding nozzles and pipes. The horizontally split turbine casing is of symmetrical construction. The exhaust casing of welded steel is vertically flanged on. Symmetrically to its center line the turbine casing rests on two lateral supports on the exhaust casing and on the front
bearing housing. By means of radial bolting with the latter a displacement of the casing center line from the bearing and rotor centres is avoided at all operating conditions. The turbine rotor, machined from one single forging, is guided axially in the front bearing housing. The exhaust flange support serves as fixed point of the machine. The turbine casings can be combined with various exhaust casing sizes depending on the exhaust volumes. Several standard blade-row combinations are available for the condensing section to optimally adapt the turbine to the given speeds and vacuum conditions.
Appendix 5B: Selecting and Sizing a Steam Turbine
185
Fig. 5B-2 Standard backpressure turbine type G, longitudinal cross-sectlon. (Source: ABB Turblnen GmbH, N0rnberg, Germany)
L1 b---~
"
~ - - - e -----~
I
g
LI .~
21"
k --
~ir
I
t
m
30 Fig. 5B-3
H
I
External dimensions, backpressure turbine type G. (Source: ABB Turblnen GmbH, N0rnberg, Germany)
The Building-Block System
expenditure and technical effort- can be largely adapted to the operating conditions of the individual turbine user.
The design of a steam turbine is dictated by many parameters (steam conditions, steam flow, output, speed). Therefore turbine models are required t h a t - w i t h a minimum of
We use a building-block system which fulfills the demand for economic production and highly developed, reliable components. The live steam valve blocks are standardized as to
186 Process Plant Machinery TABLE 5B.3
Dimensions of Figure 5B-3 Turbine Type
Dimension
G 16
G 20
G 25
G 32
G 40
G 50
G 63
G 80
LI*) H H1 b e g h i k m
1800 1800 600 600 600 1200 320 700 1200 650
2200 1800 680 750 750 1200 400 750 1300 650
2350 1900 800 800 900 1400 400 800 1500 850
2750 2500 920 1000 1000 1700 500 950 1700 1050
3250 3000 1120 1100 1100 1800 500 1100 1800 1200
4400 3000 1300 1400 1400 2000 630 1250 2000 1350
5600 3200 1700 1700 1600 2100 710 1400 2100 1500
5600 3200 1950 1800 1800 2250 1100 1600 2250 1750
*)for shortest bearing distance. All dimensions in mm and non-binding.
TABLE 5B.4
Standard Condensing Turbine Range V
Type
Max. type speed Turbine speed with generator 50 Hz Turbine speed with generator 60 Hz Nominal type output*) Max. inlet steam conditions*) (permissible deviation at cont. operation 5%/8.3 ~ Max. inlet volume via nozzles via bypass Max. inlet steam condition with bypass Max. exhaust volume 1 Max. exhaust volume 2 Nominal blading diameter inlet flange min./max. Exhaust flange, size 1 Exhaust flange, size 2 Exhaust flange, size 3
V 20
V 25
V 32
V 40
V 50
V63
V 80
V go
RPM RPM
14000 12000
12000 12000
10000 10000
8000 8000
6300 3000
5000 3000
4000 3000
3600 3000
RPM
12000
12000
10000
8000
3600
3600
3600
3600
MW
6
10
16
25
40
68
100
125
bar/~C
133/532
133/532
133/532
133/532
133/532
1331532
133/532
133/532
m3/s m3/s
0.92 1.20
2.15 3.50
3.24 5.00
5.18 6.80
7.95 10.00
12.70 13.60
12.70 17.50
bar/~ m3/s
641500 40
841500 63
841500 100
84/500 160
841500 250
841500 400
84/500 630
m3/s
50
80
125
200
320
530
840
mm
200
250
320
400
500
630
800
mm
100/150
mm mm
1000 x 390
1001250
1501300
2001350
250/2 x 350
250/2 x 400
25012 x 400
1300 x 390
1600 x 480
2000 x 580
2500x 700
3200x 900
3980 x 1130
1600 x 390
2000 x 480
2500 x 580
3200x900
mm
3980 x 1130
5300x 1570
5300x 1570
5300 x 1900
Further technical details upon request
5300 x 1900
*) Higher ratings possible with specially built turbines.
their nominal diameter and pressure and are combined with the different turbine casings of the individual type ranges. Dependent on the live steam conditions and steam flow, the steam is admitted to the casing either from the top (Figure 5B-7) or from top and bottom (Figure 5B-10). The length of the bladed section in the standard turbine casings is so selected that at each type speed and with the statistically most frequent heat drop through the turbine optimal efficiency is obtained. In the event of lower speeds (e.g. compressor drivers) or larger heat drops, the length of the bladed section and consequently the number of stages can be increased by lengthening the turbine casing in steps as shown in Figures 5B-9 and 5B-11. in the case of backpressure
turbines up to four extension rings can be incorporated into the normal standard model. The turbine models are in addition sized according to their pressure rating. The exhaust flanges can either be welded to the top or to the bottom of the fully symmetrical backpressure casing. The condensing turbine range is designed in the same manner and follows the same principle as the back-pressure turbine range (valve block and casing systems). The length of the bladed section can be increased in the high-pressure casing by incorporating extension rings, and in the lowpressure casing by using a longer and wider size of the welded rectangular exhaust casing (Figure 5B-9). Several
Appendix 5B: Selecting and Sizing a Steam Turbine 187 Pl P2 tI twe Nk D m
Inlet steam pressure Outlet steam pressure Inlet steam temp. CW inlet temp. Turbine coupling output Inlet steam flow CW/steam flows
Example Given: Inlet steam pressure Inlet steam temp. CW inlet temp. Coupling output
40 bar 400~ 30~ 8 MW
Result: with selected m ( 4 0 - 8 0 ) to two gives possible P2 from P2 horiz, to P l vert to t 1 horiz, to Nk vert down gives necessary steam flow (34 t/hr). intersection from D and P2 gives type V 40
Flg. 5B-4 Selectlon dlagram for condenslng turblnes type V, slze 1 (approxlmate values) (Source: ABB Turblnen GmbH, N0rnberg, Germany)
Fig. 5B-5 Standard condensing turbine type V, longitudinal cross-sectlon. (Source: ABB Turblnen GmbH, N0rnberg, Germany) exhaust casing sizes, each suitable for different vacuum conditions, can be supplied. Various last-row blade heights or blading cross sections are available for each exhaust nozzle. The exhaust casings can be oriented in upward direction, if required. Owing to the standard modules, the condensing turbines can also easily be designed as double-flow machines (type V ... Z, Figure 5B-12). Thus it is possible to double the exhaust volume capacity at a given type speed, which is frequently
of great advantage to the mechanical drive of high-speed compressors and pumps. The standard bearing housings are used for both the G and V type turbines as well as for other type ranges. Each bearing housing size can be used for several turbine sizes. The exhaust-end bearing housings of the backpressure turbines also serve as front bearing housings, when the machine is used as a two-end driver or does not require a pump or governor drive.
188 Process Plant Machinery
4------- g
--L1
e
i
" r
.'1
i
r
k
I
Illl
H
.,.,.
H1
,,, 9
"
I
Flg. 5B-6 TABLE 5B.5
External dlmenslons, condensing turbine type V. (Source: ABB Turblnen GmbH, N0rnberg, Germany) Dlmenalons of Flgure 5B-6
type exhaust volume for backpressure turbines or condensing turbines respectively.
,
Turbine Type Dimension
V20
V25
V32
V40
V50
V63
V80
LI") H H1 b e g h i k m
1900 1800 700 750 250 1200 320 700 1200 450
2200 2000 750 800 300 1250 400 750 1250 500
2350 2300 800 900 400 1350 400 800 1350 750
2900 3000 900 1000 400 1500 500 1000 1500 900
3400 3000 1000 1250 450 1600 500 1200 1600 1100
4700 3200 1400 1500 650 1850 630 1300 1850 1300
5450 3200 2200 1800 1000 2100 630 1500 2100 1600
*)for shortest bearing distance and exhaust flange size 1. All dimensions in mm and non-binding.
Extraction Turbines Turbines of the type ranges G and V can also be supplied with controlled extractions (types GE and VE). Through these extractions steam can be taken from the turbine at a given controllable pressure and be made available to other consumers. Since the distribution of the heat drop in the turbine ahead of the extraction point (HP section) and after it (LP section) may vary greatly depending on the pressure relationships p~, PE, P2 (live steam pressure, extraction pressure, backpressure), the extraction point must be located at different positions in the casing (Table 5B.6). The extraction point rating depends on the extraction pressure and on the maximum extraction volume which should not exceed about 45 percent or 20 percent of the maximum
The maximum possible exhaust steam flow and the maximum live steam flow are dictated by the maximum inlet and outlet volumes of the individual machine type (see Tables 5B.2 and 5B.4). The standard modules for steam extraction are fitted in the place of the extension rings (Figures 5B-9 and 5B-11) or, with alternatives 2 and 6 (Table 5B.6), ahead of the guide-blade carrier support. These standard modules for steam extraction incorporate either nozzle-group segments for a subsequent control wheel (alternatives 1,2,3 and 5, Table 5B.6) or include the throttle control elements for the low-pressure section (alternatives 4, 6 and 7). The extraction control valves can be arranged in the upper or lower casing or in both casing halves. They are actuated by laterally arranged hydraulic servomotors that are isolated from the hot parts (Figure 5B-17).
DOUBLE-EXTRACTION TURBINES Larger single-casing turbines may also be equipped with two controlled extraction points (Figure 5B-18), if a second network is to be supplied with pressure-controlled steam.
BLEED TURBINES Bleed points on turbines permit the uncontrolled extraction of steam. They are mostly provided for the supply of feedwater heaters, for the single or multistage heating of the turbine condensate, or for smaller steam consumers where variations in pressure are permissible or where the throttling losses of the succeeding reducing valves may be neglected. The bleed pressure depends on the location of the bleed point in the blading section and on the machine load or on the onflowing steam quantity, and will drop to almost backpressure level
Appendix 5B: Selecting and Sizing a Steam Turbine
189
Flg. 5B-7 Caslng cross-sectlon wlth upper nozzle chest arrangement. (Source: ABB Turbinen GmbH, N0rnberg, Germany)
Fig. 5B-8 Basic buildlng-block concept of a backpressure turbine casing. (Source: ABB Turblnen GmbH, N0rnberg, Germany) under no-load conditions. The maximum pressure is reached at full load (Figure 5B-19).
bleed pressure in the casing and the required system pressure exists (Figure 5B-20).
If, however, a network is to be supplied with constant pressure, it is possible to provide several bleed points in the blading section and to open only that point with the help of an automatically operating pressure-dependent bleed point selection station, at which the least pressure drop between the
CONTROL VALVES To obtain good steam consumptions at partial loads, the turbines are designed with flow control. Based on their long experience ABB Turbinen N0rnberg have developed a consecutive compound control valve system (Figure 5B-21) incorporating
190 Process Plant Machinery
Fig. 5B-9 Condensing turbine casing in building blocks with )ne extension ring in the HP section and two exhaust sizes. (Source: ABB Turbinen GmbH, NOrnberg, Germany)
Flg. 5B-10 Germany)
Cross-sectlon of casing, nozzle chests arranged above and below. (Source: ABB Turblnen GmbH, N0rnberg,
very few and well-protected moving parts, which ensures reliable steam flow control. Only one spindle passage must be sealed with a gland bush. The control valves are housed in tubular cast-steel casings that are most suited for high-pressure steam. Up to three of these casings are welded together in one block, dependent on the number of valves required. The wall thickness of the casings is selected in accordance with the nominal pressure. The emergency trip valve is located at one end of the valve block, the hydraulic valve actuator at the other.
When loading the turbine, valve one is first actuated by the driving rod of the main cylinder. When the first valve is fully open, the clearance to the second valve is covered and valve two starts to open. Then follows the third valve. If more valves are required for the control of the turbine, or if a second emergency trip valve is to be provided, additional valve blocks can be arranged on top of the turbine casing and connected to it by means of pipes (Figure 5B-26). Each valve spindle is supported at both ends with a short bearing span to avoid hazardous spindle breakage. The
Appendix 5B: Selecting and Sizing a Steam Turbine
191
Fig. 5B-11 Backpressure turbine casing in building blocks with two extension rings. (Source: ABB Turbinen GmbH, N0rnberg, Germany)
Fig. 5B-12 Double-flow condensing turbine type V . . Z with two upward exhaust nozzles. (Source: ABB Turbinen GmbH, N0rnberg, Germany)
192 Process Plant Machinery TABLE 5B.6 Alternative
Standard Alternatives for the Location of the Extraction Modules HP part
1 2 3 4 5
1 W+RS*) 1 W 1W+RS 1 W + RS RS
7
RS
6
1 W
LP part
Remarks with with with with with with with and
1W+RS 1W+RS 1W RS 1 W + RS RS RS
favorable heat-drop distribution small HP heat drop small LP heat drop throttle and bypass control of LP part throttle and bypass control of HP part small HP heat drop, otherwise as 4 throttle and bypass control of the HP LP parts
*)W = control wheel, RS = multistage reaction section. spindles are protected from torsion. The withdrawable valve cages are so designed that their resistance to steam flow is minimal. The valve internals are manufactured from high quality chromium-nickel steel. Valve seats and passages are hardened, the valve spindles are nitrided, and the springs, which are isolated from the steam, are made of nimonic nickel alloy or stainless spring steel. The valve actuator, arranged separately from the hot valve block (Figure 5B-21 ) is operated by pressurized oil via the oil relay. The impulse pressure P3 from the speed governor opens the inlet Pl when the turbine load is increased, or the outlet P0 when the load is gradually reduced. In case of a sudden load drop the patented quickclosing device at the bottom of the servomotor is opened by means of a second control edge so that the oil flows off and the valves are closed very quickly (Figure 5B-25). Thus the machine is prevented from reaching overspeed even in the event of a full load throw-off.
Fig. 5B-13 Double-flow condensing turbine type V50Z for driving a feed-water pump. (Source: ABB Turbinen GmbH, N0rnberg, Germany)
Extraction diagram Line AB Line AD
Limit of max. inlet steam flow Full extraction (cooling and leakage steam only for P2 network) Line CD No extraction (no steam for PE) Line CB Max. backpressure flow Intersection C Max. output at zero extraction Intersection B Peak output at 40% extraction and 60% backpressure steam The percentage refers to the inlet steam flow
Flg. 5B-14
Extractlon dlagram. (Source: ABB Turblnen GmbH, N0rnberg, Germany)
Appendix 5B: Selecting and Sizing a Steam Turbine 193
Flg. 5B-15 Backpressure turblne wlth extractlon - In bulldlng blocks. (Source: ABB Turblnen GmbH, N0rnberg, Germany)
Flg. 5B-16 Flow dlagram o an extractlon-back-pressureturblne. (Source: ABB Turblnen GmbH, N0rnberg, Germany)
Flg. 5B-17 Arrangement of extractlon valves. (Source: ABB Turblnen GmbH, N0rnberg, Germany)
194 Process Plant Machinery
Fig. 5B-18 Double-extraction backpressure turbine for 35,100 kW, live steam conditions 117 bar/517 ~ backpressure 6 bar, with a first extraction at 46 bar and a second extraction at 21 bar. (Source: ABB Turbinen GmbH, N0rnberg, Germany)
Appendix 5B: Selecting and Sizing a Steam Turbine 195 A Pz
of reducing statio n
~from
~
turbine---~,
' Operating
-I,I ..........V,3......V.2 ........ V-a/Y-e--!..
P2
2's
0
50
7S
Backpressure steam
~oOO/o flow
Bleed pressure diagram PAl, PA2, PA3-Bleed pressure at casing PA-Bleed network pressure required P]-Live steam pressure P2 -Backpressure
Flg. 5B-19 Germany)
Bleed pressure dlagram of a pressure-dependent extractlon system. (Source: ABB Turblnen GmbH, N0rnberg,
Live steampl/! 1 r_...l(~
7
'*
~3 Extraction PA-Backpressure
Fig. 5B-20 Germany)
27 1
2I 3 4 56 7
Turbinegovernor Speed Set-value adjuster Pressurecontroller Non-return valve Valveactuator Emerg.stop valve
networkP2
Pressure-dependent extraction system with pressure control station. (Source: ABB Turbinen GmbH, N0rnberg,
196 Process Plant Machinery
Flg. 5B-21 Conaecutlve compound control valve block wlth valve actuator. (Source: ABB Turblnen GmbH, N0rnberg, Germany)
Flg. 5B-22 Hydraullc valve actuator wlth ol relay. (Source: ABB Turblnen GmbH, N0rnberg, Germany)
Flg. 5B-23 Backpresaure turboaet wlth control valve block and a fourth bypass valve. (Source: ABB Turblnen GmbH N0rnberg, Germany)
Flg. 5B-24 Component parts of the consecutlve compound control valve block. (Source: ABB Turblnen GmbH, N0rnberg, Germany)
Appendix 5B: Selecting and Sizing a Steam Turbine 197
,'~ecor,~d v e l o c i t y s t ~ q e ct~sen~af:leci
I
r
~ dP]
...............,_=.p................ s e c o n d votoci'ly st~cJ~
em~a.qecl
..,-9 iif a
~ max. VSive ol3~}~ag v ~ ~ c i t y db'e to l~.~t?ited OJ~i}tltet
b
'= max: v # v e c f o s m g velo c i t y w~ihout ,~et.~ond VC't,,. OClty $ta~e d u e to} t h e c r o s s ~ectiO~ Of the o i l drattl
c
- - m a x . valve c m s l n g velo c i t y with SeCOmt v e l o c d y stage
~]
..... s p e e d
I
.
!
P3 "~ o~,Jt;el n tes,~'Ure o f S p ~ e d gOvettlr H ~ r
v~iw,;, l~avr r
-. lurt)~tle output
Flg. 5B-25
Veloclty diagram of the valve actuator. (Source: ABB Turblnen GmbH, NQrnberg, Germany)
Fig. 5B-26 Heating turbine with two control valve blocks and the corresponding emergency trip valves located ahead. (Source: ABB Turbinen GmbH, NQrnberg, Germany)
This Page Intentionally Left Blank
Chapter 6 Turboexpanders A radial expansion turbine or "expander" is generally used to recover power from steam or other gases or to provide refrigeration for petrochemical plants, hydrocarbon separation plants, and similar processes. The shaft power and process cooling is provided by the nearly isentropic expansion of the gas from the inlet pressure to the outlet pressure. With the increasing cost of power and fuel and scarcity of petrochemical feedstocks, it is increasingly important that expanders be designed for maximum efficiency and reliability. Most process-type expanders built in recent years have been of the radial type, due largely to the mechanical simplicity of the variable guide vane mechanism, the improved ability to expand high-energy streams efficiently in a single stage, and the general compactness of the design. The following discussion will deal with radial expanders, although many of the comments apply equally to axial designs. Radial expanders are usually applied when one or more of the following is a consideration: refrigeration, power recovery, and power generation. REFRIGERATION
Expanders can provide refrigeration by direct expansion of process gas, thus eliminating the need for closed-cycle refrigeration systems. Hundreds of such expanders are in operation in the cryogenic processing of hydrocarbon gases and air separation plants. Field experience has shown that, if required, substantial liquid condensation can occur in the expander without damage of any kind. POWER RECOVERY
Radial expanders can provide power recovery from pressure reduction in liquid or gas streams, such as purge gas, waste gas, fuel gas, or natural gas pressure letdown. The expander can usually be controlled in a way that does not restrict overall plant operation. POWER GENERATION
Radial expanders can form the heart of a closed- or open-cycle power generation system. Power cycles, such as the Brayton or Rankine cycle, using a working fluid * Source: Marl-Trench Corporation, Santa Maria, CA. Adapted by permission. 199
200
ProcessPlant Machinery matched to the requirements of the energy source and the expander can provide excellent full-and part-load efficiency. Several Rankine cycle geothermal power plants are presently in operation.
EXPANDER DESIGN AND CONSTRUCTION
In the typical process-type expanders, the shaft power is absorbed and used by a single-stage centrifugal compressor. The compressor is mounted on the same shaft as the expander, providing a simple and compact design. An expander of this type is shown in Figure 6-1. The expander flow enters through the flange at the upper left, flows through the inlet guide vanes (nozzles) and expander wheel, and exists through the flange at the lower left. The compressor flow enters the flange at the right, is compressed, and exits at the top center. Bearings and seals are located between the two wheels on a short rigid shaft. This arrangement is typical of expanders used for natural gas processing, petrochemical processes, and expansion of air and nitrogen in air separation and nitrogen liquefication plants. A complete dual expander system including two expanders and all required auxiliary equipment and local control panel for a petrochemical process application is shown in Figure 6-2. Radial flow expanders for process applications must be capable of reliable and efficient operation over a relatively wide range of flow rates. To accomplish this, these machines incorporate variable inlet guide vanes or "nozzles," as shown in Figure 6-3. This design uses an externally mounted pneumatic actuator to control the opening of the guide vanes and therefore the expander flow and resultant power output. An internally mounted fulcrum mechanism translates the linear motion of the pneumatic actuator into rotation of a ring that pivots each guide vane on a hardened pin. The reliable operation of this mechanism is
FIGURE 6-1
Typicalexpander/compressor assembly. (Source: Marl-Trench Corp., Santa Maria, CA.)
Turboexpanders
FIGURE 6-2
201
Expander system for a petrochemical process. (Source: Marl-Trench Corp., Santa Maria, CA.)
FIGURE 6-3
Expander inlet guide vanes. (Source: Marl-Trench Corp., Santa Maria, CA.)
critical to the control of the expander and the process, and therefore the design details and materials of construction must be carefully selected to avoid galling and excessive wear during operation. Expander and compressor wheels are usually constructed of high-strength aluminum alloy. The low-density and relatively high-strength aluminum alloy is
202
Process Plant Machinery
ideally suited to these wheels, as they operate at moderate temperature on relatively clean gas and the low-density alloy permits minimizing the weight of the wheels, which is desirable to avoid critical speed problems. A typical expander wheel is shown in Figure 6-4. This wheel was produced by machining from a solid aluminum alloy forging. This construction has been shown to be superior to wheels produced by welding, brazing, or casting. The rotor assembly (expander wheel, shaft, and compressor wheel) for a typical cryogenic process expander is shown in Figure 6-5. The expander wheel is on the fight and the compressor wheel on the left. The compact and rigid design of the rotor is apparent in this figure. Labyrinth-type seals are used between the expander and compressor wheels and the oil-lubricated bearings. These seals prevent mixing of the process gas and lube oil by injecting filtered buffer gas (seal gas) in the middle of the labyrinth and allowing the gas to flow both toward the process and the beating. The stationary and rotating elements of a stepped-type labyrinth seal for a process expander are shown in Figure 6-6.
FIGURE 6-4 CA.)
Typical process expander wheel. (Source: Marl-Trench Corp., Santa Maria,
FIGURE 6-5 Rotary assembly for process expander/compressor. (Source: Marl-Trench Corp., Santa Maria, CA.)
Turboexpanders
FIGURE 6-6
Expander labyrinth seal assembly. (Source: Marl-Trench Corp., Santa Maria,
CA.)
FIGURE 6-7 Maria, CA.)
Combination journal and thrust bearing. (Source: Marl-Trench Corp., Santa
Process expanders generally use a combination journal and thrust beating similar to that shown in Figure 6-7. An essentially identical bearing assembly is located near the expander wheel and compressor wheel, providing thrust capacity in both directions. Since both the thrust and journal beatings are hydrodynamic, there is no bearing wear during normal operation. TYPICAL APPLICATION
A simplified schematic diagram of a cryogenic turboexpander plant is shown in Figure 6-8. Hundreds of such plants are in operation throughout the world, extracting the heavier hydrocarbons from natural gas. The process is based on providing the required refrigeration by direct expansion of the process gas in a single-stage radial expander similar to that shown in Figure 6-1. In this process, the gas is first dehydrated to prevent the formation of ice or hydrates within the cryogenic portion of the plant. Next, the gas is cooled by heat
203
204
ProcessPlant Machinery
I
HIGHI~SSUFE
I RESIDIJECAS ~S,S~
FIGURE 6-8
Cryogenic turboexpander plant. (Source: Marl-Trench Corp., Santa Maria,
CA.) exchange with cold residue gas. This cooling usually results in some condensation. The condensed liquids are removed by a high-pressure separator before reaching the expander. The gas is then expanded through the expander, producing a relatively large temperature drop and substantial liquid formation. Residue gas from the cryogenic process is used to cool the expander inlet gas and is then compressed by the expander driven "boost" compressor. Additional residue gas compression is then required to increase the pressure to a level near the inlet to the plant. OPERATION
Since essentially all of the turboexpander plant flow passes through both the expander and the boost compressor, it is possible to efficiently vary the plant flow rate using the expander inlet guide vanes. The expander shaft speed is allowed to vary freely in response to changing plant flow rate. This method of control has been shown to produce excellent off-design performance. During operation at reduced flow, the boost compressor will tend to operate near surge. It is therefore essential that the compressor be provided with an adequate surge control system. If the compressor is allowed to operate in surge for any significant length of time, damage will occur to the bearings and seals due to the resulting shaft vibrations. (See the chapter on centrifugal compressors for a more detailed description of surge.) If, during operation, solid particles are carded into the expander, the centrifugal forces created by the expander wheel will tend to cause them to be centrifuged out and strike the underside of the inlet guide vanes. This can cause erosion damage to both the inlet guide vanes and the expander wheel blade tips. To minimize this type of damage, it is necessary to provide a 60- to 80-mesh screen upstream of the expander. The screen should be capable of withstanding a 100- to 200-psi differential pressure, and the differential pressure should be monitored continuously. Noncontacting shaft vibration monitoring equipment has been shown to be very useful in monitoring the mechanical condition of expanders. It is important to keep accurate logs of vibration. These data can be particularly useful in evaluating the expander condition after a major process upset or expander trip.
Turboexpanders 205 As with all high-speed rotating equipment, the cleanliness and quality of the lube oil and seal gas should be carefully maintained. Routine sampling and spectrographic analysis of the lube oil to detect the buildup of water, particulate, or trace contaminants is recommended. Any condition that causes an expander trip shuts down the expander by rapidly (in less than one-half second) closing the expander inlet trip valve. If the trip is due to an interruption in lube oil supply, the oil flow during coastdown must be provided by an accumulator. The accumulator must be adequately sized, and the precharge pressure must be properly maintained. It is possible that an expander will slowly rotate or "windmill" due to leaky process valves on the expander or compressor during shutdown. If the lube system is off during this time, damage to the bearings can occur. This type of damage is more common than might be expected because the rotating speed during windmilling may be so slow that it does not indicate on the electronic tachometers on process expanders. MAINTENANCE
Expander operating reliability has been shown to be strongly dependent on the quality of routine maintenance. As a minimum, the following items should be included in any maintenance program. 9 9 9 9 9 9
Maintain proper expander data logs to be used for trend analysis. Provide adequate surge protection for the compressor. Regularly sample and analyze the lube oil for contamination. Provide a 60-to 80-mesh expander inlet screen and differential pressure monitor. Install a compressor discharge check valve to prevent back flow on shutdown. Maintain process valves for tight sealing to prevent expander "windmilling" during shutdown. 9 Maintain proper filtration of seal gas. 9 Assure that the expander inlet trip valve closes in less than one-half second on trip. 9 Verify all expander alarm and shutdown functions at least yearly.
This Page Intentionally Left Blank
Chapter 7 Centrifugal Pumps There are two main types of centrifugal pumps in process plant service: volute pumps and turbine pumps. The former is more common. Some important pump types, applications and ratings are listed in Table 7.1. Although pump designs have not progressed as noticeably as gas turbine development, innovative manufactures are constantly finding ways to "do a dirty or difficult job for less cost". This is particularly true in applications where abrasive, corrosive, variable content (in terms of liquid versus solid percentage), and variable flows are concerned. As industry becomes more diverse, it becomes more practical to consider models that have fewer moving parts, and avoid more expensive (often larger capacity) conventional designs that have some "universal" (common to all applications) features. In "universal" applications, main components such as casing and impeller dimensions might be identical. Other components, such as special seals to handle hazardous fluids, wear plates and rings to handle erosive particles are then added to adapt the pump for the application. The reader is therefore encouraged to use Table 7.1 as the basis for exploring different manufacturers and new proven designs when plant expansion or retrofit needs so dictate. This is an increasingly competitive market, however. Hence, the operator might save on initial capital as well as operational and maintenance costs and yet retain operational flexibility for variable flows by selecting smaller, less expensive units that are tailored for very specific applications. In the case of turbine pumps, one welcome development has been the development of variable vane pumps. Large power savings of over 20% are realized, greater flow variations with an increase in upper volume transmitted are possible, giving the turbine pump access to further fields of application (see Appendix 7D). Also, the rotor dynamic problems associated with many turbine pumps that stem from configuration-related vibratory behaviour, have been more thoroughly investigated. This has helped eliminate some of the maintenance problems that formerly burdened operators and prompted selection of other types of pumps. One study into the "reed frequency" (bending mode associated) is described in Appendix 7C. As was mentioned, the innovative variable vane design means fewer moving parts and lowered required spares inventory. Its wide operating range allows it to be used in a variety of process industries (see Table 7.1). Environmental considerations also play a major role in equipment selection nowadays. Due to space considerations, this topic will be dealt with in Environmental Engineering and Management: Sustainable Development in the Power Generation, Oil & Gas and Process Industries (Butterworth-Heinemann, 1998)
207
208
ProcessPlant Machinery
TABLE 7.1
Pump Type
Typical Services
Typical Ratings
ANSI Process
Corrosive/abrasive liquids, slurries, and solids, high temperature, general purpose pumping process and transfer.
Q to 4500 GPM (1022 m3/h) H to 730 ft (222 m) T to 700~ (371 ~ H to 375 PSIG (2586 kPa)
Nonmetallic Chemical Process
Severe corrosives.
Q to 800 GPM (182 m3/h) H to 490 ft (149 m) T to 300~ (150~ P to 225 PSIG (1550 kPa)
Self-priming Process
Corrosive/abrasive liquids, slurries, and suspensions, high temperature, industrial sump, mine dewatering, tank car unloading, bilge water removal, filter systems, chemical transfer.
Q to 1500 GPM (340 m3/h) H to 375 ft (114 m) T to 500 ~F (260 ~ Suction Lifts to 25 ft (7.6 m)
Process, transfer and general service. Corrosive and volatile liquids. High temperature services.
Q to 1500 GPM (340 m3/h) H to 700 ft (207 m) T to 500 ~ (260 ~ P to 375 PSIG (2586 kPa)
Zero leakage services: toxic liquids, refrigerants, liquefied gas, high temperature heat transfer, explosive liquids, liquids sensitive to atmosphere, sphere, carcinogenic and other hazardous services.
Q to 2500 GPM (568 m3/h) H to 1400 ft (427 m) T to 700~ (371 ~ P to 450 PSIG (3103 kPa)
i
In-line Process
rill Canned motor
Centrifugal Pumps 209 TABLE 7.1
Typical Services
Pump Type
Typical Ratings
High temperature and high pressure services, offsite, transfer, heat transfer liquids.
Q to 7500 GPM (1700 m3/h) H to 1100 ft (335 m) T to 800 ~ (427 ~ P to 870 PSIG (6000 kPa)
Petrochemical, chemical, refining, offsite, gasoline plants, natural gas processing, general services.
Q to 7500 GPM (1700 m3/h) H to 750 ft (229 m) T to 650 ~ (343 ~ P to 595 PSIG (4100 kPa)
Paper stock, solids and fibrous/stringy materials, slurries, corrosive/abrasive process liquids.
Q to 28,000 GPM (6360 m3/h) H to 350 ft (107 m) T to 450~ (232 ~ P to 285 PSIG (1965 kPa)
Medium consistency (8 to 14%) paper stock.
Q to 1800 TPD (1650 MTPD) H to 400 ft (125 m) T to 250~ (120 ~
"Roto Jet" R (registered trade mark)
Meets API 610. Pulp and paper, mining, oil, steel mills, petrochemical, boiler feed and desuperheating, central cleaning systems, hydraulic systems, reverse osmosis, spraying systems, transfer, water injection
Q to 400 GPM T to 550F P to 2250 psi
Horizontal (Abrasive Slurry)
Corrosive/abrasive services. Coal, fly ash, mill scale, bottom ash, slag, sand/gravel, mine slurries. Large solids.
Q to 10000 GPM (2273 m3/h) H to 350 ft (107 m/Stage) T to 400~ (204 ~ P to 300 PSIG (2068 kPa) Sperical solids to 4" (102 mm)
Axial Flow
Continuous circulation of corrosive/ abrasive solutions, slurries and process wastes. Evaporator and crystallizer, reactor circulation, sewage sludge recirculation.
Q to 200,000 GPM (35,000 m3/h) H to 30 ft (9 m) T to 350~ (180~ P to 150 PSIG (1034 kPa) Solids to 9" (228 mm)
API Process (Horizontal)
-/l ll Ii API Process (In-line)
'J,IIUIIIII
Paper Stock/High Capacity Process
III
210
ProcessPlant Machinery
TABLE 7.1
(continued)
Pump Type Large Solids Handling (Horizontal)
Typical Services
Typical Ratings
Pumps for extra demanding municipal and industrial services; large pulpy and fibrous solids, sewage, abrasives.
Q to 100,000 GPM (22,700 m3/h) H to 240 ft (73 m) T to 202 ~ (43 ~ P to 300 PSIG (2065 kPa) Solids to 10" (254 mm)
Double suction
Cooling tower, raw water supply, booster service, primary and secondary cleaner, fan pump, cooling water, high lift, low lift, bilge and ballast, fire pumps, river water brine, sea water, pipelines, crude.
Q to 72,000 GPM (16,300 m3/h) H to 570 ft (174 m) T to 350 ~ (177 ~ P to 275 PSIG (1896 kPa)
Multi-stage
Refinery, pipeline, boiler feed, descaling, crude oil charging, mine pumping, water works ... other high pressure services.
Q to 3740 GPM (850 m3/h) H to 6000 ft (1824 m) T to 375 ~ (190~ P to 2400 PSIG (16,546 kPa)
(Vertical Dry Pit)
.It
Water, cogeneration, reverse osmosis, booster service, boiler feed, shower service. Boiler feed, mine dewatering and other services requiring moderately high heads.
Centrifugal Pumps 211 TABLE 7.1
Pump Type
Typical Services
Low Flow/High Head Multi-Stage Moderate speed
,,I, tn . llll Submersible 9 Wastewater 9 Solids Handling 9 Slurry
i
i
Typical Ratings
Reverse osmosis descaling, high pressure cleaning, process water transfer, hydraulic systems, spraying systems, pressure boosters for hi-rise buildings, all low flow applications where efficiency is critical.
Q to 280 GPM (64 m3/h) H to 2600 ft (792 m) T to 400 ~ (204 ~ P to 1100 PSIG (7584 kPa)
Flood and pollution control, liquid transfer, sewage and waste removal, mine dewatering, sump draining. Large stringy or pulpy solids. Abrasive slurries.
Q to 4000 GPM (910 m3/h) H to 210 ft (65 m) T to 140 ~ (60 ~ Solids to 2" (50 mm)
Industrial process, sump drainage, corrosives, pollution control, molten salts, sewage lift, wastewater treatment, extremely corrosive abrasive slurries, large or fibrous solids.
Q to 7500 GPM (1703 m3/h) H to 310 ft (95 m) T to 450 ~ (232 ~ Solids to 10" (254 mm)
flllllll 9
on
Vertical Submerged (Submerged Beating and Cantilever) 9 Process 9 Solids Handling 9 Slurry
_
/ m
I
m
I
212
ProcessPlant Machinery
TABLE 7.1
(continued)
Typical Services
Pump Type Vertical Turbine
a.
I
m
Typical Ratings
Irrigation, fire pumps, service water, deep well, municipal water supply, mine dewatering, cooling water, seawater and river water intake, process, utility circulating, condenser circulating, ash sluice, booster, petroleum/refiner, boiler feed, condensate, cryogenics, bilge, fuel oil transfer, tanker and barge unloading.
Q to 150,000 GPM (34,065 mm3/h) H to 3500 ft (1070 m) T to 700~ (371 ~
As for vertical turbines
Q to 570,000 GPM H to 80 ft Discharge bore to 158"
Close Coupled and frame-mounted pumps for water circulation, booster, OEM packages, irrigation, chemical process, transfer, and general purpose pumping.
Q to 2100 GPM (477 m3/h) H to 400 ft (122 m) T to 300 ~ (149 ~
I1 | ! il r
-
Vertical mixed flow variable pitch vane General Service (Frame-mounted)
(Close-coupled)
CONVENTIONAL PROCESS PUMPS* Figure 7-1 depicts a typical American National Standards Institute (ANSI) process pump, which is typical of conventional process pumps. ANSI standards for pumps are dimensional standards that facilitate pump and component interchangeability. As illustrated in Figure 7-2, this standard dimension process pump is furnished with a fully open impeller, generally preferred for solids handling and for stringy or abrasive-containing pumpage. Two different sealing arrangements are shown: soft packing above the pump centerline, and a mechanical face seal below the pump centerline. Figure 7-3 illustrates an enhanced ANSI pump with an elastomer bellows seal shown above the pump centerline and a conventional multispring seal shown below the centerline. There are literally hundreds of mechanical-seal types available to serve the numerous different pumps and pumped fluids. Some of these will be described later in this text. * Source: Goulds Pumps, Inc., Seneca Falls, NY.
Centrifugal Pumps
FIGURE 7-1
213
Standard dimension process pump (ANSI). (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
Figure 7-4 depicts a between-bearing pump with a double-flow impeller. Chosen for high-flow capability and balanced axial thrust, double-flow impellers are widely used in large or heavy-duty process pumps. Where maximum accessibility to pump parts and flanges is needed, the user may opt for vertical mounting, as illustrated in Figure 7-5. Multistage horizontal split case pumps (Figure 7-6) are used for a wide range of moderate- to high-pressure services in process plants. Their typical internal construction features are seen in Figure 7-7. The pump suction nozzle is on the left. Pumpage leaving the third impeller is routed to the suction of the fourth impeller, located at the opposite end of the pump. This internal flow arrangement results in axial rotor thrust balance by hydraulic means. The external piping ensures pressure equalization between the space to the right of the balancing drum near the entry to stage 4, and to the left of the suction eye of stage 1. A multistage centrifugal pump with barrel-type outer casing is depicted in Figure 7-8. These pumps are primarily used for high-pressure and extreme-pressure light hydrocarbon liquids, although certain boiler feedwater services often use this casing style also. Conventional low flow-high head centrifugal pumps are typically configured as shown in Figure 7-9. The multistaging is achieved by adding modular elements that are designed for maximum interchangeability and minimum spare parts i equhemeuts.
214
Process Plant Machinery
FIGURE 7-2 Pump cross section showing typical seal areas. (Source: Goulds Pumps, Inc., Seneca Falls, NY.) The particular model illustrated here achieves sealing of the casing by the use of Otings and long external tie bolts. An alternative execution, which uses a containment casing instead of the tie bolts, is shown in Figure 7-10. High-speed pumps for low-flow high-head services are substantially different from conventional low-flow high-head centrifugals and merit special coverage. These pumps are described later in this chapter. The construction features of submersible wastewater pumps are shown in Figure 7-11. These pumps obviously have to be capable of occasional solids ingestion, which makes it necessary to design and build the impeller with suitable features. Close-coupled and frame-mounted pumps are primarily designed for general purpose pumping. Quite similar to the close-coupled vertical pump, close-coupled and frame-mounted pumps have the impeller placed on the electric motor shaft. Large solids-handling pumps are manufactured in a variety of configurations. Figure 7-12 shows a horizontally arranged model. Accessibility for service and general ease of maintenance are important for this pump category. In self-priming process pumps, priming and air separation are accomplished within the pump casing. The pump is designed with two volutes; these are separate stationary channels into which the rotating impeller pushes the pumpage exiting from the impeller tip. During the priming cycle, the lower volute functions as the intake while the upper volute discharges liquid and entrained air into a separation chamber. Air is separated and expelled through the pump discharge while liquid circulates into the lower volute. Once air is completely exhausted from the suction region and liquid fills the impeller eye, the pump is primed and functions as a conventional
Centrifugal Pumps
FIGURE 7-3 Enhanced ANSI pump showing oversized seal housing and two different mechanical seals. (Source: Goulds Pumps, Inc., Seneca Falls, NY.) pump, with both volutes acting as discharges. As shown in Figure 7-13, the casing is designed so that an adequate volume of liquid for repriming is always retained in the pump, even if liquid is allowed to drain back to the source of supply from both discharge and suction. The function of a dual volute design is shown in Figure 7-14. The dual volute casing design is ideal where pumps must periodically operate at capacities above or below design capacity or at uninterrupted high head. Essentially, this design equalizes radial forces and lessens radial reaction on shaft and beatings. This equalization or balancing of radial forces is accomplished by dividing the liquid discharged by the impeller into two half-capacity volutes with two cutwaters, set 180 ~ apart. Radial forces on the shaft and bearings are equally opposed. In-line process pumps are vertically oriented pumps with the casing designed to bolt directly into the piping system. They require a miniinuin of support froIIl
215
216
Process Plant Machinery
FIGURE 7-4 Between-bearing pump with double-suction impeller (single stage). (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
FIGURE 7-5 Falls, NY.)
Vertically mounted double-suction pump. (Source: Goulds Pumps, Inc., Seneca
a relatively small foundation or similar structure and have proven to be as reliable and easy to maintain as conventional, horizontally oriented centrifugal pumps. Figure 7-15 shows an in-line pump with flexibly coupled electric motor shaft to pump shaft connection. The pump has its own bearing support whereas the so-called close-coupled in-line pump shown in Figure 7-16 uses a rigid coupling sleeve and
Centrifugal Pumps
217
FIGURE 7-6 Multistage horizontally split case process pump. (Source: Goulds Pumps, Inc., Seneca Falls, NY. )
FIGURE 7-7 Internal component arrangement of a five-stage horizontally split case pump. (Source: Goulds Pumps, Inc., Seneca Falls, NY.) has its rotor supported by the electric motor bearings only. Although the flexibly coupled and close-coupled styles are generally equally reliable, the flexibly coupled version should be preferred from an ease-of-maintenance point of view. A third variation of the in-line pump construction has the pump impeller placed on the motor shaft end. This style is found less often in process plants. American Petroleum Institute (API) process pumps get their name from an API standard (API-610) that specifies the requirements for this heavy-duty pump. While statistics show that properly applied ANSI pumps have a useful life and reliability matching that of API pumps, the latter nevertheless has some construction features that lnakt: it the p~upc~ choit:c i. tactuain high-risk applit, atiuna.
218
Process Plant Machinery
FIGURE 7-8 Multistage
barrel casing-type centrifugal pump.
(Source: Sulzer Brothers,
Winterthur,
Switzerland.)
FIGURE 7-9 Conventional low-flow, high-head centrifugal multistage pump made up of modular elements. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
FIGURE 7-10 Low-flow, high-head three-stage centrifugal pump with containment casing. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
Centrifugal Pumps 219
FIGURE
Submersible
7-11
wastewater pump.
(Source:
Goulds Pumps,
Inc.,
Seneca
Falls, NY. )
API pumps differ from ANSI in the following respects" 9 9 9 9
API pumps have greater corrosion allowance. They have higher permissible nozzle loads. API pumps have more available stuffing box space. Wear tings are furnished in API pumps. They are not always supplied with ANSI pumps. 9 API pumps are centerline-mounted; ANSI pumps are often foot-mounted. 9 Bearing housings in API pumps are generally fitted with higher load capacity bearings and higher life expectancy end seals.
Figure 7-17 illustrates a typical API process pump. Paper stock and high-capacity process pumps are typically configured as shown in Figures 7-18 and 7-19. Both pumps incorporate wear plates opposite the open side of the impeller. Maintainability and simplicity of construction are key requirements in these services. Figure 7-18 incorporates a repeller arrangement to oppose the ___~.~i
. . . . .
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ul
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_I
alull~; thC
_ t . _ ~ t . . ,x,t.: . . . ~zm, ~ . ~ . allallE,~nl~,llt
"
. ~.~,,alaLa u ~r t W O
"
Stat, lGi~a~'-j _1 . . . .
220
Process Plant Machinery
FIGURE 7-12 Large solids handling pump with vortex impeller. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
FIGURE 7-13
Self-priming process pump. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
Centrifugal Pumps
FIGURE 7-14 Principle of dual volute design. The dual volute casing design is ideal where pumps must periodically operate at capacities above or below design capacity or at uninterrupted high head. Essentially, this design equalizes radial forces and lessens radial reaction on shaft and bearings. This equalization or balancing of radical forces is accomplished by dividing the liquid discharged by the impeller into two half-capacity volutes with two cut-waters set 180~ apart. Radial forces on the shaft and bearings are equally opposed. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
FIGURE 7-15 Vertical in-line pump with flexibly coupled electric motor. (Source: Goulds Pumps, Inc., Seneca Falls. NY.)
221
222
ProcessPlant Machinery
FIGURE 7-16 Rigid spacer-coupled vertical in-line pump. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
and a rotating part. Figure 7-19 depicts a model that incorporates special inducers to accommodate difficult pumpage. In a "Roto-Jet ''R pump, liquid enters the manifold and passes into the rotating case where centrifugal force causes the liquid to accelerate and enter the rotor under pressure. The velocity energy of the liquid in the rotor is converted into additional pressure as it jets into the pick-up tube. The liquid flows through the pick-up tube and is discharged. This design results in a wide operating range with low minimum flow. The differential head can be adjusted by changing speed. The capacity can be adjusted by changing the pick-up tube size. Only two working-parts are the basis for this true one stage design: a rotating casing driven by an oversize shaft, and a stationary pick-up tube mounted in the manifold. (See also Appendix 7E) A typical slurry pump is shown in Figure 7-20. Resistance to corrosion and wear is of great importance in slurry pumps, and simple construction aids in making
Centrifugal Pumps
223
FIGURE 7-17 Single-stage back pull-out-type centrifugal process pump complying with API SPEC 610. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
FIGURE 7-18 Heavy-duty process pump for paper stock incorporating repeller arrangement to reduce load on stuffing box area. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
224
Process Plant Machinery
FIGURE 7-19
Paper stock pump incorporating special inducers. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
FIGURE 7-20
Typical slurry pump. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
the pumps maintainable. In some pumps, replaceable rubber lining is used on the wetted parts. Much of the confusion in deciding when to specify a slurry pump arises from the lack of agreement on the meaning of the word slurry. This is due in large part to the nearly infinite number of solid-liquid mixes. In place of the many academic
Centrifugal Pumps 225 slurry definitions, this broader, more functional definition is offered: A slurry is any mixture of liquid and solids capable of causing significant pump abrasion, clogging, or mechanical failure due to high loads or impact shocks. Under some circumstances, it may seem superfluous to consider the "when" of slurry pump selection. Obviously, a pump employed to move "deliberate" slurries such as mine tailings or chemical concentrates must be designed and constructed with exceptional strength, abrasion resistance, and solids-passage ability. But what about a pump employed to supply large quantities of water from a sandy fiver for cooling purposes? In such "accidental" slurries, the transport of liquids is the prime p u r p o s e - the presence of solid materials is not intended (or, sometimes, even recognized). Nevertheless, failure to use a slurry pump for this type of application can frequently result in excessive maintenance, parts usage, and downtime costs. The "when" of slurry pump selection might best be answered by a rule of thumb that says that whenever the fluid to be pumped contains more solids than are found in potable water, at least consider the use of a slurry pump. There are many features that set a slurry pump apart from a standard, general service centrifugal pump. Outwardly, there are few differences, although the slurry pump is usually larger in size. Internally, however, there are many characteristics that make a slurry pump a very specialized breed. Wall thickness of wetted-end parts (casing, impeller, etc.) are greater than those used in conventional centrifugal pumps. The cutwater, or volute tongue (the point on the casing at which the discharge nozzle diverges from the casing), in the casing is generally less pronounced in order to minimize the effects of abrasion. Flow passages through both the casing and impeller are large enough to permit solids to pass without clogging the pump. Slurry pumps are available in a variety of materials of construction to best handle the abrasive, corrosive, and impact requirements of nearly any solids-handling application. Because the gap between the impeller face and suction liner will increase as wear occurs, the rotating assembly of the slurry pump must be capable of axial adjustment to maintain the manufacturer's recommended clearance. This is critical if design heads, capacities, and efficiencies are to be maintained. Other specialized features include extra-large stuffing boxes, replaceable shaft sleeves, and impeller back vanes that act to keep solids away from the pump stuffing box. Both radial and axial-thrust bearings on the slurry pump are generally heavier than for standard centrifugals, owing to the demands imposed by slurries of high specific gravity. Although impeller back vanes (used to lower stuffing box pressures) do actually reduce axial thrust, these vanes can wear considerably in abrasive services. Consequently, the bearings must be of ample capacity to handle thrust loads by themselves. Balancing holes through the impeller should not be used to reduce axial thrust, since they can either clog or initiate excessive localized impeller wear. Nearly all slurry pumps have larger diameter impellers than units for pumping clear liquids, enabling heads and capacities to be met at reduced rotational speed. Low-speed operation is one of the most important wear-reducing features of a slurry pump. In fact, experience shows that abrasive wear on any given pump rises at least with the third power of RPM increase. An analysis of the static profile of the slurry pump will help determine the solids-passage ability, abrasion resistance, and mechanical strength required of the pump. The most important elements in the static profile can be assigned to four categories: 1. Size of the solids: What are the largest particles the pump must handle? Are these solids similar or random in size?
226
Process Plant Machinery
2. Nature of the solids: Are they pulpy or hard, light or dense, round or jagged? Are they abrasive or corrosive? 3. Nature of the liquid: How corrosive is the liquid? Will it lubricate the solids and reduce abrasion? 4. Concentration of the solids: It is the ratio of solids to liquids that determines how the characterisitcs of the solids will influence the slurry as a whole. These four static characteristics create unique demands, requiting specific pump design and construction features. For example, Figure 7-21 shows a pump model designed to handle wastes, light slurries, and random large solids. Unlike the slurry pump discussed earlier, this unit does not use wear liners. The emphasis here is on very large flow passages through the casing and impeller. Because such units are generally used for pumping sewage, light slurries, and relatively nonabrasive industrial wastes, certain wear-reducing design features can be compromised to increase hydraulic efficiency. When chemical sludges or wastes containing large solids must be pumped, a vortex pump is often the best answer. Because its impeller is fully recessed into the rear of the casing, a relatively small pump can be used to handle liquids containing very large solids. A vortex pump was shown earlier in Figure 7-12. Still other slurries may exchange the problems of large solids for the equally difficult pumping idiosyncrasies associated with high concentrations of small solids.
FIGURE 7-21 Pump designed to handle industrial and municipal wastewater. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
Centrifugal Pumps
227
More often than not, such slurries present extreme abrasion problems. Typical are those associated with lime slurry pumping, the handling of ore concentrates, kaolin clay, or cement slurries. Figure 7-22 shows an extremely heavy-duty slurry pump ideal for such applications. Figure 7-23 represents only one of numerous styles of vertical sump andprocess pumps. Also see Appendix 7D. This particular model has the discharge piping attached to the bowl assembly. Also, this model is shown with externally connected tubing for the lubrication of lineshaft bearings. Depending on the nature of the service, a process plant may be best served by vertical industrial turbine pumps similar to those shown in Figure 7-24. The vertical pump shown in Figure 7-24 has either a fabricated or cast discharge head and either a threaded or flanged column. It is designed for clean, noncorrosive liquids, at low to medium pressures. This model is often selected when lowest initial cost is of prime consideration. Its principal applications are irrigation, fire water services, service water and deep well pumping, drainage, and municipal water supply. The Use of a flanged column on pumps of the type shown in Figure 7-24 facilitates maintenance, and a good selection of additional lineshaft bearing materials is often available for these pumps. They are primarily used in low- to mediumpressure effluent, oily wastewater, and mine dewatering applications.
FIGURE 7-22 Veryheavy-duty pump for slurries with large concentrations of highly abrasive particles. (Source: Goi~Tds PUmps,rnc~, SeneCa FallS, NE.). . . . . . . . . . . . . . . . . . . . . . . . . . .
228
Process Plant Machinery
FIGURE 7-23 vertical sump and process pump. (Source: Goulds Pumps, Inc., Seneca Falls, NY.) FIGURE 7-24 Typical vertical turbine pump with principal application in water services. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
Vertical industrial turbine pumps that incorporate both a fabricated discharge head and flanged column are designed for high-pressure applications, when ease of maintenance is a prime consideration, or when alloy materials of construction are required for corrosive and/or erosive services. These pumps would also be suitable for a wide range of pumping temperatures, if used in industrial processes. However,
Centrifugal Pumps
they are primarily used in cooling water, sea and fiver water intake, utility circulating water, condensate, and ash sluice water services. A vertical can-type pump is depicted in Figures 7-25 and 7-26. Using a fabricated discharge head and barrel and a flanged column, the pump is designed for low net positive suction head (NPSH) available and subatmospheric suction pressure services. Typical applications are pipeline boosters, product unloading, refinery blending, injection/secondary recovery, ammonia transfer, condensate, cryogenic, and liquid natural gas (LNG) transfer duties.
FIGURE 7-25 Verticalcan-type turbine pump, shown with fabricated head principally used in hydrocarbon processing services. (Source: Goulds Pumps, Inc., Seneca Falls, NY.) FIGURE 7-26 Cross section of vertical can-type pump. (Source: Goulds Pumps, Inc., seneca FalIs, NE)
229
230
Process Plant Machinery
FIGURE 7-27 Verticalmarine pump with fabricated discharge head, flanged column, and rightangle gear drive. (Source: Goulds Pumps, Inc., Seneca Falls, NY.) FIGURE 7-28 Vertical industrial submersible pump for deep-well applications. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
Centrifugal Pumps
Figure 7-27 shows a vertical marine pump that uses a fabricated discharge head and flanged column. Pumps of this type are often designed to be self-priming and to efficiently unload or strip product tankers and tank barges. They have also been applied as ship firewater pumps, ballast pumps, bilge pumps, and fuel oil transfer pumps. In our next illustration, Figure 7-28, we see a vertical industrial submersible pump. This version is used for deep settings or where the use of a lineshaft pump is impractical, e.g., in irrigation, service water, and deep well supply situations. Finally, there is also a close cousin to the vertical p u m p - the vertical energy recovery turbine (Figure 7-29). The recovery turbine takes high-pressure liquid and converts it into rotating energy that can be used to power other pumps, or other rotating equipment. Virtually all of the vertical industrial pumps discussed for process plant applications include a variety of features and/or options that are of interest. Take, for instance, the bowl assembly, which is the heart of the vertical turbine pump. The impeller and diffuser-type casings are designed to deliver the head and capacity that a system requires for optimum efficiency. The fact that the vertical turbine pump can be multistaged allows maximum flexibility both in the initial pump selection and in the event that future system modifications require a change in the pump rating. Submerged impellers allow a pump to be started without priming. A variety of material options allows the selection of a pump best suited for even the most severe services. The many bowl assembly options available assure that the vertical turbine pump satisfies the users' needs for safe, efficient, reliable, and maintenance-free operation. Figure 7-30 depicts the more important ones. There exists also a large number of column options (Figure 7-31), discharge heads (Figure 7-32), coupling arrangements (Figure 7-33), and sealing flexibility options (Figure 7-34). Circulator, or axial flow pumps are shown in Figures 7-35 and 7-36. Although not strictly centrifugal flow pumps, these high-flow, low-head machines are worthy of mention.
ELECTRIC MOTOR UPPER ADJUSTABLE COUPLING
UPPER HEAD
LOWER ~ DISCHAR( HEAD
.•
TURBINE DISCHARGE FLANGE ~ " ~
ENERGY
RECOVERY
TURBINE
IRBINE SUCTION PUMP SUCTION
PUMP DISCHARGE
FIGURE 7-29 Vertical energy recovery t u r b i n e - a vertical pump in reverse rotation. (-Source: Goulu~ Pumps, Inc., Seneca r~h~, ~ )
231
232
ProcessPlant Machinery
CHOICE OF SEMI-OPEN OR ENCLOSEO
IMPELLERS
STRAINERS
Available In alloy conelructlon for a wide range of cormelve/abrasive eervtceL
Basket or cone idmlnen; a m aVllllabte to provide proteclton Item large solids.
LOW NPSH FIRST STAGE IMPELLERS For low NP8H A lppltcaUofle. Both Ilrge eye and mixed flow first stlges available; minimizes pump length.
KEYED IMPELLERS
Keyed impellers are standard on 18" and laqler slim;; furnished all pumps for Ilmperlturee a l i v e 1800F (02" C) end on cryogenic een~cee. Regardless of e~ze,keyed Impellerl provtde eate of m l l n l e n l n c t and pollllve locking under fluctuating load and temperalure conditions.
DOUBLE SUCTION FIRST STAGE
DUAL WEAR RINGS Available for enclosed impellers and bowls; permits re-eslablishtng initial running clearances and efficiency el lower cosl. Hard facing of wear earlaces ovldiable for longer life. Wear rings can be flushed when solids are present in pumpege.
FIGURE 7-30
RIFLE DRILLING/ OISCHARGE BOWL Rifle ddlling of bowl shills available for beadng prolecUon on abrasive eervicel. Oischl~ile bowl included with enclosed lineshaft construction.
The dual volute casing and doubk; suction knpeller can be inetlN~ alone or as 9first etlige with turbine sages added. In ailher event, double suction relulta in reduced NPIIH R. Another benefit is reduced axial downthrust. Goulds Vertical Double 84action (Model VDI) le Ideally euliml for steel mills, mine devtltadng and river wetmr, or. as a first etage in holwldl condensate Imcl pipeline booster applications wttere NPSH ie critical.
Assembly options for vertical pump bowls. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
Centrifugal Pumps
233
Threaded Column
Threaded column is used whenever initial cost is primary consideration. Pipe ends are machine-faced for butt fit between sections to maintain alignment, Threaded column is used where pump length requires numerous column sections such as a deep well application,
OPEN LINESHAFT ASSEMBLY
Product lubricated open llnelhafl conMrucflon using robber bwdngl. Shaft oectJon8 Joined by sleeve.type ooupllnge.
ENCLOSED LtNESHAFT ASSEMBLY 011 lubricaMd linelimit for bearing tu41~ dcaUon of long ~et pumps or water llulhlng ol llnelhllll bearings for short or
long sel pumps in abrasive service.
Flanged Column
Column sections are provided with flanged ends incorporating registered fits for ease of alignment during assembly. Facilitates disassembly where corrosion is a problem. On 12" and smaller sizes, the bearing retainer registers the fit between Ihe column section flanges. On 14" and larger sizes the bearing retainer is welded into the column section.
OPEN LINESHAFT Ranged column/procluct lubricated lineehafl is recommended for ease of maintenance or whenever II Ipeclall bearIng material is required. Keyed lineshalt coupling available in all sizes for ease of rnatnlen|nce. Various beldn9 materials available. Renewable shaft sleeve or hard lacing of lhllll llWllla134e for longer file.
ENCLOSED LINESHAFT
The l ~ f l
is pro-
IWcted by water flushing
m e encloelng tube bearings on c~oaive/ abrasive services. 011 lubricated Ilneshafl i v l i l . able on long selllngt. Alignmen! is attained by register fll between the
flange faces.
MAImIG R E T A I I ~
(4" m l r C o k . ~ ) FIGURE 7-31
Vertical pump column assembly options. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
234
Process Plant Machinery
The discharge head functions to change the direction of flow from vertical to horizontal and to couple the pump to the system piping in addition to supporting and aligning the driver. Discharge head accommodates all modes of drivers including hollow shaft and solid shaft motors, right angle gears, vertical steam turbines, etc. Optional sub-base can be supplied. Goulds offers three basic types for maximum flexibility.
FABRICATED DISCHARGE HEAD
BELOW GROUND DISCHARGE HEAD
Used wheneverVIT pump Is required to adapt to an underground discharge system.
FIGURE 7-32
CAST IRON DISCHARGE HEAD Used for low pressures (not exceeding 175 psi) and/or when low initial cost is primary consideration.
For pressures exceeding cast head limitations or ~rvlces 1hat require alloy construction such as high or low larnperalure or corrosive eervicel~ 8egmenled elbow available Ior efficiency Improvemeat. Large hlit~l holes for easy acceel;. Base flange can be mechined to maich ANSI tank flange.
Discharge head arrangements for vertical column pumps. (Source: Goulds Pumps, Inc., Seneca
Falls, NY.)
RIGID FLANGED
COUPLING (Type AR)
ADJUSTABLE
To couple pump 1o vertical hollow shaft driver. Impeller adjustment is performed on adjusting nut located on top of motor.
For vertical solid shall driver. Impeller adjustmenl made by using adjustable plate in the coupling.
FIGURE 7-33
NY.)
COUPLING
(Type A)
A D J U S T A B L E SPACER C O U P L I N G ( T y p e AS) Same function as type A coupling with addition of spacer. Spacer may be removed for mechanical seal maintenance without distutbing driver.
Coupling arrangements for vertical column pumps. (Source: Goulds Pumps, Inc., Seneca Falls,
Centrifugal Pumps
I
235
! !
I
I! !
I ! i !
i i! |
J
i
PACKED BOX Whenever packing lubrication leakage can be tolerated and the discharge pressure does not exceed 300 psi, a packed box may be used. Optional headshaft sleeve available to protect shaft.
INSIDE MOUNTED
FIGURE 7-34
Water flush tube connection is supplied when pressurized water is introduced into the enclosing tube for bearing protection on abrasive services.
OIL L U B R I C A T E D Oil lubricated option is recommended when water elevation would cause the upper Iineshaft beadngs to run without lubrication during start-up. Oil is fed thru tapped opening and allowed to gravitate down enclosln 9 tube lubricating bearings.
SEALS
Most popular method - - used for medium to high pressures. Cartridge style for ease of Installation and maintenance.
NY.)
WATER FLUSH
OUTSIDE MOUNTED SEALS Provides a method of no-leak sealing fo~ low pressure applications.
T A N D E M SEALS Two seals mounted in-line. Chamber between seals can be filled with a buffer liquid and may be fitted with a pressure sensitive annunciating device for safety.
Sealing flexibility options for vertical column pumps. (Source: Goulds Pumps, Inc., Seneca Falls,
236
Process Plant Machinery
FIGURE 7-35 Axial flow or elbow-type (circulator) pump used for high-volume, low-head applications in process plants. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
FIGURE 7-36
Principal components of axial flow pumps. (Source: Goulds Pumps, Inc., Seneca Falls, NY.
Centrifugal Pumps
CANNED MOTOR AND SEALLESS MAGNET-DRIVE CENTRIFUGAL PUMPS* Canned motor and sealless drive pumps were developed to contain hazardous, valuable, toxic, or carcinogenic pumpage. These designs avoid the use of mechanical shaft seals and confine the pumpage within a hermetically sealed space. The canned motor design (Figure 7-37) comprises a single shaft that combines the functions of a motor rotor and pump rotor in a single assembly. The motor rotor is surrounded by a stainless steel sleeve that is permeated by the magnetic flux lines generated by the surrounding stator windings. A large number of variations of the standard canned motor design are available to the user. Figures 7-38 and 7-39 give an overview that includes canned motor pump models suitable for hot fluids, abrasive liquids, and pumpage close to the vaporization temperature. A typical sealless magnet drive pump is illustrated in Figure 7-40. While unique when compared with conventional-design centrifugal pumps, magnet-drive pumps represent a combination of standard components and proven concepts. Figure 7-40 depicts a typical sealless magnet-drive pump with a separately mounted electric motor drive. In this installation, the base, the electric motor, and the motor coupling are identical to parts used in conventional pumps. The differences between these pumps and conventional pumps are concentrated in two areas: 1. Driving torque is transmitted magnetically rather than mechanically. 2. The impeller drive shaft tides in bushings housed within the pump enclosure rather than on bearings mounted externally. In Figure 7-40, the drive motor is coupled directly to the outer magnet ring by a conventional motor-to-pump coupling. The overhung load of the outer magnet ring is carried by the bearings in the bearing housing. Figure 7-40 also shows that the impeller is mounted on the same shaft as the inner magnet rotor, sometimes
FIGURE 7-37 Canned motor pump cross section. Note inducer for effective lowering of NPSH requirement. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
* Sources: The Kontro Company, Inc., Orange, MA (magnet-drive), and Goulds Pumps, Inc., Seneca Fatts;, NY- (canned motor pumps): Adapted by ly~I/lissio/l:. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
237
238
Process Plant Machinery
Suitable for handling volatile liquids; ammonia, freon and other liquified gases.
for liquids with high melting point
Same as R-Type but uses an adapter between pump and motor. Allows for greater pump/motor flexibility.
Similar to K-S Type but better suited for fluids with relatively low melting point.
FIGURE 7-38 Canned motor pump variations for special fluid conditions. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
Used for sump and unloading services.
Suitable for handling high temperature fluids; heat transfer.
FIGURE 7-39
Handles fluids with large amount Of-fine solids. S-Type with external flushing also available.
Suitable for handling ~iquids with small amount of fine solids.
Canned motor pumps for unusualfluid conditions. (Source: Goulds Pumps, Inc., Seneca Falls, NY.)
Centrifugal Pumps I
-I
239
IMPELLER /
CONTAINMENT SHELL TORQUE RING (INNER M A G N E T RING)
-=~
*~l~r====~==~==='~
MOTOR
~
COUPLING
CI
MOTOR
I
~
I
I
I
I
I
I .=,,~
BASE
FIGURE 7-40 Sealless magnet drive pump with a separately mounted electric motor. In this installation, the base, the electric motor, and the motor coupling are identical to parts used in conventional pumps. The difference between sealless pumps and conventional pumps occurs in two areas: (1) driving torque is transmitted magnetically rather than mechanically; and (2) the impeller drive shaft rides in bushings housed within the pump enclosure rather than by bearings mounted externally. (Source: Kontro Pump, Orange, MA.) called the torque ring. The impeller drive shaft is carried by one or two bushings that are within the pumping enclosure. Note that the pump enclosure is formed by the pump casing and the containment shell. The driving torque of the electric motor is transmitted to the pump impeller by the magnetic coupling of the outer magnet ring and the torque ring without breaching the pumping enclosure. It is this magnetic coupling that replaces the mechanical seals of conventional centrifugal pumps. The efficiency obtainable with canned motor and magnet drive pumps is below that of well-designed, conventional centrifugal pumps. Also, canned motor and magnetdrive pumps may not be available in size ranges much over 100 kw. Nevertheless, they represent viable options that must be evaluated on a case-by-case basis.
HIGH-SPEED CENTRIFUGAL PUMPS* Development of the High-Speed Concept The term high speed is generally used to classify equipment that operates above two pole motor speeds. Centrifugal pump designs falling into this category have gained considerable acceptance since 1960. The increasing popularity of high-speed pumps has coincided with the expanding need for higher pressures in the process and general industries since World War II. At the same time, improved technology, manufacturing techniques, and materials have facilitated the transition from design theory to production hardware. The developed head in centrifugal pumps is a function of the tip speed of the impeller and/or the number of impellers employed. There are three principal methods that can be used to achieve higher pressures, and in some cases a combination of these methods may be utilized: 9 Increasing the size of the impeller to increase its peripheral speed. This is a simple and effective method, but only to a point. The practical design limit for impeller diameters is 13 to 16 inches at two pole speeds. * ~....... .~,n.a.C.tr~.n.a. . . . . .~h,ia H~nalin,~o,.a. rx,~a~, . . . . . CO.-Adat~by p,~rrnir162 . .
.
.
.
.
.
.
240
Process Plant Machinery
Using a number of staged impellers. Although continued development of multistage pumps has resulted in hydraulic efficiencies approaching single-stage efficiencies, the complexity of the design and close impeller clearances results in high first cost, loss of performance as the clearances wear, and often high maintenance requirements if the pump is subjected to difficult service conditions. Increasing rotational speed. The practical speed limitations for pumps directly driven by induction motors is 3600 RPM with 60 cycle power and 3000 RPM with 50 cycle power. The rotational speed of electric-driven pumps can, however, be increased by using either higher frequency power or a speed-increasing gearbox. Since with few exceptions, higher frequency power has not yet become an economically viable approach, the use of gear-driven single-stage, high-speed centrifugal pumps has become widespread. This segment of our text deals primarily with a unique pump design that has been adapted for operation at higher speeds to produce high heads at low to moderate flow rates. Although high-speed pumps are now widely used in industry, comparatively few design discussions have appeared in print. An early commercial application of this design concept emerged in 1959, when a high-speed single-stage centrifugal pump was used in aircraft service to augment the thrust in jet engines during takeoff. This pump rotated at 11,000 RPM, delivering 80 gallons per minute of water to the combusters at 400 psi to increase the mass flow rate through the engine, thereby increasing the thrust by 15 percent during takeoff. The unit weighed only 8 89pounds, including step-up gearing from the 6500 RPM power takeoff pad to pump speed. Some 250 units were produced for this service. By 1962, the first integrally geared high-speed process pumps were finding their way into the petrochemical and refining industries. In subsequent decades, their use has been greatly expanded, and they are now available from 1 to 2500 HP, utilizing speeds to 32,000 RPM, and producing heads to 12,000 feet. Most commonly, these products consist of a single stage but may employ two or three stages to satisfy the need for extreme heads or the combination of high head and low net positive suction head available (NPSHA). By the early 1970s, the high-speed pump concept had been extended to include a wide variety of services across a broad spectrum of industries.
Unique Design Advantages The increasing popularity of integrally geared high-speed centrifugal pumps is due to a number of factors: 9 Shaft dynamics. As shaft speeds are increased, the size of the components required for any given condition of service grows smaller. The smaller, more compact design results in shorter shaft spans, lower shaft deflection, and improved shaft dynamics. 9 Reduced size and weight. A high-speed pump with a single six-inch impeller can exhibit the same performance as a multistage pump that uses as many as 40 stages with the same size impellers. The size and weight reduction can be as much as 5 to 1, which of course translates to smaller and less expensive mounting foundations. 9 Fewer parts required. High-speed pumps generally have one stage and occasionally two or three stages. This can mean a significant reduction in the number of "wetted" components. In processes requiting exotic metallurgies, this can provide
Centrifugal Pumps 241
FIGURE 7-41 Removablehigh-speed shaft assembly for low-flow high-head pump. (Source: Sundstrand Fluid Handling, Arvada, CO.) a substantial capital cost advantage as well as less supporting inventory and lower repair parts costs. 9 High-speed gear-driven pumps can be designed to incorporate many common parts. Seals, bearings, and housings are typically common across a given product line. Only the pump case, impeller, and/or gear ratio need to be changed to provide a wide range of performance. In a plant requiting many of the same types of pumps, the number of spare parts can be reduced. 9 Reduced maintenance. Another benefit realized by the user is simplified maintenance resulting from the reduced size and quantity of parts. Additionally, some high-speed pump designers have allowed for the complete removal and installation of all fluid end components, seals, and bearings in a single modular high-speed shaft assembly (see Figure 7-41). 9 Performance consistency, The head, flow, and efficiency of many high-speed pumps is resistant to efficiency and head degradation, since open radial bladed impellers do not require close clearances. In applications where consistency of performance over the life of the pump is desired, a high-speed pump with large running clearances can be very desirable.
Hydraulic Capabilities High-speed pumps are manufactured in both single and multistage configurations. Radial vaned open impellers are optimum for low specific speed applications (see Figure 7-42) from Ns = 150 to Ns = 850. This hardware is capable of achieving 6000 feet in a single stage. Higher flow units typically use Francis vaned impellers with wear rings when the specific speed ranges from 850 to 1860. As impeller speed is lncrease~d-to-meet a given se-e-t-o--f~pe-gafifigcoriditicifig; ttie Uppecific-Uppee--6dandpump
242
Process Plant Machinery
9~i
1
1000 GPM 500 GPM
70.
~960. 5o.
g,o.
~
.,.,...,~10 o G PM
~ ~25
GPM
/
;it/ 250
200 GPM
"~/
" 500
1000
1500
SPECIFIC SPEEDN
S
I ! !
Rs
=
2000
RPM H.75
I i
I
RADIAL
FRANCIS
FIGURE 7-42 Effect of specific speed on high-speed pump efficiency. (Source: Sundstrand Fluid Handling, Arvada, CO.) efficiency increase, while torque decreases for the same horsepower requirement. Accordingly, shaft stress, gear, and radial beating loads improve with higher speeds. The user should not be unduly concerned with increased wear or stress due to higher rotational speeds. The maximum stress level in an impeller is a function of the impeller tip speed. As shown in Figure 7-43, the stresses at 200 feet per second (FPS) are the same regardless of whether the speed is produced by a 3600 RPM twelve-inch impeller or by a 23,000 RPM two-inch impeller, and shaft deflection and bearing loading are minimized using unusually low overhang ratios and small impeller weights. As impeller speed increases, the NPSH required for stable operation also increases. Often an axial flow inducer with good suction performance is used in series with the impeller to lower NPSH requirements. It is attached directly to the shaft in place of the impeller nut. On two- or three-stage machines, the first stage can often be geared at a lower speed for lower net positive suction head required (NPSHR), while the subsequent stages do most of the work.
Mounting Arrangements The unique design of the high-speed pump lends itself to a variety of mounting configurations. Since the pump first appeared in the industrial market as a
Centrifugal Pumps
500
12 Inch
8 Inch
6 Inch
243
4 Inch
400 "
3002 Inch 200"
100-
0
3,8oo
lo:~o
20500 23,000
Pump Speed (RPM)
3o,ooo
FIGURE 7-43 Relationship of pump rotative speed and impeller diameter to tip speed. (Source: Sundstrand Fluid Handling, Arvada, CO.) single-stage vertical in-line type, the full range of possibilities has been explored. Today these pumps exist in the following forms to serve a wide range of general industrial and process markets: vertical in-line; horizontal single-stage, two-stage, and three-stage, with both single-step and two-step speed-increasing gearboxes. High-speed pumps are often available in either close-coupled or frame-mounted configurations. The close-coupled design eliminates the need for coupling alignment and occupies the least amount of floor space, while the frame-mounted units are used whenever conventional driver packages are selected.
Applications Process applications for high-speed centrifugal pumps exist wherever there is a need for medium or higher pressures. Their widespread use is based on adaptability to many diverse requirements. High-speed pumps are an essential part of processes utilized in the production of such end items as plastics, pharmaceuticals, petrochemicals, synthetic rubber, and paper. The technology incorporated in these pumps makes them especially suited to lower flow, high-head applications, displacing reciprocating two-stage and multistage pumps as the preferred product. Users apply high-speed pumps to process applications for numerous reasons, but the primary deciding factor is economics. Economic evaluations typically include first cost, installed cost, operating cost, maintenance cost, and overall evaluated cost. Each determining factor must be based on the user's specific situation. The primary reasons that high-speed pumps are often selected over reciprocating, single-stage, two-stage, or multistage centrifugals are lower first cost, lower installed cost, and occasionally lower maintenance costs. Operating costs will generally approach those of other centrifugal pumps but are almost always higher than positive displacement pumps with their inherently better mechanical efficiencies. The performance area where the high-speed pump has the greatest advantage is in the low-flow range. Pump mechanical requirements vary, depending on the critical nature and location of the particular service. In severe or hazardous applications, API-610 ~qui.~.,ents ir.ay b~ . ~ ~ y , "d l l U -' t'-'-'n ~ . - ~.p. ~. .u -1 p u t t J p ~ a t e a v a H a"'-'-'vxc n l "-'-eategory. ttn~
.
.
.
.
.
.
.
.
.
.
.
.
.
244
ProcessPlant Machinery
For less critical services, however, general service pumps should also be considered. Although many of the same design features are available in both types, the significantly lower cost associated with non-API designs encourages their use in less critical services. High-speed pumps can be used as the primary pump, as an installed spare for an existing pump, as a boost pump piped in series with another pump, or as a support pump for a seal flush or lube oil. They are utilized in both continuous and standby operation. It is not unusual to find high-speed pumps feeding a variety of systems where flow demands are constantly changing. These pumps are especially suited to highpressure washdown and shower services where multiple sprayers are turned on and off as the system demands change. The controls required to operate these systems are simple and reliable, allowing operation of the pump over most of its performance envelope down to flows as low as 15 percent of the best efficiency point flow. Process plant applications are as diverse as the industry served. Beginning with power systems, high-speed centrifugal pumps are applied in boiler feedwater, condensate return, desuperheater or attemporator, gas turbine NOX supression, and reverse osmosis applications. Process systems use high-speed pumps in a variety of services including but not limited to transfer, seal flush, waste injection or disposal, blending, sampling, recycling, descaling, metering, waste disposal, reactor feed, booster, pipeline, charge, reflux, circulation, bottoms, flare drum knockout, and high-pressure washdown. Some typical fluids pumped include water, caustics, ammonia, carbamate, fuel oil, naphtha, acids, a majority of hydrocarbons, and chemicals too numerous to mention. As evidenced by the wide variety of applications, high-speed pumps are a proven product with years of reliable operating experience.
System Controls The control of high-speed centrifugal pumps is similar to most conventional centrifugal pumps. When specifying the control system, it is important to consider the allowable operating range of the pump and its hydraulic characteristics, as well as the hydraulic requirements of the process. There are generally two objectives that need to be kept in mind when designing a control system. One is to protect the pump from damage that can be incurred from operating outside its design operating range. A second is to provide the controls that will enable the pump to meet the needs of the process. Centrifugal pumps tend to operate over a wide flow range with relatively slight variation in pressure in comparison with positive displacement pumps. The maximum and minimum operating limits for centrifugal pumps with flat performance curves are normally based on flow rather than pressure. Thus the protective controls should be designed to measure and control flow rate rather than discharge pressure (Figure 7-44). Maximum Flow Limit Volute-type centrifugal pumps have the lowest bearing radial loads at the design flow rate or best efficiency point. As the flow through the pump is increased or decreased from the best efficiency point, the radial hydraulic loads increase. Also, as flow velocities increase, the potential for impeller cavitation increases. The power also increases with flow. Operation at excessive flows can lead to bearing failures, high shaft stresses and possible failures, and cavitation damage to impellers and casings.
Centrifugal Pumps
A
i
!
l
Pressure ,
j
I
I 1
I
i
i
Flow
i
2
Power
•
NPSH Required
I
....
3
1 Minimum Flow 2 Design Flow 3 Maximum Flow FIGURE 7-44
Typical centrifugal pump performance. (Source: Sundstrand Fluid Handling,
Arvada, CO.)
Attention to the maximum flow limit of the pump and knowledge of the process hydraulic characteristics when the pump is specified can result in a process that is self limiting and without need of special controls to prevent excessive flow through the pump. Figure 7-45 shows a pump curve that has been matched to the process needs at maximum capacity. The initial startup should be carefully planned to allow pipes to empty and vessels to be filled gradually, preventing water hammer or overloading the pump.
Minimum Flow Limit As flow increases from the design point, beating radial loads generally increase and efficiency decreases. If the flow decreases enough, recirculation can occur and the
..__...=_
Pump Performance / / Process Performance (Design Capacity)
~ ~ ~ ' ~ ./ /.
~
Process Performance (Maxlmum Capacity)
Flow
FIGURE 7-45 Pump curve matched to process requirements at maximum capacity. (Source: S~ndstrand Fluid HandEng / A-FvizEa;-CO-.T ........................................................................................
245
246
ProcessPlant Machinery
pump becomes hydraulically unstable. Extensive damage can be done if a pump is allowed to operate for long periods in an unstable condition. As efficiency decreases at low flow, the rate of temperature rise of the fluid increases. This can be a concern in applications with low available NPSH. The specific minimum flow limit depends partially on the pump and partially on the process (assuming adequate beating capacity). Advertised performance curves generally show minimum flow that is expected with ideal fluid properties and proper inlet and discharge piping. Minimum flow controls should always be checked by observing the pump in operation with the minimum flow control functioning. Near the minimum flow point, most centrifugal pumps have nearly constant pressure with respect to flow. To prevent operation below the minimum flow limit, the first choice for the measured variable is flow. A control system that prevents operation above a particular maximum discharge pressure does not necessarily ensure minimum flow protection. When centrifugal pumps are operated in parallel, individual minimum flow control is necessary. A check valve should be installed in the discharge line of each pump. This is to prevent one pump from driving the other pump off its performance curve if both pumps are operating and the process is modulated to a low-capacity condition (see Figure 7-46). Modulating a bypass line is the normal method for preventing minimum flow. If the bypass line discharges immediately into the pump suction, the fluid temperature will rise because of the power being dissipated in the pump bypass loop. If prolonged operation with this arrangement is expected, then a maximum temperature trip should be considered (Figure 7-47). Suction Pressure Limit
Occasionally a process scheme may can cause the NPSH available to fall pressure to rise above the maximum system, appropriate limiting controls
allow suction pressure to vary. If variations below that required by the pump, or the inlet rated pressure for either the pump or the seal will be required. Bypass
Flow Control
"
~ By~ass
Suction
,I,
Process Capacity Control ~Discharge
FIGURE 7-46 Parallel pumps with individual flow-modulated bypasses for minimum flow protection. (Source: Sundstrand Fluid Handling, Arvada, CO.)
Centrifugal Pumps
High Temperature Shutdown
247
FIowC o n t r o l ~
Suction
Process Capaci ty Control
Discharge
FIGURE 7-47 Arvada, CO.)
Parallel pumps with individually controlled recycle lines. (Source: Sundstrand Fluid Handling,
Maintenance Considerations Routine maintenance on high-speed centrifugal pumps consists of two activities" periodic inspection and periodic service.
Periodic Inspection The items checked and the frequency of checking will vary with the specific design of the pump and its duty. The common ingredient in high-speed pumps is the integral speed-increasing gear box. Lube oil level, lube pressure, and temperature are normal inspection items on all of these pumps, requiring monitoring at least once per week. The need for periodic inspections will vary with the type of auxiliary equipment installed on the pump. Where there are additional auxiliary systems supporting the pump, there are normally automatic alarm and shutdown devices that help to simplify inspection tasks. Proper functioning of protective alarm and shutdown devices should be verified periodically. If such tests are to be made while the equipment is running, it is best to specify this provision when pump and auxiliary equipment are initially purchased. It is best if the periodic inspections include record keeping. Records can show trends that can help in the planning of service work, keeping maintenance costs low. Periodic Service
Lubrication requirements are normally identified in the manufacturer's instruction manual. The driver, coupling, and gearbox each have their own lubrication needs. Each machine is designed to be run using lubricants with certain specific properties, so the manufacturer's recommendation should be considered when lubricants are chosen. Dibasic ester and poly-alpha olefin synthetic lubes are often advantageous, ................. and know-I-ed-geab-le-gear ~afiu~r-e-rg~ql~specl-ify t ~ e mb-d-em-t~l~ ..........................
248
ProcessPlant Machinery Many high-speed gear pumps incorporate a modular high-speed rotor assembly that is easily removed for inspection and maintenance. Contaminants in the pumped fluid or in the bearing lubricant can penetrate through the film that separates the moving and stationary parts, and also may cause
wear.
Seal life is primarily determined by the seal environment. Most pumps have features that allow the user to control the seal environment maximum seal life. If seal life of less than six months occurs, then system modifications can often extend the life. If life greater than one or two years is observed, then system modifications to further improve the life will not likely be cost effective. Overall, seals and bearings are considered wearing items needing periodic maintenance. The frequency of this maintenance can vary significantly with the type of duty the pump serves. When bearings are replaced on high-speed pumps, the manufacturer's recommendations should be followed. In addition to the load capacity of high-speed ball bearings, internal clearances, contact angle, tolerance class, and retainer design are all important factors. Modem pump manufacturers specify bearings that operate well within the manufacturer's ratings, but careless substitutions can have disastrous results.
Machinery Condition Monitoring. High-speed pumps in critical service are often monitored for continuous determination of machine condition. Parameters most often monitored are vibration, lube pressure and temperature, and bearing temperature. Such monitoring is not normally considered mandatory by pump manufacturers for general pump service. However, contemporary manufacturers generally can provide optional provisions for monitoring these items when needed for pumps in critical service. Vibration Monitoring. The three types of vibration monitoring most commonly used are noncontacting proximity probes, seismic casing vibration sensors, and acceleration-spike energy transducers. Noncontacting proximity probes measure shaft displacement (peak-to-peak). These probes are normally installed inside the speed-increasing gearbox to measure displacement of the output shaft relative to the gearbox housing. They are normally installed by the manufacturer and are ordered with the pump at purchase. Casing vibration sensors normally measure housing velocity amplitude. Such instruments can be either permanently installed on the gearbox or can be obtained as portable units that are periodically taken from one machine to another. Acceleration-spike energy monitoring is often done with portable data terminals and is one of the most effective ways to obtain early warning of incipient defects in rolling element beatings. With either permanently mounted or portable types of instruments, it is best to take readings periodically and to monitor trends. Permanently installed instruments are normally connected to automatic alarm and shutdown controls. Lube Pressure and Temperature. When pressurized lubrication systems are used on speed-increasing gearboxes, lube oil pressure and temperature are often monitored. If pressures fall or temperatures rise, the equipment can be shut down automatically to prevent or minimize damage. Bearing Temperature. On pumps with journal bearings, temperature sensors can be imbedded in the bearings. Temperature sensors are usually either thermocouples or resistance temperature detectors (RTDs). Bearing temperature monitoring can provide early warning of loss of lubricating properties, reduction in lubricant flow, bearing failure, or loss of lubricant cooling.
APPENDIX 7A
Centrifugal Pump Fundamentals Head
STATIC DISCHARGE HEAD is the vertical distance in feet between the pump centerline and the point of free discharge or the surface of the liquid in the discharge tank.
The pressure at any point in a liquid can be thought of as being caused by a vertical column of the liquid which, due to its weight, exerts a pressure equal to the pressure at the point in question. The height of this column is called the "static head" and is expressed in terms of feet of liquid. The static head corresponding to any specific pressure is dependent upon the weight of the liquid according to the following formula: Head in Feet =
Pressure in pal x 2.31 Specific Gravity
A Centrifugal pump imparts velocity to a liquid. This velocity energy is then transformed largely into pressure energy as the liquid leaves the pump. Therefore, the head developed is approximately equal to the velocity energy at the periphery of the impeller. This relationship is expressed by the following well known formula: V2 H-..m 2g
Where
H = Total head developed in feet. v = Velocity at periphery of impeller in feet per sec. g = 32.2 Feet/Sec. 2
We can predict the approximate head of any centrifugal pump by calculating the peripheral velocity of the impeller and substituting into the above formula. A handy formula for peripheral velocity is: v =
RPMxD 229
Where D = Impeller diameter in inches.
The above demonstrates why we must always think in terms of feet of liquid rather than pressure when working with centrifugal pumps. A given pump with a given impeller diameter and speed will raise a liquid to a certain height regardless of the weight of the liquid, as shown in Figure 7A-1. All of the forms of energy involved in a liquid flow system can be expressed in terms of feet of liquid. The total of these various heads determines the total system head or the work which a pump must perform in the system. The various forms of head are defined as follows. SUCTION LIFT exists when the source of supply is below the center line of the pump. Thus the STATIC SUCTION LIFT is the vertical distance in feet from the center line of the pump to the free level of the liquid to be pumped. SUCTION HEAD exists when the source of supply is above the centerline of the pump. Thus the STATIC SUCTION HEAD is the vertical distance in feet from the centerline of the P_um_P to the free level of the liquid to be pumped.
TOTAL STATIC HEAD is the vertical distance in feet between the free level of the source of supply and the point of free discharge or the free surface of the discharged liquid. The above forms of static head are shown graphically in Figure 7A-2(a) and (b) FRICTION HEAD (hf) is the head required to overcome the resistance to flow in the pipe and fittings. It is dependent upon the size and type of pipe, flow rate, and nature of the liquid. VELOCITY HEAD (hv) is the energy of a liquid as a result of its motion at some velocity V. It is the equivalent head in feet through which the water would have to fall to acquire the same velocity, or in other words, the head necessary to accelerate the water. Velocity head can be calculated from the following formula: hv
=
v2 2-~
where g v
= 32.2 ft/sec. 2 = liquid velocity in feet per second.
The velocity head is usually insignificant and can be ignored in most high head systems. However, it can be a large factor and must be considered in low head systems. PRESSURE HEAD must be considered when a pumping system either begins or terminates in a tank which is under some pressure other than atmospheric. The pressure in such a tank must first be converted to feet of liquid. A vacuum in the suction tank or a positive pressure in the discharge tank must be added to the system head, whereas a positive pressure in the suction tank or vacuum in the discharge tank would be subtracted. The following is a handy formula for converting inches of mercury vacuum into feet of liquid. Vacuum, ft. of llquld =
Vacuum, In. of Hg x 1.13 Sp. Gr.
The above forms of head, namely static, friction, velocity, and pressure, are combined to make up the total system head at any particular flow rate. Following are definitions of these combined or "Dynamic" head terms as they apply to the pump. TOTAL DYNAMIC SUCTION LIFT (hs) is the static suction lift minus the velocity head at the pump suction flange plus the total friction head in the suction line. The total dynamic suction lift, as determined on pump test, is the reading of a gage on the suction flange, converted to feet of liquid and corrected to the pump centerline*, minus the velocity head at the point of gage attachment. TOTAL DYNAMIC SUCTION HEAD (hs) is the static suction head plus the velocity head at the pump suction flange minus the total friction head in the suction line. The total dynamic suction head, as determined on pump test, is the reading of the gage on the suction latiange, converted to feet of liquid and
250 Process Plant Machinery
Fig. 7A-1
Identical pumps handling liquids of different specific gravities
Fig. 7A-2 (a) Suction lift - showing static heads In a pumping system where the pump is located above the suction tank (static suction head)
corrected to the pump centerline*, plus the velocity head at the point of gage attachment. TOTAL DYNAMIC DISCHARGE HEAD (hd) is the static discharge head plus the velocity head at the pump discharge flange plus the total friction head in the discharge line. The total dynamic discharge head, as determined on pump test, is the reading of a gage at the discharge flange, converted to feet of liquid and corrected to the pump centerline*, plus the velocity head at the point of gage attachment.
TOTAL HEAD (H) or TOTAL DYNAMIC HEAD (TDH) is the total dynamic discharge head minus the total dynamic suction head or plus the total dynamic suction lift. TDH = hd + h=(with a suction lift) TDH = hd - h=(wlth a suction head)
*On vertical pumps the correction should be made to the eye of the suction or lowest impeller.
Appendix 7A" Centrifugal Pump Fundamentals 251
r ,,.,,.==..,,=' 'j ----
TOTAL STATIC HEAD
II
STATIC DISCHARGE
r
HEAD
,=,..,.. STATIC SUCTION HEAD
-7 I
Fig. 7A-2
(b) Suction head - showing static heads in a pumping system where the pump is located below the suction tank
(static suction head)
Capacity
The constant 3960 is obtained by dividing the number or foot pounds for one horsepower (33,000) by the weight of one gallon of water (8.33 pounds.)
Capacity (Q) is normally, expressed in gallons per minute (gpm). Since liquids are essentially incompressible, there is a direct relationship between the capacity in a pipe and the velocity of flow. This relationship is as follows: Q=AxVorV= A Where V
Q -A
The brake horsepower or input to a pump is greater than the hydraulic horsepower or output due to the mechanical and hydraulic losses incurred in the pump. Therefore the pump efficiency is the ratio of these two values.
Pump Eft - whp _ Q x TDH x Sp. Gr. bhp 3960 x bhp
= Area of pipe or conduit in square feet. = Velocity of flow in feet per second.
Power and Efficiency The work performed by a pump is a function of the total head and the weight of the liquid pumped in a given time period. The pump capacity in gpm and the liquid specific gravity are normally used in the formulas rather than the actual weight of the liquid pumped. Pump input or brake horsepower (bhp) is the actual horsepower delivered to the pump shaft. Pump output or hydraulic horsepower (whp) is the liquid horsepower delivered by the pump. These two terms are defined by the following formulas.
whp = ......~
Q x TDH x Sp. Gr.
3960
Specific Speed and Pump Type Specific speed (Ns) is a non-dimensional design index used to classify pump impellers as to their type and proportions. It is defined as the speed in revolutions per minute at which a geometrically similar impeller would operate if it were of such a size as to deliver one gallon per minute against one foot head. The understanding of this definition is of design engineering significance only, however, and specific speed should be thought of only as an index used to predict certain pump characteristics. The following formula is used to determine specific speed:
= 3960 Q x• TDHpumpXEfflciencySp" .Gr .....................................
N . - ~H3/4
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252 ProcessPlantMachinery Values of 8peoifir Speed, N,
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Where
Francis-Vane Area
Mixed-Flow Area
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13= < 1 e5 D=
Axial-Flow Area
D= D"'~
Rotation
1
Impeller design vs specific speed
N = Pump speed in RPM Q = Capacity in gpm at the best efficiency point H = Total head per stage at the best efficiency point
The specific speed determines the general shape or class of the impeller as depicted in Figure 7A-3. As the specific speed increases, the ratio of the impeller outlet diameter, D2, to the inlet or eye diameter, D1, decreases. This ratio becomes 1.0 for a true axial flow impeller. Radial flow impellers develop head principally through centrifugal force. Pumps of higher specific speeds develop head partly by centrifugal force and partly by axial force. A higher specific speed indicates a pump design with head generation more by axial forces and less by centrifugal forces. An axial flow or propeller pump with a specific speed of 10,000 or greater generates its head exclusively through axial forces. Radial impellers are generally low flow high head designs where as axial flow impellers are high flow low head designs.
Net Positive Suction Head (NPSH) and Cavitation The Hydraulic Institute defines NPSH as the total suction head in feet absolute, determined at the suction nozzle and corrected to datum, less the vapor pressure of the liquid in feet absolute. Simply stated, it is an analysis of energy conditions on the suction side of a pump to determine if the liquid will vaporize at the lowest pressure point in the pump. The pressure which a liquid exerts on its surroundings is dependent upon its temperature. This pressure, called vapor pressure, is a unique characteristic of every fluid and increases with increasing temperature. When the vapor pressure within the fluid reaches the pressure of the surrounding medium, the fluid begins to vaporize or boil. The temperature at which this vaporization occurs will decrease as the pressure of the surrounding medium decreases. A liquid increases greatly in volume when it vaporizes. One cubic foot of water at room temperature becomes 1700 cu. ft. of vapor at the same temperature. It is obvious from the above that if we are to pump a fluid effectively, we must keep it in liquid form. NPSH is simply a m e a s u r e of the amount of suction head present to prevent this vaporization at the lowest pressure point in the pump.
NPSH Required is a function of the pump design. As the liquid passes from the pump suction to the eye of the impeller, the velocity increases and the pressure decreases. There are also pressure losses due to shock and turbulence as the liquid strikes the impeller. The centrifugal force of the impeller vanes further increases the velocity and decreases the pressure of the liquid. The NPSH Required is the positive head in feet absolute required at the pump suction to overcome these pressure drops in the pump and maintain the liquid above its vapor pressure. The NPSH Required varies with speed and capacity within any particular pump. Pump manufacturer's curves normally provide this information. NPSH Available is a function of the system in which the pump operates. It is the excess pressure of the liquid in feet absolute over its vapor pressure as it arrives at the pump suction. Figure 7A-4 shows four typical suction systems with the NPSH Available formulas applicable to each. It is important to correct for the specific gravity of the liquid and to convert all terms to units of fleet absolute" in using the formulas. In an existing system, the NPSH Available can be determined by a gage reading on the pump suction. The following formula applies: N P S H A = Ps - Vp + G r + h v
Where
Gr = Gage reading at the pump suction expressed in feet (plus if above atmospheric, minus if below atmospheric) corrected to the pump centerline. hv = Velocity head in the suction pipe at the gage connection, expressed in feet.
Cavitation is a term used to describe the phenomenon which occurs in a pump when there is insufficient NPSH Available. The pressure of the liquid is reduced to a value equal to or below its vapor pressure and small vapor bubbles or pockets begin to form. As these vapor bubbles move along the impeller vanes to a higher pressure area, they rapidly collapse. The collapse, or "implosion" is so rapid that it may be heard as a rumbling noise, as if you were pumping gravel. The forces during the collapse are generally high enough to cause minute pockets of fatigue failure on the impeller vane surfaces. This action may be progressive, and under severe conditions can cause serious pitting damage to the impeller. The accompanying noise is the easiest way to recognize cavitation. Besides impeller damage, cavitation normally results in reduced capacity due to the vapor present in the pump. Also, the head may be reduced and unstable and the power consumption may be erratic. Vibration and mechanical damage
Appendix 7A: Centrifugal Pump Fundamentals 253
(4a) S U C T I O N S U P P L Y OPEN T O A T M O S P H E R E - - w i t h Suction Lift ,=
(4b) S U C T I O N S U P P L Y OPEN T O A T M O S P H E R E - - w i t h Suction Head
j PB I
L
NPSHA --- Pe - - (Vp + L= + hf) J.....
m _ I ml.. ~ _ _ , , . ~ " I " -----r--- - ~ ~ " - = I - - .----I ~
-
(4c) CLOSED S U C T I O N S U P P L Y
(4d) CLOSED S U C T I O N SUPPLY - w i t h Suction Head
s
I-=--:::zl=--I--- -----I
I-~__._-~'--c-_'--'T-- -- "-- /
Pe == Barometric pressure, in feet absolute. Vp == Vapor pressure of the liquid at maximum p u m p i n g temperature, in feet absolute. P == Pressure on surface of liquid in closed suction tank, in feet absolute.
Fig. 7A-4
L= == Maximum static suction lift in feet. LH --" Minimum static suction head in feet. ht = Friction loss in feet in suction pipe at required c a p a c i t y
Calculation of system net positive suction head available for typical suction conditions
such as bearing failure can also occur as a result of operating in cavitation. The only way to prevent the undesirable effects of cavitation is to insure that the NPSH Available in the system is greater than the NPSH Required by the pump.
Pump Characteristic Curves The performance of a centrifugal pump can be shown graphically on a characteristic curve. A typical characteristic curve shows the total dynamic head, brake horsepower, efficiency, and net positive suction head all plotted over the capacity range of the pump. Figures 7A-5, 6, & 7 are non-dimensional curves which indicate the general shape of the characteristic curves for the various types of pumps. They show the head, brake homepower, and efficiency plotted as a per cent of their values at the design or best efficiency point of the pump. Figure 7A-5 shows that the head curve for a radial flow pump is relatively flat and that the head decreases gradually as
the flow increases. Note that the brake horsepower increases gradually over the flow range with the maximum normally at the point of maximum flow. Mixed flow centrifugal pumps and axial flow or propeller pumps have considerably different characteristics as shown in Figures 7A-6 and 7. The head curve for a mixed flow pump is steeper than for a radial flow pump. The shut-off head is usually 150% to 200% of the design head. The brake homepower remains fairly constant over the flow range. For a typical axial flow pump, the head and brake horsepower both increase drastically near shut-off as shown in Figure 7A-7. The distinction between the above three classes is not absolute, and there are many pumps with characteristics falling somewhere between the three. For instance, the Francis vane impeller would have a characteristic between the radial and mixed flow classes. Most turbine pumps are also in this same range depending upon their specific speeds. Figure 7A-8 shows a typical pump curve as furnished by a manufacturer. It is a composite curve which tells at a glance what the pump will do at a given speed with various impellerdiametem from maximum t_Qminimum__Constant .... horsepower, efficiency, and NPSHR lines are superimposed
254 Process Plant Machinery
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Appendix 7A: Centrifugal Pump Fundamentals 255
Flg. 7A-8 Composlte performance curve over the various head curves. It is made up from individual test curves at various diameters.
Affinity Laws The affinity laws express the mathematical relationship between the several variables involved in pump performance. They apply to all types of centrifugal and axial flow pumps. They are as follows: 1. With impeller diameter, D, held constant:
Q1 N1 A. Where Q = Capacity, GPM Q2 N2 H = Total Head, Feet H1 ( NN22 1) 2 BHP = Brake Horsepower B.~--~= =
N
BHP,
c e-K~, =
= Pump Speed, RPM
IN,/3
2. With speed, N, held constant:
A. QI
Q2
D1 D2
. . .,N - (o1)' ~ BHP1
C. B__H__~2__ (DID22j/~2 When the performance (Q1, H1, & BHP1) is known at some particular speed (N1) or diameter (D1), the formulas can be used to estimate the performance (Q2, H2, & BHP2) at some other speed (N2) or diameter (D2). The efficiency remains nearly constant for speed changes and for small changes in impeller diameter.
EXAMPLE To illustrate the use of these laws, refer to Figure 7A-8. It shows the performance of a particular pump at 1750 rpm with
various impeller diameters. This performance data has been determined by actual tests by the manufacturer. Now assume that you have a 13" maximum diameter impeller, but you want to belt drive the pump at 2000 rpm. The affinity laws listed under 1 above will be used to determine the new performance, with N1 = 1750 rpm and N2 --2000 rpm. The first step is to read the capacity, head, and horsepower at several points on the 13" dia. curve in Figure 7A-10. For example, one point may be near the best efficiency point where the capacity is 300 gpm, the head is 160 ft, and the bhp is approx. 20 hp.
300 1750 Q2 2000
Q2 = 343 gpm
160 (1750)2 H2 - ~
H2 = 209 ft.
20 (1750) 3 BHP2 - ~
BHP2=30hp
This will then be the best efficiency point on the new 2000 rpm curve. By performing the same calculations for several other points on the 1750 rpm curve, a new curve can be drawn which will approximate the pumps performance at 2000 rpm, Figure 7A-9. Trial and error would be required to solve this problem in reverse. In other words, assume you want to determine the speed required to make a rating of 343 gpm at a head of 209 ft. You would begin by selecting a trial speed and applying the affinity laws to convert the desired rating to the corresponding rating at 1750 rpm. When you arrive at the correct speed, 2000 rpm in this case, the corresponding 1750 rpm rating will fall on the 13" diameter curve.
System Curves For a specified impeller diameter and speed, a centrifugal pump has a fixed and predictable performance curve. The point where the pump operates on its curve is dependent upon the characteristics of the system in which the pump is operating, commonly called the "System Head Curve". By
256 Process Plant Machinery
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plotting the system head curve and pump curve together, we can tell: 1. Where the pump will operate on its curve. 2. What changes will occur if the system head curve or the pump performance curve changes. STATIC SYSTEM HEAD
Consider the system shown in Figure 7A-10. Since the lines are oversized and relatively short, the friction head is small compared to the static head. For this example, the system head will be considered as entirely static, with the friction neglected. Assume the fluid being handled has 1.0 Sp. Gr. NPSHA is 13'. The flow requirement is 100 gpm. Since the system head is made up entirely of elevation and pressure differences, it does not vary with flow. The normal system head is 250' TDH (19' elevation difference plus 231' pressure difference). Since the discharge vessel
pressure may vary +3 psi, the system head will vary between 243' and 257'. Consider the application of a pump sized for 100 gpm at 250' TDH, with a relatively flat performance curve as shown in Fig. 11. Note that the pump will shut off at 254' TDH. At the maximum discharge tank pressure, the pump will stop delivering fluid, as the system head is greater than the pump TDH. A second consideration associated with static system head is motor overload on pump runout. Again, consider Figure 7A-11 at the minimum system head of 243'. The pump under discussion will deliver 130 gpm against 243' head. Homepower requirements will increase from 8.9 BHP at 100 gpm to 12.0 BHP at 130 gpm. A 10 HP motor could be overloaded on this service. NPSH problems may also arise when large increases in flow occur. At the rating of 100 gpm at 250' TDH the NPSHR of the pump is only 10' while the system NPSHA is 13'. At the lower system head of 243' the pump requires 13.5' NPSH and cavitation will probably occur.
Appendix 7A: Centrifugal Pump Fundamentals 257
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CAPACITY
Flg. 7A.11
caPAcilryoiql
A better selection would be a pump with a characteristic as shown in Figure 7A-12. The steeper characteristic will limit the flow to between 90 GPM at 257' TDH and 110 gpm at 243' TDH. The small increase in capacity at low head condition will mean no motor overload. Since the maximum flow is 110 gpm, the maximum NPSHR will be 12' and the pump will not cavitate.
Flg. 7A.12 Unlike the static system, the friction system is always selfcorrecting to some degree. Consider the above system with a flow requirement of 6000 gpm at 150' TDH. Also assume that the discharge tank level may drop 10'. The new system head curve will be parallel to the original one, but 10' lower as shown in Figure 7A-14. Flow under this reduced head will be 6600 gpm at 144' rather than the normal 6000 gpm at 150'. This increased flow rate will tend to raise the discharge tank level back to normal.
DYNAMIC SYSTEM HEAD In frictional systems where resistance to flow increases with flow, the system head characteristic becomes curved. The magnitude of the system head at each flow is the summation of the system static head plus the total friction losses at that particular flow rate. A typical example of this type of system is shown in Figure 7A-13.
The frictional resistance of pipes and fittings will increase as they wear, resulting in greater curvature of the system head
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Special care must be taken in selecting pumps for parallel operation. Consideration must be given to single pump operation in the system as well as parallel operation. Consider the system shown in Figure 7A-15. The NPSH available is plotted along with the system head. Since entrance and line losses increase with increases in flow, the NPSHA decreases with flow increases.
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!= t
The flow required is 16,000 gpm. We want to use two pumps in parallel, but each must be capable of single operation.
~-.-_-_.-~_ ._..--_2 /
The total system head at 16,000 gpm is 140'. Each pump must be sized for 8000 gpm at 140' TDH. NPSHR for each pump must be less than 28' for parallel operation. Consider applying two pumps each with characteristics as shown in Figure 7A-16. In order to study both parallel and single pump operation, the head-capacity curves for both single and parallel operation must be plotted with the system head curve.
STATIC ELEVATIONC H A N G E /
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The head-capacity pump curve for parallel operation is plotted by adding the capacities of each pump for several different heads and plotting the new capacity at each head. The shutoff head for the two pumps in parallel is the same as for single operation. The NPSH curve is plotted in the same manner as shown in Figure 7A-16. For example, the NPSHR for one pump at 8000 gpm is 14'. Therefore, in parallel operation 16,000 gpm can be pumped with 14' NPSHR by each pump.
Flg. 7A-14 curve. A slight drop in the pump head curve may also result from increasing pump wear and recirculation. These changes will have less effect on the flow in a dynamic system (steep curve) than in a static system (flat curve).
The curve show that each pump will deliver 8000 gpm at 140' TDH when operating in parallel. Brake horsepower for each unit will be 340 HP. NPSHR is 14'. NPSHA is 28'.
PARALLEL OPERATION It is sometimes desirable to use two or more pumps in parallel rather than a single larger pump. This is particularly advantageous when the system flow requirements vary greatly. One pump can be shut down when the flow requirement drops, allowing the remaining pump or pumps to operate closer to their peak efficiency. It also provides an opportunity for repairs or maintenance work on one unit without shutting down the entire system.
With only one pump operating, the flow will be 11,000 gpm at 108' TDH. BHP will be 355 HP. NPSHR is 26' and NPSHA is 30'. A 400 HP motor would be required. This example shows that if a 350 HP motor had been selected based on parallel operation only, the motor would have
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Appendix 7A: Centrifugal Pump Fundamentals 259 v I '
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Pump Application Data Corrosion and Materials of Construction
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been overloaded in single pump operation. The single pump operation is also critical in terms of NPSH. For example, if the system NPSHA had been in the neighborhood of 20', parallel pump operation would have been fine, but single pump operation would result in cavitation.
(1) Galvanic corrosion is the electro-chemical action produced when one metal is in electrical contact with another more noble metal, with both being immersed in the same corroding medium called the electrolyte. A galvanic cell is formed and current flows between the two materials. The least noble material called the anode will corrode while the more noble cathode will be protected. It is important that the smaller wearing parts in a pump be of a more noble material than the larger more massive parts, as in an iron pump with bronze or stainless steel trim. Following is a galvanic series listing the more common metals and alloys: Corroded End (Anodic, or least noble)
Nickel base alloy (active) Brasses
Basic Formulae and Symbols FORMULAS
0.002 • Lb./Hr. GPM = Sp. Gr. H=
2.31 x psi Sp. Gr.
H=
1.134 x In. Hg. Sp. Gr.
V2 hv = ~ = .0155V2 GPM x 0.321 A
V"-
Ns
--
=
GPM x 0.409 (I.D.) 2
GPM x H x Sp. Gr. 3960 x Eft.
BHP =
Eff.
=
GPM x H x Sp. Gr. 3960 x BHP NvrGPM H3/4 V 2
H-"
V--
2g NxD 229
DEG. C = (DEG. F - 32) x 5/9 DEG. F = (Deg. C x 9/5) + 32
GPM x psi 1715 x Eft.
SYMBOLS
GPM Lb. Hr. Sp. Gr. H psi In. Hg. hv V g
= = = = = = = = = =
A I.D. BHP Eft.
= = = =
Ns N v
= = =
D
=
gallons per minute pounds hour specific gravity head in feet pounds per square inch Inches of mercury velocity head in feet velocity in feet per second 32.16 ft/sec 2 (acceleration of gravity) area in sqare inches Inside diameter in inches brake horsepower pump efficiency expressed as a decimal specific speed speed in revolutions per minute peripheral velocity of an impeller in feet per second Impeller in inches
260 Process Plant Machinery Copper Bronzes Copper-Nickel Alloy Monel Silver Solder Nickel (Passive) Nickel Base Alloy (Passive) Stainless Steel, 400 Series (Passive) Stainless Steel, Type 304 (Passive) Stainless Steel, Type 316 (Passive) Silver Graphite Gold Platinum Protected End (Cathodic, or most noble)
hydroxide ion concentration in gram equivalents per liter, pH value is expressed as the logarithm to the base 10 of the reciprocal of the hydrogen ion concentration. The scale of pH values is from zero to 14, with 7 as a neutral point. From 6 to zero denotes increasing hydrogen ion concentration and thus increasing acidity; and from 8 to 14 denotes increasing hydroxide ion concentration and thus increasing alkalinity.
(2) Uniform Corrosion is the overall attack on a metal by a corroding liquid resulting in a relatively uniform metal loss over the exposed surface. This is the most common type of corrosion and it can be minimized by the selection of a material which offers resistance to the corroding liquid.
The pH value should only be used as a guide with weak aqueous solutions. For more corrosive solutions, temperature and chemical composition should be carefully evaluated in the selection of materials of construction.
Magnesium Magnesium Alloys Zinc Aluminum 2S Cadmium Aluminum 17ST Steel or Iron Cast Iron Stainless Steel, 400 Series (Active) Stainless Steel, Type 304 (Active) Stainless Steel, Type 316 (Active) Lead-tin Solders Lead Tin Nicket (Active)
The table below outlines materials of construction usually recommended for pumps handling liquids of known pH value. pH Value
10 to 8 to 6 to 4 to 0 to
(3) Intergrenular corrosion is the precipitation of chromium carbides at the grain boundaries of stainless steels. It results in the complete destruction of the mechanical properties of the steel for the depth of the attack. Solution annealing or the use of extra low carbon stainless steels will eliminate intergranular corrosion.
C.l.-Cast Iron, ASTM A48. D.l.-Ductile Iron, ASTM A536. Steel-Carbon Steel, ASTM A216-WCA or WCB. Brz-Anti-Acid Bronze, Similar to ASTM B143A2. 316SS-Stainless Steel, ASTM A744 Gr. CF-8M, AISI 316. GA-20-Carpenter Stainless No.20, ASTM A744 Gr. CN-7M. CD4MCu-Stainless Steel, ACI CD-4MCu. Mon-Monel Grade E. ASTM A744 Gr. M-35. NI-Nickel, ASTM A744 Gr.CZ-100. H-B-Hastelloy Alloy-B, ASTM A494. H-C-Hastelloy Alloy-C, ASTM A494. TI-Titanium Unalloyed, ASTM B367 Gr. C-1. ZI-Zirconium
(6) Stress Corrosion is the failure of a material due to a combination of stress and a corrosive environment, whereas the material would not be affected by the environment alone. (7) Erosion-Corrosion is the corrosion resulting when a metal's protective film is destroyed by high velocity fluids. It is distinguished from abrasion which is destruction by fluids containing abrasive solid particles.
Code
pH Values
A - Fully Satisfactory. C - Limited Use. B - Useful Resistance. X - Unsuitable.
The pH of a liquid is an indication of its corrosive qualities, either acidic or alkaline. It is a measure of the hydrogen or
Acetaldehyde, 70~ Acetic Acid, 70~ Acetic Acid, < 50%, To Boiling Acetic Acid, > 50% To Boiling Acetone, To Boiling Aluminum Chloride, < 10%, 70~ Aluminum Chloride, > 10%, 70 oF. Aluminum Chloride, < 10%, To Boiling
X X X A X X X
B
Corrosion Resistant Alloys All Iron Bronze fitted or Standard fitted All Bronze Corrosion Resistant Alloy Steels
This chart is intended as a guide in the selection of economical materials. It must be kept in mind that corrosion rates may vary widely with temperature, concentration, and the presence of trace elements or abrasive solids. Blank spaces indicate a lack of accurate corrosion information for those specific conditions.
(5) Crevice or Concentration Cell Corrosion occurs in joints or small surface imperfections. Portions of the liquid become trapped and a difference in potential is established due to the oxygen concentration difference in these cells. The resulting corrosion may progress rapidly leaving the surrounding area unaffected.
Corrosive
14 10 8 6 4
Materials Selection Chart
(4) Pitting Corrosion is a localized rather than uniform type of attack. It is caused by a breakdown of the protective film and results in rapid pit formation at random locations on the surface.
steel C.I. D.I.
Material of Construction
Brz.
316SS
GA-20
CD4MCu
Mon
Ni
H-B
A
A
A A A A A B B C
A A B C A C C X
A B B B A B C X
A B B B A C X X
A C X A A A A
A B X A B X X
A A B A C C X
H-C A A A A A B B X
A A A A A A A A Cont.
Appendix 7A: Centrifugal Pump Fundamentals 261
Corrosive
steel C.I. D.I.
Brz.
316SS
GA-20
CD4MCu
Mon
Ni
H-B
H-C
TI
Zl
Aluminum Chloride, >10%, To Boiling Aluminum Sulphate, 70 ~ Aluminum Sulphate, < 10%, To Boiling Aluminum Sulphate, > 10%, To Boiling Ammonium Chloride, 70~ Ammonium Chloride, < 10%, To Boiling Ammonium Chloride, > 10%, To Boiling Ammonium Fluosilicate, 70 ~ Ammonium Sulphate, < 40%, To Boiling Arsenic Acid, to 225~ 9
X X X X X X X X X X
X B B C X X X X X X
X A B C B B X C B C
X A A B B B C B B B
X A B C B C X C C C
X B X X B B C X B X
X B X X B B C X B X
A B A B
X B A B A A C C B
X A A C A A C X A
A A A B A A C X A
Barium Chloride, 70~ < 30% Barium Chloride, < 5%,To Boiling Barium Chloride, > 5%, To Boiling Barium Hydroxide, 70 ~ Barium Nitrate, To Boiling Barium Sulphide, 70 ~ Benzoic Acid Boric Acid, To Boiling Boron Trichloride, 70 ~ Dry Boron Trifluoride, 70 ~F. 10%, Dry Brine (acid), 70~ Bromine (dry), 70~ Bromine (wet), 70~
X X X B C C X X B B X X X
B B C X X X C C B B X X X
C C X A B B B B B B X X X
B B C A B B B B B A X X X
C C X A B B B B B B X X X
B B C B
B B C B B
B B C B
X B C B A
B B C A B X B C B A
A A B
B A C A B A A B
X X
C C
A A B A B B B
B A C A B A A B B X X
X X
Calcium Bisulphite, 70 ~ Calcium Bisulphite, To Hot Calcium Chloride, 70~ Calcium Chloride, < 5%, To Boiling Calcium Chloride, > 5%, To Boiling Calcium Hydroxide, 70~ Calcium Hydroxide, < 30%, To Boiling Calcium Hydroxide, > 30%, To Boiling Calcium Hypochlorite, < 2%, 70~ Calcium Hypochlorite, > 2%, 70~ Carbolic Acid, 70~ (phenol) Carbon Bisulphide, 70 ~ Carbonic Acid, 70~ Carbon Tetrachloride, Dry to Boiling Chloric Acid, 70~ Chlorinated Water, 70~ Chloroacetic Acid, 70~ Chlorosulphonic Acid, 70 ~ Chromic Acid, < 30% Citric Acid Copper Nitrate, to 175 ~F. Copper Sulphate, To Boiling Cresylic Acid Cupric Chloride Cyanohydrin, 70~
X X B C X B C X X X C B B B X C X X X X X X C X C
X X C C C B B X X X B B C B X C
B C B B C B B C X X A A A A X B X X C A B C B X B
B B B B B B B C C C A A A A B B X C B A B B B X B
B C B B C B B C X X A A A A C B
X X B A C B B C X X A B C A X
X X B A C B B C X X A B B A X
X C A B C B X B
X X C X X C C
Dichloroethane Diethylene Glycol, 70~ Dinitrochlorobenzene, 70~
C A C
B B B
B A A
B A A
B A A
Ethanolamine, 70~ Ethers, 70~ Ethyl Alcohol, To Boiling Ethyl Cellulose, 70 ~ Ethyl Chloride, 70 ~ Ethyl Mercaptan, 70~ Ethyl Sulphate, 70~ Ethylene Chlorohydrin, 70 ~ Ethylene Dichloride, 70 ~ Ethylene Glycol, 70~ Ethylene Oxide, 70~
B B A A C C C C C B C
X B A B B X B B B B X
B B A B B B B B B B B
B A A B A A A B B B B
Ferric Chloride, < 5%, 70 ~ Ferric Chloride, > 5%, 70 ~ Ferric Nitrate, 70 ~ Ferric Sulphate, 70 ~ Ferrous Sulphate, 70 ~
X X X X X
X X X X C
X X B C C
X X A B B
(dry)
X X C X C C C
B X
B
A A A
A
B C A A A A A B A B A
A B X
A B C A
X X C X X C X
A
A B A X A B C
C B A
B B A
B B A
B B A
B A A B B B B B B B B
C B A B B
X B A B B
B B B B B
B A B B B
B A B B B
B B B B
B B A A
X X B C C
X X X C C
X X X C C
X X
A X B
B
A A A A B A A A A A A A A A
A A A A B
A B A A A
A A B A A B A
A B X A A
B
X
A A A
B A A
A A A A A
A A A A A
B C A A
A A A A
A A A A
A B B B B
A B
B X
B A
B A
A
Cont.
262 Process Plant Machinery steel C.I. D.I.
Brz.
316SS
GA-20
Formaldehyde, To Boiling Formic Acid, to 212~ Freon, 70~
B X A
B C A
A X A
A A A
B A A
A C A
A A A
Hydrochloric Acid, < 1%, 70 ~F. Hydrochloric Acid, 1 - 20%, 70 ~F. Hydrochloric Acid, > 20%, 70 ~ Hydrochloric Acid, < 1/2%, 175~ Hydrochloric Acid, 1/2-2%,175 ~ Hydrocyanic Acid, 70~ Hydrogen Peroxide, < 30% < 150 ~ Hydrofluoric Acid, < 20%, 70 ~ Hydrofluoric Acid, > 20%, 50 ~ Hydrofluoric Acid, To Boiling Hydrofluorsilicic Acid, 70~
X X X X X X C X X X X
X X X X X X X B C X
C X X C X C B X X X C
B X X C
A C C C C C B B B C B
B X X X X
A A B A A
A X X X
A X X X
Lactic Acid, < Lactic Acid, > Lactic Acid, < Lime Slurries,
X X X B
B B X B
A B C B
A B B B
X C X B
C C X B
B B B B
B B B B
A A A B
A A A B
Magnesium Chloride, 70 ~ Magnesium Chloride, < 5%, To Boiling Magnesium Chloride, > 5%, To Boiling Magnesium Hydroxide, 70 ~ Magnesium Sulphate Maleic Acid Mercaptans Mercuric Chloride, < 2%, 70~ Mercurous Nitrate, 70 ~ Methyl Alcohol, 70 ~
C X X B C C A X C A
C C C A C C X X X A
B C X B B B A X B A
A B C B A B A X B A
C C C B B C X X C A
C C C A B C X C
A A B B C B
A A B B C B
A A B A B A
A A B
A
A
A
A
B C A
A
A
Naphthalene Sulphonic Acid, 70 ~ Naphthalenic Acid, To Hot Nickel Chloride, 70 ~ Nickel Sulphate Nitric Acid Nitrobenzene, 70~ Nitroethane, 70~ Nitropropane, 70 ~ Nitrous Acid, 70~ Nitrous Oxide, 70~
X C X X X A A A X C
C C X C X C A A X C
B B C B B A A A X C
B B B B B A A A C C
C C C C X B A A X X
C C X C X B A A X X
B B A
B B
B
B A A
B A A
B A B
Oleic Acid Oleum, 70~ Oxalic Acid
C B X
C X C
B B C
B B B
C X C
C X C
C B B
C B B
Palmitic Acid Phenol (see carbolic acid) Phosgene, 70~ Phosphoric Acid, < 10%, 70 ~F. Phosphoric Acid, > 1 0 - 70%, 70 ~F. Phosphoric Acid, < 20%, 175 ~ Phosphoric Acid, > 20%, 175 ~ < 85% Phosphoric Acid, > 10%, Boil, < 85% Phthalic Acid, 70~ Phthalic Anhydride, 70~ Picric Acid, 70~ Potassium Carbonate Potassium Chlorate Potassium Chloride, 70 ~ Potassium Cyanide, 70 ~ Potassium Dichromate Potassium Ferricyanide Potassium Ferrocyanide, 70 ~ Potassium Hydroxide, 70 ~ Potassium Hypochlorite Potassium Iodide, 70 ~ Potassium Permanganate Potassium Phosphate
B
B
B
A
B
B
C X X X X X C B X B B C B B C X C X C B C
C C C C C C B C X B C C X B B B C C B B C
B A A B C X B A C A A B B A B B B C B B B
B A A B B C A A B A A A B A B B A B B B B
C C C C C C B A C B C B C B B B A X B C
C C C C C C B A X B C B C B B B A X B B
B A B A B C B A
B A C A C C B A B B B B B B B B C B B B
Sea Water, 70~ Sodium Bisulphate, 70 ~
C X
B C
B C
A B
Corrosive
50%, 70 ~F. 50%, 70 ~F. 5%, To Boiling 70 ~
CD4MCu
H-B
Mon
B B B C X B
H-C
B
B A A A
B
A A
C
B B B B B B B
A B
C B X
C
A B C C C A
A B B C C A
A A A
A A A
A A
A A B A
B A A
A
A
B
B
A B
A A Cont.
Appendix 7A: Centrifugal Pump Fundamentals 263 steel C.I. D.I.
Brz.
316SS
GA-20
CD4MCu
Mon
Ni
H-B
H-C
Sodium Bromide, 70~ Sodium Carbonate Sodium Chloride, 70~ Sodium Cyanide Sodium Dichromate Sodium Ethylate Sodium Fluoride Sodium Hydroxide, 70 ~ Sodium Hypochlorite Sodium Lactate, 70 ~ Stannic Chloride, < 5%, 70~ Stannic Chloride, > 5%, 70 ~ Sulphite Liquors, To 175 ~ Sulphure (molten) Sulphur Dioxide (spray), 70 ~ Sulphuric Acid, < 2%, 70~ Sulphuric Acid, 2-40%, 70~ Sulphuric Acid, 40%, < 90%, 70 ~ Sulphuric Acid, 93-98%, 70~ Sulphuric Acid, < 10%, 175~ Sulphuric Acid, 10-60% & > 80%, 175~ Sulphuric Acid, 60-80%, 175~ Sulphuric Acid, < 3/4%, Boiling Sulphuric Acid, 3/4-40%, Boiling Sulphuric Acid, 4 0 - 6 5 % & > 85%, Boil Sulphuric Acid, 65-85%, Boiling Sulphurous Acid, 70 ~F.
B B C B B B C B X B X X X B C X X X B X X X X X X X X
C B B X X A C B X C C X C X C C C X X C X X X X X X C
B B B B B A B B C C X X B A B B C X B X X X C X X X C
B A B B B A B A C C C X B A B A B B B B B X B C X X B
B B B B B A B B C C X X B A B B C X B X X X C X X X C
B B A X
B B A X
B B B
B B B
A B A X C C X C C C C C X X X X X X X X X X
A B A X
C A
C X C C C C C X X X X X X X X X X
A A A B A B B B B X X B
C A B C B C B A B A A A B C C C B C X X B
Titanium Tetrachloride, 70 ~ Trichlorethylene, To Boiling
C B
C
C B
B B
C B
C B
B
B
Urea, 70~
C
C
B
B
B
C
C
Vinyl Acetate Vinyl Chloride
B B
B C
B B
B B
B B
C
Water, To Boiling
B
A
A
A
A
A
Zinc Chloride Zinc Cyanide, 70~ Zinc Sulphate
C X X
C B C
B B A
A B A
B B A
B B C
Corrosive
Piping Design The design of a piping system can have an important effect on the successful operation of a centrifugal pump. Such items as sump design, suction piping design, suction and discharge pipe size, and pipe supports must all be carefully considered. Selection of the discharge pipe size is primarily a matter of economics. The cost of the various pipe sizes must be compared to the pump size and power cost required to overcome the resulting friction head.
TI
Zl
A A B B
A A
B A A
B A B
A B A A C B X X X X X X X X X X A
A B
A A C C B C C B B X X B
C B
A
A
C
C
B
B
C
C
B B
A
A
A
A
A
A
B B C
B B C
B C
A B A
A B
C B B C
separation keeps the liquid from evenly filling the impeller. This upsets hydraulic balance leading to vibration, possible cavitation, and excessive shaft deflection. Shaft breakage or premature bearing failure may result. On pump installations involving suction lift, air pockets in the suction line can be a source of trouble. The suction pipe should be exactly horizontal, or with a uniform slope upward from the sump to the pump as shown in Figure 7A-17. There should be no high spots where air can collect and cause the pump to lose its prime. Eccentric rather than concentric reducers should always be used.
The suction pipe should never be smaller than the suction connection of the pump, and in most cases should be at least one size larger. Suction pipes should be as short and as straight as possible. Suction pipe velocities should be in the 5 to 8 feet per second range unless suction conditions are unusually good.
If an elbow is required at the suction of a double suction pump, it should be in a vertical position if at all possible. Where it is necessary for some reason to use a horizontal elbow, it should be a long radius elbow and there should be a minimum of two diameters of straight pipe between the elbow and the pump as shown in Figure 7A-18. Figure 7A-19 shows the effect of an elbow directly on the suction. The liquid will flow toward the outside of the elbow and result in an uneven flow distribution into the two inlets of the double suction impeller. Noise and excessive axial thrust will result.
Higher velocities will increase the friction loss and can result in troublesome air or vapor separation. This is further complicated when elbows or tees are located adjacent to the pump suction nozzle, in that uneven flow patterns or vapor
There are several important considerations in the design of a suction supply tank or sump. It is imperative that the amount of turbulence and entrained air be kept to a minimum. Entrained air will cause reduced capacity and efficiency as
The suction piping size and design is far more important. Many centrifugal pump troubles are caused by poor suction conditions.
264 Process Plant Machinery
CHECK VALVE
ECCENTRIC REDUCER
GATE VALVE
LONG RADIUS
(la) CORRECT
FOOT VALVE (IF USED)
CHECK VALVE
ECCENTRIC REDUCER
LONG RADIUS ELBOW
f
GATE VALVE
SUCTION PIPE SLOPES UPWARDS FROM SOURCE OF SUPPLY
(lb) CORRECT FOOT VALVE (IF USED) STRAINER AIR POCKET BECAUSE ECCENTRIC REDUCER IS NOT USED AND BECAUSE SUCTION PIPE DOES NOT SLOPE GRADUALLY UPWARD FROM SUPPLY
GATE VALVE
/ CHECK GATE VALVE SHOULD NOT BE BETWEEN VALVE CHECK VALVE AND PUMP
(1r WRONG
Fig. 7A-17
Air pockets in suction piping
well as vibration, noise, shaft breakage, loss of prime, and/or accelerated corrosion. The free discharge of liquid above the surface of the supply tank at or near the pump suction can cause entrained air to enter the pump. All lines should be submerged in the tank, and baffles should be used in extreme cases as shown in Figure 7A-20. Improper submergence of the pump suction line can cause a vortex which is a swirling funnel of air from the surface directly
into the pump suction pipe. In addition to submergence, the location of the pipe in the sump and the actual dimensions of the sump are also important in preventing vortexing and/or excess turbulence. For horizontal pumps, Figure 7A-21 can be used as a guide for minimum submergence and sump dimensions for flows up to approximately 3000 gpm. Baffles can be used to help prevent vortexing in cases where it is impractical or impossible to maintain the required submergence. Figure 7A-22 shows three such baffling arrangements.
Appendix 7A" Centrifugal Pump Fundamentals 265
Fig. 7A-18 Elbows at pump suction
Fig. 7A-19 Effect of elbow directly on suction
Fig. 7A-20 Keeping air out of pump
266 Process Plant Machinery
Flg. 7A-21
Mlnlmum suctlon plpe submergence and sump dlmenslons
Fig. 7A-22
Baffle arrangements for vortex prevention
Large units (over 3000 gpm) taking their suction supply from sumps, especially vertical submerged pumps, require special attention. The larger the unit, the more important the sump design becomes.
be located near the back wall and should not be subjected to rapid changes in direction of the flow pattern. The velocity of the water in the area of the suction pipes should be kept below one foot per second to avoid air being drawn into the pump.
Figure 7A-23 illustrates several preferred piping arrangements within a multiple pump pit. Note that the pipe should always
On horizontal pumps, a bell should be used on the end of the suction pipe to limit the entrance velocity to 3.5 feet
Appendix 7A: Centrifugal Pump Fundamentals 267 RECOMMENDED
NOT RECOMMENDED ~
Q
LT
0~
V, ,-, 1 fps OR LESS
V, = 2 tpS & UP
~
= t 'h TO 20
.-=
i
i
9
IF A = LESS THAN
A
0 o
NESS TO (~ DIST. .oo ROUND OR OGIVE
G
0
mmmm
WALL ENDS. GAP AT REAR OF WALL APPX. O13
@
d
,,mmm
0
~
i
0
i
i
9/16 D
-- A
PREFERED oc = 75"
Fig. 7A-23
Piping arrangements within multiple pump pits
per second. Also, a reducer at the pump suction flange to smoothly accelerate and stabilize the flow into the pump is desirable. The submergence of the suction pipe must also be carefully considered. The amount of submergence required depends upon the size and capacity of the individual pumps as well as on the sump design. Past experience is the best guide for determining the submergence. The pump manufacturer should be consulted for recommendations in the absence of other reliable data.
Stuffing B o x S e a l i n g The stuffing box of a pump provides an area in which to seal against leakage out of the pump along the shaft. Packing and mechanical seals are the two devices used to accomplish this seal.
Packing A typical packed stuffing box arrangement is shown in Figure 7A-24. It consists of: A)Five rings of packing. B)A lantern ring used for the injection of a lubricating and/or flushing liquid, and C ) A gland to hold the packing and maintain the desired compression for a proper seal. The function of packing is to control leakage and not to eliminate it completely. The packing must be lubricated, and a flow of from 40 to 60 drops per minute out of the stuffing box must be maintained for proper lubrication.
Lantern Sealing Liquid PackingGland Stufting Ring~ (Quench - ~ , ~ ,,~ Connection / Type) Box ~ i II \ ~ / Bushing ~ "~J" t , !1a ~. ~ . q 7 ~I ,....L (
J
i~\'3 . . . . . . . . . . . .
- (~
H
kZ
U
S'oUx
Throat
Fig. 7A-24 of parts)
-
~ ---~
"
Mechanical Packing
Typical stuffing box arrangement (description
The method of lubricating the packing depends on the nature of the liquid being pumped as well as on the pressure in the stuffing box. When the pump stuffing box pressure is above atmospheric pressure and the liquid is clean and nonabrasive, the pumped liquid itself will lubricate the packing Figure 7A-25. When the stuffing box pressure is below atmospheric pressure, a lantern ring is employed and lubrication is injected into the stuffing box (Figure 7A-26). A bypass line from the pump discharge to the lantern ring connection is normally used providing the pumped liquid is clean.
268 Process Plant Machinery
Lantern Ring Location F
Injected Fluid From
Thick SlurJ Including Paper Stock
Positive Fluid Pressure
Above Atmospheri, Pressure
eric re
Leakage Into _ Pump
Fig. 7A-25 Typical stuffing box arrangement when stuffIng box pressure is above atmosphere pressure
Fig. 7A-27 Typical stuffing box arrangement when pump ing slurries
Mechanical Seals The Basic Seal ,tic 3
Leakage Into _ Pump
A mechanical seal is a sealing device which forms a running seal between rotating and stationary parts. The design of liquid handling equipment with rotating parts today would include the consideration for the use of mechanical seals. Advantages over conventional packing are as follows: 1. Reduced friction and power losses. 2. Zero or limited leakage of product. 3. Elimination of shaft or sleeve wear. 4. Reduced maintenance. 5. Ability to seal higher pressures and more corrosive environments.
Fig. 7A-26 Typical stuffing box arrangement when stuffIng box pressure is below atmospheric pressure
The wide variety of styles and designs together with extensive experience allows the use of seals on most pump applications. A mechanical seal must seal at three points:
When pumping slurries or abrasive liquids, it is necessary to inject a clean lubricating liquid from an external source into the lantern ring (Figure 7A-27). A flow of from .2 to .5 gpm is desirable and a valve and flowmeter should be used for accurate control. The seal water pressure should be from 10 to 15 psi above the stuffing box pressure, and anything above this will only add to packing wear. The lantern ring is normally located in the center of the stuffing box. However, for extremely thick slurries like paper stock, it is recommended that the lantern ring be located at the stuffing box throat to prevent stock from contaminating the packing. The gland shown in Figures 7A-24 to 27 is a quench type gland. Water, oil, or other fluids can be injected into the gland to remove heat from the shaft, thus limiting heat transfer to the bearing frame. This permits the operating temperature of the pump to be higher than the limits of the bearing and lubricant design. The same quench gland can be used to prevent the escape of a toxic or volatile liquid into the air around the pump. This is called a smothering gland, with an external liquid simply flushing away the undesirable leakage to a sewer or waste receiver.
1. Static seal between the stationary part and the housing. 2. Static seal between the rotary part and the shaft. 3. Dynamic seal between the rotating seal face and the stationary seal face. Figure 7A-28 shows a basic seal with these components: 1. Stationary seal part positioned in the housing with preload on the "O" ring to effect sealing and prevent rotation. 2. Rotating seal part positioned on the shaft by the "O" ring. The "O" ring seals between it and the shaft and provides resiliency. 3. The mating faces. The faces are precision lapped for a flatness of 3 light bands and a surface finish of 5 microinches. 4. Spring assembly, rotates with the shaft and provides pressure to keep the mating faces together during periods of shut down or lack of hydraulic pressure.
Appendix 7A: Centrifugal Pump Fundamentals 269
HOUSING
_
X
/
\
_
SPRING ASSEMBLY
STATIONARY SEAL \ ROTATING SEAL PART PART / MATING FACES
Fig. 7A-28
INJECTED FLUID FROM PUMP VOLUTE /
Basic mechanical seal
FLUSH GLAND
PLAIN
~ \ ~ l . r ~
CLAMPED ,N
INJECTED FLUID .o...,.o
,J.;,~'"%mi;Ei~iS')
Flg. 7A-29
Slngle, Inslde, unbalanced seal
Flg. 7A-31
CIRCULATED FLUID INLET CONNECTION
RESTRICTING INJECTED BUSHING FLUID PLAIN
~\~-~y,~
~
~ ...~
.
.
Fig. 7A-30
.
.
.
.
.
.
.
.
.
.
CLAMPED IN
IONARY SEAT .
.
.
.
.
Slngle, Inslde, balanced seal
OUTLET CONNECTION FLUSH GLAND
~._
Single, outside, unbalanced seal
Fig. 7A-32
Double, inside, unbalanced seal
Fig. 7A-33
Tandem, Inside unbalanced seals
5. Driving member, positions the spring assembly and the rotating face. It also provides the positive drive between shaft and the other rotating parts. As wear takes place between the mating faces, the rotating face must move along the shaft to maintain contact with stationary face. The "O" ring must be free to move. A mechanical seal operates as each basic component performs its duty. Liquid pressure in the seal chamber forces the faces together and provides a thin film of lubricant between them. The faces, selected for low frictional qualities, are the only rubbing parts. These basic components are a part of every seal. The form, shape, style and design will vary greatly depending on service and manufacture. The basic theory, however, remains the same.
270 Process Plant Machinery Types Mechanical seals can be classified into the general types and arrangements shown below. Understanding these classes provides the first step in proper seal selection. (a) Single seals-Inside, outside, unbalanced, balanced (b) Double seals- Unbalanced or balanced Single Seal, Inside Unbalanced
The single inside seal mounts on the shaft or sleeve within the stuffing box housing. The pumpage is in direct contact with all parts of the seal and provides the lubrication for the faces. The full force of pressure in the box acts on the faces providing good sealing to approximately 100 P.S.I.G. This is the most widely used type for services handling clear liquids. A circulation or by-pass line connected from the volute to the stuffing box provides continual flushing of the seal chamber. Single Seals, Outside Unbalanced
This type mounts with the rotary part outside of the stuffing box. The springs and drive element are not in contact with the pumpage, thus reducing corrosion problems and preventing product accumulation in the springs. Pressures are limited to the spring rating, usually 35 P.S.I.G. Usually the same style seal can be mounted inside or outside. The outside seal is easier to install, adjust and maintain. A restricting bushing can be used to control leakage of an external sealing liquid into the pumpage. Single Seals, Balanced Balancing a seal varies the face loading exerted by the box pressure, thus extending the pressure limits of the seal. A balanced rotating part utilizes a stepped face and a sleeve. Balanced seals are used to pressures of 2000 P.S.I.G. Their use is also extensive on light hydrocarbons which tend to vaporize easily. Balanced outside seals allow box pressure to be exerted toward the seal faces, thus allowing pressure ranges to above 150 P.S.I. as compared to the 35 P.S.I.G. limit for the unbalanced outside seal. Double Seals
Double seals use two seals mounted back to in the stuffing box. The stuffing box is pressurized with a clear liquid from an external source. This liquid is circulated thru the double seal chamber at 1/4-1 GPM to cool and lubricate the mechanical
seals. Double seals are used on solutions that contain solids, are toxic or extremely corrosive. The external source fluid should be compatible with the pumpage. TANDEM SEALS
A variation of the double seal arrangement. The purpose of this seal is to provide a backup seal in the event the primary seal fails. The primary mechanical seal functions in a manner identical to that of a conventional single inside seal. The cavity between the primary seal and the backup seal is flooded with liquid to provide lubrication for the backup seal. The seal arrangement is used on toxic or hazardous chemical, and transfer and pipeline services to provide an extra measure of safety and allow equipment to operate until time to shut down. Selection The proper selection of a mechanical seal can be made only if the full operating conditions are known. These conditions are as follows: 1. Liquid 2. Pressure 3. Temperature 4. Characteristics of Liquid 1. Liquid Identification of the exact liquid to be handled provides the first step in seal selection. The metal parts must be corrosion resistant. These pads, usually available in steel, bronze, stainless steel, or Hastelloy, provide a wide choice to meet specific services. The mating faces must also resist both corrosion and wear Carbon, ceramic glass-filled Teflon, Stellite or tungsten carbide are available and offer both excellent wear properties and corrosion resistance. Stationary sealing members of synthetic rubber, asbestos and Teflon complete the proper material selection. 2. Pressure The proper type of seal, unbalanced or balanced, is based on the pressure on the seal and on the seal size. Figure 7A-34 shows the normal limits for unbalanced seals of various types. 3. Temperatures The temperature will in part determine the use of the sealing members Synthetic rubbers are used to approximately 400 F., Teflon to 500 F. and asbestos to
:I k
I
Maximum stuffing box pressures for Unbalanced Mechanical Seals on water solutions 160" max. Unbalanced seals generally limited to 200 PSIG maximum stuffing box pressure. All ratings
)400 =300 L
}=
i ~0o =,
m
.
\
9 -'~ ~
\
.~
~~S~.o r ~
~
based on one carbon face against hard face:
I
'
~
_
f .
I -CERAMIC 2-NI-RESIST ~
_
3-$TELI.ITE (not generally reeolIMtended on water sendr
4-TUNGSTEN CARBIDE
1 2 3 4 5 6 SEAL SIZE--INCHES (8Mfl or Sleeve O.D.)
Fig. 7A-34
Pressure-velocity limits, unbalanced seals
Appendix 7A: Centrifugal Pump Fundamentals 271 750 F. Cooling the liquid in the seal chamber by water cooling jackets or cool liquid flushing, often extends seal life and allows wider selection of materials.
L...
._i
,p_.._
4. Characteristics of Liquid Abrasive liquids create excessive wear and short seal life. Double seals or clear liquid flushing from an external source allows the use of mechanical seals on these difficult liquids. On light hydrocarbons balanced seals are often used to promote longer seal life, even though pressures are low.
II
Environmental Controls
Flg. 7A-36
Environmental controls are necessary for reliable performance of a mechanical seal on many applications. Pump manufactures and the seal vendors offer a variety of arrangements to combat these problems.
services, heat is provided to the jacket to melt or prevent the liquid from freezing (liquid sulfur).
1. Corrosion
DIRTY or INCOMPATIBLE ENVIRONMENTS
2. Temperature Control 3. Dirty or incompatible environments
CORROSION Corrosion can be controlled by selecting seal materials that are not attacked by the pumpage. When this is difficult, external fluid injection of a non-corrosive chemical to lubricate the seal is possible. Single or double seals could be used, depending on if the customer can stand delusion of his product.
TEMPERATURE CONTROL As the seal rotates, the faces are in contact. This generates heat and if this heat is not removed, the temperature in the stuffing box can increase and cause sealing problems. A simple by-pass flush of the product over the seal faces will remove the heat generated by the seal (Figure 7A-35). For higher temperature services, by-pass of product through a cooler may be required to cool the seal sufficiently (Figure 7A-36). External cooling fluid injection can also be used. Jacketed stuffing boxes are used on many pumps to cool the environment around the mechanical seal (Figure 7A-29). This will also allow the use of a mechanical seal on services where it would not normally function (hot heat transfer oil). For other
Mechanical seals do not normally function well on liquids which contain solids or can solidify on contact with the atmosphere. Here, by-pass flush through a filter, a cyclone separator or a strainer are methods of providing a clean fluid to lubricate the stuffing box. Strainers are effective for particles larger than the openings on a 40 mesh screen. Cyclone separators are effective on solids 10 micron or more in diameter, if they have a specific gravity of 2.7 and the pump develops a differential pressure of 30-40 psi. Filters are available to remove solids 2 micron and larger. If external flush with clean liquid is available, this is the most fail proof system. Lip seal or restricting bushings are available to control flow of injected fluid to flows as low as 1/2 GPM (Figure 7A-30). Quench type glands are used on fluids which tend to crystalize on exposure to air. Water or steam is put through this gland to wash away any build up. Other systems are available as required by the service.
Field Testing Methods A. Determination of total head The total head of a pump can be determined by gauge readings as illustrated in Figure 7A-37.
o
h
=
ha
_
Flg. 7A-35
Fig. 7A-37 readings
Determination
of total
head from
gauge
272
Process Plant Machinery
Negative Suction Pressure: TDH = Discharge gauge reading converted to feet of liquid + vacuum gauge reading converted to feet of liquid + distance between point of attachment of vacuum gauge and the centerline of the discharge gauges, h, in feet +
( Vd2 Vs2 I 2g 2g
MERCURY
Positive Suction Pressure: or TDH = Discharge gauge reading converted to feet of liquid-pressure gauge reading in suction line converted to ft. of liquid + distance between center of discharge and suction gauges, h, in feet+ Vs2 2g 2g
( V~
)
In using gauges when the pressure is positive or above atmospheric pressure, any air in the gauge line should be vented off by loosening the gauge until liquid appears. This assures that the entire gauge line is filled with liquid and thus the gauge will read the pressure at the elevation of the centerline of the gauge. However, the gauge line will be empty of liquid when measuring vacuum and the gauge will read the vacuum at the elevation of the point of attachment of the gauge line to the pipe line. These assumptions are reflected in the above definitions. The final term in the above definitions accounts for a difference in size between the suction and discharge lines. The discharge line is normally smaller than the suction line and thus the discharge velocity is higher. A higher velocity results in a lower pressure since the sum of the pressure head and velocity head in any flowing liquid remains constant. Thus, when the suction and discharge line sizes at the gauge attachment points are different, the resulting difference in velocity head must be included in the total head calculation. Manometers can also be used to measure pressure. The liquid used in a manometer is normally water or mercury, but any liquid of known specific gravity can be used. Manometers are extremely accurate for determining low pressures or vacuums and no calibration is needed. They are also easily fabricated in the field to suit any particular application. Figure 7A-38 and 39 illustrate typical manometer set ups.
B. Measurement of capacity
Flg. 7A-39
Manometer Indlcatlng pressure
is often practical when pumping into an accurately measured reservoir or tank, or when it is possible to use small containers which can be accurately weighed. These methods, however, are normally suited only to relatively small capacity systems.
c.) Venturi meter A venturi meter consists of a converging section, a short constricting throat section and then a diverging section. The object is to accelerate the fluid and temporarily lower its static pressure. The flow is then a function of the pressure differential between the full diameter line and the throat. Figure 7A-40 shows the general shape and flow equation. The meter coefficient is determined by actual calibration by the manufacturer and when properly installed the Venturi meter is accurate to within plus or minus 1%.
d.) Nozzle A nozzle is simply the converging portion of a venturi tube with the liquid exiting to the atmosphere. Therefore, the same formula can be used with the differential head equal to the gauge reading ahead of the nozzle. Figure 7A-41 lists theoretical nozzle discharge flows.
e.) Orlflce An orifice is a thin plate containing an opening of specific shape and dimensions. The plate is installed in a pipe and the flow is a function of the pressure upstream of the orifice. There are numerous types of orifices available and their descriptions and applications are covered in the Hydraulic Institute Standards and the ASME Fluid Meters Report. Orifices are not recommended for permanent installations due to the inherent high head loss across the plate.
a.) Magnetic Flow Meter
f.) Weir
A calibrated magnetic flow meter is an accurate means of measuring flow in a pumping system. However, due to the expense involved, magnetic flow meters are only practical in small factory test loops and in certain process pumping systems where flow is critical.
A weir is particularly well suited to measuring flows in open conduits and can be adapted to extremely large capacity systems. For best accuracy, a weir should be calibrated in place. However, when this is impractical, there are formulas which can be used for the various weir configurations. The
b.) Volumetric measurement Pump capacity can be determined by weighing the liquid pumped or measuring its volume in a calibrated vessel. This
TI hi
~/ H Q(GPM) = S.67 CDz= 1 - R4 C = Instrument Coefficient DI = Entrance Diameter in Inches D2 -= Throat Diameter in Inches II R -- D21D; |I~ H = Differential Head in Inches = nl -- nzL "l"-
L ~ ~ l ~ Flg. 7A-38 vacuum
Manometer Indlcstlng
Flg. 7A-40
Venturl meter
Appendix 7A: Centrifugal Pump Fundamentals 273
Theoretical Discharge of Nozzles in U.S. GPM Head
Veloc'y of , Oisch. Feet Feet per Sec.
Lbs. 10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120 125 130 135 140 145 150 175 200 250 300
I
i
23.1 38.6 34.6 47.25 46.2 54.55 57.7 61.0 69.3 66.85 80.8 72.2 92.4 77.2 103.9 81.8 115.5 86.25 127.0 90.4 138.6 94.5 150.1 98.3 161.7:102.1 , 173.2 i 105.7 1184.8 i 109.1 196.3 112.5 207.9 115.8 219.4 119.0 230.9 122.0 242.4 125.0 254.0 128.0 265.5 130.9 2 7 7 . 1 133.7 288.6 136.4 300.2 139.1 311,7 141.8 323.3 144.3 334.8 146.9 346.4 149.5 4 0 4 . 1 161.4 461.9 172.6 577.4 193.0 692.8 211.2
Diameter of Nozzle in Inches s/,6 I
0.37 0.45 0.52 0.58 0.64 0.69 0.74 0.78 0.83 0.87 0.90 0.94 0.98 1.01 1.05 1.08 1.11 1.14 1.17 1.20 1.23 1.25 1.28 1.31 1.33 1.36 1.38 1.41 1.43 1.55 1.65 1.85 2.02 1 s/z
10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120 125 130 135 140 145 150 175 i 200 i 250 i 300 ,
23.1 34.6 46.2 57.7 69.3 80.8 92.4 103.9 115.5 127.0 138.6 150.1 161.7, 173.2 184.8 196.3 207.9 219.4 230.9 242.4 254.0 265.5 277.1 288.6 300.2 311.7 323.3 334.8 346.4 404.1 461.9 577.4 692.8
38.6 47.25 54.55 61.0 66.85 72.2 77.2 81.8 86.25 90.4 94.5 98.3 102.1 105.7 109.1 112.5 115.8 119.0 122.0 125.0 128.0 130.9 13;~.7 136.4 139.1 141.8 144.3 146.9 149.5 161.4 172.6 193.0 211.2
213 260 301 336 368 398 425 451 475 498 521 542 563 582 602 620 638 656 672 689 705 720 736 751 767 780 795 809 824 890 950 1063 1163
I
s/= 1.48 1.81 2.09 2.34 2.56 2.77 2.96 3.13 3.30 3.46 3.62 3.77 3.91 4.05 4.18 4.31 4.43 4.56 4.67 4.79 4.90 5.01 5.12 5.22 5.33 5.43 5.53 5.62 5.72 6.18 6.61 7.39 8.08
$/,6 I
3.32 4.06 4.69 5.25 5.75 6.21 6.64 7.03 7.41 7.77 8.12 8.45 8.78 9.08 9.39 9.67 9.95 10.2 10.5 10.8 11.0 11.2 11.5 11.7 12.0 12.2 12.4 12.6 12.9 13.9 14.8 16.6 18.2
1% 289 354 409 458 501 541 578 613 647 678 708 737 765 792 818 844 868 892 915 937 960 980 1002 1022 1043 1063 1082 1100 1120 1210 1294 1447 1582 ,
2 378 463 535 598 655 708 756 801 845 886 926 964 1001 1037 1070 1103 1136 1168 1196 1226 1255 1282 1310 1338 1365 1390 1415 1440 1466 1582 1691 1891 2070
~,~ !
5.91 7.24 8.35 9.34 10.2 11.1 11.8 12.5 13.2 13.8 14.5 15.1 15.7 16.2 16.7 17.3 17.7 18.2 18.7 19.2 19.6 20.0 20.5 20.9 21.3 21.7 22.1 22.5 22.9 24.7 26.4 29.6 32.4
; 2%
~ I
13.3 16.3 18.8 21.0 23.0 24.8 26.6 28.2 29.7 31.1 32.5 i 33.8
135.2
! 36.4 ; 37.6 38.8 39.9 41.0 42.1 43.1 44.1 45.1 46.0 47.0 48.0 48.9 149.8 I 50.6 51.5 55.6 59.5 66.5 72.8 2Vz
479 591 585 723 676 835 756 934 828 1023 895 1106 957 1182 1015 1252 1070 1320 1121 , 1385 1172 1447 1220 1506 1267 1565 1310 1619 1 3 5 4 1672 1395 1723 1436 1773 1476 1824 1512 1870 1550 1916 1588 1961 1 6 2 1 2005 1 6 5 9 2050 1690 2090 1726 2132 1759 2173 1790 2212 1820 2250 1853 2290 2000 2473 2140 2645 2392 2955 2615 3235
]
s/z I
23.6 28.9 33.4 37.3 40.9 44.2 47.3 50.1 52.8 55.3 57.8 60.2 62.5 64.7 66.8 68.9 70.8 72.8 74.7 176.5 78.4 80.1 81.8 83.5 85.2 86.7 88.4 89.9 91.5 98.8 106. 118. 129.
%
~
I
7/,
I
36.9 53.1 45.2 65.0 52.2 75.1 58.3 84.0 63.9 92.0 69.0 99.5 73.8 106. 78.2 113. 82.5 119. 86.4 125. 90.4 130. 94.0 136. 97.7 141. ;101. 146. !104. 150. 108. 155. 111. 160. 114. 164. 117. 168. 120. 172. 122. 176. : 125. 180. 128. 184 130. 188. 1133. 192. I 136. 195. 138. 199. 140. 202. 143. 206. 154. 222. 165. 238. 185. 266. 202. I 291.
2% 714 874 1009 1128 1236 1335 1428 1512 1595 1671 1748 1819 1888 1955 2020 2080 2140 2200 2255 2312 2366 2420 2470 2520 2575 2620 2670 2715 2760 2985 3190 3570 3900 ,
3
I
3s/z
851 1041 1203 1345 1473 1591 1701 1802 1900 1991 2085 2165 2250 2330 2405 2480 2550 2625 2690 2755 2820 2885 2945 3005 3070 3125 3180 3235 3295 3560 3800 4250 4650 ,
I
72.4 88.5 i 102. 114. 125. 135. 145. 153. 162. 169. 177. 184. 191. 198. 205. 211. 217. 223. 229. 234. 240. 245. 251. 256. 261. 266. i 271. 275. 280. 302. 323. 362. 396 4
1158 1418 1638 1830 2005 2168 2315 2455 2590 2710 2835 2950 3065 3170 3280 3375 3475 3570 3660 3750 3840 3930 4015 4090 4175 4250 4330 4410 4485 4840 5175 5795 6330
~i
1 94.5 116. 134. 149. 164. 177. 189. 200. 211. 221. 231. 241. 250. 259. 267. 276. 284. 292. 299. 306. 314. 320. 327. 334. 341. 347. 354. 360. 366. 395. 423. 473. 517. 4s/z
: 1510 1915 1850 2345 2135 2710 2385 3025 2615 3315 2825 3580 3020 3830 3200 4055 3375 4275 3540 4480 3700 4685 3850 4875 4000 5060 4135 5240 4270 5410 4400 5575 4530 5740 4655 5900 4775 6050 4890 6200 5010 6350 5120 6490 5225 6630 5340 6760 5450 6900 5550 7030 5650 7160 5740 7280 5850 7410 6310 8000 6750 8550 7550 9570 ~ 8260 , 10480
1% I
120 147 169 189 207 224 239 253 267 280 293 305 317 327 338 349 359 369 378 388 397 406 414 423 432 439 448 455 463 500 535 598 655 5 2365 2890 3340 3730 4090 4415 4725 5000 5280 5530 5790 6020 6250 6475 6690 6890 7090 7290 7470 7650 7840 8010 8180 8350 8530 8680 8850 8990 91 50 9890 10580 11820 12940
1 s~ I
148 181 209 234 256 277 296 313 330 346 382 376 391 404 418 431 443 456 467 479 490 501 512 522 533 543 553 562 572 618 660 739 808 51/z 2855 3490 4040 4510 4940 5340 5710 6050 6380 6690 6980 7270 7560 7820 8080 8320 8560 8800 9030 9250 9470 9680 9900 10100 10300 10490 10890 10880 11070 11940 12770 14290 15620
1% I
179 219 253 283 309 334 357 379 399 418 438 455 473 489 505 521 536 551 565 579 593 606 619 632 645 656 668 680 692 747 799 894 977 6 3405 4165 4810 5380 5895 6370 6810 7210 7600 7970 8330 8670 9000 9320 9630 9920 10210 10500 10770 11020 11300 11550 11800 12030 12290 12510 12730 12980 13200 14250 15220 17020 18610
N o t e : - T h e actual quantities will vary from these'f,'gures, th'e amoun; of variation depending upon the shape of nozzle and size of p,pe ai the point where the pressure is determined. With smooth taper nozzles the actual discharge as about 94% of the figures given in the tables.
F l g . 7A-41
m o s t c o m m o n t y p e s are the r e c t a n g u l a r c o n t r a c t e d w e i r and the 90 ~ V - n o t c h weir. T h e s e are s h o w n in Figure 7 A - 4 2 w i t h the a p p l i c a b l e f l o w f o r m u l a s ,
i m p a c t of the f l o w i n g s t r e a m r e a d s static h e a d + velocity head, and the o t h e r reads the static h e a d only (Figure 7A-43). T h e d i f f e r e n c e b e t w e e n the t w o r e a d i n g s is the velocity head. T h e v e l o c i t y a n d the f l o w are then d e t e r m i n e d from the f o l l o w i n g
g.) P i t o t t u b e
well k n o w n f o r m u l a s .
A pitot tube m e a s u r e s fluid velocity. A small tube p l a c e d in the f l o w s t r e a m gives t w o p r e s s u r e readings; o n e receiving the full
V = C
~
w h e r e C is a c o e f f i c i e n t for the m e t e r determ i n e d by calibration, and hv = velocity head,
274 Process Plant Machinery
.....-~,.?,~.:..--.~; -,~-~~--~_ ~
~
~
t
\ ' i ~" ~ ,""~ "'". '"',-,,~~ " "
Oil-Rectangular Weir With Complete End Contractions
( 1-90" V-Notoh Weir
Q(G.P.M.) = 1495 H~/l (B-O.2H)
Q(G.P.M.) = 1140 He/2
B = Hread H Cest W iniFt"tAFb~ d h i n e e t Weir
H = Head in Feet Above Weir
Fig. 7A-42
Total head -1
i~ ~ J~
Static head
Small holes on both sides of outer tube 7 _~J,~ .....
q,,.,i/(((///.,
.......
~.
Weirs. 6. TPL (TOTAL PUMP LENGTH) - The distance from grade to lowest point of pump. Lift below discharge plus head above discharge plus friction losses in discharge line. This is the head for which the customer is responsible and does not include any losses within the pump.
7. RATED PUMP H E A D -
..............
8. C O L U M N A N D DISCHARGE HEAD FRICTION LOSS Head loss in the pump due to friction in the column Fig. 7A-43
Pltot tube
Capacity = Area x Average Velocity
Since the velocity varies across the pipe, it is necessary to obtain a velocity profile to determine the average velocity. This involves some error, but when properly applied a calibrated pitot tube is within plus or minus 2% accuracy.
Vertical Turbine Pumps Turbine Nomenclature
assembly and discharge head. Friction loss is measured in feet and is dependent upon column size, shaft size, setting, and discharge head size. Values given in appropriate charts in Data Section.
9. BOWL HEAD - Total head which the pump bowl assembly will deliver at the rated capacity. This is curve performance. 10. BOWL EFFICIENCY- The efficiency of the bowl unit only. This value is read directly from the performance
curve.
The horsepower required by the bowls only to deliver a specified capacity against bowl head.
11. BOWL H O R S E P O W E R -
BOWL HP = 1. DATUM OR G R A D E - The elevation of the surface from which the pump is supported. 2. STATIC LIQUID L E V E L - T h e vertical distance from grade to the liquid level when no liquid is being drawn from the well or source. The distance between the static liquid level and the liquid level when pumping at required capacity.
3. D R A W D O W N -
4. PUMPING LIQUID L E V E L - The vertical distance from
grade to liquid level when pumping at rated capacity. Pumping liquid level equals static water level plus drawdown.
5. SETTING- The distance from grade to the top of the pump bowl assembly.
Bowl Head • Capacity 3960 • Bowl Efficiency
12. TOTAL PUMP H E A D - R a t e d pump head plus column and discharge head loss. Note: This is new or final bowl head. 13. SHAFT FRICTION L O S S - The horsepower required to
turn the lineshaft in the bearings. These values are given in appropriate table in Data Section.
Sum of bowl horsepower plus shaft loss (and the driver thrust bearing loss under certain conditions).
14. PUMP BRAKE H O R S E P O W E R -
15. TOTAL PUMP EFFICIENCY (WATER TO WATER) - The
efficiency of the complete pump less the driver, with all pump losses taken into account. Efficiency =
Specified Pump Head x Capacity 3960 x Brake Horsepower
Appendix 7A: Centrifugal Pump Fundamentals 275
.
1
GRACe
.....J
(At
J
I
qlm
gliselm~ IPrmmm Although hydraulic balancing reduces impeller thrust, it also decreases efficiency by 1 to 5 points by providing an additional path for liquid recirculation. NOTE:
16. OVERALL EFFICIENCY (WIRE TO WATER)-The efficiency of the pump and motor complete. Overall efficiency = total pump efficiency x motor efficiency. 17. SUBMERGENCE- Distance from liquid level to suction bell.
Vertical Turbine Pumps Calculating Axial Thrust Under normal circumstances Vertical Turbine Pumps have a thrust load acting parallel to the pump shaft. This load is due to unbalanced pressure, dead weight and liquid direction change. Optimum selection of the motor bearing and correct determination of required bowl lateral for deep setting pumps require accurate knowledge of both the magnitude and direction (usually down) of the resultant of these forces. In addition, but with a less significant role, thrust influences shaft H.P. rating and shaft critical speeds. IMPELLER THRUST
Impeller Thrust in the downward direction is due to the unbalanced discharge pressure across the eye area of the impeller. See diagram A. Counteracting thi.~ load is an upward force primarily due to the change in direction of the liquid passing through the impeller. The resultant of these two forces constitutes impeller thrust. Calculating this thrust using a thrust constant (K) will often produce ony an approximate thrust value because a single constant cannot express the upthrust component which varies with capacity. To accurately determine impeller thrust, thrust-capacity curves based on actual tests are required. Such curves now exist for the "A" Line. To determine thrust, the thrust factor "K" is read from the thrust-capacity curve at the required capacity and given RPM. "K" is then multiplied by the Total Pump Head (Final Lab Head) times Specific Gravity of the pumped liquid. If impeller thrust is excessively high, the impeller can usually be hydraulically balanced. This reduces the value of "K". Balancing is achieved by reducing the discharge pressure above the impeller eye by use of balancing holes and rings. See diagram B.
Although hydraulic balancing reduces impeller thrust, it also decreases efficiency by one to five points by providing an additional path for liquid recirculation. Of even greater concern is that should the hydraulic balancing holes become clogged, (unclean fluids, fluids with solid content, intermittent services, etc.), the impeller thrust will increase and possibly cause the driver to fail. Hydraulically balanced impellers cannot be used in applications requiring rubber bowl bearings because the flutes on the inside diameter of the bearings provide an additional path to the top side of the impeller, thus creating an additional down thrust. Hydraulically balanced impellers should be used as a "last resort" for those situations where the pump thrust exceeds the motor thrust bearing capabilities. DEAD WEIGHT
In addition to the impeller force, dead weight (shaft plus impeller weight less the weight of the liquid displaced) acts downward. On pumps with settings less than 50 feet, dead weight may be neglected on all but the most critical applications as it represents only a small part of the total force. On deeper setting pumps, dead weight becomes significant and must be taken into account. NOTE:
We normally only take shaft weight into consideration as dead weight, the reason being that impeller weight less its liquid displacement weight is usually a small part of the total. SHAFT SLEEVES
Finally, there can be an upward force across a head shaft sleeve or mechanical seal sleeve. In the case of can pumps with suction pressure there can be an additional upward force across the impeller shaft area. Again for most applications, these forces are small and can be neglected; however, when there is a danger of upthrusts or when there is high discharge pressure (above 600 psi) or high suction pressure (above 400 psi) these forces should be considered. MOTOR BEARING SIZING
Generally speaking a motor for a normal thrust application has as standard, a bearing adequate for shutoff thrust. When practical, motor bearings rated for shutoff conditions are preferred. For high thrust applications (when shutoff thrust exceeds the standard motor bearing rating) the motor bearing may be sized for the maximum anticipated operating range of the pump. Should the pump operate to the left of this range for a short period of time, anti-fraction bearings such as angular contact or spherical miler can handle the overload. It should
276
Process Plant Machinery
be remembered, however, that bearing life is approximately inversely proportional to the cube of the load. Should the load double, motor bearing life will be cut to 1/8 of its original value. Although down thrust overloading is possible, the pump must never be allowed to operate in a continuous up thrust condition even for a short interval without a special motor bearing equipped to handle it. Such upthrust will fail the motor bearing. CALCULATING MOTOR BEARING LOAD
As previously stated, for short setting non-hydraulic balanced pumps below 50 feet with discharge pressures below 600 psi and can pumps with suction pressures below 100 psi, only impeller thrust need be considered. Under these conditions:
Where:
Motor Bearing Load (Ibs) 7~mp - " KHL x SG
Impeller Thrust (Ibs) K = Thrust factors (Ibs./ft.) HE = Lab Heat (ft.) SG = Specific Gravity
For more demanding applications, the forces which should be considered are impeller thrust plus dead weight minus any sleeve or shaft area force. In equation form:
(3) Shaft Area Force = Shaft area x Suction pressure * Oil Lube shaft does not displace liquid above the pumping water level and therefore has a greater net weight. CALCULATING AXIAL THRUST
Shaft Dia (in) 1 1 1 1 1 2
3/16 1/2 11/16 15/16 3/16
Shaft Dead Wt. (Ibs/ft.) Open Lineshaft
Closed Lineshaft
2.3 3.3 5.3 6.7 8.8 11.2
2.6 3.8 6.0 7.6 10.0 12.8
Shaft Area (in 2)
Sleeve Area (in)
.78 1.1 1.8 2.2 2.9 3.7
1.0 1.1 1.1 1.5 1.8 2.0
THRUST BEARING LOSS
Thrust bearing loss is the loss of horsepower delivered to the pump at the thrust bearings due to thrust. In equation form: LTB = .0075 ( BHP
where:
Motor Bearing Load = Timp + Wt O) - sleeve force (2) - shaft area force ~3~ = "It (1) Wt. = Shaft Dead Wt. x Setting In Ft. (2) Sleeve Force = Sleeve area x Discharge pressure
LTB = Thrust bearing loss (HP) BHP = Brake horsepower Tt = Motor Bearing Load (Lbs.) = Ttmp + W t (1) - - sleeve force (2) - shaft area force (3)
Vertical Turbine Bearing Material Data Material Description
Temp. and S.G. Limits
Remarks
1. Bronze-SAE 660 (Standard) #1104 ASTM-B-584-932 2. Bronze-SAE 64 (Zincless) # 1107 ASTM-B-584-937 3. Carbon Graphite Impregnated with Babbitt
- 5 0 to 250 ~ Min. S.G. of 0.6 - 5 0 to 180 ~ Min. S.G. of 0.6 - 4 5 0 to 300 ~ All Gravities
General purpose material for non-abrasive, neutral pH service. 7% TinH% Lead/3% Zinc/83% Cu. Similar to std. bronze. Used for salt water services. 10% Tin/ 10% Lead/80% Cu. Corrosion resistant material not suitable for abrasive services. Special materials available for severe acid services and for temp. as high as 650 ~. Good for low specific gravity fluids because the carbon is self-lubricatin$. Corrosion resistant except for highly oxidizing solutions. Not suitable for abrasive services. Glass filled Teflon also available. Used on non-abrasive caustic services and some oil products. Avoid water services as beatings can rust to shaft when idle. Test with bronze Bearings. Excellent corrosion resistance to a pH of 2. Good in mildly abrasive services. 80% Lead/3% Tin/17% Antimony. Use in abrasive water services. Beatings must be wet prior to start-up for TPL 50'. Do not use: For oily services, for stuffing box bushing, or with hydraulically balanced impellers. For services that are corrosive, backing material other than Phenolic must be specified. Expensive alternate for abrasive services. Hardfaced surfaces typically in the range of Rc72. Other coatings are chromium oxide, tungsten carbide, colmonoy, etc. Consult factory for pricing and specific recommendation.
4. Teflon 25% Graphite with 75% Teflon 5. Cast Iron ASTM-A-48 CL30 Flash Chrome Coated 6. Lead Babbitt
7. Rubber w/Phenolic backing (Nitrile Butadiene or Neoprene)
8. Hardened Metals: Sprayed on stainless steel shell (Tungsten Carbide)
- 5 0 to 250 OF All Gravities 32 to 180 ~ Min. S.G. of 0.6 32 to 300 ~
32 to 150 ~
All Temperatures All Gravities
APPENDIX 7B
Change of Performance*
Different industries with many different processes will have requirements for the same pump to operate at different capacities and different heads, and to have a different shape of the head-capacity curves. To ideally satisfy these requirements, one should have a variable-speed pump with adjustable vanes in the impellers. But because most of the drivers in the process industries operate at constant speed, and because the adjustable vanes cannot be produced economically, variable pump performance must be achieved by mechanical means without sacrificing efficiency. In order to provide this flexibility at minimum cost, studies were made to change pump performance within a given pump casing. This can be accomplished by varying the impeller design, cutting impellers, changing the running speed, modifying the impeller vane tips, filing the volute cutwater tip, or orificing the pump discharge. Pump users would prefer to use the same casing for a wide variation of pump performance. The pump casing is usually the most costly part of the pump. To replace a pump casing means extensive and costly work on base plate and piping. The prediction of pump performance by modifying parts other than the casing is based largely on experimentation. Many tests have been conducted by the various pump companies in such areas as: 1. 2. 3. 4. 5. 6.
Trimming the pump impellers Removing metal from the tips of impeller vanes at the impeller periphery Removing metal from the volute tip in the pump casing Providing impeller vanes of the same angularity, but different width Providing impellers with different numbers of vanes and different discharge angles Orificing the pump discharge in the pump casing
We will consider each of these means, but before we do so, we should review the so-called laws of affinity relating to centrifugal pumps. These are theoretical laws or rules that apply to the change in performance of a centrifugal pump by a change in the speed of rotation or a change in the impeller diameter of a particular pump. It should always be remembered in using these laws of affinity that they are theoretical and do not always give exact results as compared with tests. However, they are a good guide for predicting the hydraulic performance characteristic of a pump from a known characteristic. A performance change can be obtained by either the speed of rotation or the outside diameter of the impeller. I. Constant impeller diameter A. The capacity varies directly as the speed GPM1
RPM1
GPM2
RPM2
* Source: Goulds Pumps, Inc., Seneca Falls, NY. Adapted by permission. 277
278
Process Plant Machinery
B. The head varies as the square of the speed Head1 = [RPM1 ] 2 Head2
LRPM2
C. The horsepower varies as the cube of the speed BHPI = [RPM1 ] 3 BHP2
LRPM2
II. Constant speed A. The capacity varies directly with the impeller diameter. B. The head varies as the square of the impeller diameter. C. The horsepower varies as the cube of the impeller diameter. These relationships can be expressed in a simple formula:
Imp,lerioeter, Impeller diameter2
IMPELLER
[ ea l ] r"" l ] GPM2
Head2
LBHP2
CUTS
Assuming that the impeller represents a standard design and that the impeller profile is typically of average layout and not specifically designed for high NPSH, pump performance with trimmed impellers will follow the affinity laws as some vane overlap is maintained. To compensate for casting and mechanical imperfections, correction factors are normally applied to the impeller cuts (Figure 7B-1). The efficiency of the cut impellers (within a 25 percent cut) will usually drop about two points at the maximum cut. On high specific speed pumps, the performance of the cut impellers should be determined by shop tests.
r
A
~
8's
FIGURE
7B-1
~
CALCULATED DIA. IN PERCENT OF ORIGINAL DIA CORRECTION FOR IMPELLER DIA CUT
9:5
,oo
Appendix 7B: Change of Performance 279
O1 " Q x ~
Q Q1 B A
1
= = = =
Capacity Normal Capacity After Underfiling Vane Spacing Normal Vane Spacing After "Underfiling"
Metal removed
Origin
"
/
/
/ /
Head-Capacity After "Underfiling"
Vane thickness
"underfiling"
_j - 021
<
tr-
~
HEAD-CAP
100-
/
o
0u. 8 Iz ~ I.U ~n"t u W el
b
Head-Capacity Normal
80-
60-
//
w
40
/'
/
/
!
!
CAPACITY, PERCENT OF NORMAL
1
120
IMPELLER UNDERFILING
FIGURE 7B-2 REMOVING METAL FROM VANE TIPS
The pump performance can be changed by removing metal from the vane tips at the impeller periphery. Removing metal from the underside of the vane is known as underfiling. Removing metal from the working side of the vane is known as overfiling. The effect of overfiling on pump performance is very difficult to predict and to duplicate. This is because filing vanes by hand on the working side changes the discharge angle of the impeller, and nonuniformity exists between each vane. Underfiling, however, is more consistent, more predictable, and easier to apply. Underfiling is most effective at peak efficiency and to the fight of peak efficiency. Also, underfiling will be more effective where vanes are thick and specific speeds are high. Underfiling increases the area at the impeller discharge, thereby increasing the capacity at peak. This increase is directly proportional to the increase in an area due to filing, or it can be said to equal dimension "A" over "B" in Figure 7B-2. The head rating will move to the fight of peak efficiency in a straight line toward the new capacity. With underfiling, the shutoff head does not change; therefore, the change of the performance by impeller underfiling is less effective to the left of peak efficiency. REMOVING METAL FROM THE VOLUTE TIP
Capacity increase in a given pump can also be achieved by trimming of the volute tip in the pump casing. This is illustrated in Figure 7B-3. Removing metal at this point increases the total volute area. The peak efficiency and peak capacity will move to the fight as the square root ratio of the new area divided by the original area. Pump peak efficiency will normally drop one or two points.
! 140
280
Process Plant Machinery
[
= A~-7-~ x O 1
o~
Held-Caplctty Normsl H e a d . ~ l y After "Underflling" Q = Normal Capacity Q~ = Cap~ity After Volute Ch0ppir~ AI - Normal VoluteArea
Vol~e After Chipping
..................... olll .c_
Normal Volute .SI 0 O -
lion
9
~
~
VOLUTE CHIPPING
FIGURE 7B-3
. . . . . .
HEAD -CAP "~.
~.~. ~
io~
~
~.
21V2" D I A 21V2" D I A 21V=" D I A 21~" DIA
"~"" ~, ~. "~.
~.
100
r9
o
~ .
High Capacity Impeller 27"2~'4" " B A " Std. C a p a c i t y Impeller 2 " " B A " Meal. Capacity I m p e l l e r ISA6" " B A " Low Capacity Impeller I " " B A "
//" ii/
/1,1"
....
~ i
-. ""
I
$1,,"
I.'/y" 0
QAL,ONSPEwM,N FIGURE 7B-4 L O W - A N D H I G H - C A P A C I T Y IMPELLERS
In the majority of pump casings, we can install impellers of different widths for low- or high-capacity performance. Because of the variations in the design of the impeller vanes (angularity and number of vanes), it is very difficult to predict their performance. However, if we take a given impeller with a given angularity and
Appendix 7B: Change of Performance 281
i 3t I
y / /
/
//J /
/
J /
Ns
RPM
HEAD. "AI CAP. TAKEN -N AT PEAK EFF.
FIGURE 7B-5
number of vanes, we can reasonably predict the performance of the narrow, medium, and wide impellers. Figure 7B-4 shows actual test data of a two-stage fourteeninch pipeline pump with a specific speed of 1600. In this pump, the peripheral width of the normal impeller was two inches, whereas the high, medium, and low capacity impellers were 2 3/4 inches, 1 15/16 inches, and 1 inch wide, respectively. Capacities ranged from 5000 GPM to 9000 GPM, and efficiencies bracketed 82 percent to 88 percent. The performance of the different impellers in the same casing is to some degree related to specific speed and running speed. Slow running speed pumps respond better to low-capacity impellers; also, pumps of higher specific speed respond with higher efficiency to a low-capacity impeller. Figure 7B-4 shows the performance of different impeller widths and Figure 7B-5 shows the loss of efficiency for different specific speeds. IMPELLERS OF DIFFERENT NUMBER OF VANES
Certain pump applications require the pump performance curves to have differently shaped head capacity curves. For instance, to overcome friction only, as in pipeline service, the highest head per stage, or a very flat curve, is desirable. To overcome static head or to have pumps run in parallel as is customary in process or boiler feed services, a continuously rising or steep curve must be developed. A medium rising head capacity curve is usually needed for highest possible efficiency. There are two ways to vary the shape of a head capacity curve: 1. If the existing impeller has six or more vanes, removing some of the vanes and equally spacing the remaining vanes will produce a steeper head capacity curve.
282
Process Plant Machinery
HEAD.CAP
~r
)0-
,~
~" ~ .
BO-
60-
~
~
7 VANES
.....
6 VANES
.....
5 VANES
......
4 VANES
40-
20!
0
20
40
60
80
!
i
100
120
140
CAPACITY. PERCENT OF NORMAL
FIGURE 7B-6
4" PUMP-9" IMP DIA:Ns.1600 3550 RPM
120 -
?V "~70
HEAD-CAP
100-
80-
40-
o
2'o
2o
6~
;o
,;o
CAPACITY, PERCENT OF NORMAL
FIGURE 7B-7
,~o
,;o
Appendix 7B: Change of Performance 283
2" PUMP 7" DIA.-3550 RPM HEAD-CAP
170-1 ORIF.
160-
"~
;,80R,F.
518 ORIF,
PEAK EFF, LINE
:=
150140--
.70
,.n
60
130-
25-
50
120
20.
40
15-
30
lO
20
=I i
B H.P SP GR 10
,,~-t1 100
ii e
9~t
8O
4b
70
8'o
1~o
1~o 2;o GALLONSPER MIN.
FIGURE
2;0
2;o
3;~o
7B-8
2" PUMP-- 7" DIA. ~ 3550 RPM. I CHANGE,NoERMFOLRMANooCE 0/DUETTO.CE,ANGEIN ORIFICEONLY.
-tO0 -
90-~
-
70~
-~ I -~ -~ -
100 Q
80
3 v A N E IMP. ,,, . . . . . . . . . . . . . . . =,, ..,.,,..- ,,,,-,,-. "- "" "=" " " " " "=" " " 5 v A N E |MP.
~
HEAD
rr"
.,.j Q:
Q.
40
60
120 11/41 t
40
0.8~ 1 "
o~,~,c~-~ .:~~
40 2O
0.6
0.41_ 02
0 10
20
30
40
50
60
PERCENTOF NORMALCAP FIGURE
7B-9
70
80
90
100
%"
%"
o._.~
284
Process Plant Machinery
0
The fewer the number of vanes, the steeper the curve. When vanes are removed, the total discharge area of the impeller is reduced and the peak efficiency point will move to the left, as shown in Figure 7B-6. The efficiency will also drop, the lowest efficiency occurring at the least number of vanes. In a seven-vane impeller reduced to four vanes, the efficiency will drop about four points. If a different head capacity curve shape is required in a given casing and the same peak capacity must be maintained, a new impeller must be designed for each head capacity shape. The steeper the head capacity curve, the wider the impeller, and the fewer will be the number of vanes. For example, a seven-vane 27-degree exit angle will have a flat curve and a narrow impeller, whereas a three-vane 15-degree exit angle will have a steep curve and the widest impeller (refer to Figure 7B-7). In other words, to peak at the same capacity, the impeller discharge area must be the same, regardless of head capacity relationships. Also, for a given impeller diameter, the head coefficient will be the highest for the flattest curve. The efficiency of the above impeller can be maintained within one point.
ORIFICING PUMP DISCHARGE
In low specific speed pumps, where impellers are already very narrow and low capacity or narrower impellers cannot be cast, capacity reductions can be obtained by using restriction orifices in the pump discharge nozzle. Figures 7B-8 and 7B-9 illustrate these points. Figure 7B-8 shows the performance of a two-inch pump where the discharge was throttled with different size orifices. Figure 7B-9 shows the predicted performance of a throttled pump and illustrates how orifice size selection and changes in the number of impeller vanes can influence absorbed power and head developed by the pump.
APPENDIX 7C ,,
Reed Frequency Considerations for Vertical Pump Installations*
A vertical motor/pump system has experienced high vibration from the original installation, with efforts by others to reduce it unsuccessful. Often, such systems have a "Reed Frequency," or first cantilever bending mode of the above-ground system, near operating speed. Experimental vibration analysis techniques did not show a mode of the structure near operating frequency. A finite element modal analysis yielded similar results. Calculated composite support stiffness of the system structure and oil film showed the system to have a level of rotor support which led to operation near a critical speed. Modifications to the motor bearing and internal bearing supports by themselves were calculated to have a small effect. The pump head, upon which the motor is mounted, was stiffened to give increased support to the motor rotor. Upon installation, the system vibration was greatly reduced, well below company and industry standards. The field study shows, although
a system be designed for "Reed Frequency" separation, rotor support stiffness can still be inadequate. Careful attention must be given by pump manufacturers to both "Reed Frequencies" and system support stiffness to yield a suitable system.
BACKGROUND The subject machine is a 2,000 hp, 900 rpm (15 Hz) vertical motor driving a vertical water pump. This site has three Dresser-Rand Electric Machinery Division motors and one motor from a competitor. See Figure 7C-1. Vibration of the system had been an issue for a long time, reportedly from the initial startup. Several attempts to lower the vibration have been made, including foundation rework and adding dynamic vibration absorbers (DVAs) to the top of
Flg. 7C-1 Slte photograph. (Source: Dresser-Rand, Olean, NY) * Source: Dresser-Rand,Olean, NY (Paper by Morrisonand Shively). Adaptedwith permission
286 Process Plant Machinery the motor (see Unit 18, Figure 7C-1). A refurbishment of the stator and replacement rotor for Unit 17 was done to obtain a problem free system and address the vibration concerns. Dresser-Rand Electric Machinery Division was requested to examine the complete pump/motor system to determine the root cause of the excess vibration. Dresser-Rand Electric Machinery Division performed dynamic testing on site in March, 1995 on the sister units to help determine other possible vibration improvement areas. During this testing, two Operating Deflection Shapes (ODSs) of the above ground structure on Unit 18 were taken to determine vibration shape while running. The first ODS contained points along the lateral centerline on the outside surface of the machine with defining points for the pumphead openings on the sides. Figure 7C-2 through Figure 7C-4 show results of the vibration shape superimposed on the undeformed model shape. The ODS results showed large deflections in the pumphead at the side openings while the motor showed very little deflection, indicating that the primary flexibility in the system was related to the pumphead. Whether this result was true in both
directions needed to be answered with a second ODS performed while on site. The structural stiffness is very different aligned with the discharge pipe and at 90 ~ to the discharge. A second ODS was performed with the reference transducer rotated 90 ~ 1Figure 7C-5 shows the results 90 ~ from the first ODS, in line with the discharge pipe. The second ODS agreed with the first, showing the primary flexibility in the system occurring at the pump head side openings. Running tests for March, 1995, also included runup and coastdown data for Units 18 and 19. See Figure 7C-6 and Figure 7C-7. A transducer was placed on the top of the machines, measuring lateral accelerations. These measurements were useful to show that natural frequencies exist close to the operation speed. This shows that the system should be stiffened to detune it. Since these units were duplicates to Unit 17, it was assumed that it would have similar dynamic behaviour if left unchanged. In addition, it was noted that the foundation grouting was cracked and breaking up in certain places. Finally, many layers
. . ~ ~I~~'........~.i, ~ \
~
I
~ I
~t
UndeformedShape1 DeformedShapei:
~ - ~ Topof PumpHead/Bottomof MotorI - I
,,,
I NotetheFoundationMovement.] Flg. 7 C - 2 0 D S , Unlt 18, Isometrlc vlew. (Source: Dresser-Rand, Olean, NY) 1 The modal analysis software used is designed to determine "normal" modes, and looks at the responses at 90~ to the reference signal. The transducer must be rotated 90~ to see vibration in the other plane.
Appendix 7C: Reed Frequency Considerations for Vertical Pump Installations 287
~~ L i ...........
~~I
.....
1
Deformeci"Sha-pc]
+ . . . . + . . ~ / / /
..............................
' l
~Topof PumpHead/Bottomof Motor +
~--.+L
"
/~
+
........... ~+?a+,2L,+,. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
J
~x~
~N0te collapseo'i Side0pening] ....
~,,,~,%,\..+ ....
+I +'
!1 ,._
UndeformedShapeI
\
"X
Note the FoundationMovement]
Fig. 7 C - 3 0 D S of Unlt 18, slde vlew. (Source: Dresser-Rand, Olean, NY)
l Def_ormed ShaPe ]
r
,~.
,. ~
1
....... ght] t Note how the Motor Remains S!ro!gh "
Iop of PumpHead/Bottomof Motor1 ~ ~:~...." I
I Note Collapse of S,~-deopen,--ng-]
'?+ I ,;>
_I~-"1 Note the Flexibility of the Pumpheadj
ii I
Flg. 7 C - 4 0 D S of Unlt 18, end vlew. (Source: Dresser-Rand, Olean, NY)
of paint behaved as a "stress coat," flaking off in areas of greatest deflection: notably the periphery of the pumphead near the cutouts, where the ODS showed buckling.
structure, and a rotor lateral critical speed near operating speed.
There are at least two possible "above ground" sources of vibration excitation: a Reed natural frequency of the upper
Due to past experience and coastdown and runup data, it was initially believed that the above ground structure had a "Reed
REED FREQUENCY
288 Process Plant Machinery
Flg. 7C-5
2nd ODS of Unlt 18, slde vlew. (Source: Dresser-Rand, Olean, NY)
Frequency" at or near operating speed. A "Reed Frequency" is the first bending frequency for a cantilever beam. Vertical equipment, with cantilever mounting at the base, tends to have first bending or "Reed Frequencies" rather low, often near operating speed. To examine this case, a finite element modal analysis of the complete motor, pumphead, and pump assembly was conducted. Plate elements were used for the frame structure, while brick elements were used for the stator core components. Horizontal support stiffness was held constant at 10 x 106 Ib/in, while the vertical support was varied from 1 x 106 to 10 x 106 Ib/in. The rotor mass was not included since it is not dynamically coupled during rotation. Analyses showed the below ground structure behaving relatively independently from the above ground structure; minimizing its effect on the upper structure. Therefore, the mass and possible stiffening effects of the 300 psi water column were not used in the analysis. See Table 7C.1 for tabular results. The deformed shape correlated best with the ODS at 5 x 106 Ib/in vertical stiffness.
The results in Table 7C.2 show the effect of (12) 1 in. thick axial ribs welded around the periphery of the pumphead. Little change was made to the below-ground modes, while the above-ground modes were greatly impacted. The changes are likely due to the confinement of buckling, added support of the motor mounting flange on the pumphead, and general increased bending stiffness.
ROTOR/SYSTEM CRITICAL SPEED The rotor critical speed calculation changes significantly for vertical applications. Because gravity does not pull the rotor weight into the guide bearing oil films, the oil film stiffness tends to be much lower than typical horizontal applications. In addition, the bearing coefficients are more difficult to determine, since iterations with the calculated unbalance and magnetic pull forces are needed to determine the proper equilibrium position. The bearing coefficients are then determined from the equilibrium position.
These results were surprising, in that they showed the classic above-ground "Reed Frequency" significantly above the 900 rpm operating speed, at 22 Hz (1,320 cpm). See Figure 7C-8. This did not explain the response peaks found during the coastdown and runup being caused by a structural natural frequency.
The complete bearing support stiffness is made up of the following "springs" in series: the foundation, the frame/bearing structure, and the oil film. Previous methods used an experience-based composite stiffness for the bearing stiffness. The rotor critical speed was then calculated based on that composite stiffness. This works well if the structure of the pumphead and foundation don't stray far from the average geometry.
The lower pump assembly natural frequencies were also found to be well away from operating speed with the second bending mode at about 12 Hz. See Figure 7C-9. Various combinations of foundation stiffness conditions were used to show that structural resonance by itself was not responsible for the high vibration.
In this case, a FEA model was used to determine the structural stiffness at the upper and lower bearings to better estimate the rotor critical speed, shown in Table 7C.3. Models were made with: (1) original bearing/structural configuration; (2) modified bearing and bearing supports; and (3) modified bearing and bearing supports with pumphead ribs. The stiffnesses were,
Appendix 7C: Reed Frequency Considerations for Vertical Pump Installations 289
Fig. 7C-6 Runup and coastdown data, Unit 18. (Source: Dresser-Rand, Olean, NY) in some cases, considerably lower than the experience based composite stiffnesses. Figure 7C-10 shows the critical speed map with the experience based stiffness, and for the composite oil film and FEA-based stiffnesses from Table 7C.3. Note that changes to the motor itself had small effects, and that the great increase in support stiffness provided by the addition of pumphead ribs moved the system critical speed sufficiently away from operating frequency. Unbalanced response (weights in phase) at the lower bearing is shown in Figure 7C-11 and Figure 7C-12 for cases 1 and 3, respectively. In both cases, the rotor lateral critical is calculated sufficiently above operating frequency, with the multiple peaks occurring due to directional stiffness of the sleeve bearings and the difference in support stiffness at the upper and lower bearings.
Fig. 7C-7 Runup and coastdown data, Unit 19. (Source: Dresser-Rand, Olean, NY) angle were greatest in the pumphead and least in the motor, so stiffening of the motor external structure would not be very effective. Also, adding weight to detune the motor would move the "Read Frequency" closer to operating speed. For the vertical motor application, composite support stiffness is inherently low, and so increases in stiffness are advantageous. It is also possible the vibration received some contribution from a rotor critical near operating speed because of the inherently low oil film and structural support stiffnesses in a vertical application. The bearings and upper bearing support were upgraded to current design philosophies, which had changed significantly since the units were originally built. Changes to the rotor itself were calculated to be relatively ineffective for critical speed improvement.
MOTOR CHANGES
PUMPHEAD CHANGES
To reduce the potential vibration for this unit's geometry, it is desirable to stiffen the "above ground" system as much as possible. The field data showed that changes in deflection
The customer and pump manufacturer requested recommendations for possible improvements for the rest of the equipment. The foundation interface was already being improved
TABLE 7C.1 System Natural Frequencies as Function of Vertical Support Stiffness Mode
1st Bending, pump 2nd Bending, pump 1st Bending, top Discharge Direction 1st Bending, top Cutout Direction
* 1 x 106 Ib/in Vertical Support, Hz
* 5 x 10 6 Ib/in Vertical Support, Hz
* 10 x 106 Ib/in Vertical Support, Hz
1.3 11.2 16.0
1.3 11.8 21.4
1.3 11.9 22.9
16.5
22.3
23.8
*Models has 10 x 106 Ib/in support stiffness in both horizontal directions.
290 Process Plant Machinery
Fig. 7C-9 Second bending mode of below-ground components, 11.8 Hz. (Source: Dresser-Rand, Olean, NY)
Fig. 7C-8 First bending mode of above-ground components, 22.3 Hz
outer structural components at the cutout, and welded channel axial ribs on the periphery.
as much as possible through regrouting. It was suggested that stiffening of the pumphead be designed to eliminate the weaknesses produced by the large side cutouts on each side, shown in Figure 7C-13. These were shown in the ODS, FEA, and cracked thick paint acting like "stress coat" to buckle at the sides of the openings, as well as at the back of the pumphead, opposite the discharge pipe.
The net effects of modernizing the bearings and bearing support, and stiffening the pumphead are to enhance the rotor support and to increase the stiffness of the "above ground" structure. Accordingly, any "Reed Frequencies" and/or rotor lateral critical speeds will be further removed from operating frequency without bringing the below-ground natural frequencies closer to operating frequency.
Through discussions, the pumphead design was finalized as shown in Figure 7C-14, with plates connecting the inner and
TABLE 7C.2 System Natural Frequencies Before and after Addition of Pumphead Ribs Mode
Change, Percent
Original Pumphead, Hz
Pumphead with Ribs, Hz
1.3 11.8 21.4
1.4 12.0 25.8
+7.7 +7.7 +20.6
22.3
27.1
+21.5
1st Bending, pump 2nd Bending, pump 1st Bending, top Discharge Direction 1st Bending, top Cutout Direction
Models had 10 x 10s Ib/in horizontal, and 5 x 10s Ib/in vertical support stiffness.
TABLE 7C.3 Structural Support Stiffness of Above-Ground System. Location Top Bearing Bottom Bearing
* No Ribs, Original Configuration lb./in,
* No Ribs, New Upper Bearing Support, lb./in,
* Pumphead Ribs, New Upper Bearing Support, lb./in.
398,000 1,003,000
425,500 1,003,000
646,000 2,106,000
*Average of cutout and pipe discharge directions with 10 x 106 Ib/in horizontal, and 5 x 10s Ib/in vertical support stiffnesses.
Appendix 7C: Reed Frequency Considerations for Vertical Pump Installations 291
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292 Process Plant Machinery 25.0
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Fig. 7C-13 Pumphead photograph
Appendix 7C: Reed Frequency Considerations for Vertical Pump Installations 293
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POST-MODIFICATION TESTING These modifications proved to be quite successful. The vibration of Unit 17 was tested in August, 1995 at .07 in/s (0-peak) at the top of the motor, an estimated reduction of 70%. This level is very low compared with similar motors and to previous experience with these units. Dynamic testing was performed following installation of the motor onto the pumphead. Figure 7C-15 shows a plot of a coastdown from 900 rpm to 0 rpm. No significant natural frequencies are apparent in the speed range. Figure 7C-16 shows a plot of a runup from 0 rpm to 900 rpm. Note the steepness of the curve at 900 rpm. This is an indicator of how quickly the unit comes up to speed in the coupled condition. The uncoupled runup was so rapid as to make reliable measurements difficult to obtain. On the previous trip, Frequency Response Functions (FRFs) were not obtained. Those measurement problems were overcome on the following trip and driving point measurements were taken for the three Dresser-Rand Electric Machinery Division machines at standstill. Figure 7C-17 through Figure 7C-19 show the results, summarized in Table 7C.4.
MODAL ANALYSIS During the August, 1995 trip, a modal analysis of the entire refurbished unit (Unit 17), including both "above ground" and "below ground" structures, was performed while at standstill. A 2 Hydraulic Institute Standards, 13th Edition.
J single line of points down the left side of the machine (viewed from the discharge pipe) was used to define the structure. The results are shown in Figure 7C-20 through Figure 7C-23. The results do not correlate well with the FEA results due to unquantifiable coupling of the shaft mass and stiffness with the support structure while at standstill.
SUMMARY OF RESULTS The foundation interface integrity was restored, the bearings modernized, and the pumphead structure and motor bearing support was stiffened. These changes resulted in the upper structure first mode and the lower structure second bending mode being well away from operating speed, minimizing dynamics related vibration sources. A secondary effect of the changes was the increase of the system critical speed away from operating speed, further reducing dynamic amplification of the system. The post-modification testing results verify the modification steps taken within the motor and on the pumphead were very successful. Site personnel remembered Unit 17 having vibration performance similar to Unit 18 prior to the modifications suggesting roughly a factor of 3.5 reduction in vibration. Based on this apparent reduction in vibration levels, the modifications were quite successful in improving this pump system. If Unit 18 and 19 foundations are brought to an equivalent condition to Unit 17, these same modifications should also be effective for improving their vibration performance. This was confirmed recently when both Units 18 and 19 underwent similar refurbishment, resulting in vibration levels like that of Unit 17.
294 Process Plant Machinery
Fig. 7C-15 Coastdown data for Unit 17. (Source: Dresser-Rand, Olean, NY)
Fig. 7C-16 Runup data for Unit 17. (Source: Dresser-Rand, Olean, NY) There are still areas of the system that can be investigated in the future, such as the effect of the 300 psi water column, effects of rotation on system natural frequencies, field determination of rotor critical speed location when above operating speed, determining actual foundation stiffness in the field, and obtaining modal analysis calibration to FEA models. This case study shows the importance of a complete system approach to the design of vertical applications. Dynamics related vibration has long been a problem for these applications. Diagnosis of the root cause of high vibration has been very difficult prior to the development of FEA, experimental modal analysis equipment and new methods, such as the ODS, and their use. The advent of these new engineering tools
allows a much clearer understanding of the factors influencing the overall design success. Since the motor is generally placed on the top of the structure, where vibration is usually maximum, the motor is generally considered the root cause. As this case study shows, however, these systems are not that simple or vibration causes that obvious. It demonstrates that significant effort is needed to determine potential vibration causes once the system is in operation, as a structure designed solely to avoid structural "Reed Frequencies" can lack sufficient rotor support stiffness. It also shows that much design work is needed before the system is built to avoid all dynamics-related problems.
Appendix 7C: Reed Frequency Considerationsfor Vertical Pump Installations 295
Fig. 7C-17
Driving point FRF at top of Unit 17. (Source: Dresser-Rand, Olean, NY)
Fig. 7C-18
Driving point FRF at top of Unit 18. (Source: Dresser-Rand, Olean, NY)
Dynamics information is included on Dersser-Rand Electric Machinery Division vertical motor outline drawings to assist the pump manufacturer in natural frequency avoidance. In the future, it will also be advantageous for motor manufacturers to specify the required support stiffness at the top of the pumphead for acceptable rotor/system performance. It is hoped the analysis and diagnostic methods used in this case study shed light on vertical applications and the design challenges they represent.
NOMENCLATURE Bearing Coefficients: Values used for rotordynamic analysis which describe the behavior of fluid-film bearings during operation.
Coast Down Test: Allowing the rotor to coast down with the power off, observing the changes in dynamic behavior. Critical Speed Map: A plot showing the relationship of rotor lateral critical speeds as a function of composite support stiffness.
Detune: Modifying a structure to remove a natural frequency from a forcing frequency, thereby decreasing vibration. Driving Point Measurement: A frequency response function in which the impulse and response are at the same geometric location.
296 Process Plant Machinery
Fig. 7C-19 Driving point FRF at top of Unit 19. (Source: Dresser-Rand, Olean, NY) TABLE 7C.4 Modal Peaks and Operating Vibration at 900 rpm.
Unit
Frequency Response P e a k s (cpm)
17 18 19
997.5, 1030, 1100 862.5, 1040 915
Measured Operating Vibration Amplitude (In.ISec.)
Acceptable Industry Standard Limits2 (In.ISec.)
0.07** 0.24* 0.09**
0.23 0.23 0.23
* Estimated from previous coast down data. **Measured with site transducer.
Fig. 7C-20 Mode at 9.625 Hz (577 rpm, Unit 17). (Source: Dreseer-Rand, Olean, NY) DVA: Dynamic Vibration Absorber. A relatively small springmass system usually mounted to a vibrating machine to reduce vibration amplitude. The system is tuned to the natural frequency of the machine, and splits the system natural frequencies into two new frequencies different than the original natural frequency. This modification often decreases the vibration level.
Dynamics: In this paper, the term refers to vibration related characteristics of a structure. Finite Element Analysis (FEA): Computer analysis technique which divides complex structures into a system of simple geometric elements.
Appendix 7C: Reed Frequency Considerations for Vertical Pump Installations 297
Fig. 7C-21 Mode at 16.625 Hz (998 rpm, Unit 17). (Source: Dresser-Rand, Olean, NY)
Fig. 7C-22 Mode at 17.25 Hz (1035 rpm, Unit 17). (Source: Dresser-Rand, Olean, NY) Frequency Response Function: A measure of the structural
Operating Deflection Shape (ODS): A computer method of vibration analysis which produces an animation showing how a structure vibrates during operation.
quency of a shaft.
Lateral Critical Speed: Transverse (bending) natural fre-
Reed Frequency: The first bending natural frequency of a cantilever structure: It deflects in a reed shape.
Magnetic Pull: The attractive force created between the rotor and stator in motors and generators.
Rotordynamics: The engineering discipline dealing with the behavior of rotating machinery.
MiI: 1/1000th of an inch (.001").
Run up Test: Increasing the rotor speed and observing the changes in dynamic behavior.
response to a standard excitation, measured in units of acceleration per unit force over a frequency range.
Modal Analysis: Experimental and analytical computer tech-
niques for determining natural frequencies and mode shapes.
Normal Modes: Definition of resonance: motion lags force by 90 degrees, hence the term normal.
Stress Coat: A paint-like film which, when applied to a structure, displays local surface patterns corresponding to the strain in that area.
298 Process Plant Machinery
Flg. 7C-23 Mode at 18.375 Hz (1102.5 rpm, Unlt 17). (Source: Dresser-Rand, Olean, NY) Support Stiffness: A measurement of the equivalent spring stiffness supporting a rotor which may include the foundation, machine brackets, bearing oil film, etc.
Unbalance: Measurement of the moment remaining in a rotor after balance. The unbalance develops radial forces under operation. Unbalanced Response Analysls: Rotordynamic analysis which calculates the vibration characteristics of a rotor at a given operating speed.
REFERENCES 1. Ewins, D.J., Modal Testing: Theory and Practice. New York, John Wiley and Sons, 1984. 2. Vierck, R.K., Vibrations Analysis. Scranton, International Textbook Co., 1967.
APPENDIX 7D
Vertical Mixed Flow Variable Pitch Vane Pump*
The KPV-type vertical pumps have the advantages listed below (that come from the variable pitched vanes) over fixed vane models: 1. High efficiency in partial load (low flow operation) saves power consumption. 2. Wide flow control range enables operation of 0 ~, 130% capacity. 3. Vane adjustment during operation is possible depending upon fluid flow rate. 4. Starting up with vanes totally closed minimizes starting torque. 5. Operation with the vane adjustment for the flow rate eliminates vibration or cavitation of the pump. 6. Operation of the discharge valve is not necessary for control of the flow rate, which insures longer service life of the valve. This pump reduces power consumption which becomes larger as a result of the recent scale-up in various plants. Some "jumbo" pumps have delivered 130,000 m3/h, at 5,1 O0 kW and with 4,000 mm bore. Power cost savings of more than 20% have been attained in plants where variable pitched vane type pumps have been employed (compared with the cases of fixed vane types). The vane control mechanism is available in both mechanical and oil hydraulic systems, either of which is operated by the manual control provided at the unit side and by the central remote control; flow rate control is achieved with ease by manipulating a single switch.
Flg. 7D-1 Sectlonal vlew of KPV-type pump. (Source: Mltsublshl Heavy Industrles, Tokyo) * Source: Mitsubishi Heavy Industries, Tokyo
300 Process Plant Machinery
Fig. 7D-2 Performance curve of KPV type pumps. (Source: Mitsubishi Heavy Industries, Tokyo)
Appendix 7D: Vertical Mixed Flow Variable Pitch Vane Pump 301
Fig. 7D-3 Vane closed. (Source: Mitsubishi Heavy Industries, Tokyo)
Fig. 7D-4 Vane fully opened. (Source: Mitsubishi Heavy Industries, Tokyo)
302 Process Plant Machinery
Fig. 7D-5 Vane control system. (Source: Mltsublshl Heavy Industries, Tokyo)
Fig. 7D-6 Model operating ranges. (Source: Mitsubishi Heavy Industries, Tokyo)
APPENDIX 7E
Rotating Case Design* Rotating case designs fewer moving parts and lowered required spares inventory. The wide operating range means it can be used in a variety of process industries (see Table 7.1 ).
Fig. 7E-1
* Source: Enviro Pumpsystems
Basic Roto-jet design. (Source: Enviro Tech Pumpsystems, Salt Lake City.)
304 Process Plant Machinery
Flg. 7E-2 Sectlon (model RG). (Source: Envlro Tech Pumpsystems, Salt Lake Clty.)
Flg. 7E-30peratlng envelope (PG model). (Source: Envlro Tech Pumpsystems, Salt Lake Clty.)
Appendix 7E: Rotating Case Design 305
Flg. 7E-40peratlng curves (Rll) (Source: Envlro Tech Pumpsystems, Salt Lake Clty.)
Flg. 7E-50peratlng curves (RO model). (Source: Envlro Tech Pumpeysteme, Salt Lake Clty.)
306 Process Plant Machinery TABLE 7E.1 Specification (RG Model) RG
Specification Maximum Temperature (With flush) Maximum Suction Pressure Maximum Head Maximum Speed Maximum Flow Maximum Horsepower Weight
Materials of Construction
180 oF 250 oF 200 PSI 2600 Ft. 4380 RPM 400 GPM 400 HP 870 Ibs.
82 ~ 121 ~ 14 Bar 792 m 4380 RPM 91 m3hr 300 KW 395 KG
Standard
Rotor Rotor Cover Manifold Endbell Pick-up Tube Shaft
St. Steel
Ductile Iron Ductile Iron Ductile Iron Steel 17-4 PH AISI 4140
316 St. Steel 316 St. Steel 316 St. Steel Steel 17-4 PH AISI 4140
TABLE 7E.2 Specification (Rll Model) Specification
R 11
Maximum Temperature (With flush) Maximum Suction Pressure Maximum Head Maximum Speed Maximum Flow Maximum Horsepower Weight
Materials of Construction
180 oF 250 oF 200 PSI 1500 Ft. 4858 RPM 150 GPM 75 HP 350 Ibs.
Standard
Rotor Rotor Cover Manifold Endbell Pick-up Tube Shaft
Ductile Iron Ductile Iron Ductile Iron Ductile Iron 17-4 PH AISI 4140
82 ~ 121 oC 14 Bar 457 m 4858 RPM 34 m3/hr 55 KW 159 KG
St. Steel
316 St. Steel 316 St. Steel 316 St. Steel 316 St. Steel 17-4 PH AISI 4140
TABLE 7E.3 Specification (RO Model)
Maximum Temperature (With flush) Maximum Suction Pressure Maximum Head Maximum Speed Maximum Flow Maximum Horsepower Weight
Materials of Construction Rotor Rotor Cover Manifold Endbell Pick-up Tube Shaft
High Speed OpL
RO
Specification 250 ~ 550 ~ 400 PSI 2600 Ft. 4380 RPM 400 GPM 400 HP 1200 Ibs.
Steel
Carbon St. Carbon St. Carbon St. Carbon St. 17-4 PH AISI 4140
121 ~ 288 ~ 28 Bar 792 m 4380 RPM 91 m3/hr 300 KW 544 KG
St. Steel
316 St. Steel 316 St. Steel 316 St. Steel Carbon St. 17-4 PH AlSl 4140
250 oF 550 ~ 300 PSI 5200 Ft. 6321 RPM 400 GPM 400 HP 1200 Ibs.
Steel
CA-6NM CA-6NM Carbon St. Carbon St. 17-4 PH AISI 4140
121 ~ 288 ~ 21 Bar 1585 m 6321 RPM 91 m3/hr 300 KW 544 KG
St. Steel
CD-4MCu CD-4MCu 316 St. Steel Carbon St. 17-4 PH AISI 4140
Appendix 7E: Rotating Case Design 307 Typical applications for the rotating case design include: boiler feed and desuperheating, central cleaning systems, hydraulic systems, hydro-blast cleaning, oil production, mining,
petroleum-chemical, pulp and paper mills, reverse osmosis, semi-conductor manufacturing, spraying systems, steel mills, transfer and water injection.
This Page Intentionally Left Blank
Chapter 8 Positive Displacement Pumps Positive displacement pumps can be divided into two major categories: reciprocating and rotating. Reciprocating pumps include steam pumps and power pumps, as defined later. Many reciprocating pumps use a flexible membrane or diaphragm and are collectively called diaphragm pumps. Every one of the various types comes in a wide range of sizes, or with modifications, additions, and perhaps auxiliary support equipment. The same is true for the many different types of rotating positive displacement pumps. They include gear pumps, screw pumps, and peristaltic pumps, to name just a few. Each pump category, reciprocating and rotating, can be found in virtually every process plant we would typically encounter in the industrialized world. Not surprising, each has a definite application range, and the vast majority of these application ranges overlap each other.
RECIPROCATING POSITIVE DISPLACEMENT PUMPS*
Reciprocating positive displacement pumps incorporate a plunger or piston that displaces, or feeds forward, a given volume of fluid per stroke. The basic principle of a reciprocating positive displacement pump is that moving a solid component into the space occupied by a liquid will result in an equal volume of liquid being moved out of that space. To better understand reciprocating positive displacement pumps and their subgroup metering pumps, we must investigate their place within the universe of pumps. The pump universe could be organized in a variety of ways, such as by design, materials of construction, or the liquids pumped. For the purpose of this discussion, it is appropriate to organize the pump universe by classifying pumps based on the method by which the pump imparts energy to the liquid being pumped. This results in two basic classes of pumps" dynamic and displacement. Dynamic pumps encompass those shown on the left-hand side of Figure 8-1, and these impart energy to the liquid in a steady fashion. Displacement pumps encompass the remaining pumps in Figure 8-1, and these impart energy to the liquid in a pulsating fashion. The usual basic characteristics of the dynamic and displacement pumps are shown in Table 8.1. By examining this table, it is possible to identify the class of pump required for the job from its characteristics. This segment of our text is primarily concerned with the world of metering pumps, which is within the positive displacement, reciprocating class of pumps. * Source: Metering Pump Handbook, Industrial Press, Inc., New York, NY, 1984. Adapted by permission.
309
310
Process Plant Machinery
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FIGURE 8-1 Classification of the pump universe. (Source: Pump Handbook, edited by Igor J. Karassik, William J. Krutzsch, Warrent H. Fraser, and Joseph P. Messina. McGraw Hill, New York, NY, 1985)
Reciprocating pumps can be divided into two general categories: steam pumps and power pumps. Steam Pumps
Steam pumps consist of a liquid and steam cylinder joined together by a spacer cradle (Figure 8-2). These pumps may be steam or air driven. The liquid end consists of liquid inlet and outlet ports, valves, and a piston or plunger. The steam or air end consists of valve mechanisms and pistons. Normally, steam pumps are not designed to allow adjustment of the output flow while operating. This precludes their use as a metering pump in a system requiring frequent adjustment of flow rate.
FIGURE 8-2
OH, 1975)
Steampump. (Source: Hydraulic Institute Standards, 13th Edition, Cleveland,
Positive Displacement Pumps
311
Power Pumps P o w e r p u m p s consist of a liquid end and a p o w e r end (Figure 8-3). T h e s e p u m p s are generally driven by electric motors, air- or g a s o l i n e - d r i v e n motors, or any device imparting a rotary or reciprocating m o t i o n to the p u m p . T h e liquid end consists of inlet and outlet ports, valves, and pistons or plungers. The p o w e r end consists of the frame, crankshaft, bearings, c o n n e c t i n g rods, crossheads, and, s o m e t i m e s , reduction gears.
TABLE 8.1
Basic Characteristics of Modem Pumps Dynamic Rotary Centrifugal
Discharge flow Usual max suction lift (ft) Liquids handled
Usual capacity range How increased head affects: Capacity
Clean, clear; dirty, abrasive; slurries Low to high
Medium
Small to largest available
Small to medium
Decrease
Almost none Increase
Decrease
None
None
Increase
Increase
Increase
Almost none Decrease
Small increase Decrease
None
None
Steady 15
Depends on specific speed
Power input How decreased head affects: Capacity
Steam Power Pulsating Pulsating 22 22 Clean and clear
Increase
Low to highest produced Relatively small
Decrease Depends on specific speed Some External leakage Some to none Some Some Possible with added equipment Volume control Remote pumping chamber Not available Source: Metering Pump Handbook, Industrial Press, New York, 1984. Reprinted by permission. Power input
FIGURE 8-3 OH, 1975)
Metering Piston or Plunger Diaphragm Pulsating 22 Clean, clear; dirty, abrasive, slurries Low' to highest produced Relatively small
Steady 22 Viscous, non-abrasive
Steady] 15
Discharge pressure range
Displacement Reciprocating
Decrease Some 1 None Inherent in design Possible
Power pump. (Source: Hydraulic Institute Standards, 13th Edition, Cleveland,
312
ProcessPlant Machinery FIXED
CONTROLLED VOLUME
ROTARY PNEUMATIC
[ SPEED! [ADJUSTABLE I LINEAR ELECTRICAl, I
!
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Classesof metering pumps. (Source: Hydraulic Institute Standards, 13th Edition, Cleveland, OH,
Falling within this category of pumps are those designed with adjustment of output flow as well as those lacking this feature. Because a metering pump requires the adjustment of output flow in a typical process system, we will limit the remainder of this chapter to those power pumps with adjustable output flow.
Metering Pumps Figure 8-4 subdivides the metering pump class of power pumps into various types. These types delineate the methods and geometries commercially used to produce a metering pump. Metering pumps should be considered, first, as precision instruments used to feed accurately a predetermined volume of liquid into a process or system. Their secondary function is to pump, or move, a liquid from one point to another. They contain special adaptations of the conventional positive displacement reciprocating class of pumps, which are designed primarily to transfer liquid at an accurately controlled rate. They differ in that the pumping rates of metering pumps can be varied by changing the effective stroke length and, perhaps, by changing the speed. More importantly, the flow rates of metering pumps can be accurately predetermined, with repeatable flows maintained consistently to within 4-1 percent. Metering pumps come in an extremely large variety of sizes and configurations. Figures 8-5 and 8-6 give a glimpse of this variety. Ideally, a metering pump should be capable of handling a wide range of liquids, including those that are toxic, corrosive, dangerous, volatile, and abrasive, as well as those containing concentrations of suspended solids (slurries). In addition, a metering pump should be able to generate sufficiently high discharge pressures to permit injection of liquids into processes. To accomplish this wide range of requirements, many options in design must be available, including the following: 9 Size or capacity 9 Method of control 9 Materials composing the liquid-handling end 9 Valve styles 9 Primary drive requirements 9 Environmental conditions Metering pumps fall into four basic types, defined by the method used to seal the liquid end of the pump from the power end, thus preventing leakage and pumping inaccuracies: 9 Piston, packed seal 9 Plunger, gland packed seal
Positive Displacement Pumps
FIGURE 8-5
313
Three metering pumps driven from a common input shaft. (Source: LEWA, Leonberg, Germany)
FIGURE 8-6
Large diaphragm-type pumps in a chemical plant. (Source: LEWA, Leonberg, Germany)
314 ProcessPlant Machinery 9 Mechanical diaphragm seal 9 Hydraulic diaphragm seal The power end of the metering pump is common to all four types, with various designs used to generate the reciprocating movement required to power the liquid end. Most metering pump designs employ an electric motor as a power source. The motor speed can be reduced to pump design speed by the use of internal motor gears or through gearing built into the pump power end. This rotary power is converted to a linear motion through a crank mechanism producing power in a straight-line reciprocating motion. Depending on the type of adjustable output flow mechanism used, the power can be utilized on both the forward thrust of the crank and the back thrust of the crank. Other capacity controls, however, only take advantage of the power in one direction. There are also metering pumps with electromagnetic and pneumatic power ends creating linear, reciprocating motion using electromagnets and pneumatic pistons, respectively. Although they use conventional liquid end designs, they use highly specialized components to produce this linear reciprocating action. However, the over-whelming majority of metering pumps use some form of crank motion, which is then linked to a crosshead device for positive alignment of the piston or plunger to its cylinder (Figures 8-7 and 8-8).
Packed Piston Pump. The packed piston metering pump uses a power end as just described. The piston is driven by either a crank, a connecting rod, or a crosshead driven by the crank. The piston is the measuring component of the metering pump, designed to displace a measured volume of liquid with a high degree of accuracy as it reciprocates within the pump (Figure 8-9). The packing used moves back and forth with the piston to effect a dynamic seal with the inside diameter of the cylinder and a static seal with the outside diameter of the piston. The forward travel of the piston reduces the internal volume of the liquid chamber, displacing the metered liquid out the discharge check valve. The pressure required to move the liquid through the discharge check valve is also applied to the suction check valve, forcing it into a closed position, ensuring correct flow direction (Figure 8-10). ADJUSTMENT ARM ASSEMBLY EXTERNAL ADJUSTMENT HANDWHEEL
~, ~,
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FIGURE 8-7 Adjustable stroke reciprocating power mechanism. (Source: Pulsafeeder, Rochester, NY)
Positive Displacement Pumps
FIGURE 8-8
Cam-driven reciprocating power mechanism. (Source: Pulsafeeder, Rochester, NY)
FIGURE 8-9
Packed piston pump. (Source: Hydraulic Institute Standards, 13th Edition, Cleveland, OH, 1975)
The reverse travel of the piston decreases the pressure within the liquid chamber by enlarging the internal volume of the chamber. This change of pressure results in a rapid closing of the discharge check valve caused by the external pressure acting on the valve and allows the suction check valve to open because of an external pressure under the check valve that can be either above or below atmospheric pressure (Figure 8-11). An added feature of the packed piston pump is its ability to be double acting, i.e., if so designed, to provide a discharge of fluid into a system on both the forward thrust of the piston and the back thrust of the piston (Figure 8-9). This feature provides up to twice the output capacity for the same horsepower input and minimizes the typical pulsing output flow common to reciprocating pumps. The accuracy of the reciprocating metering pump is achieved by the previously described predetermined controlled piston travel of the pump, the control of the strokes per minute, and the precise opening and closing of the check valves. The inaccuracy, on the other hand, is caused by leakage past the piston packing and the check valves.
315
316
Process Plant Machinery
FIGURE 8-10 Valve action during discharge stroke of diaphragm-type metering pump. (Source: Pulsafeeder, Rochester, NY)
FIGURE 8-11 Valveaction during suction stroke of diaphragm-type metering pump. (Source: Pulsafeeder, Rochester, NY)
Positive Displacement Pumps
317
Packed Plunger Pump. The packed plunger pump is very similar to the packed piston pump except for the packing design and location. The packed plunger, unlike the packed piston, has the packing installed in a stationary gland in the inside diameter of the cylinder. As the plunger reciprocates within the pump, a dynamic seal is made between the outside diameter of the plunger and the inside diameter of the packing, and a static seal is made between the outside diameter of the packing and the inside diameter of the stuffing box (Figure 8-12). The choice between a packed piston pump or a packed plunger pump is dependent on many variables including fluid compatibility with the packing, speed of the piston or plunger, allowable leakage, and pressure requirements.
FIGURE 8-12 Packed plunger metering pump. Stroke length is adjusted by changing the position of the pivot P. (A) When the yoke is positioned at right angles to crosshead motion, the eccentricity is all directed to the crosshead and full stroke results. (B) For any intermediate position of the yoke, any interediate stroke length results. (C) When the yoke is positioned parallel to crosshead motion, the action of the crank is no longer directed to the crosshead, and minimum stroke results. (Source: Milton Roy Company, St. Petersburg, FL)
318
ProcessPlant Machinery Mechanical Diaphragm Pumps. Both the packed piston pump and the packed plunger pump allow some degree of leakage past their dynamic seals. In some cases, this is not an objectionable shortcoming; in other cases, it can be very objectionable and, in most instances, costly as well. For example, a leakage rate of only 25 drops per minute, metering a fluid costing $1.00 a gallon, can cost $200 per year in lost fluid (Figure 8-13). To be considered also is the cost of pressure flushing, drainage, contamination, maintenance, and loss of accuracy. To overcome the leakage problem, a mechanically actuated diaphragm pump can be used. The power side of the pump and the capacity control are the same as were previously described for other types of reciprocating pumps. However, in place of a piston rod or plunger, the mechanical diaphragm pump uses a connecting rod fastened to the center of a diaphragm. The configuration of the diaphragm itself can take on many forms, but the most popular designs are those illustrated in Figure 8-14 - the flat disk, the convoluted disk, and the bellows. In the mechanical diaphragm pump, the principle of positive displacement output is similar to that of the piston plunger pump except that the diaphragm becomes the displacement measuring element, as it moves back and forth in the fluid chamber (Figure 8-15). Hydraulic Diaphragm Pump. The hydraulically balanced diaphragm pump is a hybrid design that provides the principal advantages of the other three pump types. Like the other pumps, its power end and capacity are common. This, however, is
1400
/
12OC M. ,I[
100C
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t
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oi
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L E A K A 3E, I N D R O : S P E R MI 4.
FIGURE 8-13 Cost of leakage. (Source: Metering Pump Handbook, Industrial Press, Inc., New York, NY, 1984)
Positive Displacement Pumps
319
FIGURE 8-14 Mechanical diaphragm styles. (A) Flat disk; (B) convoluted disk; (C) bellows. (Source: Pulsafeeder, Rochester, NY)
OIL FILLER PLUG BALL BEARING RATE.OF.FEED INDICATOR
I ~I
!~ I
% DISCHARGE VALVE
ECCENTRIC STROKE.ADJUSTMENT LEVER
PUSH ROD
STROKE.ADJUSTMENT SHAFT
BALL BEARING DIAPHRAGM
ALUMINUM PLASTIC BASE /
"~ ~L
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.
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OI L PUMP
SUCTION VALVE
RESERVOCIRHAMBER HEAD AND FRONT COVER RET1JRN SPRING
INPUT SHAFT AND WORM
FIGURE 8-15
Mechanical diaphragm pump. (Source: Wallace and Tiernan, Belleville, NJ)
where the similarity ends, since the piston or plunger does not come into contact with the pumped fluid, and the actuation of the diaphragm is by hydraulic power instead of mechanical power (Figure 8-7). The measuring piston or plunger reciprocates within a precisely sized cylinder at an established stroke length, displacing a volume of hydraulic liquid, not the product liquid. The hydraulic liquid is stable and has excellent lubricating qualities. The piston uses the hydraulic oil to move the diaphragm forward and backward (Figures 8-10 and 8-11), causing a displacement that expels the product liquid through the discharge check valve and, on the suction stroke, takes in an equal amount through the suction check valve. The diaphragm isolates the liquid product being contained within the liquid chamber and check valves. These are the only parts that must be made of chemically compatible materials.
320
Process Plant Machinery
The only function of a diaphragm is to separate two liquids. The diaphragm normally does no work, carries no load, and pumps no liquid; rather, it serves as a moving barrier between liquids during periods of pressure imbalance. It is simply a moving partition with pressure hydraulically balanced on both sides; on one side is the liquid product and on the other side is the hydraulic oil. At full deflection, the diaphragm undergoes total combined stresses well within the endurance limit of the diaphragm material. Contoured support plates are provided on either side of the diaphragm to ensure that stresses are kept within limits. When properly installed and working within the recommended temperature range and not affected by corrosion or abrasion, the diaphragm has an unlimited life. As previously stated, the piston or plunger handles only hydraulic oil. Conventional seals are used on the piston or plunger, which does not require power flushing and complicated drain systems, as are found on conventional piston or plunger pumps handling corrosive or hazardous liquids. Even the slightest leakage past the piston is replaced on the suction stroke through the automatic functioning of a compensation system, which draws in replacement oil from the oil reservoir (Figure 8-16). Any excess pressure within the hydraulic system or the liquid product chamber is relieved through the automatic action of a pressure relief valve. This valve blows off oil, under excess pressure ahead of the piston, back into the oil reservoir (Figure 8-17). The vacuum and pressure compensator systems actually perform three important functions that the other described types of metering pumps cannot do unless auxiliary equipment is added to their piping systems. As described previously, they compensate for any leakage occurring within the hydraulic system of the pump, ensuring a balanced diaphragm movement. In addition, they serve to protect the process system from over-pressure conditions produced by the pump. For instance, the positive displacement pump, because of its design, causes excess pressure within the system to the point of damaging the pump, bursting pipes, or damaging other downstream equipment should an operator mistakenly close a shut-off valve downstream from the pump. The hydraulic diaphragm pump will, however, relieve any pump-produced pressure beyond the set pressure of the pressure relief valve, thus avoiding the dangerous buildup of pressure. The compensation system also serves to protect the pump from a closed suction line or a partially clogged strainer in the suction line. Should this occur, the backward movement of the diaphragm is
OIL LEVEL
FIGURE 8-16 Function of oil makeup valve. (Source: Metering Pump Handbook, Industrial Press, Inc., New York, NY, 1984)
Positive Displacement Pumps
OIL LEVEL
FIGURE 8-17 Function of pressure relief valve. (Source: Metering Pump Handbook, Industrial Press, Inc., New York, NY, 1984)
prevented and the vacuum relief system would automatically open to relieve the starved suction condition within the pump. In doing so, however, a surplus of hydraulic oil enters into the system between the diaphragm and piston. As the piston starts forward on its discharge stroke, the diaphragm is displaced forward and will come into contact with the contoured dish support plate in the process liquid chamber, because of the surplus oil drawn into the hydraulic chamber. At the moment of diaphragm contact with its support plate, an over-pressure condition starts to develop within the hydraulic system. The pressure relief valve now opens to relieve the surplus oil back into the hydraulic reservoir, preventing a dangerous buildup of pressure. The interaction of the two compensation systems continues stroke after stroke to activate a fluid clutch-type action to prevent overloading of the pump power end until the condition plugging the suction or discharge lines is found and corrected. As in all properly designed hydraulic systems, an air-bleed system is required to purge air from the hydraulic system either automatically or manually. Because the diaphragm is actuated hydraulically, this type of pump is highly adaptable. The product liquid chamber can be separated from the power end of the pump, using a pipe or tube to transfer the hydraulic power required. This feature is highly desirable when metering toxic, high- or low-temperature extremes. Additional diaphragms may also be added to accomplish various system requirements. Summary. In summary, the principal features and limitations of each type of metering pump are listed on page 322. ROTATING POSITIVE DISPLACEMENT PUMPS*
Extemal and internal gear pumps are prevalent not only in the process industries but in general machinery and industrial equipment as well. The external gear pump (Figure 8-18) is probably the most widely used rotary pump. It consists of two meshing gears in a close-fitting housing. Gear rotors are cut externally and, in this type of pump, fluid is carried between the gear teeth and displaced when they mesh. * Sources: As acknowledged in captions to illustrations.
321
322
Process Plant Machinery LIMITATIONS
FEATURES
Piston Pump 1. Low cost 2. Can be double acting 3. Capacities from a few cubic centimeters per hour to 1200 gallons per hour (gph) (4.5 x 103 liters/hr) 4. Packing adjustment not required 5. Accuracies to 1 percent over 10:1 range 6. Can pump high-viscosity fluids
1. Packing leakage is unavoidable, a consideration with corrosive or dangerous chemicals 2. Unsuitable for abrasive slurries 3. Maintenance needed for piston packing 4. No built-in relief features 5. Dynamic packing subject to differential pressure
Packed Plunger Pump 1. Relatively low cost 2. Capacities from a few cubic centimeters per hour to over 20,000 gph (7.6 x 104 liters/hr) 3. Accuracies to 1 percent over 15:1 range 4. Pressure capability to 50,000 psi (3.5 x 105 kPa) 5. Least effect from changes in discharge pressure 6. Can pump high-viscosity liquids
1. Packing wear requires periodic adjustment 2. Packing leakage is unavoidable, a consideration with corrosive or dangerous chemicals 3. Unsuitable for abrasive slurries 4. Maintenance needed for packing and plunger wear 5. Pressure flushing or draining of packing gland required 6. No built-in relief
Mechanical Diaphragm Pump 1. Relatively low cost 2. Zero chemical leakage 3. Can pump abrasive slurries and chemicals 4. Packing adjustment not required 5. Can pump high-viscosity liquids
1. High maintenance due to high-stress loading of the diaphragm 2. Discharge pressure limitation of 200 psi (1.4 • 103 kPa) 3. Accuracy in the 5 percent range and as much as 10 percent zero shift with pressure change from minimum to maximum 4. Limited capacity range 5. Dynamic packing subject to differential pressures 6. No built-in safety features
Hydraulic Diaphragm Pump 1. 2. 3. 4. 5. 6. 7.
Zero chemical leakage High adaptability Can pump wide range of liquids Packing adjustment not required Low maintenance Pressures to 5000 psi (3.5 x 104 kPa) Accuracies 1 percent and capacities from a few cubic centimeters per hour to 1500 gph (5.7 x 103 liters/hr) 8. Built-in safety features
1. Higher initial cost 2. Subject to a predictable zero shift of 5 percent to 10 percent per 1000 psig (6.9 x 103 kPa) 3. Limited to moderate-viscosity fluids
Positive Displacement Pumps
EXTERNAL GEAR
FIGURE 8-18
External gear pump. (Source: Dresser-Worthington, Harrison, NJ)
The gears may be one of three designs: spur, helical, or herringbone. Spur gears are used on low-capacity, high-pressure applications, since they offer line contact between the teeth. This reduces the slip through the meshing line and gives better performance at higher pressure. Spur gears are noisy at high speeds, and unless some relief is provided with grooves or ports, there is a tendency to trap liquid at the meshing point and shaft deflection could result. Single helical gears eliminate trapping by gradual engagement and disengagement of the teeth. Unfortunately, a component of the load produces an end thrust in one direction that could cause sideplate wear unless compensated for by thrust bearings. The major advantage of single helical gears over double helical or herringbone gears is cost. But herringbone gears eliminate end thrust while still retaining the advantages over spur gears of higher speed operation, pulsation-free flow, and no liquid trapping. Some controversy exists over the direction of rotation of gear pumps. The preferred direction is with the gear apexes leading so the liquid is squeezed out from the center toward both ends of the gears. However, except for extremely viscous liquids at high pressures, there is no measurable difference in capacity, power, or noise due to direction of rotation. Internal gear pumps (Figure 8-19) have one rotor with internally cut gear teeth meshing with an externally cut gear. On the outer sideplate is a stationary crescent.
323
324
Process Plant Machinery
FIGURE 8-19
Internal gear pump. (Source: Dresser-Worthington, Harrison, NJ)
FIGURE 8-20 Sliding vane pump. 1 - nonmetallic or metallic vanes; 2 - bearing area, internal or external to fluid pumped; 3 - soft packing or mechanical face seal area; 4 - safety valve; 5 - casing (could be jacketed); 6 - rotor. (Source: Foster Pump Works, Inc., Westerley, RI)
As the internal gear rotates, the idler (external) gear follows and liquid is displaced between the internal gear in the crescent and between the idler and the crescent. This type of pump is generally used for lower pressure applications at low speeds and it is generally a cantilever design. The v a n e type category of rotating positive displacement pumps consists of external vane types or sliding vanes, as shown in Figure 8-20. The vane (or vanes) may be in the form of blades, buckets, or slippers, with a cam to draw fluid into and force it out of the pump chamber. In the sliding vane pump, the rotor is slotted and a series of vanes follow the bore of the casing. Liquid is displaced between the vanes. This is a slow speed design and it does have lower viscosity limits than gear pumps. The vanes are, however, self-compensating for wear, and this pump, unlike a close-clearance gear design, can be used on mildly erosive liquids.
Positive Displacement Pumps
LOBE FIGURE 8-21 Lobe-type pump. (Source: Dresser-Worthington, Harrison, NJ) In lobe type pumps, Figure 8-21, liquid is carried between the rotor lobe surfaces from the inlet to the outlet. Since the lobes cannot drive each other, external timing gears are necessary. This type is a low-speed, high-displacement, low-pressure pump popular for marine or highly viscous, low-pressure process applications. There is little difference between lobe-type pumps and the lobe type blowers discussed in Chapter 15 of this text. In single-screw pumps (Figure 8-22), the liquid is carried between the rotor screw threads and axially displaced as the rotor threads mesh with internal threads on the stator. This design is versatile in that the rotor and the stator can be made of many different materials so the pump can be used on certain abrasive and corrosive services. Single-screw pumps are sometimes called progressive cavity pumps and are most successfully used in the food processing industries. Multiscrew pumps are self-priming rotary displacement pumps. Theoretically speaking, they operate like a piston pump with strokes of infinite length. Numerous variations are available. In Figure 8-23, a, two-flight driven spindle closely engages and rotates with a three-flight driven spindle. Both are located in a close-fitting
FIGURE 8-22 Single-screw pump and progressive cavity detail. (Sources: Dresser-Worthington, Harrison, NJ, and Netzsch Inc., Exton, PA)
325
326
Process Plant Machinery
FIGURE 8-23 T w i n - s c r e w p u m p . 62* - m e c h a n i c a l seal; 6 9 + - l o c k i n g s c r e w ; 72 +* 75 - g l a n d ; 76 + - s c r e w ; 81 - a d j u s t i n g s c r e w ; 82 + - c a p nut; 1 7 0 +* - b a l l b e a r i n g ; 172 - s u p p o r t i n g disk; 173 +* - c i r c l i p - d r i v i n g spindle-drive side; 1 7 4 + - circlip-
gland;
cover-drive-side; + - spare parts; * - DIN parts. (Source: Leistritz Pump Corporation, Allendale, N . J )
casing. In combination, screws and casing form perfectly sealed chambers. The material confined in these chambers is continuously advanced without undue shear force and turbulence.
Peristaltic Pumps Peristaltic pumps are positive displacement pumps incorporating a flexible hose. The model shown in Figure 8-24, essentially a development of the conventional peristaltic pumping concept, is glandless and valveless. The liquid passes straight through the pump without restrictions. Three main components produce the pumping a c t i o n - a rotor with three rollers and a smooth-bored flexible hose tube element with the addition of a flexible rubber separator element. The separator is attached to the top of the pump casing, creating a vacuum on the suction side. The hose tube passes directly through the pump between inlet and outlet interposed between the
Positive Displacement Pumps
327
FIGURE 8-24 Peristaltic pump, exploded view and sequence of operation. The rotor (1) with three rollers (2) rotates within a stationary separation belt (3) (see exploded view on the top). The separation belt (3) seals the outer space (4) on the suction side from the inner space (5). The outer space (4) is sealed on all sides and during rotation, the intake hose (8) is surrounded by a vacuum that increases the hose volume, causing a suction effect. The hose volume in space (6) remains constant and the hose volume in space (7) is reduced, which effects a pumping action. The sealed housing holds approximately 1 1/2 quarts of light paraffin-base oil that acts as a lubricating, blocking, and cooling agent. The vacuum gauge indicates the vacuum existing in space (4) and also inside the intake hose (8). The hose support (9) is an elastomer selected to withstand abusive action of foreign matter. (Source: National Mastr Pump, Inc., Houston, TX) separator and the bottom of the casing. Each roll/squeeze movement of the roller displaces the pumped liquid while the vacuum surrounding the tube on the suction side ensures that it regains shape instantly when the pressure is released. This rapid tube recovery feature enhances and accelerates the normal suction effect of the roll/squeeze peristaltic movement and confers on the pump its exceptional suction characteristics. It allows the use of thin-walled tubes that create a high RPM/volume range. Peristaltic pumps are suitable for liquids, slurries, and gases. Different hose materials accommodate the various feed streams.
This Page Intentionally Left Blank
APPENDIX 8A
Principles of Operation of Reciprocating Pumps How a Reciprocating Pump Works A reciprocating pump is a positive displacement mechanism with liquid discharge pressure being limited only by the strength of the structural parts. Liquid volume or capacity delivered is constant regardless of pressure, and is varied only by speed changes. Characteristics of a reciprocating pump are 1)positive displacement of liquid, 2)high pulsations caused by the sinusoidal motion of the piston, and 3)high volumetric efficiency.
Plunger or Piston Rod Load Plunger or piston "rod load" is an important power end design consideration for reciprocating pumps. Rod load is the force caused by the liquid pressure acting on the face of the pis-, ton or plunger. This load is transmitted directly to the power frame assembly and is normally the limiting factor in determining maximum discharge pressure ratings. This load is directly proportional to the pump gauge discharge pressure and proportional to the square of the plunger or piston diameter. Occasionally, allowable liquid end pressures limit the allowable rod load to a value below the design rod load. IT IS IMPORTANT THAT LIQUID END PRESSURES DO NOT EXCEED PUBLISHED LIMITS.
Calculations of Volumetric Efficiency Volumetric efficiency (Ev) is defined as the ratio of plunger or piston displacement to liquid displacement. The volumetric efficiency calculation depends upon the internal configuration of each individual liquid body, the piston size, and the compressibility of the liquid being pumped.
Tools for Liquid Pulsation Control, Inlet and Discharge Pulsation Control Tools ("PCT", often referred to as "dampeners" or "stabilizers") are used on the inlet and discharge piping to protect the pumping mechanism and associated piping by reducing the high pulsations within the liquid caused by the motions of the slider-crank mechanism. A properly located and charged pulsation control tool may reduce the length of pipe used in the acceleration head equation to a value of 5 to 15 nominal pipe diameters. The pulsation control tools are specially required to compensate for inadequately designed or old/adapted supply and discharge systems.
the inlet valve chamber. The magnitude of the surges and how they will react in the system is impossible to predict without an extremely complex and costly analysis of the system. Since the behavior of the natural frequencies in the system is not easily predictable, as much of the surge as possible must be eliminated at the source. Proper installation of an inlet pulsation control PCT will absorb a large percentage of the surge before it can travel into the system. The function of the PCT is to absorb the "peak" of the surge and feed it back at the low part of the cycle. The best position for the PCT is in the liquid supply line as close to the pump as possible, or attached to the blind flange side of the pump inlet. In either location, the surges will be dampened and harmful vibrations reduced.
REQUIRED FORMULAE AND DEFINITIONS Acceleration Head ha-
LVNC Kg
V
GPM (2.45)(D)
where ha = Acceleration head (in feet) L = Length of liquid supply line (in line) V = Average velocity in liquid supply line (in fps) N = Pump speed (revolutions per minute) C = Constant depending on the type of pump C = 0.200 for simplex double-acting = 0.200 for duplex single-acting = 0.115 for duplex double-acting = 0.066 for triplex single or double-acting = 0.040 for quintuplex single or double-acting = 0.028 for septuplex, single or double-acting = 0.022 for nonuplex, single or double-acting K = Liquid compressibility factor K = 2.5 For relatively compressible liquids (ethane, hot oil) K = 2.0 For most other hydrocarbons K = 1.5 For amine, glycol and water K = 1.4 For liquids with almost no compressibility (hot water) g = Gravitational constant = 32.2 ft/sec 2 d = Inside diameter of pipe (inches)
Stroke One complete uni-directional motion of piston or plunger. Stroke length is expressed in inches.
AcceleraUon Head
Pump Capacity (Q)
Whenever a column of liquid is accelerated or decelerated. pressure surges exist. This condition is found on the inlet side of the pump as well as the discharge side. Not only can the surges cause vibration in the inlet line, but they can restrict and impede the flow of liquid and cause incomplete filling of
The capacity of a reciprocating pump is the total volume through-put per unit of time at suction conditions. It includes both liquid and any dissolved or entrained gases at the stated operating conditions. The standard unit of pump capacity is the U.S. gallon per minute.
330 Process Plant Machinery
KI i . / : ~
100e/o
~
o.b'r 0
.,.
6O
0~ I,,"1 0
.
.
.
. DUPLEX DOUBLE-ACTING Average Flow Maximum Flow Minimum Flow Total Flow Var.
,,,
120
180
240
300
-----
100% 100% + 24% 100% - 22% 46%
360
Crankshaft Angle (Degree)
[ ,'
100~
B ~
i I i ?'.I, 60
J I I ,;I
120
180
240
TRIPLEX SINGLE-ACTING
I,, ~,,,_~_.., 300
Average Flow Maximum Flow Minimum Flow Total Flow Var.
-----
100% 100% + 6% 100% - 17% 23%
360
Crankshaft Angle (Degree) F i g . 8A-1
Reciprocating
pumps
flow characteristics
100o,tl QUINTUPLEX SINGLE-ACTING =D
o,
0
6o
J
120
m,
L,,
Average Flow Maximum Flow Minimum Flow Total Flow Var 180
240
300
-----
100% 100% + 2o/o 100o/o - 5% 7%
360
Crankshaft Angle (Degree) 100% SEPTUPLEX SINGLE-ACTING Average Flow Maximum Flow Minimum Flow Total Flow Var. 0
60
120
180
240
300
-----
100o/o 100% + 1 2 % 100% - 2.6% 3.8%
360
Crankshaft Angle (Degree) 100% ~ NONUPLEX SINGLE-ACTING Average Flow - Maximum F l o w Minimum Flow - Total Flow Var. - -
o. i::i 0
60
120
180
240
300
100% 100% + 0.6% 100% - $1.5% 2.1%
360
Crankshaft Angle (Degree) Fig. 8 A - 2
Pump Displacement
For d o u b l e - a c t i n g piston p u m p s :
(D)
T h e d i s p l a c e m e n t of a r e c i p r o c a t i n g p u m p is the v o l u m e s w e p t by all p i s t o n s or p l u n g e r s p e r unit time. D e d u c t i o n for piston rod v o l u m e is m a d e on d o u b l e acting piston t y p e p u m p s w h e n c a l c u l a t i n g d i s p l a c e m e n t . T h e s t a n d a r d unit of p u m p d i s p l a c e m e n t is t h e U.S. g a l l o n p e r m i n u t e . For s i n g l e - a c t i n g p u m p s : D =
Asnm 231
D=
(2A - a)snm 231
Where A = P l u n g e r or piston area, s q u a r e inch a = Piston rod c r o s s - s e c t i o n a l area, s q u a r e inch (double-acting pumps) s = S t r o k e length, inch
Appendix 8A: Principles of Operation of Reciprocating Pumps 331 n = RPM of crankshaft m = Number of pistons or plungers
Plunger or Piston Speed (v) The plunger or piston speed is the average speed of the plunger of piston. It is expressed in feet per minute. V--
ns 6
Pressures The standard unit of pressure is the pound force per square inch.
Discharge Pressure ( P d ) - The liquid pressure at the centerline of the pump discharge port.
Volumetric Efficiency (r/v)- The ratio of the pump capacity to displacement. Q r/v=
Plunger Load (Single-Acting Pump) The computed axial hydraulic load, acting upon one plunger during the discharge portion of the stroke is the plunger load. It is the product of plunger area and the gauge discharge pressure. It is expressed in pounds force.
Piston Rod Load (Double-Acting Pump) The computed axial hydraulic load, acting upon one piston rod during the forward stroke (toward head end) is the piston rod load.
Suction Pressure (Ps) - The liquid pressure at the centerline
It is the product of piston area and discharge pressure, less the product of net piston area (rod area deducted) and suction pressure. It is expressed in pounds force.
Differential pressure (Ptd) - The difference between the liquid discharge pressure and suction pressure.
Liquid pressure
of the suction port.
Net Positive Suction Head Required (NPSHR)-The amount of suction pressure, over vapor pressure, required by the pump to obtain satisfactory volumetric efficiency and prevent excessive cavitation. The pump manufacturer determines (by test) the net positive suction head required by the pump at the specified operating conditions. NPSHR is related to losses in the suction valves of the pump and frictional losses in the pump suction manifold and pumping chambers. Required NPSHR does not include system acceleration head, which is a system-related factor.
Slip (S) Slip of a reciprocating pump is the loss of capacity, expressed as a fraction or percent of displacement, due to leaks past the valves (including the back-flow through the valves caused by delayed closing) and past double-acting pistons. Slip does not include fluid compressibility or leaks from the liquid end.
Cylinder force (pounds) - pressure (psi)x area (square inches)
Cylinder speed or average liquid velocity through piping (feet/second)= flow rate (gpm) 2.448 x (inside diameter (inch))2
Reciprocating pump displacement (gpm) = rpm x displacement (cubic in/revolution) 231
Pump Input horsepower - see horsepower calculations Shaft torque (foot-pounds) = horsepower x 5252 shaft speed (rpm) 120 x frequency (Hz) number of poles
Electric motor speed (rpm) =
Power (P) Pump Power Input (Pi) - The mechanical power delivered to
a pump input shaft, at the specified operating conditions. Input horsepower may be calculated as follows: Pi =
Q x Ptd 1714 x r/p
Pump Power Output (Po) - The hydraulic power imparted to the liquid by the pump, at the specified operating conditions. Output horsepower may be calculated as follows: Po =
Q x Ptd 1714
The standard unit for power is the horsepower. Efflclenclee (n)
Pump
(pounds) or (psi) = square inch Cylinder area (square inches) = (3.1416) x (Radius(inches))2 = (3.1416) x (diameter(inches)) 2 4
Efficiency (r/p) (also called pump mechanical efficiency) - The ratio of the pump power output to the pump power input. Po OP = -~-
Three phase motor horsepower (output) =
1.73 x amps x volts x efficiency x power factor 746
Static head of liquid (feet) =
2.31 x static pressure (psig) specific gravity
Velocity head of liquid (feet) =
liquid velocity 2 g = 32.2 ft/sec 2
Absolute viscosity (centipoise) = specific gravity x Kinematic viscosity (centistrokes) Kinematic viscosity (centistrokes) =
180 0.22 x saybolt viscosity (ssu) - Saybolt viscosity (ssu)
Absolute pressure (psia) =
Local atmospheric pressure + gauge pressure (psig)
Gallon per revolution = Area of plunger (sq in) x length of stroke(in) x number of plungers 231
332 Process Plant Machinery Barrels per day = gaPrev x pump speed (rpm) x 34.3
Specific gravity (at 60~
Ptd = Discharge pressure minus suction pressure in psi S = Slip, expressed in decimal value
141.5 = 131.5 + API gravity (degree)
Example: Find the volumetric efficiency of a reciprocating pump with the following conditions:
Bolt clamp load (Ib) = 0.75 x proof strength (psi) x tensile stress area (in 2)
Type of pump
0.2 (or 0.15) x nominal diameter in Bolt torque (ft-lb)= 12 inches x bolt clamp load (Ib)
Liquid pumped Suction pressure Discharge pressure Pumping temperature c d S
0.2 for dry 0.15 for lubricated, plating, and hardened washers
Calculating Volumetric Efficiency for Water
Find/51 from Table of Water Compressibility (Table 8A. 1).
The volumetric efficiency of a reciprocating pump, based on capacity at suction conditions, using table of water compressibility, shall be calculated as follows:
Vol. Eft. =
1 -Ptd/~t 1 +
/~t = 0.00000305 at 140 F and 1800 psia Calculate volumetric efficiency:
c
1 - Ptd ~t
Vol. Eff. =
- S
1 - [Ptd~t (1 + d ) ] - S : 1 - Ptd~t
Where = Compressibility factor at temperature t (degrees Fahrenheit or centigrade). (See Tables 8A. 1 and 8A.2). c = Liquid chamber volume in the passages of chamber between valves when plunger is at the end of discharge stroke in cubic inches d = Volume displacement per plunger in cubic inches
TABLE 8A.1
3 in diam plunger x 5 in stroke triplex Water Zero psig 1785 psig 140 F 127.42 cu in 35.343 cu in 0.02
1 - [(1785 - 0)(.00000305)] 1 +
[ 127.42 ] 35.343 J _ 0.02
1 - (1785 - 0)(0.00000305) = 0.96026 = 96 per cent
Water Compressibility Compressibility Factor pt x 10 -6 = Contraction in Unit Volume Per Psi Pressure Compressibility from 14.7 Psia, 32 F to 212 F and from Saturation Pressure Above 212 F TEMPERATURE
Pressure 0 C 20 C 40 C 60 C 80 C 100 C120 C140 C160 C180 C200 C220 C240 C260 C280 C300 C320 C340 C360 C Psia 32 F63 F104 F140 F176 F212 F248 F284 F320 F356 F392 F428 F464 F500 F536 F572 F608 F644 F680 F 200 400 600 800 1000
3.12 3.06 3.11 3.05 3.103.05 3.103.04 3.09 3.03
3.06 3.05 3.05 3.04 3.03
3.12 3.11 3.10 3.09 3.09
3.23 3.22 3.21 3.21 3.20
3.40 3.39 3.39 3.38 3.37
3.66 3.64 3.63 3.62 3.61
4.00 3.99 3.97 3.96 3.95
4.47 4.45 4.44 4.42 4.40
5.11 5.09 5.07 5.04 5.02
6.00 5.97 5.93 5.90 5.87
7.27 7.21 7.15 7.10 7.05
8.95 8.85 8.76
11.6 11.4
16.0
1200 1400 1600 1800 2000
3.083.02 3.073.01 3.063.00 3.052.99 3.042.99
3.02 3.01 3.00 3.00 2.99
3.08 3.07 3.06 3.05 3.04
3.19 3.36 3.18 3.35 3.17 3.34 3.16 3.33 3.15 3.32
3.60 3.59 3.58 3.57 3.56
3.94 3.92 3.91 3.90 3.88
4.39 5.00 4.37 4.98 4.35 4.96 4.34 4.94 4.32 4.91
5.84 5.81 5.78 5.75 5.72
7.00 6.95 6.91 6.87 6.83
8.68 8.61 8.53 8.47 8.40
11.2 11.1 10.9 10.8 10.7
15.4 15.1 14.8 14.6 14.3
23.0 21.9 21.2 20.7
36.9 34.7
2200 2400 2600 2800 3000
3.032.98 3.022.97 3.01 2.96 3.002.95 3.002.94
2.98 2.97 2.96 2.96 2.95
3.04 3.03 3.02 3.01 3.00
3.14 3.14 3.13 3.12 3.11
3.31 3.30 3.29 3.28 3.28
3.55 3.54 3.53 3.52 3.51
3.87 3.85 3.85 3.83 3.82
4.31 4.29 4.28 4.26 4.25
4.89 4.87 4.85 4.83 4.81
5.69 5.66 5.63 5.61 5.58
6.78 6.74 6.70 6.66 6.62
8.33 8.26 8.20 8.14 8.08
10.6 10.5 10.4 10.3 10.2
14.1 13.9 13.7 13.5 13.4
20.2 19.8 19.4 19.0 18.6
32.9 31.6 30.5 29.6 28.7
86.4 69.1 61.7 57.2 238.2 53.8 193.4
3200 3400 3600 3800 4000
2.992.94 2.982.93 2.972.92 2.962.91 2.952.90
2.94 2.93 2.93 2.92 2.91
3.00 2.99 2.98 2.97 2.97
3.10 3.09 3.09 3.08 3.07
3.27 3.26 3.25 3.24 3.23
3.50 3.49 3.48 3.47 3.46
3.81 3.80 3.79 3.78 3.76
4.23 4.22 4.20 4.19 4.17
4.79 4.78 4.76 4.74 4.72
5.55 5.53 5.50 5.47 5.45
6.58 6.54 6.51 6.47 6.43
8.02 7.96 7.90 7.64 7.78
10.1 9.98 9.89 9.79 9.70
13.2 13.0 12.9 12.7 12.5
18.3 17.9 17.6 17.3 17.1
27.9 27.1 26.4 25.8 25.2
51.0 48.6 45.4 44.5 42.8
161.0 138.1 122.4 110.8 101.5
4200 4400 4600 4800 5000
2.952.90 2.942.89 2.932.83 2.922.87 2.91 2.87
2.90 2.90 2.89 2.88 2.87
2.96 2.95 2.94 2.94 2.93
3.06 3.05 3.05 3.04 3.03
3.22 3.21 3.20 3.20 3.10
3.45 3.44 3.43 3.42 3.41
3.75 3.74 3.73 3.72 3.71
4.16 4.70 4.14 4.68 4.13 4.66 4.12 4.64 4.10 4.63
5.42 5.40 5.37 5.35 5.32
6.40 6.36 6.32 6.29 6.25
7.73 7.68 7.62 7.57 7.52
9.62 9.53 9.44 9.36 9.28
12.4 12.2 12.1 12.0 11.8
16.8 24.6 16.5 24.1 16.3 23.6 16.0 23.2 15.8 22.7
41.3 40.0 38.8 37.6 36.6
93.9 87.6 82.3 77.7 73.9
5200 5400
2.902.85 2.87 2.92 3.02 2.902.85 2.86 2.91 3.01
3.18 3.17
3.40 3.39
3.69 3.68
4.09 4.07
5.30 5.27
6.22 6.19
7.47 7.41
9.19 9.12
11.7 11.6
15.6 15.3
35.6 34.6
70.3 66.9
4.61 4.59
22.3 21.9
Appendix 8A: Principles of Operation of Reciprocating Pumps 333 TABLE 8A.2 Water Compressibility Compressibility Factor/~t x 10 -6 = Contraction in Unit Volume Per Psi Pressure Compressibility from 14.7 Psia at 68 F and 212 F and from Saturation Pressure at 392 F
Temperature
Temperature
Pressure Psia
20 C 68 F
100 C 212 F
200 C 392 F
Pressure Psia
20 C 68 F
100 C 212 F
200 C 392 F
6000 7000 8000 9000 10000 11000 12000 13000 14000 15000 16000 17000 18000 19000 20000 21000
2.84 2.82 2.80 2.78 2.76 2.75 2.73 2.71 2.70 2.69 2.67 2.66 2.65 2.64 2.63 2.62
3.14 3.10 3.05 3.01 2.96 2.92 2.87 2.83 2.78 2.74 2.69 2.65 2.60 2.56 2.51 2.47
5.20 5.09 4.97 4.87 4.76 4.66 4.57 4.47 4.38 4.29 4.21 4.13 4.05 3.97 3.89 3.82
22000 23000 24000 25000 26000 27000 28000 29000 30000 31000 32000 33000 34000 35000 36000
2.61 2.59 2.58 2.57 2.56 2.55 2.55 2.54 2.53 2.52 2.51 2.50 2.49 2.49 2.48
2.42 2.38 2.33 2.29 2.24 2.20 2.15 2.11 2.06 2.02 1.97 1.93 1.88 1.84 1.79
3.75 3.68 3.61 3.55 3.49 3.43 3.37 3.31 3.26 3.21 3.16 3.11 3.07 3.03 2.99
Calculating Volumetric Efficiency For Hydrocarbons
Vol. Eft. = Volumetric efficiency expressed in decimal value.
The volumetric efficiency of a reciprocating pump based on capacity at suction conditions, using compressibility factors for hydrocarbons, shall be calculated as follows: V~ Where Pd " - density at discharge pressure p8 = density at suction pressure c = Fluid chamber volume in the passages of chamber between valves, when plunger is at the end of discharge strike, in cubic inches d = Volume displacement per plunger, in cubic inches P = Pressure in psia (Ps = suction pressure in psia; Pd = discharge pressure in psia) Pc = Critical pressure of liquid in psia Pr -" Reduced pressure Actual pressure in psia P Critical pressure in psia Pc Ps Prs = Reduced suction pressure = Pcc erd-
el = - - x (o x 62.4 = density of liquid in Ib per cu ft (O1
s d P1 (ol
= Density in Ib per cu ft at suction pressure = Density in Ib per cu ft at discharge pressure = Expansion factor of liquid = Characteristic constant in grams per cubic centimeter for any one liquid which is established by density measurements and the corresponding values of (See Table 8A.3).
Example: Find volumetric efficiency of the previous reciprocating pump example with the following new conditions: Type of Pump Liquid pumped Suction temperature Discharge temperature Suction pressure Discharge pressure
3 inch dia plunger x 5 inch stroke triplex Propane 70 F 80 F 242 psig 1911 psig
P~ Reduced discharge pressure = ~cc
S t
= Slip expressed in decimal value = Temperature, in degrees Rankine = Degrees F + 460 (ts = suction temperature in degrees Rankine; td = discharge temperature in degrees Rankine) Tc = Critical temperature of liquid, in degrees Rankine (See Table 8A.3) ]'r = Reduced temperature actual temp. in degrees Rankine critical temp. in degrees Rankine t = T-~ (See Fig. 1) Trs = Reduced suction temperature ts Tc "l'rd = Reduced discharge temperature td Tc
TABLE 8A.3 Carbon Atoms 1 2 3 4 5 6 7 8 9 10 12 14
Name
Tc Degrees Rankine
Pc Lb Per Sq in
pl /(ol Grams Per cc
Methane Ethane Propane Butane Pentane Hexane Heptane Octane Nonane Decane Dodecane Tetradecan e
343 550 666 766 847 915 972 1025 1073 1114 1185 1248
673 717 642 544 482 433 394 362 332 308 272 244
3.679 4.429 4.803 5.002 5.128 5.216 5.285 5.340 5.382 5.414 5.459 5.483
334 Process Plant Machinery 1.0
0.9
0.8
I..J uJ n..
0.7
I-rr" uJ Q. =E LU I--
a LU (..) ::::) Q UJ
0.6
).5
.14
.13
.12
.11
.10
.09
.08
.07
.06
EXPANSION F A C T O R - Fig. 8 A - 3
Find density at suction pressure: ts Trs = Tc
460 + 70 666 - 0.795
Ps 257 Prs = Pcc = 642 = 0.4 P l = 4.803 ( F r o m Table 8A.3, propane) (Ol (0 = 0.1048
Expansion factor vs. reduced temperature
ps=--
pl (.01
x(0x62.4
= 4.803 x 0 . 1 0 4 8 x 62.4 = 31.4 Ib per c u f t Find density at discharge pressure: td Trd -- Tc -
4 6 0 + 80 666 = 0.81
.05
.04
).4
Appendix 8A: Principles of Operation of Reciprocating Pumps 335 Pd 1926 P r d = Pcc = 642 = 3 " 0 = O. 1089 pd=
pl ~-~ x o~ x 62.4
Therefore Vol. Eft.
1 Is ~(1 ~)1 -1 Ioo2 12'42(13s343~)J
= 4 . 8 0 3 x 0.1089 x 62.4
= 0.8376
= 3 2 . 6 4 Ib per cu ft
= 8 3 . 7 6 per cent
This Page Intentionally Left Blank
Chapter 9 Vacuum Pumps An amazing number of process plant applications require vacuum systems for continuous or intermittent services. Central vacuum systems are often found in power stations and in industrial and marine installations for the purpose of priming large centrifugal pumps; similar systems provide vacuum for surgical and clinical purposes in hospitals, dentist's offices, and laboratories. In sterilization processes, vacuum pumps remove air to permit rapid penetration of steam and/or gas into articles requiring sterilization. Food processing plants utilize vacuum systems for anything from poultry evisceration to final packaging in bags, windows, or blister packs for enhanced appearance and marketability. Vacuum filtration is used for sewage, chemicals, foods, and many mined or fibrous products. Many condensing processes utilize vacuum pumps to evacuate condensers and to remove air leakage and noncondensibles. Smaller installations use liquid ring vacuum pumps in a dual role, removing condensates as well as noncondensibles, sometimes utilizing the cooled condensate as a service liquid. Vacuum pumps are also used in impregnation and metallurgical treatment processes where air and gas have to be removed prior to impregnation with appropriate chemical fluids; or prior to the application of diffusion coatings, or in advanced ion and plasma-type surfacing techniques. Vacuum pumps are used for drying, distillation, and evaporation. Lower boiling temperatures attained under vacuum preserve nutrients and improve taste, quality, and shelf life of products such as candies, jams, pharmaceuticals, and many mild products. Deaeration is needed for products such as meat pastes, sauces, soups, cellulose, latex, bricks, tiles, sewer pipes, and pottery clay. Also, vacuum conveyance of dangerous, viscous, contaminated, powdery, flaky, bulky, or simply hard-to-handle materials or products is used. The author remembers the ease and utter simplicity with which laminated plastic toothpaste tubes are transferred in partially evacuated, transparent plastic pipes from the forming machine at one end of the plant to the filling equipment at the other end of the building. With a profusion of processes and applications thus benefiting from vacuum pumps, it is not surprising that many different types and styles, sizes and models, configurations and variations of vacuum producing machinery are available to the user. The familiar steam, gas, and fluid jet injectors/eductors must be acknowledged as prime vacuum producers; however, we will only mention them in passing because they lack moving parts and thus do not fit our definition of "machinery." Vacuum pumps are often classified in two broad categories: dry type and liquid type. Dry types include lobe, rotary piston, sliding vane, and even diaphragm pumps. Liquid vacuum pumps include liquid jet and liquid ring pumps. Figures 9-1 and 9-2 show the operating ranges for many of these pumps. It should be noted that there is considerable overlap among ranges. * Source: As acknowledged in captions to illustrations. 337
338
Process Plant Machinery
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AIR EJECTOR/LIQUID RING COMBINATION COMPOUND LIQUID RING PUMP WITH OIL SEALANT
,.-.---+ - - - - - - - ~ - ~ ,
~
, ~ 1 - -
BOOSTER/LIQUID RING OR =~1lilt JET IC'T"P/"~&Ar'tlklATI/"%kl BOOSTER/LIQUID COMBINATION
BOOSTER/AIR EJECTOR/LIQUID RING OR BOOSTER/BOOSTER/LIQUID RING COMBINATION
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I I LIQUID JET PUMP
I
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0001
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I t
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100
1
10
100
l
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P R E S S U R E IN T O R R ( m m Hg A b s o l u t e )
FIGURE 9-1
Typical pressure ranges for various vacuum pumping devices. (Source: Stokes Division of Pennwalt Corporation, Philadelphia, PA.)
The most important vacuum producers and their respective operating modes and features are of interest to us in the order listed in Figure 9-1.
SINGLE-STAGE LIQUID RING PUMPS Figure 9-3 depicts the operating principle of a liquid ring pump. Its circular pump body (A) contains a rotor that consists of a shaft and impeller (B). Shaft and impeller center-lines are positioned parallel, but eccentrically offset relative to the centerline of the pump body. The amount of eccentricity is related to the depth of the liquid ring (C). The liquid ring is formed by introducing service liquid, normally water, via the pump suction casing (L) and through the channel (D) positioned in the suction port plate (E). The centrifugal action of the rotating impeller forces the liquid toward the periphery of the pump body. By controlling the amount of service liquid within the pump body where the impeller blades are completely immersed to their root at one extreme (F) and all but their tips exposed at the other extreme (G), optimum pumping performance will be attained. When this pumping action is achieved, the vapor to be handled is induced through the suction port (H) when the depth of impeller blade immersion is being decreased. Then as the immersion increases, the vapor is compressed and discharged through the discharge port (J) in the intermediate port plate (K). As there is no metalto-metal contact between the impeller and the pump body and intermediate plates, the need for lubrication is eliminated and wear is reduced to a minimum.
Vacuum Pumps
339
FIGURE 9-2 Operating range of large mechanical vacuum pumps. (Source: Stokes Division of Pennwalt Corporation, Philadelphia, PA.) During the compression cycle heat is being imparted to the liquid ring. In order to maintain a temperature below the vapor point, cooling must be applied. This cooling is achieved by continuously adding a cool supply of service liquid to the liquid ring. The amount of coolant added is equal to that discharged through the discharge port (J) together with the compressed vapor. The mixture of vapor
340
Process Plant Machinery
FIGURE 9-3
Island, NY. )
Operating principle of liquid vacuum pumps. (Source: SIHI Pumps, Inc., Grand
FIGURE 9-4 Liquid ring vacuum pumping system with full sealant recovery. (Source: Kinney Vacuum Company, Boston, MA.) and liquid is then passed to subsequent stages and eventually through the pump discharge for separation. An entire vacuum pumping system is shown in Figure 9-4. This so-called full sealant recovery system is used to conserve sealant and/or where suitable or compatible sealant is not available from an outside source. Periodic sealant makeup
Vacuum Pumps
FIGURE 9-5
341
(A) Typical liquid jet vacuum pump. (B) Cutaway view of liquid vacuum pump. (Source: Kinney Vacuum Company, Boston, MA.)
and/or purge may be required. Full recirculation of sealant is provided from the discharge separator tank. Cooling is provided by running recirculated sealant through a heat exchanger. Separate cooling liquid or gas is required.
LIQUID JET VACUUM PUMPS A typical liquid jet pump is illustrated in Figure 9-5. A centrifugal pump circulates water (the usual hurling liquid) through the multijet nozzle and venturi and returns it to the separation chamber. The water, forced at high velocity across the gap between the nozzle and venturi, entrains the air and gases in multiple jet streams, creating a smooth, steady vacuum in the air suction line and vacuum system. This mixture is discharged through the venturi tangentially into the separation chamber, causing the water in the separation chamber to rotate, which results in a centrifugal action that forces the water to the periphery of the chamber, while the air is separated and discharged. When the hurling liquid is water, it is cooled by a continuous flow of cooling water into the separation chamber. Where process requirements allow and economy is an important factor, automatic controls and other cooling methods are often utilized.
AIR EJECTOR AND/OR BOOSTER LIQUID RING PUMPS Air ejectors or mechanical booster pumps are often used upstream of liquid ring pumps for applications requiring higher pumping speeds and lower pressures. These systems are particularly suited for processes where freedom from oil in the pumping system is required, such as oxygen handling, or where pump oil contamination makes the use of oil-sealed pumps impractical or expensive.
342
Process Plant Machinery Ejector/Liquid Ring
Booster/Liquid Ring
Motive Air Inlet
SystemInlet
~ 9
End view of air ejector backed by a liquid ring vacuum pump.
End view of mechanical booster pump backed by liquid ring vacuum pump.
Typical Combinations of liquid Ring Pumps with Mechanical Booeters and Air Ejectors. P U M P I N G C A P A C I T Y (SPEED) VS. P R E S S U R E
Sealing Water Inlet Temperature 6 0 ~ _
g. 300 L)
z
A C3 UJ I.U
~,oo L) ,,r
(9
_z
o.
~ o o - -
n
o
to
too
~Go
P R E S S U R E IN T O R R (mmHg Absolute)
FIGURE 9-6
Liquid ring pumps with air ejector and mechanical booster. (Source: Kinney Vacuum Company,
Boston, MA.)
Figure 9-6 shows how the addition of air ejectors or mechanical booster pumps can extend the range of liquid ring vacuum pumps.
MULTISTAGE COMBINATION UNITS Multistage vacuum units consist of one or two rotary blowers, the necessary intermediate gas coolers, possibly a gas ejector, and a liquid ring vacuum pump. All this equipment is typically mounted on a single-base frame.
Vacuum Pumps
343
The entire system consists of the rotary blower(s), gas ejector, and liquid ring vacuum pump are connected in series, whereby the extracted gases flow first through the blower, then the gas ejector, and finally the liquid ring vacuum pump. The latter can be operated with fresh water or with a closed ring-liquid circuit. This is always indicated when the evacuated gases and vapors are not allowed to be routed to the atmosphere or discharged into the sewage system for environmental protection reasons, or if the vapors contain valuable raw materials that have to be recovered by means of condensation. In such cases, the self-priming liquid ring vacuum pump draws the liquid out of the tank. The gas and liquid are then separated and cooled in the heat exchanger. The liquid returns to the tank and the gas passes through the exhaust gas cooler, where the condensibles are recovered before leaving the system. ROTARY OIL-SEALED V A C U U M PUMPS
Of the variety of mechanical vacuum pumps described earlier and illustrated in Figure 9-2, the unit most commonly used for high-vacuum work is the rotary piston, oil-sealed, single-stage mechanical vacuum pump (Figure 9-7). This pump type, sometimes called a cam and piston pump, is typically capable of producing ultimate pressures below 10 microns. Its normal operating range is between 0.05 Torr and 100 Torr, and typical evacuation rates range from 10 to 1000 cubic centimeters per minute. The rotary oil-sealed pump employs an eccentrically mounted piston-slide assembly, rotating at speeds of 400 RPM and higher, depending on pump size.
EXHAUST ' I
~
OtL SEPARATOR
POPPET VALVE
] N
SHAFT
i
iNLET
HINGE BAR PISTON a SLIDE
ECCENTRIC
.[
FIGURE 9-7 Rotary piston oil-sealed vacuum pump. (Source: Stokes Division of Pennwalt Corporation, Philadelphia, PA.)
344
ProcessPlant Machinery
The vacuum "seal" is maintained by a "wedge" of oil ahead of the piston. The oil wedge prevents blow-by of gases and provides lubrication. In operation, the gas to be evacuated enters the rotary oil-sealed vacuum pump freely through the intake port of the piston-slide. As the oil-sealed piston revolves, it closes the inlet port from the vacuum system and traps ahead of it all air or gas that has entered. With each revolution of the pump, the gas is compressed and discharged through the valve ports to the atmosphere or to a collecting system.
Chapter 10 Cooling Water Supply Systems The machinery most closely associated with many cooling water supply systems consists of pumps and fans. In a modem process plant, machines can take on a variety of configurations and range from small "office-size" and inexpensive to extremely large and expensive. We have covered pumps in Chapter 7 and will, therefore, concentrate on cooling tower fans and their drive systems. Mechanical draft towers use either single or multiple fans to provide flow of a known volume of air through the tower. Thus their thermal performance tends toward greater stability and is affected by fewer psychrometric variables than that of the atmospheric towers. The presence of fans also provides a means of regulating air flow, to compensate for changing atmospheric and load conditions, by fan capacity manipulation and/or cycling. Mechanical draft towers are categorized as either forced draft (Figure 10-1), on which the fan is located in the ambient air stream entering the tower and the air is blown through, or induced draft (Figure 10-2), wherein a fan located in the exiting air stream draws air through the tower. Forced draft towers are characterized by high air entrance velocities and low exit velocities. Accordingly, they are extremely susceptible to recirculation and are therefore considered to have less performance stability than the induced draft.
Water
e......
Air T
Out
T
Air In
,, Water Out
FIGURE 10-1
Forced draft, counterflow, blower fan tower (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.) * Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO. Adapted by permission.
345
346
Process Plant Machinery
FIGURE 10-2 Induced draft, crossflow, propeller fan tower. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.) Furthermore, located in the cold entering ambient air stream, forced draft fans can become subject to severe icing (with resultant imbalance) when moving air laden with either natural or recirculated moisture. Usually, forced draft towers are equipped with centrifugal blower type fans, which, although requiting considerably more horsepower than propeller type fans, have the advantage of being able to operate against the high static pressures associated with ductwork. Therefore, they can either be installed indoors (space permitting) or within a specially designed enclosure that provides significant separation between intake and discharge locations to miminize recirculation. Induced draft towers have an air discharge velocity of from three to four times higher than their air entrance velocity, with the entrance velocity approximating that of a 5-mph wind. Therefore, there is little or no tendency for a reduced pressure zone to be created at the air inlets by the action of the fan alone. The potential for recirculation on an induced draft tower is not self-initiating and, therefore, can be more easily quantified purely on the basis of ambient wind conditions. Location of the fan within the warm air stream provides excellent protection against the formation of ice on the mechanical components. Widespread acceptance of induced draft towers is evidenced by their existence on installations as small as 15 gallons per minute (gpm) and as large as 700,000 gpm. Hybrid draft towers (Figure 10-3) can give the outward appearance of being natural draft towers with relatively short stacks. Internal inspection (Figure 10-4), however, reveals that they are also equipped with mechanical draft fans to augment air flow. Consequently, they are also referred to as fan-assisted natural draft towers. The intent of their design is to minimize the horsepower required for air movement, but to do so with the least possible stack cost impact. Properly designed, the fans may need to be operated only during periods of high ambient and peak loads. In localities where a low-level discharge of the tower plume may prove to be unacceptable, the elevated discharge of a fan-assisted natural draft tower can become sufficient justification for its use.
Cooling Water Supply Systems
FIGURE 10-3 Fan-assisted natural draft tower. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.)
FIGURE 10-4 Cutaway of fan-assisted draft tower. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.) CHARACTERIZATION BY AIR FLOW
Cooling towers are also classified by the relative flow relationship of air and water within the tower, as follows: In counterflow towers (Figure 10-5), air moves vertically upward through the fill, counter to the downward fall of water. Because of the need for extended intake and discharge plenums, the use of high pressure spray systems, and the typically higher air pressure losses, some of the smaller counterflow towers are physically higher, require more pump head, and utilize more fan power than their crossflow counterparts. In larger counterflow towers, however, the use of low-pressure, gravity-related distribution systems, plus the availability of generous intake areas and plenum spaces for air management, is tending to equalize, or even reverse, this situation. The enclosed nature of a counterflow tower also restricts exposure of the water to direct sunlight, thereby retarding the growth of algae. Crossflow towers (Figure 10-6) have a fill configuration through which the air flows horizontally, across the downward fall of water. Water to be cooled is delivered to hot water inlet basins located atop the fill areas and is distributed to the
347
348
ProcessPlant Machinery Air ~
,
.
.
A
A
Wate~ In
Air
In
FIGURE 10-5 Induced draft, counterflow tower (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.) Air
,.=
9
_
Water In
Air In
Water Out
FIGURE 10-6 Double-flow,crossflow tower. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.) fill by gravity through metering orifices in the floor of those basins. This obviates the need for a pressure-spray distribution system and places the resultant gravity system in full view for ease of maintenance. By the proper utilization of flow-control valves, routine cleaning and maintenance of crossflow tower distribution systems can be accomplished sectionally, while the tower continues to operate. Crossflow towers are also subclassified by the number of fill "banks" and air inlets that are served by each fan. The tower indicated in Figure 10-6 is a doubleflow tower because the fan is inducing air through two inlets and across two banks of fill. Figure 10-7 depicts a single-flow tower having only one air inlet and one fill bank, the remaining three sides of the tower being cased. Single-flow towers are customarily used in locations where an unrestricted air path to the tower is available from only one direction. They are also useful in areas having a dependable prevailing wind direction, where consistent process temperatures are critical. The tower can be sited with the air inlet facing the prevailing wind, and any potential for recirculation is negated by the downwind side of the tower being a cased face. Spray-filled towers have no heat transfer (fill) surface, depending only upon the
Cooling Water Supply Systems Air Out
Y
Water In
Air
. .
Water Out
'
9
' -
~
J
, ...........
FIGURE 10-7 Single-flow tower. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.) water break-up afforded by the distribution system to promote maximum water-toair contact. Removing the fill from the tower in Figure 10-5 would also make it "spray-filled." The use of such towers is normally limited to those processes where higher water temperatures are permissible. They are also utilized in those situations where excessive contaminants or solids in the circulating water would jeopardize a normal heat transfer surface. CHARACTERIZATION BY CONSTRUCTION
Field-erected towers are those on which the primary construction activity takes place at the site of ultimate use. All large towers, and many of the smaller towers, are prefabricated, piece-marked, and shipped to the site for final assembly. Labor and/or supervision for final assembly is usually provided by the cooling tower manufacturer. Factory-assembled towers undergo virtually complete assembly at their point of manufacture, whereupon they are shipped to the site in as few sections as the mode of transportation will permit. A relatively small tower would ship essentially intact. Larger, multicell towers are assembled as "cells" or "modules" at the factory and are shipped with appropriate hardware for ultimate joining by the user. Factoryassembled towers are also known as "packaged" or "unitary" towers. CHARACTERIZATION BY SHAPE
Rectilinear towers (Figure 10-8) are constructed in cellular fashion, increasing linearly to the length and number of cells necessary to accomplish a specified thermal performance. Round Mechanical Draft (RMD) towers, as the name implies, are essentially round in plan configuration, with fans clustered as close as practicable around the centerpoint of the tower. Multifaceted towers, such as the octagonal mechanical draft (OMD) depicted in Figure 10-9, also fall into the general classification of "round" towers. Such towers can handle enormous heat loads with considerably less site area impact than that required by multiple rectilinear towers. Additionally, they are significantly less affected by recirculation.
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Process Plant Machinery
FIGURE 10-8 Multicelled, field-erected, crossflow cooling tower with enclosed stairway and extended fan deck to enclose piping and hot water basins. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.)
FIGURE 10-9 Octagonal mechanical draft (OMD) counterflow cooling tower. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.) CHARACTERIZATION BY METHOD OF HEAT TRANSFER
All of the cooling towers heretofore described are evaporative-type towers, in that they derive their primary cooling effect from the evaporation that takes place when air and water are brought into direct contact. At the other end of the spectrum is the dry tower (Figure 10-10), where, by full utilization of dry surface coil sections, no direct contact (and no evaporation) occurs between air and water. Hence the water is cooled totally by sensible heat transfer. In between these extremes are the plume abatement (Figure 10-11) and water conservation (Figure 10-12) towers, wherein progressively greater portions of dry surface coil sections are introduced into the overall heat transfer system to alleviate specific problems, or to accomplish specific requirements.
Cooling Water Supply Systems
351
Ill "--" VENT MANIFOLD SYSTEM
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t; FIGURE 10-10 Dry-type cooling tower, cross-sectional elevation. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.) 30UTLs AIR
SUPERSATUUTION (Foe) AREA
j~' ~/"
3
SUPER HEAT (Non-Fog) AREA
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r H HC
~ ER
Z
t 1 INLETAIR
DRY BULB TEMPERATURE (OF)
1 INLET AIR FILL
-1L
.... t
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COLDWA rER 4 ~
FIGURE 10-11 Plume abatement tower and psychrometrics (coil before fill). (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.) MECHANICAL COMPONENT REVIEW
Cooling tower mechanical equipment is required to operate within a highly corrosive moisture-laden atmosphere that is unique to the cooling tower industry, and the historical failure rate of commercially available components caused reputable tower manufacturers to undertake their own production of specific components some years ago. Currently, the low failure rate of manufacturer-produced components reinforces that decision. Purchasers also benefit from the advantage of single-source responsibility for warranty and replacement parts. Exclusive of motors, the mechanical components basic to the operation of the cooling tower are fans, speed reducers, drive shafts, and water flow control valves.
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P, - - Q
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----4
FIGURE 10-12 Water conservation cooling tower. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.) Fans
Cooling tower fans must move large volumes of air efficiently and with minimum vibration. The materials of manufacture must not only be compatible with their design, but must also be capable of withstanding the corrosive effects of the environment in which the fans are required to operate. Their importance to the ability of mechanical draft cooling towers to perform is reflected in the fact that fans of improved efficiency and reliability are the object of continuous development.
Cooling Water Supply Systems
Propeller Fans Propeller-type fans predominate in the cooling tower industry because of their ability to move vast quantities of air at the relatively low static pressures encountered. They are comparatively inexpensive, may be used on any size tower, and can develop high overall efficiencies when "system designed" to complement a specific tower structure- f i l l - fan cylinder configuration. Most-utilized diameters range from 24 inches to 10 meters (Figure 10-13), operating at horsepowers from 1/4 to 250+. Although the use of larger fans, at higher power input, is not without precedence, their application naturally tends to be limited by the number of projects of sufficient size to warrant their consideration. Fans 48 inches and larger in diameter are equipped with adjustable pitch blades, enabling the fans to be applied over a wide range of operating horsepowers. Thus the fan can be adjusted to deliver the precise required amount of air at the least power consumption. The rotational speed at which a propeller fan is applied typically varies in inverse proportion to its diameter. The smaller fans turn at relatively high speeds, whereas the larger ones turn somewhat slower. This speed-diameter relationship, however, is by no means a constant one. If it were, the blade tip speeds of all cooling tower fans would be equal. The applied rotational speed of propeller fans usually depends on best ultimate efficiency, and some diameters operate routinely at tip speeds approaching 14,000 feet per minute. However, since higher tip speeds are associated with higher sound levels, it is sometimes necessary to select fans turning at slower speeds to satisfy a critical requirement. The increased emphasis on reducing cooling tower operating costs has resulted in the use of larger fans to move greater volumes of air more efficiently. Much research has also gone into the development of more efficient blade, hub, and fan cylinder designs. The new generations of fans are light in weight to reduce parasitic energy losses, and have fewer, but wider, blades to reduce aerodynamic drag. Moreover, the characteristics of air flow through the tower, from inlet to discharge, are analyzed and appropriate adjustments to the structure are made to minimize obstructions; fill and distribution systems are designed and arranged to promote maximum uniformity of air and water flow; and drift eliminators are arranged to direct the final pass of air toward the fan. This is recognized as the "systems" approach to fan design, without which the best possible efficiency cannot be obtained.
FIGURE 10-13 Typical large-diameter fan utilized on cooling towers. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.)
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Process Plant Machinery The intent of good propeller fan design is to achieve air velocities across the effective area of the fan, from hub to blade tips, that are as uniform as possible. The most effective way to accomplish this is with tapered and twisted blades having an airfoil cross section. Historically, cast aluminum alloys have been the classic materials used for production of this blade type. Cast aluminum blades continue to be utilized because of their relatively low cost, good internal vibration damping characteristics, and resistance to corrosion in most cooling tower environments. Currently, lighter blades of exceptional corrosion resistance are made of fiberglass-reinforced plastic, cast in precision molds. These blades may be solid, formed around a permanent core, or formed hollow by the use of a temporary core. In all cases, they have proved to be both efficient and durable as long as the design avoided aerodynamically induced vibration resonance. Fan hubs must be of a material that is structurally compatible with blade weight and loading, and must have good corrosion resistance. Galvanized steel weldments, gray and ductile iron castings, and wrought or cast aluminum are in general use as hub materials. Where hub and blades are of dissimilar metals, they must be insulated from each other to prevent electrolytic corrosion. Smaller diameter fans are customarily of galvanized sheet metal construction with fixed-pitch nonadjustable blades. These fans are matched to differing air flow requirements by changing the design speed.
Automatic Variable-Pitch Fans These are propeller fans on which a pneumatically actuated hub controls the pitch of the blades in unison (Figure 10-14). Their ability to vary airflow through the tower in response to a changing load or ambient condition- coupled with the resultant energy savings and ice control - make them an optional feature much in demand.
Centrifugal Fans These are usually of the double inlet type, used predominantly on cooling towers designed for indoor installations (Figure 10-15). Their capability to operate against
FIGURE 10-14 Automatic, variable-pitch fan used to adjust air flow and fan horsepower, (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.)
Cooling Water Supply Systems
FIGURE 10-15 Blower-type cooling tower fan. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.) relatively high static pressures makes them particularly suitable for that type of application. However, their inability to handle large volumes of air, and their characteristically high input horsepower requirement (approximately twice that of a propeller fan), limit their use of relatively small applications. Three types of centrifugal fans are available: (1) forward curved blade fans, (2) radial blade fans, and (3)backward curved blade fans. The characteristics of the forward curved blade fan make it the most appropriate type of cooling tower service. By virtue of the direction and velocity of the air leaving the fan wheel, the fan can be equipped with a comparatively small size housing, which is desirable from a structural standpoint. Also, because the required velocity is generated at a comparatively low speed, forward curved blade fans tend to operate quieter than other centrifugal types. Centrifugal fans are usually of sheet metal construction, with the most popular protective coating being hot-dip galvanization. Damper mechanisms are also available to facilitate capacity control of the cooling tower. Fan Laws
All propeller-type fans operate in accordance with common laws. For a given fan and cooling tower system, the following is true: 1. The capacity (cfm) varies directly as the speed (RPM) ratio, and directly as the pitch angle of the blades relative to the plane of rotation. 2. The static pressure (hs) varies as the square of the capacity ratio. 3. The fan horsepower varies as the cube of the capacity ratio. 4. At constant cfm, the fan horsepower and static pressure vary directly with the air density. If, for example, the capacity (cfm) of a given fan were decreased by 50 percent (either by a reduction to half of design rpm, or by a reduction in blade pitch angle at constant speed), the capacity ratio would be 0.5. Concurrently, the static pressure would become 25 percent of before, and the fan horsepower would become 12.5 percent of before. These characteristics afford unique opportunities to combine cold water temperature control with significant energy savings.
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Selected formulas, derived from these basic laws, may be utilized to determine the efficacy of any particular fan application:
Symbols Q = Volume of air handled (cfm). Unit: cu ft per min. A -- Net flow area. Unit: sq ft. V -- Average air velocity at plane of measurement. Unit: ft per sec. g = Acceleration due to gravity. Unit: 32.17 ft per sec per sec. D ----Density of water at gauge fluid temperature. Unit: lb per cu ft. d = Air density at point of flow. Unit: lb per cu ft. hs -- Static pressure drop through system. Unit: inches of water. hv -- Velocity pressure at point of measurement. Unit: inches of water. ht = Total pressure differential (= hs + hv). Unit: inches of water. Vr = Fan cylinder velocity recovery capability. Unit: percent. Thermal performance of a cooling tower depends on a specific mass flow rate of air through the fill (pounds of dry air per minute), whereas the fan does its job purely in terms of volume (cubic feet per minute). Since the specific volume of air (cubic feet per pound) increases with temperature, it can be seen that a larger volume of air leaves the tower than enters it. The actual cfm handled by the fan is the product of mass flow rate times the specific volume of dry air corresponding to the temperature at which the air leaves the tower. This volumetric flow rate is the "Q" used in the following formulas, and it must be sufficient to produce the correct mass flow rate or the tower will be short of thermal capacity. Utilizing appropriate cross-sectional flow areas, velocity through the fan and fan cylinder can be calculated as follows:
V
Q Ax60
It must be understood that "A" will change with the plane at which velocity is being calculated. Downstream of the fan, "A" is the gross cross-sectional area of the fan cylinder. At the fan, "A" is the area of the fan less the area of the hub or hub cover. Velocity pressure is calculated as follows"
hv
V2•
12xd
2xgxD
If V represents the velocity through the fan, then hv represents the velocity pressure for the fan itself (hvf). Moreover, if the fan is operating within a nonflareddischarge fan cylinder, this effectively represents the total velocity pressure because of no recovery having taken place. However, if the fan is operating within a flared, velocity-recovery-type fan cylinder (Figure 10-16), hv must be recalculated for the fan cylinder exit (hve), at the appropriate velocity, and applied in the following formula to determine total velocity pressure: hv = hvf - [(hvf - hve) x Vr] Although the value of Vr will vary with design expertise and is empirically established, a value of 0.75 (75 percent recovery) is normally assigned for purposes of anticipating fan performance within a reasonably well-designed velocity-recovery cylinder.
Cooling Water Supply Systems
) FIGURE 10-16 Cutawayview of a velocity-recovery-type fan cylinder. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.) The power output of a fan is expressed in terms of air horsepower (ahp) and represents work done by the fan: QxhtxD ahp = 33,000 x 12 Static air horsepower is obtained by substituting static pressure (hs) for total pressure (ht) in the formula. A great deal of research and development goes into the improvement of fan efficiencies, and those manufacturers that have taken a systems approach to this research and development effort have achieved results that, although incrementally small, are highly significant in the light of current energy costs. Static efficiencies and overall mechanical (total) efficiencies are considered in the selection of a particular fan in a specific situation, with the choice usually going to the fan that delivers the required volume of air at the least input horsepower: Static Efficiency -
static ahp input hp
ahp Total Efficiency = . mput hp It must be understood that input hp is measured at the fan shaft and does not include the drive-train losses reflected in actual motor brake horsepower (bhp). Input hp will normally average approximately 95 percent of motor bhp on larger fan applications.
Speed Reducers The optimum speed of a cooling tower fan seldom coincides with the most efficient speed of the driver (motor). This dictates that a speed reduction, power transmission unit of some sort be situated between the motor and the fan. In addition to reducing the speed of the motor to the proper fan speed (at the least possible loss of available power), the power transmission unit must also provide primary support for the fan, exhibit long-term resistance to wear and corrosion, and contribute as little as possible to the overall noise level. Speed reduction in cooling towers is accomplished either by differential gears of positive engagement or by differential pulleys (sheaves) connected through Vbelts. Typically, gear reduction units are applied through a wide range of horsepower ratings, from the very large down to as little as 5 hp. V-belt drives, on the other hand, are usually applied at ratings of 50 hp or less.
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Gear Reduction Units Gear speed reducers are available in a variety of designs and reduction ratios to accommodate the fan speeds and horsepowers encountered in cooling towers (Figure 10-17). Because of their ability to transmit power at minimal loss, spiral bevel and helical gear sets are most widely utilized, although worm gears are also used in some designs. Depending on the reduction ratio required and the input hp, a gear speed reducer may use a single type gear or a combination of types to achieve "staged" reduction. Generally, two-stage reduction units are utilized for the large, slower-turning fans requiting input horsepowers exceeding 75 bhp. The service life of a speed reducer or speed increaser is directly related to the surface durability of the gears, as well as the type of service imposed (i.e., intermittent versus continuous duty). The American Gear Manufacturers Association (AGMA) has established criteria for the rating of geared speed reducers, which are subscribed to by most reliable designers. AGMA Standard 420 defines these criteria and offers a list of suggested service factors. The gear speed reducer manufacturer should have established service factors for an array of ratios, horsepowers, and types of service, commensurate with good engineering practice. The Cooling Tower Institute (CTI) Standard 111 offers suggested service factors specifically for cooling tower applications. Long, trouble-free life is also dependent on the quality of bearings used. Bearings are normally selected for a calculated life compatible with the expected type of service. Bearings for industrial cooling tower gear speed reducers (considered as continuous duty) should be selected on the basis of a 100,000-hour L-10 life. L-10 life is defined as the life expectancy in hours during which 90 percent or more of a given group of bearings under specific loading condition will still be in service. Intermittent duty applications provide satisfactory life with a lower L-10 rating. An L-10 life of 35,000 hours is satisfactory for an 8- to 10-hour per day application. It is equivalent, in terms of years of service, to a 100,000-hour L-10 life for continuous duty. Lubrication aspects of a gear speed reducer, of course, are as important to longevity and reliability as are the components that compose the gear speed reducer. The lubrication system should be of a simple, noncomplex design, capable of lubricating equally well in both forward and reserve operation. Remote oil level indicators
FIGURE 10-17 Gear speed reducer used for applied horsepowers above 75 HP. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.)
Cooling Water Supply Systems
and convenient location of fill and drain lines simplify and encourage preventive maintenance. Lubricants and lubricating procedures recommended by the manufacturer should be adhered to closely. Synthetic lubricants may greatly increase both gear life and lubricant drainage intervals.
V-Belt Drives These are an accepted standard for the smaller factory-assembled cooling towers, although most of the larger unitary towers are equipped with gear speed reducers. Correctly designed and installed, and well maintained, V-belt drives can provide very dependable service. The drive consists of the motor and fan sheaves, the bearing housing assembly supporting the fan, and the V-belts. V-belts (as opposed to cog belts) are used most commonly for cooling tower service. A variety of V-belt designs is available, offering a wide assortment of features. Most of these designs are suitable for cooling tower use. In many cases, more than one belt is required to transmit power from the motor to the fan. Multiple belts must be supplied either as matched sets, measured and packaged together at the factory, or as a banded belt having more than one V-section on a common backing. Various types of bearings and bearing housing assemblies are utilized in conjunction with V-belt drives. Generally, sleeve bearings are used on smaller units and ball or roller bearings on the larger units, with oil being the most common lubricant. In all cases, water slinger seals are recommended to prevent moisture from entering the beating, and oil mist is often used as a purge medium to prevent moisture condensation in bearing housings. Belts wear and stretch, and belt tension must be periodically adjusted. Means for such adjustment should be incorporated as part of the motor mount assembly. Stability and strength of the mounting assembly is of prime importance in order to maintain proper alignment between driver and driven sheaves (Figure 10-18). Misalignment is one of the most common causes of excessive belt and sheave wear. Manually adjustable pitch sheaves are occasionally provided to allow a change in fan speed. These are of advantage on indoor towers, where the ability to adjust fan speed can sometimes compensate for unforeseen static pressure.
FIGURE 10-18 Adjustable motor mount for a V-belt driven fan (belt guard removed). (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.)
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Process Plant Machinery
Drive Shafts The drive shaft transmits power from the output shaft of the motor to the input shaft of the gear speed reducer. Because the drive shaft operates within the tower, it must be highly corrosion resistant. Turning at full motor speed, it must be well b a l a n c e d - and capable of being rebalanced. Transmitting full motor power over significant distances, it must accept tremendous torque without deformation. Subjected to long-term cyclical operation and occasional human error, it must be capable of accepting some degree of misalignment. Drive shafts are described as "floating" shafts, equipped with flexible couplings at both ends. Where only normal corrosion is anticipated and cost is of primary consideration, shafts are fabricated of carbon steel, hot-dip galvanized after fabrication (Figure 10-19). Shafts for larger industrial towers, and those that will be operating in atmospheres more conducive to corrosion, are usually fabricated of tubular stainless steel (Figure 10-20) or epoxy-coated wound carbon filament. The yokes and flanges that connect to the motor and gear speed reducer shafts are of cast or welded construction, in a variety of materials compatible with that utilized for the shaft. Flexible couplings transmit the load between the driveshaft and the motor or gear speed reducer, and compensate for minor misalignment. A suitable material for use in the saturated effluent air stream of a cooling tower is neoprene, either in solid grommet form (Figure 10-21) or as neoprene-impregnated fabric (Figure 10-22), designed to require no lubrication and relatively little maintenance. Excellent service records have been established by the neoprene flexible couplings, both as bonded bushings and as impregnated fabric disc assemblies. These couplings are virtually impervious to corrosion and provide excellent flexing characteristics. It is very important that drive shafts be properly balanced. Imbalance not only causes tower vibration, but it also induces higher loads and excessive wear on the mechanical equipment coupled to the shaft. Most cooling tower drive shafts operate at speeds approaching 1800 RPM. At these speeds, it is necessary that the shafts be dynamically balanced to reduce vibrational forces to a minimum.
FIGURE 10-19 Driveshafi in a relatively small fan drive application. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.)
Cooling Water Supply Systems
FIGURE 10-20 Close-upof a larger driveshaft showing guard. (Source: The Marley Cooling Tower Company, Mission, KS, and Kansas City, MO.)
FIGURE 10-21
FIGURE 10-22
Safety Considerations
Fan cylinders less than six feet high must be equipped with suitable fan guards for the protection of operating personnel. Drive shafts must operate within retaining guards (Figure 10-20) to prevent the drive shaft from encountering the fan if the coupling should fail. The motor shaft and outboard drive shaft coupling should either be within the confines of the fan cylinder or enclosed within a suitable guard.
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Chapter 11 Centrifugal Compressors OVERVIEW OF GAS COMPRESSION MACHINERY
The purpose of compression is simply to increase the pressure of a gas from one level to another. Depending on a host of circumstances and situations, the pressure increase imparted to a gas will be from a fraction of a pound per square inch (psi) (or a few pascals) in laboratory equipment to literally tens of thousands of psi in hypercompressors used for the manufacture of polyethylene. Before we embark on our more thorough consideration of centrifugal compressors, we should examine gas compression machinery in general. In a typical process plant, compression services include instrument and plant air, combustion air for burners and furnaces, gas circulation, or simple elevation to pressure conditions that will allow chemical reactions to take place. Gas volumes will vary from laboratory quantities to flows well in excess of a million cubic feet per minute (~2 million m3/hr). Two principal methods are used to compress gases. The first method is to trap a volume of gas and displace it by the positive action of a piston or rotating member; we call these machines positive-displacement compressors. The second method uses dynamic compression; it is accomplished either by the mechanical action of contoured blades, which impart velocity and hence pressure to the flowing gas, or by the entrainment of gases in a high velocity jet of the same or another gas (e.g., steam). The entrainment principle is generally found in ejectors that operate with inlet pressures below atmospheric pressure, but since our definition of process machinery implies the existence of moving parts, we have elected not to cover ejector devices in this text. The two major groupings of compression machinery, positive displacement and dynamic, can be further subdivided as shown in Figure 11-1. Positive-displacement units are those in which successive volumes of gas are confined within a closed space and elevated to a higher pressure. Reciprocating compressors are positivedisplacement machines in which the compressing and displacing element is a piston that moves back and forth within a cylinder. Rotary positive-displacement compressors are machines in which compression and displacement is effected by the close meshing of rotating elements. Sliding-vane compressors are rotary positive-displacement machines in which axial vanes slide radially in a rotor. This rotor is eccentrically mounted in a cylindrical casing. Gas trapped between the vanes is compressed and displaced. Liquid-piston compressors are rotary positivedisplacement machines in which water or other liquid is used as the piston to compress and displace the gas handled. Roots| straight-lobe compressors are rotary positive-displacement machines in which two mating lobed impellers trap gas and displace it from intake to discharge. There is no internal compression. Helicalor spiral-lobe compressors are rotary positive-displacement machines in which two intermeshing rotors, each with a helical form, compress and displace the gas. 363
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Process Plant Machinery COMPRESSORS
DYNAMIC
POSITIVE--DISPLACEMENT
ROTARY
RECIPROCATING
ROTATING
EJECTOR (NON-ROTATING)
ROOTS-TYPE SLIDING-VANE
STRAIGHT-LOBE
LIQUID-PISTON
CENTRIFUGAL
HELICAL-LOBE
MIXED-FLOW
AXIAL-FLOW
FIGURE 11-1 Major compressor types Dynamic compressors are continuous-flow machines in which either the rapidly rotating element or ejector nozzle accelerates the gas as it passes through the element. The velocity head is converted into pressure, partially in the rotating element and partially in stationary diffusers or blades. Centrifugal compressors are dynamic machines in which one or more rotating impellers, usually shrouded on the sides, accelerate the gas. Main gas flow is radial. Axial compressors are dynamic machines in which gas acceleration is obtained by the action of the bladed rotor shrouded on the blade ends. Main gas flow is axial. Mixed-flow compressors are dynamic machines with an impeller form combining some characteristics of both the centrifugal and axial types. Here we will treat centrifugal compressors. Other types of compressors are covered in succeeding chapters.
CENTRIFUGAL COMPRESSORS* Centrifugal compressors are employed in numerous fields" chemical and petrochemical industries, refineries and fertilizer plants, nuclear reactors and air separation plants, iron and steelworks, production of liquefied natural gas (LNG) and substitute natural gas (SNG), cryogenic and refrigeration plants, mining, transportation and storage of gas, onshore and offshore installations. The range of applications can be expanded still further by combining these centrifugal compressors with other compressor types such as axial-flow or reciprocating compressors. The wide range of processes in which centrifugal compressors are employed makes varying demands on these machines. Compressor design is dependent on such factors as the fluid handled, the pressure ratio, the volume flow, the number of interstage coolers, injection and extraction of the medium, and the type of shaft sealing. * Source: Mannesmann-Demag, Duisburg, Germany. Reprinted and adapted from copyrighted material, by permission.
Centrifugal Compressors
FIGURE 11-2 Horizontally split compressors in a modem process plant. (Source: Mannnesmann Demag, Duisburg, Germany.) Taking all these factors into consideration, the major compressor manufacturers have developed series of centrifugal compressors offering an optimum engineering solution implemented by the use of standard components. These series include the two basic types, distinguished by horizontally or vertically split casing (see Figures 11-2 and 11-3), compressors with two or three pairs of main nozzles, and compressors with additional sidestream nozzles. A few of the many possible nozzle arrangements are depicted in Figure 11-4. Horizontally split casings with nozzles in the lower half permit simple removal of the rotor and facilitate the checking of labyrinth clearances and O-rings. As pressure levels rise and gas molecules become smaller, vertically split casings are employed.
Horizontally Split Centrifugals Centrifugal compressors with horizontally split casings typically permit internal pressures of 70 bar with small volume flow rates and volume flow rates of up to 300,000 ma/hr at low pressures. Drive ratings of 30 MW for single-casing machines have already been implemented. Figure 11-5 shows a cross section of a six-stage horizontally split centrifugal compressor. Standardized components ensure high availability and easy fitting. The two halves of the casing are sealed and bolted together. The rigid structure is supported at the centerline, thus preventing vertical shifting of the compressor shaft as a result of thermal expansion. For erection and dismantling purposes, the top half of the casing, complete with the associated stationary components, can be handled as a single unit. All types of drivers can be employed, for example, gas turbines, steam turbines, and electric motors. Many processes require compression of a fluid in one process stage only, i.e., continuous compression from the first to the final stage with constant mass flow rate. Most major centrifugal compressor manufacturers build machines for this field
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Process Plant Machinery
FIGURE 11-3 Vertically split (barrel-type) centrifugal compressor. (Source: Mannesmann Demag, Duisburg, Germany.) of application with up to nine or as many as twelve compression stages. Aerodynamic matching of the individual compression stages is by means of diaphragms with diffuser channels and vaned return passages. Following the final stage, the compressed gas enters a collecting chamber in the form of a volute before it reaches the discharge nozzle. The shaft is supported in bearings outside the compression space. Shaft sealing is by means of tried and proven systems such as labyrinth, floating ring, mechanical contact or non-contact seals. The wide range of possible variations in the materials used and in the selection of the sealing system render compressors of this series suitable for virtually all fields of application in industry, chemical and petrochemical processes, and for almost all gases and mixtures of gases.
Sidestream Compressors In multistage refrigeration processes, different mass flows pass through the various refrigeration stages. Sidestreams therefore have to be injected into or extracted from the main flow in the compressor at process-dependent intermediate pressures and
Centrifugal Compressors
~ .
R..... -I"....l....... E..... F-..~..ll~
~---r-1
~
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FIGURE 11-4 Nozzle arrangements and impeller lineups typically available in centrifugal compressors. (A) Straight-through, (B) back-to-back, (C) back-to-back with sidestream entry, (D) straight-through (barreltype), (E) back-to-back (barrel-type), (F) future addition to stage possible, (G) single stage. (Source: Mannesmann Demag, Duisburg, Germany.)
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,3
FIGURE 11-5 Horizontally split compressor with two sidestream entry nozzles. (Source: Mannesmann Demag, Duisburg, Germany.) temperatures. Injection or extraction is by means of additional nozzles. (The flow path through a sidestream compressor is illustrated in Figure 11-6.) In the case of gas injection, the sidestream is mixed with the main stream in the return channel. Mixing takes place over the entire periphery. When a sidestream is extracted, a separating volute removes part of the main stream. The compressor stages are designed to correspond to the stages of the refrigeration process. The working media for refrigeration processes are primarily ethylene, propylene, ethane, and propane. Refrigeration processes usually form a closed cycle, rather less frequently a semi-open cycle. In centrifugal compressors with two main nozzle pairs, the two process stages can be arranged back to back, i.e., flow in the two process stages is in opposite directions, or they can be arranged in series. In the double-flow version, the compression process of both stages terminates in a common discharge nozzle. Back-to-back arrangement of the first and third stages or series arrangement of the stages is also possible. Typical back-to-back arrangements are shown in Figure 11-7. Tried and proven labyrinth seals separate the individual process stages. The choice of shaft seal system is dictated by the service. Whereas casings with two main nozzle pairs are widely employed for a variety of media and processes, casings
! I ! | i il~
FIGURE 11-6 Flow path through horizontally split straight-through compressor (left) and sidestream compressor (right). (Source: Mannesmann Demag, Duisburg, Germany.)
Centrifugal Compressors
FIGURE 11-7 Compressors with back-to-back oriented impellers. (Source: Mannesmann Demag, Duisburg, Germany.)
with three process stages are mainly employed for air, oxygen, and nitrogen. The medium is normally cooled outside the compressor. Interstage cooling produces an almost isothermal compression process. This requires the least compression work. Intercooling also becomes necessary when the temperatures produced by compression have to be limited. In most compression systems, the coolers are mounted separately, permitting a high degree of freedom in design and layout. However, compressors with internally arranged coolers are available from some manufacturers and may merit consideration when the ultimate in compactness must be achieved.
Vertically Split Compressors Vertically split (barrel-type)centrifugal compressors are the preferred, and sometimes mandatory, design for high pressures or for compressing gases rich in hydrogen. The cylindrical casing ensures good stress distribution and extremely good gas-tightness. Unlike the casing, the stationary internal components of the compressor, with the exception of seal components, are horizontally split. During assembly of the compressor they are mounted together with the rotor and inserted axially into the casing. The end covers are retained by shear ring segments. A cross-section view is shown in Figure 11-8. The inlet and discharge nozzles are welded to the cylindrical casing or, where heavy wall thicknesses are involved, are integral with the casing; the pipework is bolted to these nozzles. These compressors are also built for two process stages; in this case they feature two main nozzle pairs. The main fields of application for barrel-type compressors are in handling gases rich in hydrogen; hydrogenation cracking; synthesis of ammonia, urea, and methanol; gas lift and reinjection; and transportation of gas in pipelines. A compressor with one stage is often adequate for compression applications involving a low head. For such applications, the user may choose from compressor types that may be vertically or horizontally split. As mentioned earlier, the vertically split version is particularly suitable for high pressures and for compressing gases of low molecular weight. Depending on the operating conditions involved, the tried and proven systems employing labyrinth, mechanical contact, or floating ring seals are used for shaft sealing. Other seal systems may go under the names of trapped bushing seal, gas phase mechanical seal, cone seal, etc., but each type represents merely a variation of the three principal configurations. Fields of application for single-stage process compressors include phthalic acid anhydride plants and cupolas, oxosynthesis, and water treatment. In addition, they
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are suitable for handling all gases and gas mixtures in the main and auxiliary processes in the chemical and petrochemical industries. They also find application in the transportation of natural gas and mineral oil gas as well as in closed-loop systems in nuclear technology.
Compressor Trains Large pressure ratios cannot be handled by one casing alone. Similarly, it is not possible to split the compression cycle into more than two or three process stages within one casing. The major compressor manufacturers therefore build compressor trains that may consist of up to four separate casings. A train with three separate casings is shown in Figure 11-9. These separate compressors, which need not be of the same type, are interconnected by couplings; they can be powered by a common driver. When additional transmission gearing is used, the compressor casings may also be run at different speeds. The train is designed so that a minimum of dismantling is necessary for maintenance, i.e., when a vertically split casing is used, it is located at the opposite end to the driver. The compressor train can be arranged on a common baseframe. In special applications, for instance, offshore installations, this baseframe can be of torsionally stiff design.
Centrifugal Compressors
FIGURE 11-9
Centrifugal compressor train. (Source: Mannesmann Demag, Duisburg,
Germany.)
Construction and Mode of Operation Each application requires its own casing configuration. In spite of this, the intemal design and construction of a given manufacturer's centrifugal compressor is often essentially the same. This allows the use of standard components. The components that are important for the compression function are the rotor and the energy-converting parts, as illustrated in Figures 11-10 and 11-11. The rotor consists of the shaft and the impellers. The number of impellers is determined by the thermodynamic operating conditions, but it is limited by the mechanical and dynamic behavior of the rotor. The shaft is carried in pressurelubricated tilting-pad bearings; one of these is purely a radial bearing, while the other is either a separate or a combined radial and thrust bearing. The shaft is generally provided with a balance piston to reduce axial thrust. Shaft seals separate the gas spaces from the oil-lubricated bearings and the atmosphere. Simple labyrinth seals, multiported labyrinths with buffer gas, mechanical contact, or floating ring seals are employed, the choice being dictated by the process involved and the fluid handled.
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FIGURE 11-10 Cutaway view ofhorizontally split centrifugal compressor. (Source: Mannesmann Demag, Duisburg, Germany.)
FIGURE 11-11 Barrel-type compressor casing and internals. (Source: Mannesmann Demag, Duisburg, Germany.)
Centrifugal Compressors The materials for the rotor, internals, and casing are selected on the basis of their mechanical properties and compatibility with the fluid to be compressed. A lube oil system and a seal liquid system supply the bearings and the liquid seals with the required volumes of oil and seal liquid. The fluid to be compressed passes through the inlet nozzle and aerodynamically designed inlet channel into the first impeller. The first impeller may be preceded by an adjustable inlet guide vane unit. Impellers and the diffusers following them are designed so as to provide an optimum low-loss compression cycle. After the diffuser channel, the gas enters the return bend, which guides it to the vaned return channel, and it then reaches the impeller of the next stage with the correct angle of incidence. Immediately following the final stage or an intermediate stage, after which the compressed gas leaves the compressor, the diffuser opens out into a volute that widens gradually in the direction of flow to match the increase in volume.
Impellers The head to be produced and the volume flow to be handled provide the design criteria for impellers. The total energy y that can be transferred by the impeller is a function of the peripheral velocity u2 and the pressure coefficient gr: y -- r
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An upper limit is set to the permissible peripheral velocity by the type of impeller and the mechanical properties of the materials from which it is made. The pressure coefficient increases with increasing blade outlet angle ~2. However, the impeller geometry is not exclusively determined by the maximum energy that can be transferred. Further criteria include good overall efficiency of the stage and a broad envelope of curves with as steep a rise as possible toward the surge limit, thus ensuring good partial load controllability and stable operation within the plant (immunity to pressure fluctuations). The standard impeller in many centrifugal compressors is the type with backward-leaning blades and cover (Figure I 1-12). It represents a good compromise to meet the aforementioned requirements. This type of impeller may feature blades curved in two or three dimensions. Impellers with three-dimensional blading (Figure I l-13) have high capacity limits, due to the large eye diameter relative to the outside diameter and a large outlet width. Three-dimensional impellers are among the standard features available for modem centrifugal compressors.
FIGURE 11-12 Impeller with backward-leaning blades, shown before cover assembly (Source: Mannesmann Demag, Duisburg, Germany.)
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FIGURE 11-13 Impeller with three-dimensional blades. (Source: Mannesmann Demag, Duisburg, Germany.)
Manufacture of Impellers. Five principal manufacturing methods are typically available for the efficient production of impellers: milling and riveting, milling and brazing, milling and welding, welding and welding, and casting. Impellers fabricated by milling and riveting have the gas passages milled from the solid impeller disc. The cover is riveted in place. This long-established method combines great mechanical strength and reliability with high aerodynamic quality, i.e., dimensional accuracy (particularly important for narrow gas passages) and good surface finish. More recent mechanical analyses have especially demonstrated the system-damping properties of the riveted fixing, properties that afford enhanced protection against fracture due to alternating stress cycles. Milled and brazed impellers are used when thin blades and narrow gas passages with good aerodynamic properties are necessary. As in the case of riveted impellers, the gas passages are milled from the solid impeller disc; the cover is then brazed into place with a gold-nickel brazing alloy, using a high-temperature, high-vacuum process. The strength of the brazed joint is equal to that of the parent metal. It is totally immune to HES stress corrosion cracking. When access to the gas passages is good, i.e., when the passages are wide, the impellers are of milled and welded or welded and welded construction. The blades are either milled from the solid impeller disc or welded to it. In both instances, the cover is welded to the blades by a continuous weld. If a number of identical impellers are required, casting is the most economical method of manufacturing them. Large impellers are cast in sand molds. Precision casting processes that ensure unusually good quality of the product are being used to an increasing extent for small impellers. Impeller Testing. Tests are carried out on material specimens prior to and after manufacture to prove the required properties. On completion of manufacture, the dimensions of the impeller are checked, after which the impeller is then tested for cracks, using dye penetration or magnetic particle methods. The impeller is then initially balanced and run at overspeed. This speed may be so high that flow and cold strain hardening may take place at the stress peaks. After overspeeding, the impeller is once again checked for dimensional accuracy and tested for cracks. On successful completion of these tests, the impeller is rebalanced. Rotor The shaft carries the impellers and the balance piston, as illustrated in Figure 11-11. It is supported in tilting-pad, plain, or modified (contoured) sleeve beatings. The
Centrifugal Compressors 375 impellers and balance piston are shrunk onto the shaft. Multipart rings or similar components locate the impellers in the direction of the axial thrust. The shrink fit offers the advantage of uniform stress distribution over the whole circumference and a constant self-centering effect. This shrink fit is designed so that after the bore has expanded due to centrifugal force at maximum speed, sufficient shrinkage effect still exists to transmit the torque and the axial thrust. At the same time, the impellers can be removed whenever necessary without damage. The balance piston balances part of the axial thrust produced by pressure differentials across the impellers. Part of the axial thrust is automatically balanced with a back-to-back layout. All component parts such as shaft, impellers, balance piston, and couplings are separately balanced, after which the complete rotor is assembled. Each time another component is added, concentricity of running is checked. The stresses produced in the shaft by the shrink-fit method of mounting the impellers and balance piston are relieved by running the complete rotor up to operating speed.
Rotor Dynamics. Detailed vibration analyses as early as the preliminary design stage ensure the operational availability of modem, well-engineered centrifugal compressors. These analyses investigate the following vibration phenomena: resonance behavior, stability, and torsional analysis. Investigation of the resonance behavior includes computation of the first and second lateral critical natural frequencies based on the bearing stiffness and damping for each casing. Precise determination of the critical natural frequencies, which are principally lateral vibrations of the rotor caused by unbalance forces, allows multistage compressors to be operated at a speed between the first and second lateral critical speeds. The stability behavior describes the behavior of the rotor when the vibrations are self-excited. This essentially includes vibrations excited by the specific spring and damping characteristic of the bearings (oil whip) and rotor-dynamic effects of labyrinths and sealing elements. The torsional analysis covers torsional vibrations produced by electric motors when switching occurs in the supply network and when a synchronous motor is run up to operating speed asynchronously. A precise knowledge of these interrelated factors ensures trouble-free operation and a high degree of operational availability of major, unspared centrifugal compressors.
Bearings Tilting-pad bearings support and locate the compressor rotor. They employ the hydrodynamic principle and are designed in the light of the most recent scientific and engineering knowledge in this field. The running surfaces of the bearings are divided into segments and inserted into the horizontally split bearing bracket. This bearing bracket is positioned in the bottom half of the bearing housing (Figure 1 l-14) and is typically secured in place by a bolted bearing retainer. Properly designed compressor bearings can be inspected without the compressor casing having to be opened. All bearings have pressure-oil lubrication. A lubricating oil system supplies them with a flow of oil sufficient to form an oil film on the running surfaces and to dissipate the heat produced by friction. Oil is piped centrally to the bearings. Retaining rings fitted at both sides control the rate of discharge of oil from the bearing via the gap set.
Radial Bearing. The radial bearing-often specially developed by capable compressor manufacturers - is a multipad bearing with four or five pads arranged so that the stationary shaft rests on one of them or, in some designs, between pads. In a
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FIGURE 11-14 Compressorbearing housing with combination thrust and radial bearings. (Source: Mannesmann Demag, Duisburg, Germany.)
FIGURE 11-15 Tilt-pad radial bearing. (Source: Mannesmann Demag, Duisburg, Germany.) multipad bearing, the reaction forces act over the entire circumference and therefore stabilize the position of the shaft. In addition, the tilting pads adjust automatically to suit the operating conditions; optimum load distribution is therefore achieved at all times. With the exception of certain special designs, these pads are symmetrically supported and therefore are unaffected by the sense of rotation. Figure 11-15 depicts a radial tilting-pad beating. Thrust Bearing. At the nondrive end of the compressor, the radial bearing is combined with a thrust bearing- also frequently specially developed by the
Centrifugal Compressors compressor manufacturer. This bearing employs the principle of a double adjustable pad-thrust bearing and absorbs the residual rotor thrust resulting from the unbalanced gas forces acting on the impellers and balance piston. The tilting pads are sometimes asymmetrically supported. The axial thrust is determined with the most modem methods available, taking into consideration all the aerodynamic effects that arise. A certain amount of residual thrust has a stabilizing effect, since the rotor is then in contact with a specific side of the beating. The design chosen and the load-bearing capacity of the thrust beating must ensure the operational readiness of the compressor even if thrust reversals and sudden loads occur during the widely varying operational phases the compressor may experience. Figure 11-16 illustrates a combination radial/thrust tilting-pad bearing.
Sealing Elements The long-term, reliable operation of centrifugal compressors must be ensured by thoroughly well-proven sealing elements. Labyrinth seals (Figure 11-17) minimize the flow around the impellers and hence also minimize leakage losses. These labyrinths are located over the rim of the impeller eye on the inlet side and close to the shaft over the hub at the back of the impeller on the discharge side. Centrifugal compressors with several nozzle pairs employ labyrinth seals to separate the individual process stages. Labyrinth seals that prevent lube oil and oil mist from escaping from the bearing chambers are also used to seal the bearing housings. Thoroughly proven sealing systems seal the shaft exists. Depending on specific requirements, multiported labyrinth, floating ring, or mechanical contact systems are employed for this purpose. Multiported labyrinth seals (Figure 11-18) with buffer gas injection are used when buffer gas can be allowed to mix with the process gas or when leakage of process gas is permissible. Provision for ejection can be made in order to avoid excessive leakage of process gas. In all other instances, floating ring or mechanical non-contact or contact seals are used, the former being employed for use with high pressures, the latter being used with clean gases.
Labyrinth SeaL The labyrinth seal is a noncontacting seal. It consists of a number of sealing strips in an insert in the stationary part of the compressor. This insert is horizontally split and easily replaced. Sealing strips are also sometimes fitted to the shaft. "Straight" or stepped labyrinths may be employed, depending on the specific need. The labyrinth seal forms a series of throttling points, at each of which the pressure differential is decreased. The smaller the clearance - the distance between
FIGURE 11-16 Combined radial and thrust tilt-pad bearing for centrifugal compressor. (Source: Mannesmann Demag, Duisburg, Germany.)
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FIGURE 11-17 Labyrinth seals surround each impeller of this centrifugal compressor rotor. (Source: Mannesmann Demag, Duisburg, Germany.)
FIGURE 11-18 Multiported oxygen compressor labyrinth seal. (Source: Mannesmann Demag, Duisburg, Germany.)
the labyrinth strip and the surface of the s h a f t - the more the leakage is reduced. The turbulence zones between the strips enhance the throttling effect. The pressure difference involved dictates the number of labyrinth strips. If absolutely reliable separation of the oil and process gas spaces is necessary, a multiported labyrinth seal (Figure 11-19) with injection of buffer gas under controlled pressure is used.
Centrifugal Compressors 379
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111 FIGURE 11-19 Multiportlabyrinth seal, cross-section view. (Source: Mannesmann Demag, Duisburg, Germany.) Part of this gas flows outward and prevents atmospheric air from entering, while the remainder passes into port 2 at lower pressure. This space is connected to a leak-off line. Ports 3 on the suction and discharge sides of the compressor are interconnected and can be controlled and monitored jointly. The pressure is adjusted so that, although it is lower than the pressure of the buffer gas in port l, it is higher than the pressure in port 2, causing some gas to flow into that port and mix with the buffer gas there. This arrangement ensures that, in the event of a fault in the discharge line, neutral buffer gas may enter the compressor via port 2, although the gas being compressed inside the centrifugal compressor cannot escape to the atmosphere. Ingress of buffer gas into the compression spaces is normally prevented by the pressure drop across port 2. The multiported labyrinth system is very adaptable. For instance, a seal employing buffer gas only and without port 2, with a smaller flow of buffer gas into the compressor, is just as feasible as a seal without buffer gas from which a mixture of atmospheric air and the compressed medium are evacuated via port 2.
Floating Ring Seal. In conjuction with a seal liquid introduced under pressure, the floating ring seal illustrated in Figure 11-20 prevents process gas from escaping at even the highest of operating pressures. It operates without mechanical contact and therefore without wear. This type of seal consists of the inner ring between the process gas space and the seal liquid space, and the outer ring that permits enough seal liquid to escape outward to ensure adequate cooling of the seal. The form of the intermediate ring leads to intensive heat dissipation away from the inner ring. Sealing is maintained by the controlled seal liquid pressure being above the process gas pressure at all times. Clearance between the inner ring and the shaft is such that only very little seal liquid passes through to the gas side. Since this clearance is smaller than the bearing clearance and also because of rotor dynamics
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considerations, the sealing rings are designed to float, i.e., they can follow any radial shaft deflections without acting like bearing supports. The seal liquid enters the seal via supply pipe 1. Most of it is discharged outward through the radial gap formed by the outer ring. A small volume flows into port 5 via the radial gap formed by the inner ring. Seal liquid is prevented from entering the compressor by a constant flow of buffer gas from the supply pipe 7 to port 5 via labyrinths. The mixture of seal oil and buffer gas is led via connection 6 to an automatic separator. Mechanical Contact Seal The mechanical contact seal shown in Figure l 1-21 employs a stationary carbon ring in sliding contact with a rotating ring manufactured from high-quality material with a special finish. A seal liquid is employed. This type of seal is also effective when the compressor is at standstill and the oil pumps have been shut down. The main components are the carbon ring and the rotating ring for inward sealing. In the outboard direction, a floating ring controls the flow of the seal liquid that cools the seal. The seal liquid enters the seal via the supply pipe 1 and flushes the seal ring components via the holes in the distributor ring. The pressure of the liquid is higher than that of the gas, so that the carbon ring, under constant spring pressure, is always kept in sliding contact with the rotating ring. Some of the liquid wets the sliding surface 5 and reduces wear. Only a very small proportion of this liquid passes to the gas side. A controlled flow of buffer gas flowing from the supply pipe 10 through a labyrinth to port 8 entrains this leakage liquid and leads it via outlet 9 to the separator.
Centrifugal Compressors
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FIGURE 11-21 Mechanical contact seal, cross section view. (Source: Mannesmann Demag, Duisburg, Germany.) O-rings fitted externally and within the seal reliably separate the buffer gas and seal liquid spaces. Mechanical Non-Contact Seals ("Gas Seals"). These novel seals are very similar in function and geometry to mechanical contact seals. Instead of a sealing liquid, they use a small quantity of a clean sealing gas. The seal faces operate without actually making contact. The escaping seal gas separates the seal faces by a fraction of a thousandth of an inch.
Casings The two halves of horizontally split compressors are joined together by hydraulically pretensioned bolts. The joint is sealed by a suitable sealant, where necessary of the string type. The vertically split casings are provided with end covers retained by shear tings. O-rings or O-ring joints between inner casing and barrel casing provide proper sealing of the gas zone from the atmosphere. Casings are supported at the centerline. The pedestal supports at the drive end are fixed points, so that axial thermal expansion is in the direction of the free end of the shaft. Lateral alignment is provided by guide lugs in the vertical center plane of the compressor. Thermal expansion of shaft and casing is compensated by the rotor being located by the thrust bearing at the free end of the casing. Positioning of the suction and discharge nozzles can be arranged to suit requirements. Compressor casings are hydraulically tested at 1.5 times the maximum operating pressure.
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Stationary Components The term "stationary components" refers to the inlet channel, the diffuser, and the return channel. Impellers achieve optimum efficiency and operating characteristics only if the inlet flow is free of disturbance, i.e., unswirled, and exhibits a uniform velocity profile. Factors such as these are extremely important for impellers of high suction capacity. Single-shaft process compressors have radial intakes, and the flow therefore has to be deflected through 90 ~ prior to entry into the first impeller. High-quality compressors are produced with inlet channel designs that meet the above requirements. These incorporate standard, stationary blades, which, distributed over the periphery, exhibit defined angular settings. Similarly, these centrifugal compressors would have as a standard vaneless annular diffusors that have parallel walls or profiled cross sections. This feature ensures a wide range of regulation with almost constant optimum efficiency. Leading manufacturers have developed special cross-sectional profiles that produce wider ranges of regulation and better efficiencies than parallel-wall diffusors- especially with high-capacity impellers. As the gas passages become smaller in cross-sectional area, the surface finish of the diffusor has a decisive effect on the stage efficiency. In special cases, the surfaces are therefore coated. Particular importance is attached to smooth inflow into the following impeller in the design of the return channels, the blading of which is generally profiled. Like the impellers, they are normally milled from the solid material to ensure a superior surface finish and dimensional accuracy.
Auxiliary Equipment The driver, suitable gearing, and the couplings have to be matched, with the compressor, to the specific application requirements of the plant and to the conditions on the site. It is only then that the maximum of operational reliability is ensured. One of the requirements for smooth and economical compressor operation is control and regulation to match operational needs plus inherent functional reliability of the coolers and lubrication and seal liquid systems. Monitoring of the beatings via the oil temperature, for example, and measurements of shaft vibrations and shaft position ensure that potential trouble during compressor operation is recognized at an early stage. The monitoring equipment, together with the alarm systems and controls, is accommodated in control panels.
Drive Components Couplings interconnect the various units of a compressor train. Diaphragm or curved-tooth gear couplings are employed as a rule. Couplings of these types allow angular deflection and axial deflection, but they are nevertheless torsionally stiff. Gear couplings with convex tooth flanks require precision manufacture. The lube oil system supplies them with lubricant. Coupling hubs are located and secured by keys or oil-hydraulic means to the shaft. Torsionally elastic couplings are employed if shock torque loads occur, for example, when starting up synchronous motors or if there is a brief drop in voltage. The shock torque load is absorbed by the deformation of elastic components such as rubber buffers. Owing to the lower torque transmission capacity of these components, torsionally elastic couplings are larger. Spur gears or epicyclic gears are employed between driver and compressor or between two casings in order to drive centrifugal compressors at the optimum speed.
Centrifugal Compressors 383 Coolers Process coolers are employed for a variety of reasons: 9 to limit the maximum temperature of the process gas 9 to limit the maximum temperature of the components for safety reasons 9 to minimize the compressor power consumption by aiming at isothermal compression Process coolers are employed as intercoolers and aftercoolers. The principal cooling medium used is water, including seawater. There are various criteria for deciding whether the gas is to flow through or around the tubes of the cooler. These are the quality of the gas, the quality of the cooling medium, and the pressure level involved. If either the process gas or the cooling medium contains contaminants, the aim is to pass these through the tubes and the relatively clean gas or cooling medium around the externally finned tubes. This is because internal cleaning of the tubes is much easier. When pressures are high, the process gas will be routed through the tubes because of mechanical strength considerations. The individual tubes form a bundle, the tube ends of which are expanded into the tube end plates, brazed or welded to produce a gas-tight fit. Depending on the design used, the tubes are combined into bundles or single elements. The singleelement design is a lightweight, easily portable unit often allowing economical maintenance of spares. Baffle plates determining the flow of the process gas are provided inside the cooler housing. Air-cooled process gas coolers are used in special applications. The process gas flows through the tubes around which cooling air flows boosted by fans. Air-cooling requires much larger equipment, and a number of elements are therefore arranged in parallel.
Control Centrifugal compressors always match the process requirements. They form an integral part of the process plant, the operating characteristics of which are a function of volume flow and pressure. Centrifugal compressors can be controlled so that they maintain constant pressure at the intake or at a preceding process point, or at the discharge nozzle or following that point. When process control is staged, intermediate pressure can also be maintained. Requirements for constant volume flow can also be met. Daily and annual meteorological influences are taken into consideration. In addition, the compressors have to be protected against unstable operating conditions. If these requirements are to be met, the operating characteristics of the compressor have to be controlled. This can be done either by varying impeller speeds, by varying the angle of incidence of the gas, or by throttling.
Speed Variation. When speed variation is used, all velocity components are equally affected. This requires a speed-controllable driver or suitable intermediate element. The principal advantage of speed regulation is that only as much energy is required as is needed for the process. Variation of Angle of Incidence. The compressor characteristic for constant speed can be directly influenced by using an inlet guide vane unit to vary the angle of
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ProcessPlant Machinery incidence. Apart from providing economical operation under partial load, this also allows an extension of the characteristic above normal. With constant conditions at the impeller outlet, the changed angle of incidence causes the specific work to be influenced by the relative and absolute velocities at the inlet. With a positive guide vane setting, the total energy is reduced because of the lower relative speed at the inlet. A negative setting has the opposite effect. The influence of a negative guide vane setting generally reaches a peak value between 20 ~ and 30 ~ After this setting is exceeded, the angled guide vanes choke the flow cross section to such a degree that throttling with local separation phenomena occurs, and this causes losses to increase rapidly. Toward the lower end of the volume flow range, a positive guide vane setting has an advantageous effect, particularly in the region close to the design point. The direction of the relative velocity is turned to coincide with the design angle of the leading edge of the blade, and the shock loss otherwise occurring due to an unfavorable approach angle is reduced. In the remainder of the partial load range, a throttling effect resulting from constriction of the flow cross section by the angled guide vanes is superimposed on the aerodynamic effect. One special advantage of guide vane adjustment is that the stable operating range is extended toward smaller volume flow rates, although this applies to the particular stage only. With multistage compressors, the effect is reduced in proportion to the number of uncontrolled stages.
Throttling. With throttling, the compressor characteristic remains unchanged. When throttling is performed on the suction side, a throttle valve reduces the compressor inlet pressure when the volume flow rate is increased. Downstream of the compressor, a lower discharge pressure is produced, which corresponds to the pressure ratio associated with the point on the compressor characteristic or compressor performance curve. In relation to the useful compression work, throttling requires maximum specific energy. Throttling on the suction side and variation of inlet guide vane angle are also used to expand the operating range of compressor plants with speed-controlled drivers. Antisurge Control.
In order to protect the machines from excessively high mechanical loading due to unstable operating conditions (surging), which primarily stress the bearings, compressors can be equipped with antisurge control. One such control is depicted in Figure 11-22. If the delivery required is below the minimum delivery volume of the compressor, the surplus is led away as a sidestream, via a valve. Depending on the nature of the gas, the surplus is either discharged to the atmosphere or it is cooled and returned to the suction side. In the example shown, the valve is operated by a controller that uses the volume flow rate and the discharge pressure as input parameters. The blow-off or recycle limit in the compressor curve envelope is normally an approximate simulation of the surge limit. Pneumatic, hydraulic, electropneumatic, or electrohydraulic control systems may be employed. Depending on the cross-sectional area of the discharge, control valves or flaps are employed as regulating units.
Discharge Pressure ControL The need for constant compressor discharge pressure can be met, irrespective of the delivery volume, by using a guide vane unit. The signal for the actual value is taken from the discharge line and fed to a PI controller. After comparing this reading with the set value, the controller adjusts the setting of the guide vanes via a servo-cylinder.
Centrifugal Compressors
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Simple example o f combined antisurge and discharge pressure control. 1 - inlet; 2 - inlet guide vane unit; 3 - compressor; 4 - nonreturn valve; 5 - discharge line; 6 - pressure transmitter; 7 - discharge pressure controller; 8 - servocylinder; 9 - orifice measurement; 10 - differential pressure transmitter; 11 - computer; 12 - surge limit controller; 1 3 - blow-off valve; 1 4 - b l o w - o f f line. (Source: M a n n e s m a n n Demag, Duisburg, Germany.)
Lube Oil System The lube oil system supplies oil to the compressor and driver bearings and to the gears and couplings. Figure 11-23 illustrates a typical lube oil system. The lube oil starts off in the reservoir, from where it is drawn by the pumps and fed under pressure through coolers and filters to the bearings. On leaving the bearings, the oil drains back to the reservoir. The reservoir is designed to permit circulation of its entire contents between eight and twelve times per hour. Oil level and temperature are constantly monitored. The oil can be preheated electrically or indirectly by steam for starting up at low temperatures. A thermostat with surface temperature limiter prevents overheating of the oil. The reservoir is vented. Oil is normally circulated by the main oil pump. An auxiliary pump serves as a standby. These two pumps generally have different types of drive. When both are driven electrically, they are connected to separate supply networks. On compressors with step-up gearboxes, the main oil pump may be driven mechanically from the gearbox. The auxiliary pump then operates during the start-up and run-down phases of the compressor plant. Relief valves protect both pumps from the effects of excessively high pressures. Nonreturn valves prevent reverse flow of oil through the stationary pumps. Heat generated by friction in the bearings is transferred to the cooling medium in the oil coolers. The return temperature is monitored by a temperature switch. Aircooled oil coolers may be employed as an alternative to water as coolant. The former have long been used in regions where water is in short supply. Twin coolers with provision for changeover have filling and venting connections so that the standby cooler can be filled with oil prior to changing over. This eliminates the
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FIGURE 11-23 Lube oil schematic for a centrifugal compressor. 1 - Reservoir; 2 - safety valve; 3 - main oil pump; 4 - auxiliary oil pump; 5 - cooler; 6 - pressure-regulating valve; 7 -filter; 8 - overhead tank; 9 - d r i v e r and other users; P l - p r e s s u r e indicator; D P l - d i f f e r e n t i a l pressure indicator; P S - p r e s s u r e switch; T I - temperature indicator; T S - temperature switch; L I - level indicator; L S - level switch; H - heater; A - reservoir vent. (Source: Mannesmann Demag, Duisburg, Germany.)
possibility of disturbances and damage due to air bubbles in the pipework system. Twin oil filters with provision for changeover have the same facilities. A pressure-regulating valve is controlled via the pressure downstream of the filters and maintains constant oil pressure by regulating the quantity of bypassed oil. The auxiliary oil pump is switched on by a pressure switch if the oil pressure falls. A second pressure switch shuts down the compressor plant if the pressure still continues to fall.
Centrifugal Compressors 387 The filters clean the lube oil before it reaches the lubrication points. A differential pressure gauge monitors the degree of fouling of the filters. An overhead oil tank can be provided to ensure a supply of lubricant to the bearings in the event of faults while the compressor is being run down. A continuous flow of oil through an orifice maintains the header oil constantly at operating temperature. Should the pressure in the lube oil system fall, the nonreturn valve beneath the tank opens to provide a flow of oil. The flow of oil to each bearing is regulated individually by orifices, particularly important for lubrication points requiring different pressures. Lube oil for the driver and other users is taken from branch lines. When a hydraulic shaft position indicator is used, this is supplied with oil from the lube oil system. Temperatures and pressures are measured at all important locations in the system; the readings can be taken locally or transmitted to a monitoring station. Except for a few components, the lube oil system is a conveniently installed packaged unit supplied complete and ready for installation. Oil pumps, coolers, and filters are grouped around the oil reservoir on a common baseplate. Design and construction of the lube oil system must take into account the relevant regulations and any special requirements. One such requirement might be blanketing with inert gas; another might be the on-stream purification of lube oil by modem vacuum dehydrator units.
Seal Liquid System The seal liquid system supplies the mechanical contact and floating ring seals with an adequate flow of seal liquid at all times, thus ensuring that they function correctly (Figures 11-24A and B). An effective seal is provided at the settling-out pressure when the compressor is not running. Starting in the main oil reservoir, the medium passes to the seals via the pumps, the twin oil coolers, and the twin filters. Instruments for monitoring the oil level and temperature are mounted on the reservoir. If necessary, the seal oil is heated; a thermostat with surface temperature limiter protects against excessively high temperatures. Every system has a main oil pump and an auxiliary oil pump with independent drives. They are designed for a higher delivery rate than is actually needed by the seals. To protect the pumps and downstream equipment, safety valves are fitted. Nonreturn valves after each pump prevent seal oil from flowing back to the reservoir through the stationary pumps. The coolers dissipate the heat transferred to the seal oil. A temperature switch monitors the permissible temperature range. The filters retain all impurities, the pressure drop across them being checked by a differential pressure indicator. The floating ring seals are supplied with seal oil at a defined differential pressure above the reference gas pressure (pressure within the inner seal drain). This is schematically represented in Figure 11-24A. Figure 11-24C shows a lube oil system with seal oil booster arrangement; a seal oil system with pressure reduction for the lube oil portion is illustrated in Figure 11-24D. The flow of seal oil is regulated by a differential pressure-regulating valve, which, if there are changes in the reference gas pressure, regulates the pressure of the seal oil, or, as shown in the diagram, by a level-control valve that maintains a constant level in the overhead tank. The oil in the overhead tank is in contact with the reference gas pressure via a separate line. The static head provides the required pressure differential. In addition, the oil in the overhead tank compensates for pressure fluctuations and serves as a run-down supply if pressure is lost. If the level in
388
Process Plant Machinery
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A FIGURE 11-24 (Continued on following pages) Seal liquid system schematics. A - Seal oil system for floating ring seals; B - seal oil system for mechanical contact seals; C - lube oil system with seal oil booster system for floating ring seals; D - seal oil system for floating ring seals with pressure reduction for the lube oil system (combined system); I - oil tank; 2 - relief valve; 3 - main oil pump (high pressure); 4 - auxiliary oil pump (high pressure); 5 - oil cooler (high pressure); 6 - regulating valve; 7 - high-pressure filter; 8 - seal oil overhead tank; 9 - demister separator; I 0 - condensate trap; 1 1 - degassing tank; 1 2 - sour oil tank; 1 3 - l e v e l control valve; 14-pressure-reducing valve; 1 5 - l u b e oil overhead tank; 1 6 - m a i n oil pump (low pressure); 17-auxiliary oil pump (low pressure); 1 8 - o i l cooler (low pressure); 19-low-pressure filter; O1- seal oil supply; 0 2 - buffer gas supply; 0 3 - outer drain; 0 4 - inner drain (seal oil/buffer gas); 05 - buffer gas supply; 07 - reference pressure line; 08 - lube oil supply; PI - pressure indicator; PS - pressure switch; DPI-differential pressure indicator; TI-temperature indicator; T S - temperature switch; L I - l e v e l indicator; L S - l e v e l switch; L I C - l e v e l controller; H - h e a t e r ; A - reservoir vent. (Source: Mannesmann Demag, Duisburg, Germany.)
Centrifugal Compressors 389
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the tank falls excessively, a level switch shuts down the compressor plant. There is a constant flow of oil through the overhead tank, and this heats the oil at all times. For the mechanical contact seal, the seal oil is kept at a constant differential pressure with respect to the reference gas by a regulating valve (Figure 11-24B). As the name indicates, the mechanical contact seal provides a mechanical seal when the compressor plant is shut down. To prevent oil from gaining ingress to the compressor, the space between the oil drain and compression space is sealed by a flow of gas. The pressure of this sealing gas or buffer gas is above the pressure of the reference gas. A differential pressure indicator monitors the pressure differential. The flow of seal oil divides in the compressor seals. Most of the flow returns under gravity to the reservoir. A small quantity passes through the inner seal ring to the inner drain, where it is exposed to the gas pressure. This oil, mixed with the buffer gas, is led to the separator system. On each side this consists of a separator and a condensate trap. The separated gas is led either to the flare stack or to the suction side of the compressor. The oil flows into a tank for degassing.
390
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If oil is used as sealing liquid and can be used again, degassing is accelerated by exposure to a partial vacuum and by heating. The oil is then returned to the reservoir. If the oil becomes unusable, it is led away for separate treatment. The quantities of oil passing through the inner drain in well-designed centrifugal compressors are very small. Temperature and pressure measuring points with local or remote reading are provided at all major points of the seal liquid system. The seal oil system may be combined with the lube oil system if the gas does not adversely affect the lubricating qualities of the oil or provided the oil made unserviceable by the gas does not return into the oil system. There are two methods of combining lube oil and seal oil systems. In the first of these, the oil can be raised to the pressure required for lubrication purposes and part of it then raised further to the pressure needed for sealing (booster system). Alternatively, all the oil is initially raised to the seal oil pressure and the flow of oil required for lubrication is then reduced in pressure (combined system).
Centrifugal Compressors 391
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Compressor Monitoring and Safety Equipment In order to ensure the operational reliability and safety of centrifugal compressors, all major mechanical and thermodynamic parameters must be constantly scanned and evaluated. Only fast detection of changes makes possible timely intervention, thus protecting the compressor plant and the process from potentially damaging effects. Prominent manufacturers therefore design and supply complete monitoring and instrument panels for compressor and driver. These contain all the instrumentation for monitoring, including control equipment, an alarm system, and interlock and shutdown systems. Moreover, these manufacturers often furnish fully automatic compressor plants with automatic start/stop control for single compressor sets or, if required, including the higher-level station control facility. The latter switches the individual compressors on or off as required when such machines are operating in parallel. The control equipment needed for compressor and driver, fittings and auxiliaries is fully taken into consideration. Installations of this kind already supplied by major manufacturers include those with air compressors in large compressed air systems, sewage treatment plants, and offshore platforms with gas compressors.
392
ProcessPlant Machinery Bearing Monitors Smooth running of the shaft can be guaranteed only by bearings that function properly. The simplest check is monitoring of the lube oil draining from the bearings; wear of or damage to the running surfaces or insufficient beating clearance mean greater friction and therefore an elevated oil discharge temperature. The high temperatures present another hazard, since they reduce the viscosity of the lube oil and hence its lubricating effect. Thermocouples, mercury thermometers, and resistance thermometers may be used for monitoring oil temperature. The permissible temperature range is limited by switching contacts; if the set limits are exceeded, an alarm is initiated or the compressor is tripped. It is also possible to measure the bearing temperatures directly in the bearing pads.
Measurement of Vibration and Shaft Position Mechanical vibration of shaft and casing and axial shaft displacements are important indications of possible hazard to compressor operating reliability. Modem technology renders it possible to detect even the slightest of changes in these parameters. Shaft vibrations may be due to a number of causes: deposits or other impeller unbalances; shaft distortion as a result of thermal stresses or shock loads; changed bearing condition; effects emanating from the driver or gearing. Measurements are performed by probes that are calibrated for a certain distance between shaft and probe head. These probes produce an electrical signal proportional to the amplitude of vibration. The permissible shaft vibration depends on the speed and the rotating mass. Low speeds allow higher vibration. Apart from measuring the total level, which is the sum of the amplitudes at different frequencies, it is also possible to analyze the vibrations on the basis of frequency. This may give a clue to the exciting frequencies or other causes of vibration. Seismic pickups are employed for measuring casing or bearing pedestal vibrations. This method does not supply such precise results as shaft vibration measurement, but it offers the advantage of permitting measurements to be performed on running machines without having to make modifications first. The shaft position indicator monitors the axial position of the shaft relative to the casing. The cause of axial displacement of the shaft may be wear of the thrust bearing or sudden loads that may occur when the compressor is operating in the unstable region. Detection of axial shaft shift is usually by electrical, or less frequently, by hydraulic means. Electrical measurement basically involves the same probes as are used for contactless vibration measurement. All devices can be fitted with switches that trigger alarms or initiate shut-down of the plant to prevent damage when limit values are exceeded. Modem shut-down devices use the input from several sensors into a logic module. The machine would be automatically taken off-line if, for instance, a temperature sensor and a vibration sensor would independently confirm a violation of two setpoints. SELECTION OF A CENTRIFUGAL COMPRESSOR Compressor selection topics are covered in the appendices on barrel and isotherm compressors in this chapter, and also in the appendix to the chapter on axial compressors. Since environmental considerations play a major role in equipment selection today, this topic will be dealt with in more depth in Environmental Engineering and Management: Sustainable Development in the Power Generation, Oil & Gas and Process Industries (Butterworth-Heinemann, 1998).
APPENDIX 11A Compressor Design* The operating conditions of a plant determine the compressor design. The impellers are selected on the basis of the thermodynamic data. Selection is influenced by peripheral conditions such as materials data, aerodynamic and thermodynamic limits, or the operating characteristics. THERMODYNAMICS Centrifugal compressors are designed to convert velocity energy into pressure energy. They thus change work introduced through the impellers into the required gas-side operating conditions; this primarily means the required discharge pressure. The minimum driver size or rating needed is dictated by the energy required to accomplish this conversion, in which the losses should obviously be as small as possible. The work performed to accomplish the change in volume results in an equivalent quantity of heat energy being transferred to the gas, causing an increase in temperature. Compressor design takes into account the fact that critical temperatures should not be exceeded if they might initiate polymerization or reaction of the medium with the compressor materials. AERODYNAMICS The compression cycle is designed to prevent the emergence of near-sonic velocities. These could initiate shock waves at the leading edges of the impeller blades, which then choke the flow cross sections and manifest themselves as irregular pressure rises. Near-sonic velocities may easily occur with compression media of high molecular mass, small isentropic exponent, and low temperature. OPERATING CHARACTERISTICS Optimized centrifugal compressors are designed so that their characteristics match the process requirements. The total required energy is divided among the various stages in accordance with a number of criteria. Lightly loaded stages favor a stable working range that is as wide as possible, although the number of stages becomes larger and the work involved in building the compressor therefore more extensive. If a compressor is designed for a narrow volume flow range, highly loaded impellers are more economical. This enables the number of stages and hence the size of the compressor to be drastically reduced. In addition to fluctuations in pressure and temperature, changes in the composition of the gas also occur. If the change is toward greater molar mass, * Source: Mannesmann Demag, Duisburg, Germany. 393
394
ProcessPlant Machinery a speed-controlled driver can be employed to reduce the drive speed and prevent surge in the final stages.
Design Diagrams The principal characteristics of a compressor can be approximated from the diagram reproduced in Figure 11A-1 for the power, number of stages, and speed in conjunction with the auxiliary diagram for quantity relationships (Figure 11A-2). The first of these diagrams is valid for isentropic, i.e., uncooled, compression. The specific energy Ys and compressor efficiency are not shown, but they are implicit as a function of their influences. Compressors with pressure ratios of up to 7:1 are covered by the diagram, since cooling is almost always necessary for higher pressure ratios. In order to determine the compressor data in such cases, the compression process is split at the point where intermediate cooling takes place; the two separate sections of the compression process can then be summed after each has been determined separately.
Calculation Example Using metric units, the prescribed operating data are as follows: Medium: a gas mixture consisting of inert gases and hydrocarbons Molar mass Mid
= 36.5
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Appendix l lA: Compressor Design 395
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1.31
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=
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= 134,046 kg/h
Intake pressure Pl
= 5.14 bar
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= 303 Kelvin (K)
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= 19.45 bar
Maximum tip speed (assumed) UEmax
= 240 m/s
The design procedure is as follows: 1. From the intake pressure Pl and discharge pressure P2, the following pressure ratio is obtained: Jr = P2/Pl = 19.45/5.14 = 3.784
396
Process Plant Machinery
2. Correction of the ideal molar mass Mid, using the real gas factor Z: M = Mid/Z = 36.5/1 = 36.5 3. The data coveting the volume flow rate ~' and the mass flow rate rh for any quantity can be taken from the adjacent auxiliary diagram for the quantity relationship for the given operating conditions. Taking M, p, and T into consideration, the following volume flow rate is obtained: "~/= 5 m3/s 4. Starting from pressure ratio Jr -- 3.784, the arrow points vertically upward to the point of intersection with the curve for the isentropic tc -- 1.31; from there the next arrow leads horizontally toward the point of intersection with the curve for the same molar mass M -- 36.5. Vertically above that point, the temperature line T1 = 303 K is reached, after which the next arrow points horizontally toward the range of influence of the volume rate V. The heavy line then runs parallel to the family of curves until the specified volume flow rate V = 5 m3/s is reached, after which the heavy line runs horizontally again until it intersects with the corresponding mass flow rate line rh = 37.24 kg/s. Vertically above this point of intersection the power requirement P = 5150 kW can be read off. 5. The heavy line is now extended horizontally through the range of influence of the volume flow rate V until the line reaches the volume-dependent lines of opposite inclination. From this point it is extended vertically downward as far as the stage number line z. For the assumed tip speed of u2 = 240 rn/s we then obtain z = 4 as the number of stages. 6. Starting from the volume flow rate scale ~' = 5 m3/s, the heavy line in the direction of the arrow then impinges vertically on a specific size of compressor casing. From that point the heavy line is taken horizontally to the fight until it intersects the line u2 = 240 rn/s established under step 5 (above). Vertically beneath this point of intersection, the speed n - 8200 RPM is read off.
Calculation Procedure power
"~/--+ rh --+ p[kW]
zr ---~ x --+ M ---~ TI Number of impellers
----> u 2 ----~ z
~' ~
Speed
Calculation Steps 1. rr = 1 9 . 4 5 / 5 . 1 4
= 3.784
2. M = 36.5/1.0
= 36.5
3. rh = 134,064/3600
= 37.24 kg/s
~ / = 5 m3/s 4. P = 5150 kW 5. z = 4 u2 = 240 rn/s 6. n = 8200 RPM
~
u2 --+ n[RPM]
Appendix 11A: Compressor Design
397
Influence of Intercooling Energy saving is the main advantage and aim of cooled compression. Since the energy required is proportional to the intake temperature, the compression process is divided into a number of steps; the intake temperature of each of these steps is reduced in an intercooler. The advantage of intercooling becomes apparent when the thermodynamically attainable energy savings are contrasted with expenditure on the cooling medium and coolers. The thermodynamic study must take into account the flow resistance of the cooler. This pressure loss is compensated by additional compression work. The number of intercooling stages therefore depends on the overall pressure ratio Jr, the isentropic exponent K, which is determined by the temperature rise during compression, and the temperature differential At between the intake temperature of the first stage and the recooling temperature in the subsequent stages. The recooling temperature is determined by the temperature of the cooling medium and the heat exchange surface of the cooler. The possible advantage of intercooling can be visualized with the help of Figure 11A-3. Example: Pressure ratio Jr = 5.35 : 1 Isentropic exponent K = 1.38 Number of intercooling stages j = 2 Result: Power factor f = 0.9, i.e., for the case in question, intercooling twice would result in a power saving at the compressor coupling of about 10% For further information on compressor calculations, see Appendix 1 lB.
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APPENDIX 11B
High-Speed Centrifugal Compressors As mentioned at the beginning of this chapter, dynamic gas compression is achieved by the mechanical action of a rotating impeller that imparts velocity energy to the gas. This velocity energy is then converted to pressure rise of the gas in a diffusion section that can be of a partial emission, vaneless, or vaned diffuser design. The rotating element or impeller of the centrifugal compressor was historically limited in rotational speed by the motor or turbine driver speed capability. With the use of separate speed-increasing gearboxes or high-speed drivers, the rotor speed was gradually increased to a maximum of 10,000 to 12,000 RPM. These are primarily multistage, medium-speed, high-flow centrifugal compressors. In the late 1950s, commercial work was compelled on centrifugal compressor designs that used an integral speed-increasing gearbox with the single impeller of the compressor rotating at speeds of 34,000 RPM and more. This provided the process industries with a design for low flows with high head capabilities. Flow rates for high-speed centrifugal compressors can range from 10 ACFM to more than 100,000 ACFM. Power capability ranges from 15 to 2500 HP or greater. See Figure 11 B- 1 for the composite performance envelope of high-speed centrifugal compressor designs. Integrally geared high-speed compressors (Figure 11B-2) are available for use with any type of gas that a process design might involve. These gases can range from air or nitrogen to such exotic gases as hydrogen bromide. To define where a high-speed centrifugal compressor is primarily used, the concept of specific speed and specific diameter should be understood and applied. Specific speed, Ns, is a dimensionless index number for the impellers or rotors of various types of compressors and pumps. Using English/US values, the definition is the same for both pumps and compressors: s--
N4(H)3/4
Another dimensionless quantity for impellers or rotors is termed the specific diameter, Ds, defined by: D(H) 1/4 Ds -In both formulas:
H = Head in ft-lbf/lbm, or foot-pounds (force) per pound (mass) Q = Flow capacity in ft3/second at inlet conditions N -- Rotational speed in revolutions/minute D --- Diameter of impeller or rotor in feet
* Source: Sundstrand Fluid Handling, Arvada, CO, and Larry E. Glassburn (National Business Consultants, Inc.). 398
Appendix 11B: High-Speed Centrifugal Compressors N, W K g =ooooo
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The process or unit size determines the head (H) and capacity (Q) for a given system design. Components N and D are defined by the availability of mechanical designs from the various manufacturers. For high-speed centrifugal compressors, the normal range for these variables is N from 5,800 rev/min to 50,000 rev/min and D from 5 inches (127 mm) to 36 inches (914 mm). After the process design requirements and the need for the compression equipment have been defined, it is necessary to determine the equipment best suited for the process by using the quantities of specific speed and specific diameter. Using Balje's chart (Figure 11B-3) with specific speed and specific diameter calculated for a particular application, it is possible to determine if the application is a good fit for a single-stage, high-speed, centrifugal compressor. High-speed centrifugal compressors can be of partial emission, radial flow, or mixed flow design. By adding additional parallel stages or series stages of compression, the performance of a single-stage machine can be increased in flow or head. This can be done integrally within the same compressor frame, or it can be done by separate compressor units. The detailed sizing of a high-speed centrifugal compressor is no different than sizing any centrifugal compressor (for further information see previous chapter and also Suggested Reading, Lapina). Virtually every compressor manufacturer has literature available that will assist in sizing the specific unit that would be required for a given process. The available literature usually contains performance envelopes showing the manufacturer's capabilities based on a single gas for a range of flows (Q) in ACFM or cubic meters per hour and heads (H) in foot-pounds per pound mass or newton-meters per kilogram. This enables the user to compare anticipated flow rates and heads with the machinery offered by a given manufacturer. To obtain a budgetary quotation or further information from the manufacturer, the user will have to provide the following data: inlet pressure, P1, either in gauge or absolute; discharge pressure, P2, either in gauge or absolute; inlet temperature, TI, ~ or ~ required capacity or flow rate, Q, preferably in pounds per hour or any weight
400
ProcessPlant Machinery
FIGURE 11B-2 Verticalhigh-speed centrifugal compressor with integral gearbox. (Source: Sundstrand Fluid Handling, Arvada, CO.)
401
Appendix 11B: High-Speed Centrifugal Compressors
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__Jl.___ compressor , 1,000
3,000
10,000
S p e c i f i c speed,/Ms
FIGURE 11B-3 Specific speed versus specific diameter for initial selection of a type of single-stage compressor. (Source." O.E. Balje. A Study on Design Criteria and Matching of Turbomachines- Part B.) flow unit; and the molecular weight, MW, of the total process gas or the individual components of the total process gas mixture. The specific heat ratio, K, is also required, as is the inlet compressibility factor, ZI. The compressor manufacturer can define compressibility factors and specific heat ratios for the majority of gases or gas mixtures. Proprietary or nonstandard gases must have the properties defined for proper selection. Adiabatic head for single-stage compression is usually calculated from the following expression:
HAD=ZI
(1545)(T1)( K ) MW K-1
I(
__P2 Pl
K
-1
ft-lbf/lbm
Dividing the discharge pressure (absolute) by the inlet pressure (absolute), one obtains the pressure ratio. Selection charts from the various manufacturers display either head or pressure ratio. The weight flow can be converted to volume flow rate by using the following formula: Q
(W)(10.729)(TI)(Z~) actual ft3/min (ACFM) (P~)(MW)
With these two values, the head and flow, it is normally possible to determine if the available compressor meets the user's needs. The design horsepower will be BHP
=
(W)(HAD) + Frictional horsepower = Total horsepower (33,000)(r/AD)
The chart in Figure 11B-4 is an approximating method for high-speed centrifugal single-stage compressor selection based on head and flow. The chart applies to two
402
ProcessPlant Machinery MINIMUM CAllING INLET FLANGE l l I Z E (150 F T / l l l C MAX.)
I _
I
"" I
PI " 14.0 p l l l
I
"
I
'"
1~ /6
! !!!!
MW., 4.0
.<
TI - 70oF
---Td',,
TIAD- 75~ p.,Idi=l Flow Oellgn
.,,,. F
I
16" 118"
I'"'1'"'
D-8"
I
D- 10"
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i~.t/
D= 16"
/%
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I"1 O=20"
'/ N.~" /
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/~
..,~,, ~ K / ~>, i x
D=24"
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D- 28"
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/;><,. / ~
"
,
-
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./ "
,'<,/
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>.,i" >Z"
/,~
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700
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/[
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)
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I~AD ,= Ar Efficiency [from C h i r r depending on de6*gn~ HPA " Horsepower ot &Opllcat*on
1
'..>"
t(X~
2 [.-~:~
0 1 FLOW (ACFIk~
P1 = Inlet Pressure (D$,II P2 = Discharge Preesuve (l:lelal 1"1 ,, Inlet Temperetuee("R) '9R ,, 4 6 0 + "'F MW " Molecular We,ght ol G e l ZJ " Inlet C o m p r e l l s b i l i t y K = Specific I l e a l Rit~o i t Inlet T e m p e r l t u r e ~V - Weight Flow(Ibe/m,nutel T2 A" OilCharge Teml:)erlliU re (" RI o / A p p l i c a t i o n
,i'Y"v "
~
, ~o~176176
7i
ISource : E. J. Flril/' H. E, Grlvellel
[xlmple Given MW-20; K~. 1.25
Pl,,3101~lm;, P2-3501g/i; TI e l ( l l ' F Z1 = 0 . 9 4 1 ; W1096.6 IIWmn~
CiIcutete: H A g - 5000.76 lt.ll)/Ibm O A C F M " 999.98 AGFM Enler chaff with theee valueL I n t e m e c t l l ~ pOklll c o m p r e e l o r i I z i n g ell RPM - 1S000. (approximelely) D ,, 8.5 inches (llx~roxln~ltely] HPCHART - 2. (apl)roxlmetell# Cak::ulete i p p l i c a l i o n r
T2A"
+
DOtal ;
580,,578.3"R-
118.3"F
Minimum Calling Inlet flenge IlZe " 6"
F I G U R E 11B-4 High-speed centrifugal compressor estimating selection chart~procedure. (Source: Sundstrand Fluid Handling, Arvada, CO.)
different impeller/diffuser designs from Balje (Figure 11B-3), the partial emission design for low flows, and the radial/mixed flow design for higher flows. To use Figure 11B-4, calculate the required head and flow for a given application, and enter the chart locating the intersecting point. This point will define the approximate speed, impeller diameter, and power for the compressor. The chart will
Appendix I1B: High-speed Centrifugal Compressors 403
OAcFM.
Ivh
I107291 1111 1211
.
ISCFW I002831 I T i l lZi1
IPIIIMW
lPl1
Enter C n m and Define Inleraeolmp Poml Chan Oupul
FIGURE. 118-4 (continued)
404
Process Plant Machinery
also provide the discharge pressure (P2) and temperature (T2) by the formulas on the chart. The power curves in the chart are based on MW -- 4. To calculate power for other MW gases, use the correction formula as shown. When calculating the value of T2 for another MW gas, note the two different values based on each design type (efficiency). Power values are based on aerodynamic work only. Frictional losses are not included, but they will not greatly affect the estimate. For a given mechanical design of a high-speed centrifugal compressor (speed, impeller diameter, etc., are all fixed), there will exist only one performance curve of head versus flow based on the inlet conditions (P1, T1, MW, K, and Z1) of the process gas selected. The head/flow curve will relate discharge conditions to the given inlet condition. If inlet conditions change for the fixed mechanical design, the head/flow curve remains the same; only the discharge conditions will change. This head/flow curve is fixed by the mechanical design of the high-speed centrifugal compressor to accommodate the worst combination of operating conditions. Any other off-design conditions that do not fall on the head/flow curve must be met by control methods as described later. The head generated by a single impeller stage of a high-speed compressor is quite high for a centrifugal design. The high-speed compressor design will not match the head capability of a positive displacement design. However, unless high heads are required, the benefits of the centrifugal design, such as flow variation at constant pressure, often outweigh those of positivedisplacement machines. Typical applications for the high-speed centrifugal compressor involve those that require pressure ratios ranging from approximately 1.005 to 3.5. These are classified as recycle-, regeneration-, or booster-type applications. Pressure ratios in excess of 3.5 can be achieved by series arrangement of the high-speed centrifugal design. These arrangements can be either individual compressor units or multiple impeller stages mounted on the same speed-increasing element. Typical applications for the high-speed centrifugal compressor are molecular sieve absorption regeneration systems, vapor recovery systems, gas injection systems, gas recycle systems, booster compressors, chlorine transfer, and numerous other types of applications that require low pressure rise. High-speed centrifugal compressor designs are also used for inert gases such as air and nitrogen. Available metallurgies for high-speed centrifugal compressors include carbon steel, 316 stainless steel, and 17-4 PH stainless steel, which are all quite compatible with normal gases. When the gases become more corrosive or erosive, as might be the case with chlorine, bromine, hydrogen sulfide, etc., special materials such as Hastelloy B or C, Inconel | or titanium may be required and are available from most manufacturers. Any type of process gas can be effectively handled by the high-speed centrifugal compressor design, from hydrogen to Freon | because the components of the high-speed centrifugal compressor that come in contact with the process gas are relatively few and small. Corrosive or erosive gases can also be handled quite effectively through the use of special metallurgy. Similarly, these dynamic compressors can be built with gas containment seals suitable to meet environmental and loss prevention concerns. A number of seal designs are available for high-speed centrifugal compressors. They can range from simple labyrinth design for nonhazardous low-pressure applications to the typical mechanical contact face-type seals in either single, double, or tandem arrangements. The seals can be either gas- or liquid-type. Any type of seal arrangement must have some leakage across the seal face for it to function properly. This leakage can be either the process gas or a buffer fluid (gas or liquid), depending on the environmental requirements. All high-speed centrifugal compressors consist of three major components: the compressor, the speed-increasing element, and the main driver. The main driver can
Appendix 11B: High-Speed Centrifugal Compressors 405 be an electric motor, steam turbine, air motor, gas turbine, combustion engine, or some other prime mover. Driver selection is generally related to the utility balance of the process plant installation. The compressor or process fluid element is fairly typical for all high-speed centrifugal compressors and consists of an impeller and diffuser contained within a pressure casing. The speed-increasing element between the driver and the compressor impeller is typically a unique integral gearbox that is provided by the same manufacturer (Figure 11B-2). The gearbox has been specifically designed to provide reliable and efficient operation of the unit. High-speed centrifugal compressors are manufactured with either vertical or horizontal in-line, or horizontal axial inlet, and with top or side discharge. Other arrangements are possible and depend on physical size restrictions of the installation. The following control methods are available for high-speed centrifugal compressors: 9 Speed variation 9 Suction pressure throttling 9 Inlet guide vanes 9 Flow control 9 Variable diffuser 9 Discharge pressure throttling Control and operation of individual series and/or parallel units is essentially managed the same as with other dynamic compressors. Surge control must also be provided to protect the high-speed centrifugal compressor from serious mechanical damage. Further details on these subjects can be found in the chapter on axial compressors. As with any mechanical device, routine maintenance of the high-speed centrifugal compressors is required. Depending on the process application and local environment where the high-speed centrifugal compressor is installed, yearly inspection and/or replacement of lubricating oil, oil filtration cartridges, ball bearings, and mechanical seals is recommended. Some applications may require lube oil analysis and/or replacement at a six-month interval. This can be determined by operating experience or discussion with the individual compressor manufacturer. The high-speed centrifugal compressor does not require any more maintenance than any other piece of mechanical equipment within the typical process plant. Experience shows that with proper selection, installation, and operation, the high-speed centrifugal compressor will provide many years of satisfactory operation and performance for any application in a gas compression system. If a high-speed centrifugal compressor has been defined as a potential candidate for a given process, the design advantages may be quite numerous: oil-free operation (no contamination of the process gas), compact design (requires very small floor space and has minimal foundation requirements), simple process piping arrangements are allowed, smooth continuous flow (no pulsations into the downstream process), high inlet pressure capability, simple control systems are required, a wide range of metallurgies to handle various process gases is readily available, and there are no close clearances between rotating and stationary components impeding the passage of small contaminants. Potential concerns with the high-speed centrifugal compressor design should be reviewed when these designs are compared with the conventional designs described earlier in the chapter. An inherently narrow, stable operating range must be considered, as must the fact that like all centrifugal compressors, high-speed machines must be designed for the worst combination of operating conditions. Also, depending
406
Process Plant Machinery
on the specific process conditions, high-speed centrifugal compressors may exhibit lower efficiencies than comparable types of equipment that may be available for a particular application. The user must judge if these characteristics are significant to the potential installation. SUGGESTED READING Ingersoll Rand Co.: Compressed Air & Gas Data. Lapina, Ronald O.: Estimating Centrifugal Compressor Performance. Gulf Publishing Company, Houston TX, 1982. Balje O.E.: A study on Design Criteria and Matching of Turbomachines- Part B. Trans. ASME, J Engr Power, January 1962.
APPENDIX 11 C Barrel Compressors
Barrel compressors were pioneered in the early 1970s. Their specific advantage over high-pressure reciprocating compressors is high delivery volume. Each manufacturer incorporates a combination of design features that are specific to their models of barrel compressor, as well as some features adopted from their regular product line of centrifugal compressors. The operating rotor speeds and pressures involved in each application dictate these selections, as is indicated in the information in this appendix.*
Standard barrel assembly Compressor Selection Although the same aerodynamic design and impeller families are used in the axially split R, RZ, RS series and in the radially
Flg. 11C-1 Standard barrel assembly. 1 - L o w e r half of Inner caslng and r o t o r - restlng on the l a b y r l n t h s - are mounted; 2 - Autoclave and front cover bolted to lower caslng half, bearlngs and seals mounted; 3 - Mountlng of the upper half of the Inner caslng and of the plston-rlngs; 4 - Inner assembly or compressor cartrldge Is llfted from support and deposlted on maln and auxlllary ralls. Rollers on Inner casing on auxlllary, rollers on front cover on maln ralls; crane hook removed; 5 - Cartrldge pushed Into barrel untll Inner caslng rollers are supported on the barrel Inslde; auxlllary ralls removed. Cartrldge further pushed Into barrel untll It Is centerlng Itself; 6 - Wlth erectlon bolts cartrldge Is tlghtly pulled Into barrel. Front cover bolts are then tlghtened. (Source: Sulzer Burckhardt, Swltzerland). ------,i
Inner caslng assembly support Main railing support; auxiliary railing is supported on main railing Railing Permanent roullers on Inner caslng and removable rollers on front cover
* Source: Sulzer Burckhardt. Adapted with permission.
408
Process Plant Machinery
split RB, RBZ, RBS series, the equipment selection in the field of barrel compressor applications is often complex due to the high pressures and powers involved. AXIALLY OR RADIALLY SPLIT?
VOLUME, HEAD, SPEED The performance of a multistage compressor can be represented in terms of dimensionless coefficients by the following equations:
The decision as to whether an axially or radially split casing should be used, depends on:
V ( l ) - D2.u
(11C.1)
Hp = ~# u92 ~ z. u2#
Tightness The radially split design has circular casing joints or flanges with a perfectly even load distribution. The leakage of gas at the two covers can be prevented most effectively. Besides metal-to-metal contact, endless O-rings are inserted in grooves on the two covers. By monitoring the pressure between two adjacent rings the tightness can be controlled. For toxic, flammable and explosive gases the barrel design is therefore always of advantage. For this reason, the API 617 Standard specifies the radially split casing construction for gases containing hydrogen if the hydrogen partial pressure exceeds 13.8 bar (200 psig).
Material stress The cylindrical design with the smallest possible inner diameter is obviously the most suitable construction. With axially split casings the available space for bolting is moreover restricted at the two shaft penetrations. In order to achieve the required tightness, a high contact pressure at the joints is required. The necessary forces in the bolts are often higher than would be required by the static gas forces if the casing flanges were perfectly rigid and flat. For large compressor frame sizes the radially split design can therefore be the only possible solution even at moderate pressures.
Nozzle locatlon and malntenance For operating pressures where an axially split design would be perfectly adequate, barrel compressors are often preferred since the nozzles can be arranged in any radial direction. If the necessary space in the axial direction at the nondriven shaft end is available for the horizontal pull-out of the inner casing cartridge, inspections, rotor changes or complete cartridge replacements can be accomplished quickly without removing any process piping to and from the compressor. For tandem units, on the other hand, the first or low-pressure casing should be of the axially split design up to the highest possible operating pressure. A barrel compressor coupled at both shaft ends has to be removed for overhaul or replacement of the rotor. Seals and bearings of Sulzer barrel compressors can, however, always be serviced and replaced with the barrel casing remaining in place.
s-'1
-
41
0583 9O2O
Flg. 11C-2 Arrangement of a tandem compressor unlt wlth an axlally split LP body and a vertically spllt HP body. (Source: Sulzer Burckhardt, Swltzerland)
u-
(11C.2)
~.n.D 60
(11C.3)
He - Hp
(11C.4)
0p
V u Hp He p~ n r/p D z
Dimensionless flow coefficient Volume flow at inlet (m3/s, acfm) Circumferential or tip speed at impeller outlet (m/s, fps) Polytropic head (J/kg, Btu/Ib) Effective head (J/kg, Btu/Ib) Impeller polytropic head coefficient Mechanical speed (rpm) Polytropic efficiency Impeller outlet diameter (m, ft) Number of impellers
The graph (Fig. 11C-3) shows the relationship between the flow coefficient 9and the efficiency of a multistage compressor. At high ~)-values, the gas velocities in the impeller flow channels increase and lead to a fall in efficiency with increasing ~. At low ~-values, the efficiency falls rapidly because of the increasing influence of the shroud leakage loss and of the unavoidable friction loss in the boundary layers of the impeller disks and blades. It can be seen that the efficiency of a multistage compressor can be improved substantially if the (t)-values of the impellers can be kept in the range 0.03 to 0.09. If we eliminate u from equations (11C.1) and (11C.2) above we obtain ~=
~
,/-~ 9
02
v
~
For given operating conditions V/v/~p is approximately constant. Although Hp remains nearly the same, the power required is related to the effective head and, as seen in equation (11C.4), depends on the efficiency. In high-pressure applications, the high-density gases have an intake volume which gives rise to a low (t)-value with correspondingly low efficiency. A larger flow coefficient (t) and hence a more efficient compressor can then only be obtained by: 9Decreasing the diameter, but keeping z and the tip speed u constant, the latter within the admissible stress limit with due regard to the impeller type, method and material of construction. In this case, the mechanical speed increases in direct proportion, since from (11c.3) n ~. u/D. At the high powers required in barrel compressors, the mechanical speed is often restricted from the driver side. If the driver is a turbine, the mechanical speed is often too low for direct coupling with high-pressure centrifugal compressors. The use of a gearbox between compressor and driver gives additional mechanical losses of the order 2 to 2.5%, but the cost can often be recovered with a smaller and more efficient compressor designed for optimum ~. If a gearbox is used, the admissible mechanical speed is again limited for design reasons (pitch line velocity, bending stress on the pinion, bearing speeds and loads).
Appendix l lC: Barrel Compressors 409
Fig. 11C-3 The graph shows the Influence of the flow coefficient # on the efficiency. 1, 2, 3 ... are the individual Impellers or stages and their efficlencles in a multistage compressor designed for optimum speed and overall efficiency ( O , O , O ) and, alternatively, for reduced speed (V, V, V). (Source: Sulzer Burckhardt, Switzerland) 9Increasing the number of stages without changing the diameter. This results in a speed reduction proportional to Vrz--+--x/~/i where x is the number of added stages. This approach is also limited for reasons related to the design of slender rotors (lower critical speeds). 9A combination of both methods normally allows the best adaption of a compressor to the given mechanical speed of the driver. In practice there are also differences between the head coefficient, axial length and diameter of the different stages that can be used; the polytropic head is not exactly constant, and sophisticated computer methods must be used for the optimization. The above discussion illustrates an important design consideration for multistage compressors in general, but especially barrel compressors for high-pressure applications. For a given tip speed or a given mechanical speed, any attempt to design a more efficient compressor by using impellers with higher flow coefficients leads invariably to a more slender rotor. The design of the rotor is therefore of utmost importance, especially in consideration of the fact that modern turbo-machinery must often be designed for operation close to or above the second critical speed. STABILITY LIMIT At lower gas densities, a compressor may operate smoothly at the specified head, volume and speed range. At higher densities and a correspondingly increased shaft input, the compressor rotor may become unstable. This rotor instability manifests itself by unacceptably high subsynchronous vibration amplitudes. Design measures aimed at a greater insensitivity, such as a more rigid rotor, by increasing the shaft diameter and reducing the bearing span or improved damping of the journal bearings, were helpful, but did not
eliminate the causes which determine the stability limit of a given rotor or compressor. These became the subject of intensive research and testing programs at Sulzer*. It was demonstrated that rotor instability is caused by the swirling leakage flow around the rotor, especially pronounced at the impeller shaft seals. Swirl brakes mounted at these locations virtually eliminated the causes, known also as cross-coupling forces. The stability limit could be shifted way beyond the highest gas densities as required today for gas reinjection projects, thus solving one of the major problems encountered in modern turbocompressors technology for all practical purposes. Measurements on compressors equipped with and without swirl brakes were moreover in full agreement with theoretical stability limit predictions. Hysteresis of shrink fits or of sleeves on the rotor another potential cause of rotor instability, is eliminated by Sulzer's sleeveless rotor construction and the radial dowel pin impeller fixation.
Application Chart The chart (Fig. 11C-4) is a guide in the first planning phase of projects involving high-pressure compression equipment. It allows determination of the approximate number of stages *The results were presented by Sulzer in the following papers: 1. Laterala Vibration Reduction in High-Pressure Centrifugal Compressors, presented and published in the Proceedings of the 9 th Turbomachinery Symposium, December 9-11, 1980. Turbomachinery Laboratories, Mechanical Engineering Department, Texas A & M University, College Station, Texas. 2. Prediction of Stiffness and Damping Coefficients for Centrifugal Compressor Shaft Seals, contributed by the Gas Turbine Division of the American Society of Mechanical Engineers and presented at the 29th International Gas Turbine Conference and Exhibit, Amsterdam, June 4-7, 1984. Transactions of the ASME, paper No. 84GT-86.
4 lO Process Plant Machinery
Fig. 11C-4 The chart shows the volume and the corresponding standard Impeller diameters versus the head. The parameter is the number of Impellers. As a measure of the casing frame size Sulzer uses the Impeller diameter. An RB 28-4 is a barrel compressor with a "nominal" Impeller diameter of casing 280 mm and four Impellers. Such a casing can, however, accommodate also impellers up to 315 mm and down to 250 mm in diameter. (Source: Sulzer Burckhardt, Switzerland)
required to develop the specified head, the approximate frame size, speeds and powers involved.
Barrel Compressor Design
The selection chart is calculated for impeller tip speeds not exceeding 260 m / s - this value corresponds to about 80% of the admissible speed for impellers in standard hightensile strength steels - (t)-values at inlet not exceeding 0.07 and, finally, a scaling down of the impeller diameters in the rear stages. The chart is therefore typical for highpressure barrel compressor selections where the speed is often restricted for mechanical reasons or due to the use of corrosion resistant materials. The scaling down of the impeller diameter is standard practice in order to increase the ~-values of the last stages and, therefore, the overall efficiency. This aspect is particularly important at high powers.
FUNCTIONS OF A CASING
The mean polytropic stage efficiency would extend from 0.83 for a well-utilized compressor, but designed for a moderate total head or a gas with a low density increase from suction to discharge, down to 0.70 for poorly utilized compressors - inlet ~-values below 0.07 - but designed for maximum head and/or a high density increase from suction to discharge. The proper selection of barrel compressors requires special skills and expertise in many fields, such as: 9Reynolds number influence at high gas densities 9Mach number influence, especially for gases of high molecular weight 9Properties of gases and of gas mixtures. Different equations of state, such as published by Soave- Redlich- Kwong, Peng Robinson or Starling-Han BWR, may have to be compared. It is, therefore, recommended that projects requiring high-pressure and high-power compression equipment should be discussed with the manufacturer at an early stage.
1 To serve as a pressure containing vessel: Note: Except for special requirements, such as circulators in nuclear reactors, the construction, manufacture and testing of compressor casings is normally exempt from the approval, inspection and certification procedures of the authorities in the state or country of installation. This practice is evidenced, e.g. in the American National ASME Standard, where it is stated that pressure containers which are an
integral part of rotating devices, such as compressors, are not considered to be within the scope of the ASME Boiler and Pressure Vessel Code (Section VIII, Division 1 and 2).
The reference to this code in API 617 implies, however, that casing stress values shall conform to the values recommended for the selected materials of construction at the specified operating conditions. 2 To be gastight: A casing must be technically gastight for compressors handling flammable, explosive or toxic gases.
3 To accommodate and position as mounting base all other compressor components: Diaphragms with return channels and diffusors, shaft seals, bearing housings with bearings, supporting the rotating element. 4 To provide aerodynamically smooth passages: For the acceleration and the deceleration of the gas velocities between suction, discharge and all interstage nozzles of the respective stage or stage groups an optimum aerodynamic design is required.
Appendix 11 C." Barrel Compressors 411
Fig. 11C-5 The Identical elements of R and RB compressor types are: 1 - Rotor with impellers; 2 - Seals; 3 - Bearings; 4 - Return channels; 5 - Diffusors. The different element between the two compressor types is the casing. It will also be noted that the axial casing guide of axially split casings is located on the bearing housings, respectively on the barrel casing of radially split compressors due to the higher forces on the nozzles. * Z is a compressor with interstage "out" and "in" nozzles for external intercooling and liquid separation and/or external mixing with a sidestream or extraction; S is a compressor with an "in" or "out" nozzle only between the respective stage groups for introduction or extraction of a sidestream. (Source: Sulzer Burckhardt, Switzerland)
For axially split compressors these four functions are all met by cast or fabricated casings consisting of an "upper" and "lower" half, bolted together at the "horizontal" joint or dividing flanges. For reasons of maintenance the nozzles are preferably located in the lower half. For radially split compressors two concentrically arranged casings accomplish the above four functions, ensuring furthermore maximum ease of assembly and disassembly. The inner casing is again of the axially split type. Its design entirely fulfills the 3rd function, and together with the outer casing the 4th function. The compressor is assembled in exactly the same way as an axially split type. The design provides unobstructed access for checking all clearances and the correct alignment of the parts. Besides accommodating the nozzles, the outer casing serves exclusively as gastight pressure container for the complete compressor mounted in the inner casing.
BARREL CASINGS The outer casing of a barrel compressor allows optimum utilization of the material and therefore minimum casing weight. It is of a forged construction with welded-on nozzles, or a casting. The materials of construction are almost exclusively low-alloy steels of high tensile strength. The inner casing is subject to the gas differential pressures only, with compression being the dominant type of stress. The
,!
t
.1
Fig. 11C-6 1 - I n n e r casing; 2 - Cover; 3 - S p r i n g s ; 4-Clearance (variable). (Source: Sulzer Burckhardt, Switzerland) material of construction is cast steel or nodular cast iron which can be used up to high pressures. For special applications, castings in stainless steel or nonferrous materials can be used.
Diaphragms The bending stress on the diaphragms between the individual stages is proportional to the gas pressure. For high operating
412 Process Plant Machinery
Fig. 11C-7 pressures the diaphragms are therefore of a reinforced construction, the material being nodular cast iron, cast steel or plate steel. At low pressures cast iron can be used.
Covers The front cover and the autoclave cover are forgings of the same material as the outer casing.
Bearing housings Standard bearing housings for axially and radially split centrifugal compressors are made of nodular cast iron.
AXIAL AND RADIAL CLEARANCE During start-up of the cold machine, the inner axially split casing warms up faster than the outer barrel casing. This requires adequate clearances between the two casings and minimum friction resistance. The piston-rings on the inner casing ensure this in the radial direction besides facilitating the pushing in and pulling out of the compressor cartridge. For the required clearance in the axial direction the inner casing is connected to either one of the two covers, the autoclave or the front cover, so that the parts can move axially against each other while having no freedom of movement in any other direction. The inner casing with the two covers mounted thus forms a rigid and self-supporting assembly except for the required axial movement. On small and medium frames this design requirement is achieved by spring-type connection as shown in the sketch. The springs are precompressed in order to obtain an axially sufficiently rigid cartridge for the final assembly. For large barrels, guiding rods with restricted axial float are used.
Casing Design Pressures On the preceding pages it became evident that the application limits of high-pressure centrifugal compressors are set by the rotating element; i.e. the critical speeds and the stability limit. Beyond these limits the only method to design a reliable compression unit for a given duty is a multiple casing train. Barrel compressor casings can be built for almost any required final discharge pressure. As an example, Sulzer investigated the feasibility of barrel compressors for 2400 bar (35 000 psia). These compressors would replace hypercompressors of the reciprocating type. It is for this reason that the upper design pressure limit of Sulzer multistage axially split centrifugal compressors is of greater interest. At pressures above those applying for these standardized axially split casings, the radially split barrel casing is generally preferred. For special applications and for inert, nonflammable gases axially split casings for higher pressures than those shown on the chart below have been adopted. Examples: An oxygen compressor unit, model RZ 28-6, with a working pressure of 73.50 bar (1070 psia). This machine is the HP compressor of a tandem unit supplying the oxygen for the gasification of heavy residual oils in a refinery. An air compressor, model RZ271-6, discharging at 62 bar (900 psia). This machine is the HP body of an axial- centrifugal unit for the CAES (Compressed Air Energy Storage) plant at Huntorf, Federal Republic of Germany. The casings of both the above installations were cast in ductile nodular cast iron GGG 40.
Appendix 11C: Barrel Compressors 413
Fig. 11C-8 Design study of a barrel compressor for a final discharge pressure of 2400 bar (35000 psla). For the rotor an electron-beam-welded construction would be used. Shaft and Impeller hub would be an Integral forging. For the covering disc and the blades the Sulzer standard construction with gold/nickel brazing would be used. 1 - Solid electron-beam-welded rotor; 2 - Impeller, Integral with shaft; 3 - Inner casing; 4 - Shrink rings Instead of bolting for Inner casing assembly; 5 - Piston seal rings; 6 - Front cover; 7 - Centering piece; 8 - Balance piston stationary labyrinths; 9 - Labyrinths; 1 0 - Combined bearing-seals. (Source: Sulzer Burckhardt, Switzerland).
Fig. 11C-9 The limits are for information only. The design material stress has been calculated for an assumed casing temperature of 200~ The design pressure of fabricated axially split casings is determined by the actual operating conditions. For medium and large axially split casings segmented case hydrostatic testing can be used. (Source: Sulzer Burckhardt, Switzerland). Standard for cast iron axially split casings Standard for ductile nodular cast iron axially split casings ,,,-. 9 ---
Standard for low-alloy cast steel axially split casings Maximum standard casing design pressure for forged or cast steel radially split "barrel" compressors
414 Process Plant Machinery
Fig. 11C-10
Exploded view of barrel compressor showing outer casing, autoclave cover, Inner casing and main cover.
(Source: Sulzer Burckhardt, Switzerland)
Barrel Compressor Design Features Exclusive for RB, RBZ, RBS Series EXCENTRIC INNER BORE WITH OPTIMUM WALL THICKNESS This standard SULZER design results in an optimum utilization of the material due to equal levels of the hoop stress as
Fig. 11C-11
well as the stress at the nozzle locations as determined by finite-element stress calculation and measurements. Barrel casings can therefore be designed with minimum weight. AUTOCLAVE COVER The principal advantage of the Sulzer autoclave cover barrel construction is the complete external assembly of the compressor, including the seals and bearings at both shaft ends. Alignment, inspection, control and immediate correction of all clearances can indeed be accomplished as if it were an axially
Finite-element grids for three-dimensional stress analysis of a barrel casing and its discharge nozzle. (Source:
Sulzer Burckhardt, Switzerland)
Appendix 11C: Barrel Compressors 415 TABLE 11C.1 Centrifugal Compressor Design Features Common for R, RZ, RS and RB, RBZ, RBS Series Barrel In-line or single-flow Impeller arrangement Sulzer pioneered the in-line impeller arrangement with controlled residual thrust under off-design conditions for high-pressure duties. In 1958, the first high-pressure barrel compressors for 150 bar (2175 psia) and a gas molecular weight of 29 were designed with in-line impeller arrangement. The in-line design produces the lowest alternating thrust at surge conditions and improves the rotor stability.
Controlled Impeller thrust By influencing with small directional blades or other means, the tangential and radial velocity vectors of the gas leakage stream on the outside of the impeller, the static pressure can be controlled. A patent was received in 1970 (Baumann, Eggmann).
Swirl brakes The same patent was the basis for theoretical and experimental investigations on the rotor stability of the influence of swirl brakes in the gas leakage flow. The exact effects of swirl brakes were measured in the late seventies. The influence of cross-coupling forces could be virtually eliminated. Swirl brakes on part or all stages are now standard at high pressures.
High-efficiency low-flow Impellers Besides the most advanced aerodynamic design the high efficiency of SULZER low-flow impellers is also due to the smoothness of the surfaces, minimizing boundary layer losses. The gold/nickel brazing construction technique, introduced by Sulzer in 1968, is a prerequisite for narrow impellers with smooth surfaces.
Solid quill-shaft thrust bearing The thrust bearing located directly on the coupling shaft is a standard design option for tandem units. It allows the installation of the largest possible thrust area and facilitates maintenance. It also results in minimum bearing spans, lowest critical speeds and minimum overall length of tandem units.
TC thrust collar-type high-capacity turbogears Though Sulzer does not manufacture gearboxes, the close cooperation with gear manufacturers allowed Sulzer to take an active part as a consultant in the development of high-capacity thrust collar gears and to develop relevant technical specifications.
416 Process Plant Machinery TABLE 11C.1
(continued)
Impeller dowel pin The Sulzer standard impeller mounting with a shrink-fit, secured by the symmetrically arranged radial dowel pins, is used exclusively on all multistage compressors since 1974. This design feature allows the use of the largest possible diameters and hence the stiffest shafts. It assures perfect concentricity and balance.
Labyrinths rotating and stationary Sulzer's standard is the use of rotating labyrinths for reasons of safety whenever possible. For very high pressure applications where the gas forces could bend the thin rotating strips, stationary labyrinths are used, especially on the balance piston with the large pressure drop from final discharge pressure to suction. For minimum clearance and leakage losses, special abrasive materials can be used for the stationary portion.
split design. This saves time and effort for routine inspections and maintenance work, but also for rotor changeout and for uprating or adjusting the compressor performance map to different operating conditions.
PISTON-RING INNER CASING SEALS Piston-ring seals instead of O-rings guarantee perfect internal sealing of the assembled compressor during operation. They are not subject to aging. They allow liberal clearances between inner casing and barrel casing. They adapt ideally to the thermal expansion of the inner casing. The friction resistance during assembly is minimal, and the risk of damaging O-rings in grooves is nonexistent.
BOLTED MAIN COVER With bolted covers the barrel casing weight and its length can be kept at a minimum. The bolted cover also enables the complete external cartridge assembly. The inner surface of the barrel remains perfectly even, without grooves for inner rings or seals where O-rings could be distorted or damaged during assembly. The bolted front cover, used up to the highest pressures, also facilitates maintenance as it allows easily accessible flanged connections at the circumference. Removal of high-pressure seal oil and seal gas piping is not necessary in order to gain access to bearings and seals or for cartridge removal.
Shaft seal functions CENTRIFUGAL COMPRESSOR SEALS Compressor seals are of the labyrinth type between the individual stages, on the balance piston and at the shaft ends. For these noncontact seals the sealant is the process gas. Its quantity depends on the number of labyrinths, their clearance and the differential pressure between each stage, the balance piston and the suction side and, as regards the leakage to the outside, the suction pressure itself. The internal leakage is continuously recompressed and must therefore be kept as low as possible. At the shaft seals, a process gas leakage can only be prevented by either using a buffer gas or by adding a secondary gastight seal Dry gas seals can be used for flammable or toxic gases. These gastight seals use
the process gas or sealant rendering thus a seal oil system superfluous.
LIQUID SEALING SYSTEMS Except for air, inert gases and applications at low pressures, gastight seals are standard equipment for toxic, flammable and explosive gases or gas mixtures. The sealant can be furnished by the lube oil system, by a booster system taking the oil from the lube oil system or by a separate sealing system. For applications where any traces of oil in the process gas cannot be tolerated, demineralized water or other liquids can be used as sealant. Separate sealing oil systems are often preferred for the compression of gas mixtures at high suction pressures. They are normally required for any compressor handling gases which by entrainment and dissolution in the oil would reduce the flash point and especially the required minimum viscosity for gearboxes and thrust bearings. This risk could only be avoided by introducing on the labyrinth portion of the shaft seal an inert or a sweet and light buffer gas. For gas mixtures containing hydrogen sulfide and other acidforming components, the use of a separate sealing oil system is normal practice. For such cases the contaminated or sour oil returning from the shaft seal can be collected in a separate tank for periodic or permanent reconditioning in an oil reclamation unit. Since the low-pressure side of the balance piston is always connected to the compressor suction side, both shaft seals operate at roughly equal pressures. In order to prevent any ingress of the sealant liquid into the compressor and hence into the process gas, the preceding labyrinth seal is always purged via a port by a sealing gas stream at a somewhat higher pressure. Unless a buffer gas from an external source is used, this purging or sealing gas is tapped off from the compressor. Most of it flows directly back into the compressor, where it is recompressed together with the balance piston leakage. A much smaller amount, adjusted to the required minimum velocity in the labyrinth seal by the flow orifices from the seal oil traps and a pressure reducing valve in the feed line, flows towards the gastight liquid-film or mechanical seal, where it mixes with the gas side sealant liquid. This small amount of gas entrained in the sealant is either flared or recuperated in a small gas-oil separator and then returned to the compressor suction.
Appendix 11 C: Barrel Compressors 417
Fig. 11C-12 The schematic illustrates the large number of options for compressors equipped with gastight shaft seals. Increasing costs for gas, oil and energy require close cooperation between client and compressor supplier in order to determine and design the most appropriate sealing system. Special care is also needed for the design, enginsering and procurement of the equipment. (Source: Sulzer Burckhardt, Switzerland) The supply pressure of the sealant must necessarily be kept above the buffer or purge gas pressure in order to ensure positive sealing and a minimum flow between the stationary and rotating faces of the seal for cooling. The design of the shaft sealing system depends also on: 9Block-in time 9Coast-down time 9Settling-out pressure and starting pressure 9Sealant supply back-up sources
Centrifugal Compressor Shaft Seal Applications DRY SEALS Labyrinth noncontact seals Application. Is generally used for inert and nontoxic gases. It can be used for flammable and toxic gases at lower pressures, provided an inert buffer gas is available (steam, nitrogen). At pressures close to atmospheric pressure, a scavenging system can be used as an alternative or in addition to a buffer gas. Labyrinth seals are also preferred for extreme
temperature applications, where a liquid sealant could freeze or evaporate. An application are boil-off compressors in LNG regasification and storage plants. Variations. Labyrinth seals can be designed with extremely small, initially negative clearances. Carbon rings and other special abrasive material are then used on the stationary side. For oxygen service, selected materials of construction with maximum resistance to ignition are used.
Gastight dry gas seal Application. for flammable, toxic and pure gases up to temperatures of 200 ~ One single seal is capable of sealing a pressure differential of 100 bar. For higher levels the number of seals is increased accordingly in order to reduce the pressure step by step. Process gas tapped from the compressor discharge is fed to the seal through a fine (2 #) filter. The small leakage is led to the stack. A fan ensures that an ample air flow dilutes at all times any accumulation of gas/oil vapor behind the seal to a safe level, well below an explosive mixture. Variations. An additional seal serving as standby can be installed (tandem arrangement). Installation of a labyrinth seal with inert gas injection between two dry gas seals is possible. WET SEALS Gastight liquid-film seal Application. for flammable and toxic gases. Liquid-film seals are most suited for difficult gases, high pressures and speeds
418 Process Plant Machinery
Fig. 11C-13 Dry, noncontact labyrinth seals with or without ports. Ports are used for discharging the process sealing gas into a low-pressure compressor for recompression, for introducing a buffer gas and/or scavenging. (Source: Sulzer Burckhardt, Switzerland)
and, in general, all applications where uninterrupted service for several years is required. Their disadvantage is the larger amount of sealant in contact with the gas side of the seal. For sour gas applications, separate sealing oil systems should therefore be employed, unless a sweet buffer gas is available. Variations. At high pressures the floating ring on the atmospheric side must be of the balanced type. Windback-type floating rings on the gas side permit a substantial reduction of the contaminated sealant return flow. This variation is therefore preferred for gases containing hydrogen sulfide. It requires gravity control of the sealant supply pressure. Gastight mechanical seal Gastight double-mechanical seals Application. The mechanical seal is an option to the liquidfilm seal. Due to its advantage (minimum gas side sealant flow) it is often the preferred seal for clean gases and also for "dirty" gases, except for applications at high pressures (above approximately 100 bar, 1450 psia) and where the required shaft diameter at the seal would result in face velocities in excess of approximately 90 m/s. Variations. As for liquid-film seals, the floating ring on the atmospheric side must be of the balanced type for high operating pressures. Up to medium pressures the double-mechanical seal is preferred. Besides a lower sealant flow the doublemechanical seal performs as a standstill seal. An overhead run-down tank is not required.
Fig. 11C-14 Dry gas seal of tandem arrangement. A fan ensures an adequate air flow avoiding any accumulation of gas/oil vapor mixture behind the seal. Dry gas seals require no seal oil system, hence capital costs, energy consumption (no pumps) and requirements are reduced. (Source: Sulzer Burckhardt, Switzerland). m
Atmospheric or air side
m
Process gas, also sealing or buffer gas Sealant liquid, seal oil
n
Sealing gas and sealant liquid mixture: seal gas to flare or back to suction via separator and/or active carbon filter; contaminated sealant to separator or collector tank to waste or to oil reclamation unit Gas/oil vapor mixture diluted with air
Couplings, Bearings For Barrel Compressor COUPLING SELECTION The low density of gases in comparison to liquids requires compressors and turbines operating at the highest possible mechanical speeds. Couplings must be carefully selected with due regard to 9normal, transient and especially pulsating torques; 9movements of the coupled rotors and their casings due to thermal expansion and other causes; 9maximum combined axial thrust of each component under extreme conditions, such as e.g. surging or loss of vacuum in case of condensing turbine drivers.
Appendix 11C: Barrel Compressors 419
Fig. 11C-15 Wet liquid-film seal with floating rings on gas and atmospheric side. For minimum liquid sealant flows towards the gas side, the gas side rings can be of the windback type (upper half), in order to reduce the "sour" seal oil quantity. The threaded windback grooves act as the stator of a pump. The atmospheric side ring can be of the "unbalanced" or of the hydraulically balanced type (lower half). The shaft at the seals is protected by chromium plating or by replaceable sleeves. (Source: Sulzer Burckhardt, Switzerland)
Fig. 11C-16 Wet single-mechanical (contact) seal. On the atmospheric side the seal is of the liquid-film type. The floating ring can be of the "unbalanced" or of the hydraulically balanced type (lower half). R Reference pressure (connected to the seal at the compressor discharge side) O Sealant overflow for cooling of seal. (Source: Sulzer Burckhardt, Switzerland)
At the high speeds and powers dry, nonlubricated couplings of the quill-shaft or of the metallic-diaphragm type are generally preferred to all other coupling types. The requirements for low weight and high balance quality often exclude the use of other coupling types. Toothed-type coupling depend moreover on the steady supply of lube oil, free from solid particles. Another element requiring careful selection and dimensioning are the thrust bearings. Whenever possible, the total number of thrust bearings per shaft string should be kept at a minimum. For this reason the order of preference is: 9Solid couplings with a flexible quill-shaft as per API1671, 1st edition, 1979, paragraph 2.1. The quill-shaft allows the elimination of high-speed thrust bearings of geared units and, for direct driven units, the use of one single-thrust bearing. 9Diaphragm-type couplings as per API671, 1st edition, 1979, paragraph 1.1.3. Diaphragm couplings are necessary when the two connected rotors or gearbox shafts are or must be equipped with their own thrust bearing. The diaphragm acts as a soft spring in the axial direction. Besides the required misalignment capability, it must be designed for compensating the maximum combined axial displacement of the two coupled shafts from their fixed points.
Flg. 11C-17 Wet double-mechanlcal seal. Thls seal requires the least quantlty of sealant llquld. Most of the sealant, enterlng below, Is used for coollng and Is therefore flowlng dlrectly back to the reservolr. The appllcatlon of the double-mechanlcal seal Is, however, restrlcted to low and medlum pressures. (Source: Sulzer Burckhardt, Swltzerland)
Both coupling types should also be given preference at lower speeds, where other coupling types may, however, offer special advantages. For motor drivers, toothed-type couplings with
limited axial end float and, in special cases, maximum torque limiting shear bolts can be used on the motor side besides
420 Process Plant Machinery
Fig. 11C-18 Solid quill-shaft coupling. (Source: Sulzer Burckhardt, Switzerland)
Fig. 11C-21 Four-tilting-pad radial bearing, type R 4K, for high speeds and normally used in barrel compressors. (Source: Sulzer Burckhardt, Switzerland)
Fig. 11C-19 Tilting pad self-leveling or self-equalizing thrust bearing with Individual oil Injection for each pad for flexible shafts. For stiff shafts, such as gearbox shafts, self-leveling pad support is not required. (Source: Sulzer Burckhardt, Switzerland)
Fig. 11C-20 Two-lobe radial bearing, type R 2F, for heavy rotors and speeds normally below 7000 rpm. (Source: SuIzer Burckhardt, Switzerland)
couplings with transient and oscillating torque dampening capabilities. The hubs for high-speed couplings should be tapered and designed for hydraulic fitting without keys. At low speeds and powers, straight bores and keyed shrink tiffing can be used.
Fig. 11C-22 The main and intermediate gearboxes of the below-shown compressor unit are of the single-helical thrust collar type, allowing the use of solid quill-shaft couplings with one common low-speed thrust bearing on the main gearbox. This compressor unit serves a nitrogen reinjection scheme for enhanced oil recovery (EOR) in the United States. The nitrogen is compressed from 2.46 to 58.00 bar (35.7 to 840 psla) injection pressure. For injection pressures above approximately 80 bar (1160 psla), the HP casing would be a radially split barrel compressor. The compressor casings are cast in ductile cast iron. (Source: Sulzer Burckhardt, Switzerland)
Appendix llC: Barrel Compressors 421
Shaft String Layout
Fig. 11C-23 Shaft string layout. (Source: Sulzer Burckhardt, Switzerland)
422 Process Plant Machinery RECOMMENDED STANDARD SHAFT STRING LAYOUT
be reviewed carefully and jointly by the user and the equipment manufacturer.
Layouts 1 to 8 show tandem units. If the second or HP compressor only is of the radially split design, it should be located at the end for axial pull-out of the cartridge. Layouts with a single compressor are similar as indicated by the dotted line. The thrust bearing on the gear of layouts 3 and 4 would then be on the direct driven compressor. Layout 2 shows the preferred thrust bearing location on the intermediate quill-shaft for tandem compressors operating at the same speed.
The certification that the ultimately selected materials conform to their specifications is of utmost importance. The quality assurance measures and methods of testing should be clearly defined, and all parts for which certificates, witnessed or observed tests are required should be listed. It is Sulzer's practice to submit, together with the compressor quotation, defined quality assurance and testing programs for all vital parts and to comment explicitly on the deviations, if any, from the requirements of the material requisition.
~ Compressor
i
Drlver The arrow indicates the required sense of rotation. R and RB compressors always rotate clockwise in the direction of the gas flow.
~TC
(Thrust Collar) type gearbox
If single-helical gearboxes without thrust collar are used, layouts 3, 4, 5 and 6 would require one, layouts 7 and 8 two additional thrust bearings on the high-speed side.
H H
The governing material specification is the one in general use in the country of the material supplier. For SULZER compressors produced in Europe, forgings and castings are normally procured from European suppliers. The temperature limits for which the materials as listed below can be used, depend on the actual requirements as regards tensile and impact strength. Materials for very low temperatures have not been listed.
Solid quill-shaft coupling Diaphragm coupling
(Material, Table 11 C.2 on facing page).
A low speeds other coupling types with axial displacement capability could be used, often designed with limited actual float for electric motor drivers, equipped with journal bearings only.
Barrel Compressor Applications
',1', Thrust bearing, radial bearing Layout 9 shows a double-helical speed increaser. The arrow shows an alternate thrust bearing location on the gearbox. With single-helical gearboxes, the layout would be similar except that the gearwheel would also require a thrust bearing. Layout 10 shows an epicyclic gearbox allowing the same or opposite rotation of the output shaft. Layout 11 is an arrangement with a multiple-shaft gearbox with two outputs. This layout can be of advantage at high speed ratios of multiple-casing barrel compression units and where space is restricted.
Materials of Construction For Barrel Compressor The gas or gas mixture to be compressed and the required physical properties determine the most important selection of the materials of construction. Besides the recommendations of industry standards and of the suppliers of the materials, the experience of the user and of the compressor manufacturer should be taken into consideration. This is especially important for corrosive and wet gas service. An important example in this regard are gases containing hydrogen sulfide. In such cases the materials of construction should conform to the NACE Standard MR-01-15 and subsequent revisions. The materials of SULZER compressors as tabulated below are limited to the most frequently used specifications and the most important components: impellers, shafts, SULZER solid couplings and the casings. Quite often materials of different qualities can be proposed. Examples are casings in ductile nodular cast iron as an alternative to steel. In other cases both stainless and nonstainless materials are to be considered. For the impellers, martensitic stainless steels are often preferred since they offer a higher strength, allowing somewhat higher impeller tip speeds. In such cases the most appropriate selection should
Applications for barrel compressors have grown very significantly since their introduction, as is indicated by the figures below.
OIL AND GAS PRODUCTION, SYNFUEL VIA OXYGEN, REFINERIES, SYNGAS
The fact that oil and natural gas are precious, nonrenewable resources contributes to a diversified and fastgrowing market for centrifugal compressors of both the axially and, especially, the radially split barrel type. The selection of the most suitable compression equipment and the drivers depends on the conditions, differing from site to site, and the resultant technical and economical considerations, such as 9the type and geographic location of the field, the products- oil, commercial natural gas, LPG, NGL, L N G as well as the markets and the mode of transportation which are in turn related to 9the methods of production: reinjection of gas for conservation and for maintaining reservoir pressure, EOR recovery with inert gases, gaslift operation, insitu combustion. Successful operation of a barrel design for the compression of pure oxygen up to 140 bar (2030 psia) was a major step toward high-pressure oxygen plants, as required for the production of synfuels based on coal conversion. For cracking and upgrading processes in refineries, barrel compressors for hydrogen-rich gases over a wide range of pressures and flows are required. An impeller layout aimed at a wide stable operating range is often mandatory. The barrel-type centrifugal compressor competes with positive displacement-type compressors. For the latter, standby units are normally required. It is for this reason that a single barrel compressor unit is often preferred. For the compression of the light synthesis gas mixtures in tonnage methanol and ammonia plants, multicasing barrel compressor trains are standard equipment.
Appendix 11C." Barrel Compressors 423 TABLE 11C.2
Part
Sulzer/DIN specification form
Comparable or equivalent U S specification
Notes
9For welded impellers
Impeller covers and hubs
25 CrMo4 forged low-alloy steel
AlSl 4130
Impeller covers and hubs
23 CrNiMo 747 forged low-alloy steel
ASTM A 471 Class 1 upto7
9For welded impeller constructions
Impeller covers and hubs
X 5 CrNi 134 forged martensitic stainless steel
ASTM A 182 Grade F 6NM
9For welded and gold/nickel vacuum-brazed impellers 9Impeller material if stainless steel is required 9NACE (editorial revision 1984) accepts this material with 23 HRC
Impeller covers and hubs
X 6 CrNiMoCu 145 precipitation hardening forged stainless steel
14-5 PH ASTMA 705 Type XM-25 UNS S 45 000
9For welded and gold/nickel vacuum-brazed impeller constructions 9NACE (editorial revision 1984) accepts this material with 31 HRC which allows substantially higher tip speeds for corrosive gas service
Impeller blades
Dye-pressed blades are of the same material in plate form as the hub and cover. Blades can also be machined out of the integral hub or cover forging
Shaft
25 CrMo 4 forged low-alloy steel 34 CrNiMo 6
AISI 4130
GGG 40 ferritic nodular cast iron GS-C 25 carbon cast steel
ASTM A 395
Casing axially split
GS-17 CrMo 55 low-alloy cast steel
ASTM A 217 Grade WC 6
Casing axially split
GS-17 CrMo 9 10 low-alloy cast steel G-X5 CrNi 134 stainless cast steel
ASTM A 757 Grade D1 Q1
Solid quill-shaft coupling Casing axially split Casing axially split
Casing axially split
9Good resistance against stress cracking corrosion (sour gas service) if heat-treated to a maximum yield strength not exceeding 620 N/mm 2 (90 000 psi). Admissible rated impeller tip speed is then limited to approx. 285 m/s, depending on wheel type
AISI 4340
ASTM A 216 Grade WCB
ASTM A 757 Grade E 3N ASTM A 487 Class CA 6 NM
Casings radially split
For cast steel casings: same material specifications as above apply
Casings radially split
25 CrMo 4 forged steel ASTMA 336 Class X2CrNiMoN225 forged F 12 A182, F51 steel For cast casings: same material specifications as above for axially split casings apply
Inner casings of radially split barrel compressors Bearing housing
GGG 40 ferritic nodular cast iron
9Standard material for high tip speeds up to 360 m/s, depending on impeller type and heat treatment
ASTM A 395
Duplex
9For special applications fabricated or cast steel housings are used
424 Process Plant Machinery
Fig. 11C-24 Aerial view of the El Morgan complex in the Gulf of Suez. The platform on the right shows the gaslift compression modules engineered and furnished by Sulzer. The weight of each module is approximately 250 t. The five identical compressor units are driven by SULZER type $3 gas turbines. After a total of over 200 000 accumulated running hours at the end of 1984, the average utilization factor was 91.05%, the availability factor 96.3%. and the reliability factor 99.3%. (Source: Sulzer Burckhardt, Switzerland)
Fig. 11C-25 Reinjection compressor unit for the Moomba Field in Australia. The tandem barrel compressor unit is designed for a hydrocarbon gas mixture with a molecular weight of 32 and a final discharge pressure of 202 bar (2930 psla). The driver is an industrial gas turbine with a site power rating of 2890 kW and an Integrated gearbox. (Source: Sulzer Burckhardt, Switzerland)
Appendix llC: Barrel Compressors 425
Fig. 11C-26 Syngas compressor unit for an ammonia plant during shop assembly. The recycle compressor is on the left, the feedgas compressor on the right. The discharge pressures are 227 and 236 bar (3300 and 3450 psla) respectively. For this plant, a 13 000-kW electric motor driver had been chosen. The required two-stage gearbox, rated 13 MW, with a total speed ratio of 1490/14 530 rpm also provides the driving shafts for the mechanical lube and seal oil pumps. (Source: Sulzer Burckhardt, Switzerland)
Fig. 11C-27 Axially split medium- and radially split high-pressure compressor of a gas Injection compressor train for an enhanced oil recovery (EOR) project in Texas, USA. The plant consists of three injection trains with a total Installed power of 45 000 kW. The low-pressure compressor, model RZ 71-7, is coupled to the steam turbine driver. The medium- and high-pressure compressors as shown below, models RZ 35-8 and RB 28-7, are driven via an intermediate gearbox of the thrust collar type from the low-pressure compressor. The gas is compressed from 1.0 to 130 bar (14.5 to 1885 psla) Injection pressure. (Source: Sulzer Burckhardt, Switzerland)
426 Process Plant Machinery
Fig. 11C-28 Offshore modules for the Bombay High South oll and gas field. The modules with a weight of approximately 1300 t were engineered by IHI and completed in the IHI Aichi works In Japan In December 1981. The modules are for associated gas service and are equipped with tandem barrel compressor units, type RBZ 45-6 plus RB 31-7, driven via thrust-collar-type gearbox by an aeroderivative gas turbine with an ISO rating of 14780 kW. The compressors were built by the SULZER licensee IshikawaJima-Harima Heavy Industries Co. Ltd. (IHI). Each unit compresses 2.4 MMNm3/d gas from 7.5 to 87.8 bar (109 to 1270 psla). (Source: Sulzer Burckhardt, Switzerland)
Fig. 11C-29 One of three identical units for an offshore gaslift operation in the Persian Gulf. The tandem units are driven via gearbox by a 10 360 kW industrial gas turbine. The gas is compressed from 13 to 152 bar (188 to 2200 psla). The compressor casings, model R 28-7 for the LP duty and RB 28-8 for the HP duty, are equipped with mechanical seals. The barrel compressor, located at the end of the shaft string, is shown with the Inner cartridge removed for maintenance. (Source: Sulzer Burckhardt, Switzerland)
Appendix llC: Barrel Compressors 427
Fig. 11C-30 High-pressure reinJection unit for the Alrar Field in the Sahara. The three identical units compress natural gas from 81.7 to 322.7 bar (1185 to 4680 psla) and are driven by industrial-type gas turbines. For the work tests at full load, nitrogen was used up to a discharge pressure of 388 bar (5630 psla) in order to verify the rotor stability. The gearboxes were furnished by Maag. They are of the thrust-collar design with a throughput of 21 830 kW at the rated speed ratio of 4670/12 924 rpm. (Source: Sulzer Burckhardt, Switzerland)
Fig. 11C-31 Tandem unit for gas compression duty on a platform in the North Sea. For this plant three identical units were furnished, mounted on three-point support skids with motor and gearbox. The lube and seal oil systems are located on separate skids. The picture shows the two compressors with acoustic enclosures, ready for shipment. (Source: Sulzer Burckhardt, Switzerland)
428 Process Plant Machinery
Fig. 11C-32 Syngas compressor train in a 1000t/d methanol plant. The unit Is driven by a high-efficiency, "low-speed"
(8000 rpm) condensing-extraction turbine. The four barrels, operating in series, are arranged in a H-type configuration on
the two sides of the multiple shaft gear-box. Compressors and the single driver can therefore be designed for optimum speeds. The resultant gain in efficiency Improves the overall energy consumption besides compensating for the relatively large mechanical losses of the 25 MW gearbox. (Source: Sulzer Burckhardt, Switzerland)
Fig. 11C-33 For gas transmission projects, multistage barrel compressors in pipeline configuration - horizontally opposed
nozzles - are used, allowing direct coupling and optimum efficiency. The picture shows the Ruswil station, equipped with a Suizer barrel compressor, driven by a Sulzer split-shaft gas turbine, type $3, with four combustors. Sulzer gas turbines are used for crude oil and gas pipelines in Romania, lran, Egypt, Libya, Germany, and Switzerland. (Source: Sulzer Burckhardt, Switzerland)
Appendix llC: Barrel Compressors 429
Fig. 11C-34 Refinery gas compressor, type RB 31-5, on the Sulzer (Zurich) compressor test stand. For this performance test, the job lube and seal oil systems as well as the test stand gearbox and motor were used. The compressor is designed for compressing a hydrogen-rich gas with a molecular weight of 4.7 from 156 to 188 bar (2260 to 2730 psia). The job driver as a Siemens high-speed condensing turbine. After the performance test, a string test with the job turbine at full speed and power was conducted on the Siemens turbine test stand in Wesel. (Source: Sulzer Burckhardt, Switzerland)
APPENDIX 11D
Isotherm Turbocompressors
Common Features Common Features are incorporated to best satisfy the requirements of the end-user, such as: 1 Low power consumption
6 Easy erection 7 Simple maintenance 8 Minimum space requirement 9 High reliability
2 Resistance against corrosion
And this is how it has been achieved and proved in hundreds of industrial applications:
3 High rotor stability - low vibration level
1 Low power consumptlon
4 Low noise level 5 Simple shaft-string configuration
9Intensive intercooling of the gas reduces the inlet temper-
ature into the subsequent stage and therefore its power requirement. (continued on page 432)
Fig. 11D-1 Air compressors, type RIK 80 and RIK 56. Transportation as a slngle-Iift package. (Source: Sulzer Turbo Ltd, Z0rlch, Switzerland) *Source: Sulzer Turbo Ltd, ZOrich. Switzerland. Adapted with permission.
430
Appendix 11D." Isotherm Turbocompressors 431 TABLE 11D.1 Type Range
432 Process Plant Machinery
Fig. 11D-2 Influence of flow coefficient of the first Impeller on the efficiency of subsequent stages. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland) 9Optimization of the distribution of the total cooling surface within the individual cooling stages with respect to heat load, cooling effect and air-side pressure drop adds to the high overall efficiency.
9The skillfully designed short flow path achieved by the single-shaft in-line design with cooler tube bundles integrated in the casing, reduces the pressure losses on the gas side as there is no external piping. 9The staggered high-flow impellers ensure optimal combined performance of all stages, avoiding the lower range of flow coefficients which exhibit a progressing sharp drop of efficiency (Figure 11D-2). All impellers are of fully welded or welded and brazed construction with the blades shaped in three dimensions (Figure 11D-3). 9In the case of very dirty cooling water, an automatic cleaning system for the cooler tubes can be installed.
This would extend time between overhauls without impairing long term efficiency.
2 Resistance against corrosion 9Most of the carefully designed flow path is handling hot superheated air;, the not quite saturated air after the cooler is taken by the shortest raise to the next impeller (Figure 11D-4). 9Thanks to the vertical position of the built-in lateral cooler tube bundles the inertia-type water separators fitted at the outlet of the coolers (except for the RIO types) show a very high separation efficiency enhanced by the effective condensate removal by mere gravity (Figure 11D-5 and 6). Due to this and owing to the subcooling effect along the tube fins, the air entering the following stage has a mean temperature just slightly above the dew point, which again inherently reduces the proneness to erosion and corrosion. The condensate is drained by automatic traps.
3 Rotor stability In conjunction with the radial bearings and the special coupling technique, turbocompressor rotors are designed for inherently high stability under all practical operating conditions. This is achieved by the following main features (Figure 11D-7):
4 Low nolse level The sturdy radial casing with the built-in coolers have an attenuating effect on the noise generated by the active highvelocity parts embedded in this compact outer package. The same applies to the double-casing axial part of the ARI series. The noise level is therefore lower than that of a centrifugal compressor with separate external coolers and the necessary interconnecting piping. In case of severe noise level restrictions, the compactness of the compressor results in a simple and inexpensive noise hood covering compressor and gear.
5 Simple shaft-string configuration Fig. 11D-3 Welded Impeller of high flow coefficient. (Source: Sulzer Turbo Ltd, ZUrich, Switzerland)
The single-shaft in-line concept allows a very simple configuration of a complete motor or steam turbine driven
Appendix liD: Isotherm Turbocompressors 433 TABLE 11 D.2 Deslgn
Advantages
1 Solid, sturdy rotor with shrunk-on dummy piston
Minimum sensitivity to critical speeds and unbalance due to higher rotor stability; reduction of rotor internal damping
2 Shrink fit secured by symmetrically arranged radial dowels for impellers
No need for keys and distance bushings; fixation assures concentricity and perfect balance under extreme operating conditions; allows larger shaft diameters; reduces stress on shaft and impeller
3 No shaft sleeves between stages
Reduces rotor hysteresis and increases running stability
4 Labyrinths always on the rotating element. Stainless steel strips caulked into the shaft and impeller grooves
No distortion of rotor due to local heating-up in case of rubbing; labyrinths can be refitted easily
5 Nickel plating or other coating of shaft portions exposed to corrosion, if necessary
Plating instead of shaft sleeves is a more direct protection; allows larger shaft diameters
6 Tilting-pad radial bearings for higher speeds
Improves running stability; no oil-whip; higher external damping
7 Solid coupling, tightly bolted to flexible intermediate shaft
Improves reliability due to elimination of high-speed thrust bearing and toothed-type couplings; no torque lock thrust on high-speed thrust bearing
8 High-flow impeller at suction
Improves overall efficiency
Fig. 11D-4 Temperature and humidity conditions: m hot a cool, but well above dew point M cold, but not yet saturated 9 cold and condensing (Source: Sulzer Turbo Ltd, Z0rich, Switzerland)
compressor train with a minimum of shafts and bearings. The common solutions generally are:
load on thrust bearing due to torque lock. Standard for ARI and semi-packaged RIK types.
a Steam turbine direct drive with one common axial thrust bearing in the steam turbine and solid coupling with flexible
b Steam turbine direct drive with individual thrust bearings.
intermediate shaft between turbine and compressor. Four journal bearings. Axial thrust compensated. No additional
Four journal bearings. Axial thrust not compensated. Additional load on thrust bearing due to torque lock caused by thermal expansion of shafts taken up by conventional
434 Process Plant Machinery gear coupling. Depending on magnitude of transient thermal expansion of compressor and turbine rotor, a diaphragmtype coupling can be used, reducing the additional axial thrust and requiring no lubrication. Alternative for ARI and semi-packaged RIK types.
c Motor drive with speed increasing gear of conventional design. Axial thrust bearing on compressor and gearwheel shaft. Eight journal bearings including gear and motor. Solid coupling and flexible shaft between compressor and gear; no torque lock. Solution suitable for high-power lowspeed compressors. Preferred alternative for the large ARI types.
d For higher speeds with motor drive, the technically and economically superior solution, described in Figure 11D-9 to 11D-11 is applied. SOLID COUPLINGS WITH THRUST COLLAR GEARS Particularly for electric motor driven high-speed compressors of medium to high power and/or pressure, it is normal practice to make extensive use of solid couplings allowing the use of only one axial thrust bearing for single- or multiple casing arrangements. An intermediate shaft, flexible enough to allow for considerable misalignment, is inserted between the two shaft ends of the machines to be coupled together (Figure 11 D-9). In the case of motor-driven units, the normal technique is to use single helical gears provided with thrust collars on the pinion shaft, as shown in Figures 11D-10 and 11. The thrust collars not only neutralize the axial thrust created by the meshing of the teeth cut at an angle to the axis of the shaft, but also transmit the residual axial thrust of the high-speed rotor train to the thrust bearing on the low-speed wheel shaft. Good gear meshing requires parallelity of gear and pinion shaft and automatically assures parallelity of the contact surfaces of thrust collar and wheel rim. The slight tapering of the thrust collars is responsible for the formation of a wedge-type oil film creating a pressure zone spread out on an enlarge surface with a pressure distribution very similar to that of a standard oil-lubricated journal bearing. The relative motion between the two contact surfaces of the thrust collar system is a combination of rolling and sliding and takes place near the pitch circle diameter, resulting in a very small relative velocity. The thrust transmission is therefore effected with almost no mechanical losses. The considerably reduced losses of the single thrust bearing on the low-speed shaft as compared with the high losses of individual thrust bearings on the high-speed train lead to a substantial power saving. Moreover, this low-speed bearing can be more amply dimensioned to provide a much higher overload capacity. This coupling arrangement avoids heavy overhung gear couplings which are usually responsible for not clearly defined lower critical speeds and for the phenomena of torque lock leading to additional loading of the axial thrust bearing. The resulting axial friction forces can become quite substantial if insufficient attention is given to the cleanliness of the lubricating oil. This arrangement is, therefore, the preferred solution. Its strict application is clearly visible on the air compressor train (Figure 11D-8).
6 Easy erection on site and dismantling for inspection due to: -
Package construction
-
One single horizontal plane of the axis
-
Vertical cooler bundles easily withdrawable
-
No heavy and cumbersome cross-over piping between compressor and external intercoolers
7 Simple and low-cost maintenance, because all components are easily accessible. 8 Minimum space requirement through compact single-shaft design with integrated coolers. Low elevation of operating floor for ARI types; skid-mounted single-lift package with integrated gear and lube oil system for RIK and RIO types. Fig. 11D-5 Intercooler outlet with water separators. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland)
9 High reliability by using well-known elements proven in hundreds of machines of the same basic design.
Appendix 11D: Isotherm Turbocompressors 435
Fig. 11D-6 RIK model with intercooler tube bundles in the casing bottom half. The vertical water separators at the cooler outlet ensure effective drainage of the condensate. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland) JOURNAL AND AXIAL BEARINGS 9Two-lobe journal bearings are used on the larger frame sizes of the ARI series running at a moderate speed (Figure 11D-12). 9Tilting-pad journal bearings are incorporated in the RIK and RIO series operating in a higher speed range. They contribute to the high rotor stability at high rotational speeds (Figure 11D-13).
The horizontally split journal bearings are white-metal-lined and forced-feed-lubricated. Adjusting plates with a slight curvature in axial direction allow the bearings to be set accurately on erection. Shims placed between the plates and the bearing shell make a possible corrective realignment very easy. Thermoelement connections for white metal temperature measurement are fitted. 9The axial thrust bearing is normally located on the low-speed shaft of the gear. In multicasing arrangements with no gears it is normally located on the intermediate shaft. The thrust bearing is fitted with a load equalizing system. The pads are individually lubricated (Figure 11 D-14).
NP = reference point (100 %) = design point
(x = angular position of the inlet guide vanes (RIK models) or the adjustable stator blades (ARI models) Valid for air at constant inlet data. Depending on the specific process requirements, such as higher over-load capacity, a certain pressure rise to surge, maximum efficiency at design point or rather at a certain part load, the process design point NP may be placed differently in the characteristic curve.
Design Features (RIK) SKID-MOUNTED SINGLE-LIFT CONCEPT The various frame sizes are skid-mounted units with built-in intercoolers and integrated lube oil system. They are suitable for mounting at grade. The erection on site is reduced to mere setting of the skid on a simple foundation slab, lining up with driver and connecting gas and cooling-water piping (Figure 11 D-22). Alternatives to the standard motor-driven concept are possible. For example: 9Compressor directly coupled to the steam turbine driver or expander
436 Process Plant Machinery
Fig. 11D-7 Design principles of Sulzer turbocompressor rotors. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland)
9A booster coupled to the compressor, with or without intermediate gear 9Separate freestanding lube oil system 9Suction nozzle facing downwards
passages to and from the coolers is the result of exhaustive model tests and ensures that the gas flow in each stage is equally distributed between the parallel cooler elements (Figure 11 D-24). For the hydrostatic tests the casing is divided into several chambers and submitted to a water pressure of 1.5 times the maximum possible operating pressure of the corresponding compartment.
CASING
SEALS
The horizontally split casing contains the five centrifugal stages and three pairs of vertical intercoolers (Figure 11 D-23). The axial inlet ensures ideal flow conditions through the inlet guide vanes into the first stage. The bearings can be inspected without having to disconnect the gas or oil piping nor to disturb the casing top half. The efficient aerodynamic form of the flow
The shaft and interstage seals are of the labyrinth type. The stainless steel strips are fixed in grooves in the rotating parts (shaft, impeller hub and cover disc) and have a very small radial clearance against stationary plastic ring segments in the corresponding partition walls. Internal and external leaks are thus kept to a minimum (Figure 11D-7).
Appendix liD: Isotherm Turbocompressors 437
Fig. 11D-8 Typical shaft string configuration of a motor-driven ISOTHERM compressor with booster. Axial thrust transmission according to Figures 11D-9 to 11 with one single thrust bearing on the low-speed side of main gear. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland)
Fig. 11D-9
Fig. 11D-10 Method of axial thrust transfer in a single helical gear with thrust collar F u force; u - Peripheral speed; p - Pressure. (Source: SuIzer Turbo Ltd, ZOrich, Switzerland)
-
Peripheral force; FA - Axial
438 Process Plant Machinery The axial thrust of the impellers is almost entirely compensated by a balance piston at the discharge end of the compressor. The piston is provided with labyrinth strips rotating against a white-metal-lined steel ring. With the in-line arrangement of the impellers the resulting total axial thrust of the rotor is, contrary to a back-to-back arrangement, always acting in the same direction and of the same order of magnitude under all operating conditions (normal, reduced load, surge and rotating stall). The balance piston is dimensioned in such a way that the compensated residual thrust is reduced to a minimum, but still always acting in the same direction. With this method the axial thrust bearing need not to be oversized, and the bearing losses are reduced accordingly without any risk of overloading it under abnormal operating conditions. On the suction side, a special sealing system prevents any oil or oil mist of the bearing space seeping into the surrounding suction ducts thus contaminating, for example, the air of an oxygen or nitrogen plant (Figure 11D-25) The sealing air introduced in the middle of the shaft seal is discharged to atmosphere on the bearing side in order to avoid building up
Fig. 11D-13 Multiseqment journal bearing with four tilting pads. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland)
of pressure in the confined bearing space connected with the oil tank.
ROTOR AND IMPELLERS (FIGURE 11D-26)
Fig. 11D-11 Transfer of external forces. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland)
The advantages of the rotor design of Sulzer turbocompressors were described in detail on Figure 11D-7. The impeller and shaft materials undergo a number of metallurgical tests. Further tests are carried out during manufacture. The finished impellers are then balanced at low speed, subjected to an overspeed test. The assembled rotor is dynamically balanced over the whole speed range up to full speed. Impellers with a medium to high flow coefficient are of the fully welded construction with the blades shaped in three dimensions (see Figure 11D-3). Small and narrow impellers are of the combined welded/brazed construction. The sense of rotation of the rotor is clockwise, seen from the suction side.
INLET GUIDE VANES To obtain the characteristics as shown in Figure 11 D-18 with an infinite number of operating points between maximum performance and surge line, inlet guide vanes are fitted ahead of the first stage. They are actuated by a pneumatic servomotor which may be connected to an automatic pressure or flow controller (Figure 11D-27). The guide vanes are pivoting in self-lubricating bushes and are connected to an adjusting ring by a maintenance-free linkage system (Figure 11 D-28). Another link connects the ring with the pneumatic actuator. The absence of a lubricant avoids contamination of the process gas. When starting the compressor, the guide vanes are in an interlocked closed position to reduce the starting torque to a minimum.
COOLERS
Fig. 11D-12 Two-lobe journal bearing. (Source: Sulzer Turbo Ltd, ZOrich, Switzerland)
Three pairs of intercooler tube bundles are mounted in a vertical position on each side of the casing and bolted to the water box. They can expand freely. Round finned tubes ensure excellent heat transfer on the air side. The gap between tube bundle and casing at the exit of the cooler is sealed with a rubber membrane to avoid bypassing of uncooled air.
Appendix 11D: Isotherm Turbocompressors 439
Design Features (RIO)
THE ADVANTAGE OF THE RIO
RIO SERIES FOR COMPRESSING OXYGEN
compressor series lies in its advanced design and the thorough cooling of the medium by built-in intercoolers. The compression approaches the ideal of efficient isothermal compression. Result: lowest possible energy consumption and a very compact machine. There are no external coolers, no crossover piping, no expansion joints. Its simple, lightweight package requires less space, has low overall profile, is easy to erect and results in minimum installation cost.
9Compact five-stage centrifugal inline design 9Three pairs of vertical cooler bundles integrated in casing 9Nominal discharge pressure up to 20 bar 9Directly coupled booster compressor available for high discharge pressure
(see Figure on page 450)
Fig. 11D-14 Kingsbury-type axial thrust bearing with self-equalized pads with directed lubrication. (Source: SuIzer Turbo Ltd, ZOrich, Switzerland)
Fig. 11D-15 Determination of the absolute humidity x and the molecular mass Mf of the wet air. (Source: Sulzer Turbo Ltd, ZOrich, Switzerland)
440 Process Plant Machinery TABLE 11D.3
Performance Data RIK and ARI
Type designation
Compressor slze Operating condition
Indices
Nominal diameter Power input
D (cm) P (kw)
Mass flow Suction pressure Suction temperature Relative humidity of the air or gas Discharge pressure Molecular mass
rn(kg/s) Pl (bar abs) TI(K)/tl(~ ~1(%) p2(bar abs) M (kg/kmol)
The following factors and symbols are also used for the calculation: Suction volume (actual) Absolute humidity Isothermal efficiency
~/l(m3/h)
Suction branch Discharge branch Dry Wet
1 2 t f
Design Features ARI AXIAL-RADIAL CONCEPT Where the ideal concept of the centrifugal ISOTHERM compressor reaches the economical limits with regard to size, weight and specific cost, the axial compressor design offers an attractive and technically convincing solution for large volume flows. For equal aerodynamic loading (Mach number) of the machine and for the same tip diameter of the rotor, an axial stage will handle a volume flow about twice as large as the one of a
(-) ,1,,o(%)
x
wide centrifugal impeller. If, therefore, the centrifugal section of an isotherm compressor is preceded by an axial booster with a pressure ratio of about 2, the centrifugal section is correspondingly reduced in size and its optimum speed will coincide with that of the axial part. Therefore, the two sections can be combined in one machine with one single rotor running in two bearings only, a design principle well known from the industrial single-shaft gas turbine. Stage and cooler optimization for the predominant pressure ratios for air between 6 and 8 led to a compact axial-centrifugal compressor with six axial and three centrifugal stages and three pairs of intercoolers. This configuration results in an excellent overall isothermal efficiency due to the higher
Appendix 11D" Isotherm Turbocompressors 441 TABLE 11D.3 (continued) compressor
Selection and performance calculation of an ISOTHERM
Given
Mass flow(dry) Suction pressure Suction temperature Relative humidity Discharge pressure Cooling-water temperature Dry molecular mass
Example 1 Type RIK
Example 2 Type ARI
rht = Pl = T1 = ~p~ = P2 = tw = Mt =
mt = 136.48 kg/s Pl = 1 bar T~ = 308 K; t~ = 35~ (pl = 60% P2 = 7.6 bar tw = 20 ~ Mt = 28.96 kg/kmol
17.29 kg/s. 1 bar 308 K; t~ = 35~ 60% 9.8 bar 20 ~ 28.96 kg/kmol
Calculation instructions
1 Determination of the absolute humidity (x) and the wet molecular mass using diagram, 8315 wet gas constant Rf Mt
x = 0.021 Mf = 28.58 kg/kmol
x = 0.021 Mf = 28.58 kg/kmol
Rf = 290.94 J/kgK
Rf = 290.94 J/kgK
2 Calculation of the wet mass flow rhf = rhf(1 + x)
mf = 17.29.1.021 = 17.65 kg/s
rhf = 136.48.1.021 = 139.35 kg/s
volume V1 = mf. Rf. T1 3600 9 P1 105 9 (m3/h)
k/1 = 56950 m3/h
V1 = 449520 m3/h
4 Determination of the discharge temperature t2
t2 = 98 oC
t2 = 63 oC
5 Selection of the compressor frame size and power input P
RIK 56 P = 5.0 MW
ARI 90 P = 20.3 MW
3 Determination of the actual suction
Conversion factors
1000 Nm3/h (1.013 bar, 273 K, dry) = 0.3592 kg/s 1000 scfm (14.7 psia, 60 F, dry) = 0.5774 kg/s 1000 scfm (14.7 psia, 70 F, dry) = 0.5665 kg/s 1 m3/h = 0.5886 cfm/1 bar = 14.5 psi
Fig. 11D-16
Determination of the discharge temperature. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland)
442 Process Plant Machinery
Flg. 11D-17 Compressor selectlon dlagram. (Source: Sulzer Turbo Ltd, ZOrlch, Swltzerland)
Flg. 11D-18 RIK compressor wlth Inlet gulde vanes and constant speed drlver. (Source: Sulzer Turbo Ltd, ZOrich, Swltzerland)
Appendix l lD" Isotherm Turbocompressors 443
Fig. 11D-19 RIK compressor without inlet guide vanes, but running at variable speed. (Source: Sulzer Turbo Ltd, ZQrich, Switzerland)
Fig. 11D-20 ARI compressor with adjustable stator blades and constant speed driver. (Source: Sulzer Turbo Ltd, ZOrich, Switzerland)
444 Process Plant Machinery efficiency of the axial part and the high stage efficiency of the subsequent three wide impellers (see Figure 11D-2).
CASING The horizontally split casing consists of six major components (Figure 11 D-33). The cast axial inlet (1) and center part (2) cylinder are flanged to the welded centrifugal casing (3). The inlet casing alone or together with the centre part can be lifted for the inspection of their internals while leaving the centrifugal casing in its place. The cast blade carrier (7) and diffuser wall (11 ) are also bolted to the centrifugal casing. The discharge volute (4) flanged to the centrifugal casing need not be dismantled when lifting the top half of the latter for the purpose of removing the intercooler tube bundles or inspecting the impellers. For all maintenance operations the external pipe connections remain undisturbed. The bearings are easily accessible by simply lifting the bearing housing top (6) on the discharge side or the top half of inlet casing (1)and bearing housing (5)on the suction side. The oil and sealing-air connections are located in the bottom half of the respective casings, and any inspection or maintenance work leaves them unaffected. The discharge nozzle - forming part of the bottom half discharge volute ( 4 ) - is normally pointing downwards, but can also be directed horizontally. Two pendulum-type feet and two additional feet attached laterally to the centrifugal cooler casing support the machine on the foundation. The flow passages to and from the intercoolers follow the same design principle as for the RIK series.
BLADE CARRIER AND STATOR BLADES The blade carrier (7) and the short diffuser wall (11) bolted together are flanged to the cooler casing (3)and can expand freely towards the suction side (Figure 11D-33). The double-casing design with outer casing and blade carrier offers various advantages: 9Rigid casing construction; the clearances in the blade duct are not influenced directly by external pipe forces.
Fig. 11D-22 RIK skid completely shop-erected, as transported to site as a single-lift package. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland) 9Simple fitting of the blades and assembly of the casing parts; the top half of the casing can be raised without dismantling the blade-adjusting mechanism. 9Possibility of fitting different blade carriers, for adapting the blade channel and thus the compressor characteristics to greatly changed operating conditions. 9Optimal protection of the adjusting mechanism in the space between the casing and blade carrier; the space is kept under suction pressure in order to safeguard the adjusting mechanism against condensation and corrosion attack. Each of the adjustable stator blades (9)is made out of one piece with a cylindrical shaft. The latter is seated in a bearing bush in the blade carrier (Figures 11D-37 and 38). The high damping characteristics of this seating arrangement practically excludes the occurrence of dangerous vibration amplitudes associated with the stator blades.
STATOR BLADE ADJUSTING MECHANISM The adjusting mechanism is located in the annular space between casing and blade carrier. It is maintenance-free and does not require any lubrication. The adjusting mechanism is operated by means of two hydraulic servomotors (10)which are affixed laterally to the bottom half casing. One of the servomotors is equipped with a positioning transmitter and the second operates hydraulically in parallel (Figure 11 D-39).
Fig. 11D-21 Casing Internals Impeller(7), cast iron partition walls (4), vaned diffusors (5). (Source: Sulzer Turbo Ltd, Z0rich, Switzerland)
The linear movement of the servomotor piston rods is transmitted directly to the adjusting cylinder (8) by way of two ball and socket joints. The adjusting cylinder of welded design can
Appendix liD: Isotherm Turbocompressors 445
Fig. 11D-23 Section through an RIK series ISOTHERM compressor (above: vertical section; below: horizontal section) 1 -Casing; 2 - Inlet housing; 3 - Discharge volute; 4 - Partition walls; 5 - Diffusors; 6 - Shaft; 7 - Impellers; 8 - Balance piston; 9 - Shaft seals; 1 0 - Bearing housing, discharge end; 1 1 - Bearing housing, Intake end; 1 2 - Journal bearings; 13 - Inlet guide vanes; 14 - Vane adjusting mechanism; 15 - Cooler tube bundle; 16 - Water separator; 1 7 - Coupling flange. (Source: Sulzer Turbo Ltd, ZOrich, Switzerland)
move in the axial direction and is dry-seated. There is no restriction of heat expansion in any direction. U-shaped guide rings are provided on the inner side in which the adjusting levers are engaged. These rings facilitate the assembly of the adjusting cylinder with the stator blade linkages. The adjusting levers provided on the end of each stator blade shaft are connected to the guide rings of the adjusting cylinder
by means of pivoting sliders. The axial movement of the cylinder is converted into a rotating movement of the stator blades (Figure 11D-40). The self-lubricating bearing bushes of the blade shafts are seated in the radial holes of the blade carrier. O-ring packings prevent the ingress of contaminants into the stator blade seating.
446 Process Plant Machinery
Fig. 11D-24 TABLE 11D.4
Cross-section through diffuser and return channel. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland)
Materlals of Constructlon
Part
Material
DIN Standard
Comparison ASTM Standard
Casing Compressor inlet Bearing housing Partition walls
Nodular cast iron*** Nodular cast iron Nodular cast iron Cast Iron
GGG-40*** GGG-40 GGG-40 GG-20
A A A A
Diffusers: Discs Blades
Nodular cast iron Carbon steel plates
GGG-40/1693 HI/17155
A 536 A 515, Grade 55
Shaft Impellers Dummy piston Journal bearings
Low-alloy steel Forged steel Alloy steel Steel with white metal
28 NiCrMoV 8 5 * 34 CrNiMo 6 CK 15 + WM
A 470 * AlSl 4340 AlSl 1015 + WM
Inlet guide vanes
Stainless steel
X 20 Cr 13
Cooler tubes
Copper
alternatively:**
Aluminium brass Copper nickel alloy Copper nickel alloy
SF-CuF 20 CuNi 10Fe CuZn 20 A1 F34/1785 CuNi 10 F 29 CuNi 30 F 36
395*** 536 536 48, Class 30
B 111/687 B 111 C 70600 B 111 C 71500
Fins**
Copper
For all tube alternatives
Tube plates top alternatively: Tube plates bottom
Carbon steel plate Muntz metal Muntz metal
HI/17155 CuZn 38 SNAL CuZn 38 SNAL
A 515, Grade 55 B 171 C 36500 B 171 C 36500
Water separators Water boxes
Stainless steel Cast iron
X 5 CrNi 18 9 GG-20
A 167, Grade 304 A 48, Class 30
,,
*On request, depending on application. **Other alternatives, such as Duplex designs, on request adapted to prevailing cooling-water properties and air contamination. ***Frame sizes RIK 90 and above welded design (carbon steel plate).
Appendix llD: Isotherm Turbocompressors 447 SUCTION SIDE BEARING SEAL By means of double compressed-air sealing on the suction side combined with a double-walled bearing housing (5) which vents to the atmosphere, preventing any sacking in of oil mist in the event of subatmospheric pressure at the machine inlet (Figure 11D-41).
POWER OIL SUPPLY A separate high-pressure control oil unit (Figure 11 D-42) actuates the hydraulic servomotor of the adjustable axial stator blades. This control oil unit comprises an oil tank, two motordriven pumps, a changeover-type twin oil filter, two bubble accumulators, a regulating valve for constant pressure and the necessary instrumentation. All components are mounted on a bedplate and piped up accordingly. In case of failure of the control oil pumps the two accumulators will supply enough oil for a quick and safe reaction of the control elements.
ROTOR The basic design principles with their distinct advantages are shown on page 6, Figure 7. Both axial and centrifugal part are forming a single forged monobloc shaft (see Figure 11D-44). The axial blades have rhomboidal fir-tree roots and are firmly
Fig. 11D-25 Bearing housing suction side with special sealing system. (Source: Sulzer Turbo Ltd, ZOrich, Switzerland)
Fig. 11D-26
Fig. 11D-27 Inlet guide vanes with pneumatic actuator. (Source: Sulzer Turbo Ltd, ZQrich, Switzerland)
Fig. 11D-28 Inlet housing with inlet guide vane linkage system. (Source: Sulzer Turbo Ltd, ZOrich, Switzerland)
Rotor. (Source: Sulzer Turbo Ltd, ZOrich, Switzerland)
448 Process Plant Machinery braced in an exactly defined position in peripheral grooves of the shaft (Figure 11D-38). The three centrifugal impellers of a high flow coefficient and therefore inherent high efficiency (Figure 11D-2) are of the fully welded construction with three-dimensional blades (Figure 11D-3). They are balanced at low speed and subjected to an overspeed test. The assembled and bladed rotor is then dynamically balanced over the whole speed range up to full speed. The sense of rotation of the rotor is clockwise, seen from the suction side.
Flg. 11D-30 For mere Inspectlon of rotor and caslng Internals, the Intercoolers need not necessarlly be wlthdrawn. (Source: Sulzer Turbo Ltd, Z0rlch, Swltzerland)
Flg. 11D-29 For Inspectlon of the Intercoolers, each tube bundle can easlly be wlthdrawn Indlvldually. (Source: Sulzer Turbo Ltd, Z0rlch, Swltzerland)
Design Features COOLERS The three pairs of intercooler tube bundles are mounted in a vertical position on each side of the centrifugal casing and bolted to the lower water box. They can freely expand upwards. The upper water box is fixed to the upper tube plate and guided in the top water chamber cover (Figure 11D-43). Round finned tubes ensure excellent heat transfer on the air
Flg. 11D-31 Compressor startlng torque wlth closed Inlet gulde vanes (Source: Sulzer Turbo Ltd, Z0rlch, Swltzerland) side. The gap between tube bundle and casing at the exit of the cooler is sealed with a rubber membrane to avoid bypassing of uncooled air. The cooling-water connection and condensate drains are located at the bottom.
Appendix l lD: Isotherm Turbocompressors 449 TABLE 11D.5
Technical Data and Dimensions RIK 56
RIK 63
RIK 71
RIK 80
RIK 90
RIK 100
RIK 112
RIK 125
RIK 140
7230 5270 6600*
8450 5750 7300*
9300 6200 8100"
10400 6500 9000*
A B C
4600 3800 3400
4480 4000 4100
5900 4200 3700
6500 4400 4000
6700 4900 5700*
D E
1840 500
2250 560
2040 560
2140 630
3200 630
3350 630
3800 630
4500 710
5000 710
a b
800 300
900 350
1000 400
1100 450
1200 500
1400 600
1600 700
1800 800
2000 900
4800 6700 5300
5600 8200 6500
5400 7400 6300
5700 7900 6700
7200 10100 9000
7750 10850 9650
8700 12100 10800
9900 13700 12100
11000 15200 13400
H1 H2 H3 G1 G2 G3 G4 m~ Q 4 passes 2 passes F
8.2 1.1 25 36 18 325 690 3350
11.1 1.4 33 48
14 1.6 40 53
27 325 690 3700
45 430 900 4500
24 2.3 69 100
17.3 2.0 50 66 85 540 1150 5600
150 593 1186 5600
*incl. servomotor for guide vane drive Technical data, dimensions and weights
(dimensions in ram, weights G in metric tons) E = Gear centre distance (average) H1 = Crane height to lift casing over rotor H2 - Crane height to lift cooler tube bundles over casing H3 -- Crane height to lift casing top half over cooler tube bundles G1 = Casing top half G2 = Heaviest single cooler tube bundle
Main components 1 Compressor 2 Base frame with integrated lube oil system 3 Main oil pump 4 Auxiliary oil pump 5 Twin oil filter 6 Oil cooler 7 Gear 8 Motor 9 Instrument rack l0 Oil mist fan
G3 = Bare compressor Ga = Complete skid, max. mr 2 = Compressor rotor mass moment of inertia in kg m 2, referred to compressor speed, max. Q = Cooling-water rate of compressor intercoolers in m3/h F = Maximum oil filling of oil tank in liters
Pipe connections 11 Compressor suction 12 Compressor discharge 13 Cooling-water inlet 14 Cooling-water outlet 15 Compressor condensate drain
29 2.5 85 120 280 680 1360 8000
41 3.2 120 155 500 760 1520 8000
57 4.5 165 205 880 900 1800 10000
80 6.3 230 270 1570 1025 2050 10000
450
Process Plant Machinery
1 Casing 2 Inlet housing 3 Discharge volute 4 Partition walls 5 Diffusers 6 Shaft 7 Impellers
8 Balance piston 9 Seals 10 Bearing housing, discharge end 11 Beating housing, intake end 12 Journal beatings 13 Cooler tube bundle 14 Coupling flange
Appendix 11D: Isotherm Turbocompressors 451
452 Process Plant Machinery
Flve frame slzes, provldlng a complete coverage over the wlde flow range from 12 000 to 90 000 m3/h Source: Sulzer Turbo Lid, ZOrlch, Swltzerland.
Source: Sulzer Turbo Ltd, ZOrlch, Swltzerland.
Appendix l lD: Isotherm Turbocompressors 453
Fig. 11D-32 Centrlfugal casing Internals. (Source: Sulzer Turbo Ltd, ZOrlch, Swltzerland)
454 Process Plant Machinery
Fig. 11D-33 Section through an ARI series ISOTHERM compressor (above: vertical section; below; horizontal section) 1 - Inlet casing of axial section; 2 - Centre part of axial section; 3 - Casing of radial section; 4 - Discharge volutel 5 - Double-walled bearing housing, suction side; 6 - Bearing housing, discharge side; 7 - Blade carrier; 8 - Blade adjusting cylinder; 9 - Adjustable stator blades; 1 0 - Servomotor; 11 - Short-diffuser wall; 12 - Bladed diffusers; 13 - Partition walls; 14 - Cooler bundles; 15 - Water separator; 16 - Water chamber covers; 17 - Shaft; 18 - Rotor blades; 1 9 - Impellers; 2 0 - Journal bearings; 2 1 - Position of thrust bearing, if fitted; 2 2 - Balance piston; 2 3 - Shaft seal. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland)
Fig. 11D-34 Functional and compact design with easy access to all vital parts characterize the ARI series ISOTHERM compressors. (Source: Sulzer Turbo Ltd, Z(irich, Switzerland)
Appendix l lD: Isotherm Turbocompressors 455 TABLE 11D.6
Materials of Construction
Part
Material
Valid for air or similar gases DIN Standard
Comparison ASTM Standard
Axial flow casing Inlet casing with bearing housing Radial flow casing Discharge volute with bearing housing
Nodular cast iron Nodular cast iron
GGG-40/1693 GGG-40/1693
A 395 A 395
Carbon steel plate Nodular cast iron
HI/17155 GGG-40/1693
A 515, Grade 55 A 395
Blade carrier Adjusting cylinder Stator blades Discharge guide vanes
Nodular cast iron Carbon steel plate Martensitic stainless steel Martensitic stainless steel
GGG-40/1693 HI/17155 X 20 Cr 13 X 20 Cr 13
A 395 A 515, Grade 55 AISI 420 AISI 420
Shaft Impellers Rotor blades Dummy piston Labyrinth strips Journal bearings
Low-alloy steel Forged steel Martensitic stainless steel Low-alloy steel Stainless steel Steel with white metal
28 NiCrMoV 85 * X 20 Cr 13 34 CrNiMo 6 X6CrMo 17 1 CK 15 + WM
A 470 * AISI 420 AISI 4340 AISI 434 AISI 1015 + WM
Diffusers Diaphragm Partition walls
Carbon steel Cast iron Cast iron
HI/17155 GG-20/1691 GG-20/1691
A 515, Grade 55 A 48, Class 30 A 48, Class 30
Cooler tubes alternatively:** Fins**
Copper Aluminium brass Copper nickel alloy Copper nickel alloy Copper
SF-CuF20 CuZn 20 AI F34 CuNi 10 Fe F29 CuNi 30 Fe F36 For all tube alternatives
B B B B
Tube plates top alternatively: Tube plates bottom
Carbon steel plate Muntz metal Muntz metal
HI/17155 CuZn 38 SNNL CuZn 38 SNNL
A 515, Grade 55 B 171 C 36500 B 171 C 36500
Water separators Water boxes
Stainless steel Cast iron
X 5 CrNi 18 9 GG-20/1691
A 167, Grade 304 A 48,Class 30
111 111 111 111
C C C C
12200 68700 70600 71500
**Other alternatives, such as Duplex designs, on request adapted to prevailing cooling-water properties and air contamination.
Fig. 11D-35 Flow passage between axial outlet and entry into the adjacent pair of intercoolers. (Source: SuIzer Turbo Ltd, ZQrich, Switzerland)
Fig. 11D-36 Flow passage from Impeller outlet to the adjacent pair of intercoolers. (Source: SuIzer Turbo Ltd, ZQrlch, Switzerland)
456 Process Plant Machinery
Fig. 11D-37 Adjustable stator blade, rotor blade and fixed stator blade with Intermediate piece. (Source: Sulzer Turbo Ltd, ZQrich, Switzerland)
Fig. 11D-38 Fixation of adjustable stator blades. (Source: SuIzer Turbo Ltd, ZQrich, Switzerland)
Flg. 11D-39 Hydraullc servomotor. (Source: Sulzer Turbo Ltd, ZQrlch, Switzerland)
Fig. 11D-40 Stator blade adjusting mechanism. (Source: Sulzer Turbo Ltd, ZQrich, Switzerland)
Fig. 11D-41 Double compressed-air sealing system on the suction side bearing housing. (Source: SuIzer Turbo Ltd, Z0rich, Switzerland)
Appendix l lD: Isotherm Turbocompressors 457
Fig. 11D-42
Power oil unit. (Source: SuIzer Turbo Ltd, ZQrich, Switzerland)
Fig. 11D-43 Two-pass version of the intercooler. Finned tubes and highly effective water separators contribute to the excellent efficiency of the compressor. (Source: Sulzer Turbo Ltd, ZOrlch, Switzerland)
458 Process Plant Machinery
Fig. 11D-44 The single monobloc rotor running in only two journal bearings ensures high rotor stability and low vibration level. (Source: Sulzer Turbo Ltd, ZQrich, Switzerland)
Fig. 11 D-45 Inspection of the Internal parts Is made easy by simply lifting the centrifugal casing top half without disturbing coolers and adjacent casing parts. (Source: Sulzer Turbo Ltd, ZOrlch, Switzerland)
Appendix 11D: Isotherm Turbocompressors 459
Technical Data and Dimensions (ARI) TABLE 11 D.7 ARI 71
AR180
AR190
A B C
9300 4400 3500
10300 4800 3500
11300 5200 3700
12800 5500 3700
D E
1150 800
1300 900
1400 860
1400 860
a b
1900 800
2100 900
2400 1000
2700 1100
H1 H2 H3
5900 9300 7600
6500 10200 8400
6900 10800 8900
6900 10800 8900
G1 G2 G3
40 4.3 150
m~ Q 4 passes 2 passes
55 5.8 210
685 900 1800
1210 1025 2050
ARI 100
70 7.8 290 2200 1300 2600
Technical data, dimensions and weights
(dimensions in mm, weights G in metric tons)
E = Gear centre distance (average) H1 = Crane height to lift casing over rotor H2 = Crane height to lift cooler tube bundles over casing
Fig. 11D-46
Fig. 11D-47
H3 = Crane height to lift casing top half over cooler tube bundles G1 = Casing top half G2 = Heaviest single cooler tube bundle
G3 -- Bare compressor mr 2 = Compressor rotor mass moment of inertia in kg m 2, referred to compressor speed Q = Cooling-water rate of compressor intercoolers in m3/h
Dimension drawing. (Source: Sulzer Turbo Lid, ZOrlch, Switzerland)
Compressor starting torque with closed stator blades (Source: Sulzer Turbo Ltd, ZOrlch, Switzerland)
90 10 390 3750 1600 3200
460 Process Plant Machinery
Typical Plant Layout
Fig. 11D-48 Typical plant layout. Main components 1 - Compressor; 2 - Gear; 3 - S y n c h r o n o u s motor; 4 - E x c i t e r ; 5 - Weather hood; 6 - Roll-o-matic filter (1st stage) supply unit; 7 - Bag filter (2nd stage); 8 - Bypass doors; 9 - Suction silencer; 10 - Rubber expansion joint; 11 - Mechanically driven main oil pump; 12 - Motor rotor withdrawal; 13 - Motor cooler; 14 - Discharge silencer; 15 - Nonreturn valve; 16 - Discharge Isolating valve; 17 - Discharge air header; 18 - Blow-off valve; 19 - Blow-off silencer; 20 - Lube oil; 21 - Power oil supply unit. (Source: Sulzer Turbo Ltd, ZOrich, Switzerland)
Appendix l lD: Isotherm Turbocompressors 461
Control system and oil supply
shutdown system must protect the compressor train from possible damage under such conditions.
SAFETY SYSTEMS AND PROCESS CONTROL Two main requirements are to be met by the control system of turbo-compressors:
Auxiliary component control assures a safe supply of lube and control oil.
9Compressor safety - to prevent the compressor from operating in an unstable range or at other hazardous conditions 9Process requirements - t o adjust the compressor to the demands of the process The engineering and supply of complete compressor control and safety systems ensures the optimum protection of the compressor and the plant. Thanks to the strict use of standard signals, the control system can be integrated into other systems without difficulty. It allows remote control, automation of starting and stopping and can be linked with distributed control systems (process computers). SAFETY SYSTEMS Antisurge control The stable operating range is defined by the characteristic curves C and limited by the surge line S (Figure 11D-49). Operation under surge conditions, occuring on the left side of this surge line, is avoided by an antisurge control system. It measures flow and pressure and is designed to closely follow the actually measured surge line with a predetermined safety margin. As soon as the operating point approaches the response line (L), the controller progressively opens the antisurge valve as a function of the difference between the minimum stable flow of the compressor at a given pressure and the flow required by the process. This valve is either a blow-off valve (air) or a bypass valve (nitrogen, oxygen) followed by a bypass cooler. As a protecting device the system has to act independently of any other control system and must not be used for pressure or flow control. Limitation of temperature, pressure, axial displacement Under certain circumstances, external influences or other irregularities may lead to undesired changes of the normal level of gas and bearing temperatures, pressures, axial displacement of the rotor, etc. A reliable interlock, alarm and
Fig. 11D-49 Characteristics of a turbo-compressor p Discharge pressure; V - Flow rate; C - Compressor characteristic curves; S - Surge line; L - Limit flow. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland)
Auxiliary component control
PROCESS CONTROL Suction pressure Constant suction pressure to adapt the compressor flow to an upstream production unit. Discharge pressure Constant discharge pressure in cases where chemical reactions or physical processes have to take place at a clearly defined pressure, or where the compressor flow has to be adapted to a fluctuating down-stream demand. Flow Constant mass flow control corresponding to a constant plant output. The above-mentioned process control systems can also be combined via selecting relays, e.g. flow control with discharge pressure limitation. EXAMPLE The typical P & I diagram shows all the necessary control and monitoring equipment for the safe operation of an isotherm compressor. Additional control and monitoring equipment can be added on request. OIL SUPPLY For the RIK series the lube oil system is normally integrated in the compressor skid and comprises the same components as the Sulzer standard oil supply unit shown in Figure 11D-50. The separate standardized lube oil units for the RIO and ARI series comprise the oil tank, auxiliary oil pump, twin oil filter, selectively single or twin oil cooler, oil mist fan, oil filling sieve, electric heater, automatic start-up testing device, oil pressure control valve, and all the valves, apparatus and instruments necessary for the safe operation of the compressor set (Figure 11D-50). The main oil pump is normally mechanically driven by and attached to the speed increasing gear or the steam turbine. The lube oil system is designed to supply the lube oil for compressor gear and driver.
Fig. 11D-50 Standard lube oil supply Sulzer Turbo Ltd, Z0rich, Switzerland)
unit. (Source:
462 Process Plant Machinery
Typical Compressor Plant P &l Diagram
Valve m~ a4rpmmt~ ~
Instrument symbols
-~
Me~~uC. "~iable~s ([~st lettex): B Ant,.~uzge F Flmv
G:obe[;:)all/disc valve Valve vn:h ccn:rol Irma 'I ~ee-way valve
H L P P~
Hand
bevel P:ummzu -I,~-r Noruret~n valve wi:h born I Su~[iOll 8LlaineI P,"(~.~"~Ir(~ d:fference 9l Am~I~bc dra~; trap T 'lemperature X Vih~'~,non ~ ResL:iction cr~'toe Z i~:tton ,.b D:amp'ug Output ;~.cLiun (fc[lowm,;j Leith): Double acting p~ton actmtor A .~larm DiaphraGm actuator C Contrcl ,~lennid G Sigl'.t-g:~.~ ,,~,thout meammement I [r,dica'ion S Svatch flmction ( ~ Electric r~tm,~r uh Bellows compensat=,r T 'l~'~r:uttmg Water sepaxatoz Y Rely. c~mpu:e t~l Electnopne'.Lm,at.c conve~te~ Ex'rem.=.vah~e~ Oaat le.m'): ~] Elect~c-hydzaukc convert=,r G Closed 0 O~.en I.,ow-s~g~m] seectmg relay H Eigk [~ T~a~fez u~,kcl: L Low Fu~qion generato: Rise time delay relay
~-
Non,etch ;~I;~
Signal procmmtng Compr~r
journal boating ~ : Thrust bo~inq ~ _ 1 Gea~ with thn:st collar
d.nvor start :~onu'~moion
<~> Compres~r driver '&.utdc,,m ~> Cornpres,.-~r dr:vez start-up interlock b'tart of auxihary oil pump
Line symbols
---- bube os letup:, --C o o l . [ : ~ w~te: .... C~ndc[~tc Ir._~runlcn: a~r Elc~tric/,~lcc~rc,mcaiqnal Mochan ic~,l co:,,1~:,-.r Pav,~r c,i] supply Power eli ren:rn
--
_I ...
"v~nt:ine
Seahngair C,.I mist
Abbreviations D V
Drain ','bnI
Appendix l lD: Isotherm Turbocompressors 463
Applications
Flg. 11D-51 Alr compressor, type RIK 56, at the East Drlefonteln mlne of the Gold Flelds of South Afrlca Ltd., Johannesburg, compressing 51 000 m3/h from 0.825 to 9.5 bar. Power Input 4000 kW. (Source: Sulzer Turbo Ltd, Z0rlch, Swltzerland)
Flg. 11D-52 Alr compressor, type RIK 71, Installed In one of AGA's alr separatlon plants In the USA. 65 000 Nm3/h are compressed from 0.97 to 6.76 bar; power Input 5535 kW. (Source: Sulzer Turbo Lid, Z0rlch, Swltzerland)
464 Process Plant Machinery
Fig. 11D-53 Nitrogen compressor train consisting of an Isotherm compressor, type RIK 56, with RZ 3 5 - 6 booster, Installed at Union Carbide's air separation plant Prentiss, Canada. The 7900-kW electric motor driven set compresses 37 420 Nm2/h of nitrogen from 0.91 to 47.1 bar. (Source: Sulzer Turbo Ltd, ZQrich, Switzerland)
Fig. 11D-54 Compact Isotherm process air compressor, type RIK 63, driven by a tall gas expander E 40 and a steam turbine with a combined power output of 8200 kW. The set compresses 50 000 Nm3/h of air to 10 bar in a South Korean nitric acid plant. (Source: Sulzer Turbo Ltd, Z(irich, Switzerland)
Appendix l lD: Isotherm Turbocompressors 465
Flg. 11 D-55 Oxygen plant Burns Harbor, USA, of Union Carblde's Llnde Dlvlslon, each alr separatlon llne supplylng 2000 st/d of oxygen to nelghbourlng steel works. The two motor-drlven alr compressors, type ARI 80, dellver 300000 Nma/h each at 7.2 bar, absorbing 25.2 MW. The two oxygen compressors supply 55 000 Nm3/h each at 34.2 bar; power Input 9050 kW. The maln Items of thls compressor plant are: 1 - Oxygen compressor I; 2 - A l r compressor I; 3 - A l r filter house I; 4 - Alr aftercooler for alr compressor I; 5 - Oxygen compressor II; 6 - Control room II; 7 - Alr aftercooler for alr compressor II; 8 - Electrlc motor II; 9 - Nolse attenuatlng enclosure for gearbox; 10 - A l r compressor II; 11 - O11 reservolr II; 12 - Alr filter house II. (Source: Sulzer Turbo Ltd, Z0rlch, Swltzerland)
Fig. 11D-56 These two air compressors, type ARI 56, Installed ad Anglo American's Vaal Reefs mine, South Africa, are the biggest mining compressors of the world. Each machine supplies 170 000 m3/h at 9 bar and is driven by a 15-MW synchronous motor. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland)
466 Process Plant Machinery
Fig. 11D-57 Skid-mounted air compressor unit, type RIK 56, with R 35-5 and steam turbine arriving at an ammonia plant in Cartagena (Spain). Compressor capacity 34 900 Nm3/h, discharge pressure 34 bar. (Source: Sulzer Turbo Ltd, Z(lrich, Switzerland)
Fig. 11D-58 One of two Isotherm compressors, type ARI 63, Installed In the steelworks of Diefflen/Dillingen, FRG, shown during erection on site. The compact units with Integrated coolers allow their Installation also at locations where space is very limited, still ensuring easy access to all vital parts for maintenance. Each compressor delivers 180 000 Nm3/h of atmospheric air at 7.1 bar. Power Input 15.4 MW. (Source: Sulzer Turbo Ltd, ZQrich, Switzerland)
Fig. 11D-59 This motor-driven. 12.6-MW compressor, type ARI 63, delivers 139420 Nm3/h of air at 7.17bar to Airco-Cryoplants' air separation installation supplying oxygen for the Coolwater coal gasification plant at Daggett, Cal., USA. The eyngaa produced serves as fuel for a 100-MW cogeneration power plant. (Source: Sulzer Turbo Ltd, Z0rich, Switzerland)
APPEN DIX 11 E
Gas Seal Design
Gas seal designs are a critical feature in compressor design, operation and maintainability. Manufacturers constantly seek improvements in this area as is indicated by the following extracts from a technical paper. The purpose of this appendix is to emphasize two of the main technical features incorporated in the design of the D-R Gas Seal: 9 The groove pattern (generates high film stiffness and optimum pressure distribution) 9 The "L" Sleeve design (reduces the risk of "hang up" and therefore cuts operational costs).
Gas Seal Principle Figure 11E-1 shows a simplified cross-section drawing of a D-R gas seal where the main active parts are: 9 The rotating seat, also called tungsten carbide ring 9 The stationary seat, also called the carbon ring 9 The pusher sleeve, also called "L" sleeve Figure 1 I E-2 shows, in parallel, the essential parts of a D-R gas seal, and the detailed representation of the closing and opening forces which determine the equilibrium of the seal faces. When calculating the balance of forces on the carbon face for a given gap, we take into account: 9 the opening forces (OF): sum of (pressures inside the gap x surface) 9 and compare that to the closing forces (CF): sum of (pressures at the back of
the carbon face x surface on which they apply + spring force (Fs) -friction force (FI") If the gap increases, then the opening forces will decrease; if the gap decreases, then the opening forces will increase. The film of gas acts just like compression springs: If the speed increases, then the opening forces will increase as the grooves will generate more lift. Conversely, we can also adjust the closing forces. If we increase the hydraulic diameter, then the closing force will diminish; and if we lower the hydraulic diameter, the closing forces will increase. This is called 'playing with the balance ratio'. * Source: Dresser-Rand, Olean, NY. 467
468
ProcessPlant Machinery
FIGURE 11 E-1
(Source:Dresser-Rand, Olean, NY)
FIGURE 11E-2 (Source: Dresser-Rand, Olean, NY) This can be put in simplified equations as follows CF = Pin • S l + Pout X 52 "+" Fs OF = k (rpm, press, groove geometry, temperature, gap, gas characteristics) In the steady-state situation O F - - C F . Thus the solution of these equations is to calculate the gap (operating gap) which exactly balances the opening forces and the closing forces
Groove Pattern and Pressure Profile Considering the above equations, while it is simple to calculate the closing forces, the calculation of the opening forces is a complex function of several parameters, which are not independent. Therefore Dresser-Rand have developed a computer code that will calculate pressure, speeds and temperature distributions at every point within the interface. This program iterates until the operating parameters are established and stable; in fact it finds the gap at which the opening forces exactly balance the closing forces. Of course various groove geometries can be analyzed and then optimized in order to provide the highest reliability and performance to the gas seal.
Appendix 11E: Gas Seal Design Ideally, a gas seal would like to have a minimum gap in order to minimize the leakage rate (the gas seal leakage is approx, a function of the gap raised to the power 3). However, a smaller gap gives a higher risk of accidental contact between the two faces of the seal. A small gap as well as the highest possible gas film stiffness will be the optimum. A uniform pressure distribution between the faces is important, since it reduces the local deformation of the parts. Therefore, let's discuss pressure profiles, balance of forces, and gas film stiffness. It is clear that the layer of gas trapped between the rotating seat and the carbon face of the gas seal changes pressure as it goes from the OD (at seal supply pressure to the ID (at flare pressure or atmospheric pressure). The grooves in the rotating seat alters the normal pressure decay and generates zones of overpressure. Using the pressure at each point of the interface, calculated with the above mentioned calculation code, it is possible to generate a 3D representation of the pressure profile. One example is shown in Figure 11E-3. From the same computation output it is also possible to show the isobar curves (Figure 11E-4) which highlight the overpressurized areas responsible for the separation of the seal faces. Comparing the various pressure profiles and isobar curves obtained with different groove patterns, it is possible to select a groove shape which provides high pressure rise and as uniform as possible pressure distribution. The patented groove pattern used by Dresser-Rand satisfies these two important criteria. As an example, Figures 11E-5 and 6 show isobar curves at different operating conditions (high speed/low speed and high pressure/low pressure)
Gas Film Stiffness Gas film stiffness is a major parameter for gas seals as it can be used to evaluate the ability of a gas seal to resist sudden positional changes (surge for instance), or also to compare two seals with different groove geometry Let's define what the gas film stiffness is: At a given operating gap an opening force (OF1) is generated within the interface, if the gap is forced to close (or open) by say 1/100 of its value (gap/100), a new opening force will develop (OF2).
FIGURE 11E-3 (Source: Dresser-Rand, Olean, NY)
469
470
ProcessPlant Machinery !:'::.':i i
52.3 50.5 48.8
47 45.2
-0.05
::i. i).. i:':i
41.7 39.9 38.2 36.4
34.6 i:'i :::i~ :i.
0.05
31.1 29.3 27.5 25.8 24 . . . . . . .
0.05
0
-0.05
9. .
20.5
18.7
16.9
15.2 13.4
FIGURE 11E-4
(Source: Dresser-Rand, Olean, NY)
In a similar way, springs are calculated with the formula: dF = S x dX, hence S = dF/dX in this case S = stiffness; dF = ( O F 2 - OFI); and dX = gap/100. The gas film stiffness is ( O F 2 - OF1) x 100/gap. As an order of magnitude, stiffness of more than 3 kN/micron for a medium size of seal (4.875 inch diam) at 50 Bar and 11500 rpm are typical. See Figure 11E-7. In imperial units, this would be about 175 x 106 lbs/in (at 725 psi). This is a very high value, but necessary to avoid any contact between the two faces which at the same time are separated by about 4 microns (0.16 thousandth of an inch). At higher pressure the stiffness values are even higher, which is necessary as the gap will also be smaller. In fact, the stiffness value gives an indication of how the seal can withstand axial forces (especially abnormal forces due to vibrations or upset conditions.) Incidentally, it is possible to use the same program to calculate the behavior of seals having a symmetrical groove pattern (bi-directional). The results demonstrate a significant decrease of gas film stiffness together with a smaller gap. This is why bi-directional seals are not as forgiving as uni-directional seals. The Hang-up Syndrome
Considering again the seal equilibrium, let us analyze now the event of a very low sealing pressure (for instance start up conditions before pressurizing the compressor). See Figure llE-8. When the seal is depressurized, only the spring force can close the seal gap. In this situation, if the shaft of the compressor has a small axial displacement (for
Appendix 11E: Gas Seal Design
0
'PRESdr.RT'I" 54.1 52.3 50.5 48.8 0.04 47 45.2 43.5 0.05 41.7 39.9 0.06 38.2 38.4 34.6 0.07 32.9 31.1 29.3 0.08 27.5 25.8 0.09 24 22.2 20.5 18.7
-0.05
16.9
15.2 13.4
11500 rpm, 50 Bar inlet, ISOBAR 'PRESBV.R~ 52.3 50.5 48.7 48.8 0.04 45 43.2 41.3 0.05 ~.5 37.7 0.06 ~.8 34 0.07 0.08 0.09 1
0
I
-0.05
~.2 ~.3 29.5 ~.7 ~.8
21.2 19.3 17.5 15.7 13.8 12 10.2
500 rpm, 50 Bar inlet, ISOBAR FIGURE 11E-5 (Source: Dresser-Rand, Olean, NY)
instance differential thermal expansion between shaft and compressor casing) the friction between the O-ring and the seal housing or the deformation of the O-ring itself, may prevent the spring closing the seal. A much larger gap (50 to 100 times the normal gap) may then appear between the two faces of the seal. The leakage through this interface could then be so high that pressurizing the unit becomes impossible.
471
472
ProcessPlant Machinery
'PRES150.RT'r' 157 151 146 141 0.04 136 131 125 0.05 120 115 0.06 110 105 99.5 0.07 94.3 0.08 0.09 0
-0.05
89.1 84 78.8 73.6 68.4 63.2 58 52.9 47.7 42.5 37.3
11500 rpm, 150 Bar inlet, ISOBAR 'PRESOO2.RTT' 3.47 3.06 2.65 2.24 0.04 1.63 1.42 0.05 0.06 0.07
0.08 0.09 0
-0.05
11500 rpm, 2 Bar inlet, ISOBAR FIGURE 11E-6 (Source: Dresser-Rand, Olean, NY)
This is known as the 'hang-up syndrome'. The only remedy is to disassemble the compressor end, remove the seal cartridge, fix the gas seal and reinstall it. This, of course involves unnecessary down time and high maintenance costs, usually at a time when there is an urgent need to have the compressor up and running. Dresser-Rand has developed and incorporated in the D-R Gas Seal, an "L" sleeve design (patent pending) which drastically reduces the risk of "hang up".
Appendix 11E: Gas Seal Design
FIGURE 11E-7 (Source: Dresser-Rand, Olean, NY)
FIGURE 11E-8 (Source: Dresser-Rand, Olean, NY)
The 'L' Sleeve Design As explained above, the hang-up situation is caused both, by the tendency of the O-ring to stick or to extrude, and by the simultaneous lack of gas pressure, pressure which in normal operating conditions, is the major contribution to the closing force. The Dresser-Rand "L" sleeve design addresses these two causes:
(a) Reduced risk for extrusion The balance diameter O-ring is located remote from the hot area of the seal, thus it is subject to a lower temperature. The clearance may be better adjusted (smaller)
473
474
ProcessPlant Machinery
since the housing and the sleeve are of the same material (same thermal expansion coefficient and therefore less risk or tendency for extrusion).
(b) Increased closing force due to the gas pressure (small but existing) The installation of the O-ring in the sleeve (as opposed to its installation in the housing) tends to decrease the balance diameter, thus slightly improves the closing forces due to the gas pressure (see Figure 11E-9).
FIGURE 11E-9 (Source: Dresser-Rand, Olean, NY)
Chapter 12 Axial Flow Compressors* As stated in the introduction to the preceding chapter, dynamic compressors are machines in which air or gas is compressed by the mechanical action of rotating components imparting velocity and pressure to the air or process gas. In an axial compressor, as the name implies, flow is in the axial direction, i.e., parallel to the axis of rotation. Axial compressors are basically high-flow, low-pressure machines, in contrast to the lower flow, high-pressure centrifugal compressors. Figure 12-1 shows the performance characteristics of a centrifugal and an axial compressor at constant speed for the same operating conditions. From this figure, a direct comparison of the characteristics is easy. The "turndown" capability of the centrifugal is much larger than that of the fixed geometry axial. The range of operation is greatly increased through the use of variable geometry.
FIELD OF APPLICATION Axial flow compressors have found wide use in refineries, petrochemical plants, and steel mills. Particularly in refineries, applications formerly handled by centrifugal units are now handled by axial flow compressors. This is due to several trends: first, plant sizes are growing dramatically, which brings the air requirements up into a desirable range for axial compressors; second, due to rising energy costs, there exists an increasing trend toward higher efficiencies; and third, technological improvements have made axial compressors more reliable than ever before. Axial compressors are generally more efficient than centrifugal compressors in the common flow range, depending on conditions. An axial compressor will also generally be smaller than a centrifugal compressor designed for the same flow rate. Although the axial flow compressor requires more stages due to the lower pressure rise per stage, the diametral size is much greater in a centrifugal compressor in order to pass the required air flow. The axial compressor must operate at significantly higher speeds for the same condition and is usually more costly than a comparable centrifugal compressor. In applications where speed is not a major consideration, an efficiency and size versus cost evaluation must be made.
Petroleum Refineries These are probably among the largest current users of axial flow compressors for providing the air for catalytic cracking. This service requires 50,000 cfm to 300,000 cfm at discharge pressures from 25 to 50 pounds per square inch gauge (psig). Figure 12-2 depicts a typical installation. * Source: Dresser-Rand Company, Phillipsburg, N.J. 475
476
Process Plant Machinery
FIGURE 12-1 Comparison of axial centrifugal characteristic curves. (Source: Dresser-Rand Co., Phillipsburg, N.J.)
FIGURE 12-2 The first power recovery train in fluid catalytic cracking service in a refinery and the equipment train. The equipment train comprises tandem motor-axial compressor-steam turbine-hot-gas expander units. The compressor is rated at 127,900 inlet cfm with a discharge pressure of 34 psig. The train power rating is approximately 15,000 HP. (Source: Dresser-Rand Co., Phillipsburg, N.J.)
Axial Flow Compressors
Butadiene Plants Air capacities from 60,000 to 150,000 inlet cfm and discharge pressures from 20 to 30 psig with atmospheric intake enable the axial compressors to perform well in this application.
Nitric Acid Plants In plants with capacities in excess of 500 tons per day, axial-centrifugal compressor combinations are frequently used to handle flows from 20,000 to greater than 80,000 cfm (approximately 100 cfm per ton-day) and discharge pressures from 110 to 130 psig for the combined compressor string. Power recovery expanders utilizing process "tail gas" usually drive the compressor string.
Air Separation Plants Axial compressors are used almost exclusively for higher flow air services. Services range to 100,000 cfm and discharge pressure up to 100 psig.
Blast Furnaces Axial compressors are replacing many of the older, less efficient centrifugal blowers. The compressors provide air at discharge pressures from 30 to 90 psig and flows from 125,000 to 350,000 cfm. In addition to the common processes described above, axial compressors are also often used for wind tunnel service, waste treatment facilities, specialized testing facilities, and are being developed for co-generation combustion service.
BASIC AXIAL COMPRESSOR PERFORMANCE CAPABILITIES As described earlier, the axial flow compressor is a machine with wide ranges of capacity, i.e., flow, pressure, and horsepower required. Axial compressor flow capabilities on the low end of the range, say 20,000 to 75,000 cfm, obviously overlap the higher range of centrifugal compressor coverage. It is within this low flow region where cost and size evaluation, along with driver considerations, must control the selection. Above this range, however, axial compressors are often the obvious choice. The physical size of the axial compressor is far smaller than the comparable centrifugal machine that would be required. In many high-flow situations, the axial is a better match for the drivers that will probably be selected. To increase the pressure capability of the axial flow compressors, multiple casing designs have also been developed. These are known as biaxials and triaxials. As their names imply, these are two- or three-body axial compressor trains capable of pressure ratios up to approximately 12 to 1. The machines were developed for use in nitrogen injection services. Horsepower requirements for axial flow compressors range from 3000 HP to 65,000 HP for single casing units. Horsepower inputs vary with the flow and pressure requirements of the service. A simple formula for the approximate power
477
478
Process Plant Machinery requirement of an axial flow compressor would be: Horsepower = where W W R T PR
= = -=
W W x R x T x (PR ~
- 1)
182.5
Wet weight flow, lb/min Gas constant, f t - l b f / l b m - ~ Inlet temperature, ~ Pressure ratio
FUNDAMENTALS OF AXIAL COMPRESSOR DESIGN A multistage axial flow compressor has two or more rows of rotating components operating in series on a single rotor in a single casing. The casing includes the stationary vanes (the stators) for directing the air or gas to each succeeding row of rotating vanes. These stationary vanes, or stators, can be fixed- or variable-angle, or a combination of both. A typical axial flow compressor cross section is shown in Figure 12-3; Figure 12-4 shows an axial compressor with the top half removed. The major components and their nomenclature are depicted in Figure 12-3 for reference use throughout this chapter. There are two basic types of blading that are employed in an axial flow compressor; these are obviously rotating and stationary. A brief overview of these parts is presented next.
Stationary Blades Inlet Guide Vanes T h e first r o w o f s t a t i o n a r y blades is u n i q u e . T h e s e blades are r e f e r r e d to as i n l e t g u i d e vanes. T h e s e vanes are d e s i g n e d to p r o v i d e p r e r o t a t i o n to the air o r gas stream p r i o r to e n t r y i n t o the r o t o r blades. B l a d e p r o f i l e s have a i r f o i l - s h a p e d cross sections.
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ING
Typical axial flow compressor cross section. (Source: Dresser-Rand Co., Phillipsburg, N.J.)
Axial Flow Compressors
479
FIGURE 12-4 Axial compressor, type AV 100-16, during erection Two identical steam-turbine-driven machines are supplying air to the blast furnace of a British steel works. Suction volume 560 000 Nm3/h discharge pressure 6.2 bar power input 52 000 kW each. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland.) Stator Vanes
The majority of the stationary blades within the compressor are simply called stators. There exist two types of stator vanes, variable and fixed. Variable Stator Vanes. Variable stator vanes fit through the stator casing or a blade carrier of some kind (depending on the manufacturer's design) and are attached to a drive mechanism that moves the vanes with respect to the air flow. A more detailed description of the actuating system is provided below. The inner end of each vane can be shrouded to improve the stress condition and to reduce the interstage losses through sealing strips mounted in the inner shroud. The actuation system used to move the variable stator section is usually a combination of linkages designed to move the vanes simultaneously. One type of linkage system is shown in Figure 12-5. Each variable stator vane is connected to a driving ring by a small link. These rings, one for each stage of blades, are individually connected to a main driving shaft so that the stages move simultaneously. The drive shaft is connected to a hydraulic (or pneumatic) power piston, which, through push-pull effect, opens and closes the stator vane. Different designs are chosen for this system by the various manufacturers. Desirable features would include the following:
1. Solid or "tight" linkage systems to prevent slow or inefficient actuation of the vane position.
480
Process Plant Machinery
FIGURE 12-5 A typical variable stator vane actuation system and linkage arrangement. (Source: Dresser-Rand Co., Phillipsburg, N.J.) 2. A minimum number of "joints" in the system that can wear with time and become loose or seize due to the presence of dirt. 3. Dual power cylinders (one on each side of the unit) to provide even movement of all vanes. Fixed Stator Assembly. The fixed stators are typically welded assemblies comprising the vanes and inner and outer shrouds. These assemblies are fitted into machined grooves in the stator casing. The fixed stator assembly is also fitted with sealing strips for leakage reduction.
Rotating Blades The rotating blades within the axial compressor are appropriately called rotor blades. These are National Aeronautics and Space Administration (NASA)-developed tapered and contoured airfoil sections. The rotor blades have an attachment on one end to allow for assembly within the rotor. A simplified partial section of an axial flow compressor flow path is shown schematically in Figure 12-6. The basic components would typically include the following: 9 An inlet duct to collect and accelerate the gas toward the inlet guide vanes with minimum pressure losses. 9 A row of inlet guide vanes to impart prewhirl to the gas stream in the direction of rotation for smooth entry to the rotor blades and for the control of the inlet relative Mach number. 9 A multiplicity of stages, each consisting of a row of rotor blades and a row of stator vanes of airfoil shape, to increase the static and/or total pressure of the flow.
Axial Flow Compressors
BASIC FLOW PATH
INLET INLET GUIDEVANE
STAGE ~I ~ ROTOR BLAD
--
N
STAGE ~1 STATOR BLADE
STAGE'~n ROTORBLADE~
N
STAGE ~n STATORBLADE
~ EXIT GUIDEVANE I TO DISCHARGE DIFFUSER ROTATION
=--
FIGURF 12-6 Schematic presentation of an axial flow compressor flow path. (Source: Dresser-Rand Co., Phillipsburg, N.J.) The total energy transfer to the gas stream is accomplished by the rotor blades. The hub stagger (the angle between the blade chord and the axis of rotation) is fixed, thereby fixing the amount of work done by each stage and consequently fixing the number of stages necessary to achieve the required discharge pressure. The standard frame design is adjusted to meet the required air flow by varying the rotor and stator blade heights and the unit operating speed. 9 A row of exit guide vanes, oriented to remove the whirl component from the flow leaving the last stage stator vanes, and to begin deceleration of the flow. 9 A discharge diffuser to further decelerate the flow and to convert the residual velocity energy into static pressure rise.
Natural Frequencies and Resulting Stresses Due to the inner and outer shrouding of the fixed stators and the internal shrouding and casing support of the variable stators, vibration occurring at component natural frequencies and the resulting stresses in these components are not of major concern. If a specific manufacturer does not use shrouds on both inner and outer surfaces,
481
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Process Plant Machinery
it is important to review the frequency analysis of the stationary vanes in the same manner as discussed below for the rotor blades. Since the rotor blades are mounted in a cantilever beam arrangement, i.e., one end unsupported, natural frequency excitation and the resulting stresses must be thoroughly analyzed by the designer. Clear definition of the natural frequencies of the blading, possible sources of excitation within the unit, and the resulting stress levels should be provided by the manufacturer. This should be presented in the form of Campbell and Goodman Diagrams for the compressor blading. (Refer to Chapter 5 for a discussion of these diagrams.) Evidence of the accuracy of the information, in the form of test data or successful long-term operating experience, should be made available by the manufacturer and should be reviewed by the purchaser. Similar scrutiny is appropriate for thrust beating designs. OPERATIONAL LIMITATIONS The successful prediction of the performance of an axial compressor requires knowledge of the flow behavior likely to be encountered in the machine. In certain areas of the operating characteristic, accurate prediction of the performance is not possible. Violent overall instabilities identified as stall, surge, and choke complicate these predictions.
Stall Stall is a commonly used term with regard to axial flow turbomachinery. It is too often incorrectly used as a cause of problems, mainly due to a misunderstanding of the true flow behavior during a stalled condition. Stall occurs when the flow separates from the surface of the blade. An aerodynamic disturbance is formed downstream of the point of separation. Stall is a generalized term. It is often used with additional descriptors to explain the flow condition that causes the separation. For instance, the separation can occur on either side of the airfoil. High positive incidence stall, caused when the angle between the flow and the inlet to the rotor blade is too large, causes separation of the airflow from the suction, or convex, side of the airfoil. High negative incidence stall causes separation of the airflow from the pressure, or concave, side of the airfoil. Conditions of rotating stall can be established when a group of blades becomes stalled. This is a phenomenon of stall cells being created within a stage. As the first blade stalls, it causes a disturbance to the airflow of the adjacent blade, eventually causing it to become stalled. As each subsequent blade becomes stalled, the last blade in the patch of stalled blades begins to recover. Thus, the effect is a rotating patch of stalled blades. Whether a few blades experience stall or a rotating stall cell is formed, the mechanical damage possibly caused by this unstable aerodynamic condition can be significant. It is difficult to determine the actual loads induced on the blades during such an event. Laboratory testing has shown that loading levels can reach ten times normal levels. Surge As with any dynamic compressor, surge occurs when the slope of the pressure ratio versus capacity curve becomes zero. It is associated with the complete breakdown
Axial Flow Compressors
of flow through the machine, and it takes place when several adjacent stages are subjected to high positive incidence stall. At any given speed, as the inlet flow is reduced, a point of maximum discharge pressure is reached. As flow is further reduced, the pressure developed by the compressor tends to be lower than the pressure in the discharge line and a complete flow reversal of an oscillatory nature results. The reversal of flow tends to lower the pressure in the discharge line and normal compression resumes. If no change to either the system back pressure or the operation of the compressor occurs, the entire cycle is repeated. This cycling action is an unstable condition varying in intensity from an audible rattle to violent shock, depending on the energy level of the machine. Intense surges are capable of causing serious damage to the compressor blading and seals. The uncertainties surrounding this oscillating flow are cause for concern.
Surge Control It is standard procedure at process plants to install reliable antisurge control equipment in the compressor piping to prevent operation in the surge region. A typical surge control system should incorporate or encompass recycle loops, i.e., valved bypass piping to provide sufficient flow through the compressor to keep it away from surge. Experience points to the following requirements: 9 The system is to be electronic rather than pneumatic for the fastest response time, and the surge valve must be interlocked with the trip circuit such that it immediately opens on a train trip. 9 The control system logic requires flow, pressure, and stator vane position input. 9 The surge valve positioner-operator system must be capable of driving the valve fully closed to fully open in one second and fully open to fully closed in ten seconds. 9 The surge valve should open on the loss of any input signal or operator medium. 9 The surge valve should be sized to pass full flow at any point along the surge line with the valve at 60% open and with full consideration given to the downstream system pressure drop. The point on the surge line requiring the largest valve and discharge system is normally at maximum speed with stators full open, but a point at lower flow with a lower discharge pressure may in some instances dictate size. 9 A check valve should be installed close to the compressor discharge just downstream of the surge valve connection in the discharge line. In addition to the antisurge control system discussed above, a thermocouple detection system can also be employed to determine the presence of the surge recycling effect. When an axial flow compressor experiences surge, it essentially undergoes a momentary internal gas flow reversal. This flow reversal will slightly elevate the stator casing inlet gas temperature. The increase results from the intermediate gas, which has been heated by compression, flowing back into the inlet area. This detection system should be implemented in addition to a primary antisurge control system. It consists of thermocouples installed in the airstream just upstream of the inlet guide vanes. Upon reaching a predetermined setpoint, which indicates surge, a signal is sent commanding the surge control valves to go to the full open position. Upon reaching an acceptable level of temperature, i.e., the compressor has recovered from the surge, the surge valves are permitted to return to the normal position and control of the surge valve is returned to the primary surge control system.
483
484
Process Plant Machinery ,~
OPERATIONAL LIMITATIONS ABC DEF EBF DGH GHA
c r-w
-
ROTATING STALL, FR()NT STAGES SHOCK STALL, FRONT STAGES CHOKING STALL, DISCHARGE STAGES ROTATING STALL, DISCHARGE STAGES SURGE
SURGE
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CAPACITY
FIGURE 12-7 Operating envelope for axial compressors. (Source: Dresser-Rand Co., Phillipsburg, N.J.) Choked Flow Choking occurs when the slope of the pressure ratio versus capacity curve approaches infinity. It occurs at the point where a further increase in mass flow through the cascade is not possible. Choked flow is associated with the flow velocity reaching a Mach number of 1.0 at some cross section within the machine. Unlike with surge, there is no accompanying increase in noise level or machine vibration amplitude. Choke is a "quiet" phenomenon, which, when operation continues for extended periods of time, can cause damage to the rotating blades, with eventual failure possible.
Choke Control A choke control system is needed to avoid operation within the detrimental region. A typical system would encompass a choke valve designed for minimum pressure drop in the open position and capable of full response in ten seconds. The valve would open in the event of signal failure and would respond to a control system using differential pressure (flow) and discharge pressure as inputs. Surge and choke conditions are affected by geometry, speed, and ambient conditions. Each of the aerodynamic instabilities is most likely to occur within a particular primary region of the performance characteristic. Figure 12-7 shows these various regions. STANDARD MAINTENANCE CONSIDERATIONS Like all machinery, axial flow compressors require both periodic preventive and corrective maintenance. A daily review of key operating data is quite often the best preventive maintenance strategy for axial flow compressors. These data should include various ambient and process parameters that define the operation of the unit.
Axial Flow Compressors
In addition, machinery vibration and bearing temperature data should be logged. While these daily (or weekly) records may fail to tell the entire operating history, they can nevertheless establish trends of operation, i.e., has the present problem been progressing slowly for several days or weeks, or is it a sudden change? Quite often, these records can provide the information essential to avoiding unscheduled shutdowns simply through the establishment of normal trends. Just as important, an abrupt reading might be cause for immediate investigation. Such a situation may call for an unscheduled shutdown to reduce risking catastrophic failure. To maximize the benefits available from these records, it will be necessary to be consistent. Records must be maintained daily, since without the proper parameters being accurately recorded, the validity of the record could rightly be questioned. A typical listing of the parameters that should be included in the daily log is outlined in Table 12.1. It is recommended that the operators discuss the list with the equipment manufacturer and add any items that both parties feel are critical to the analysis of operational fitness of the machinery. Data logging could be done manually or automatically, using either a process computer or a dedicated machinery condition computer. Typical items to visually inspect on a regular (daily) basis are listed below. 9 Check the unit for oil leaks at flanges, instrumentation outlets, vane actuator connections, etc. 9 Observe the operation of the stator vane linkage during a change of its position. Check for binding of any components. Ensure smooth movement. "Jumping" or sporadic movement usually indicates mechanical binding or foreign matter in linkage joints. 9 Listen to the unit for unusual sounds, such as rubbing (seals or blades) or leaking gaskets. If the sound intensity is significant, investigate the probable causes. 9 Check bearing oil drain sight flow indicators to ensure good oil flow through the beatings. Note any significant change in the running level, e.g., much fuller than normal, less flow than normal. TABLE 12.1
Parameters for Condition Monitoring of Axial
Compressors Inlet temperature Inlet pressure Inlet relative humidity (dew point temperature) Unit flow (at inlet or delivered) Discharge pressure Discharge temperature Stator vane position Unit speed Radial vibration amplitude Axial position Surge valve position Choke valve position Radial bearing pad temperature Thurst bearing pad temperature Lube oil supply temperature Lube oil supply pressure Lube oil drain temperatures Lube oil total flow
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Process Plant Machinery
Compressor Internal Cleaning Modem axial flow compressors will operate for long periods between shutdowns. Well though-out metallurgy is essential in the design and manufacture of rotor blade components. External surface coatings are applied to protect the blades from corrosive and erosive attack. Along with the addition of coatings, compressor blade life may be increased by installing inlet air filtration systems. On-line cleaning of internals is usually considered after severe degradation of compressor performance is noted. With a properly sized and operating inlet filtration system, fouling should be minimized. However, on-line cleaning is treating a symptom rather than curing the cause. The problem is more directly addressed through investigation of the air quality entering the compressor. Proper design of the inlet filtration system has always been important to the manufacturer: proper maintenance of the system must become similarly important to the operator. There are several on-line cleaning methods employed by operators depending on the process and the available cleaning methods. The question of whether or not axial flow compressors should be cleaned during on-line operation is a complex one. Due to possible problems with each of the cleaning methods currently used, it is appropriate to consult the manufacturer. Below is a short description of some of the systems currently in use. The possible problems, from both the process operation and machine reliability points of view, are also highlighted. The most effective cleaning procedure for the compressor is a low-speed water/kerosene wash. It is to be performed at approximately 30% of normal speed. The compressor is essentially soaked in the cleaning solution for approximately thirty minutes. A rinsing cycle removes residual cleaning fluid, and returning to full speed effectively dries the intemal components of the compressor. This is technically the safest and most effective cleaning method. It is, however, the least desirable from the operational view, since it requires the unit to be removed from the process for approximately one hour. This is the approximate time required to complete the procedure. A second process employed today is a water/solvent spray wash system. It is to be used at full speed and, in most cases, is not detrimental to the process. This system requires the addition of a spray nozzle assembly into the inlet casing of the compressor unit. Commercially available cleaning fluids are used. Possible problems include the incomplete atomizing of the fluid prior to entry into the compressor. The blading could be damaged by the impingement of large water particles. In addition, pulsations created on the rotor blades from the spray nozzles add another excitation to be considered in the natural frequency and stress analysis. Most importantly, while this system is somewhat effective on the first stage on the machine, the cleaning efficiency is drastically diminished at each successive stage. The cleaning medium is simply centrifuged to the casing wall and has little effect on either the rotating or stationary blades downstream. The third method of on-line cleaning is one that has been employed in similar equipment for some time. This procedure entails the introduction of crushed walnut shells (or apricot pits) into the airstream. These "cleaning media" are introduced into the piping upstream of the inlet casing. They are fed at a rate of approximately 50 pounds every two minutes. The possible problems with such a system are very similar to those encountered with the spray wash system. The solids are centrifuged to the casing wall so quickly that the effectiveness of the system is severely diminished beyond the first stage. In units employing blade coatings, nut shells or pits can cause accelerated erosion of the blade coatings. In any event, strict control of the injection rate will be required.
Axial Flow Compressors 487 Corrective Maintenance Corrective maintenance may become necessary every three to five years. At that time, inspection and replacement of wearing parts is often appropriate. Prior to the inspection shutdown, an inventory of the available spare parts should be performed. Discussions with the manufacturer should take place to determine that proper quantities of spares are available to ensure a complete and timely turnaround of the machine. A typical inspection and replacement shutdown may require approximately two weeks. During that time, typical inspections would include the following: 9 Rotor blade cleaning and magnetic particle nondestructive testing to ensure the integrity of parts if a complete spare rotor is not available. 9 Nondestructive testing of rotor discs, particularly in the area of blade attachment, to check for evidence of stress-related damage. The normal method is dry powder magnetic particle inspection. 9 Liquid penetrant nondestructive testing of stationary blading to ensure the integrity of the welded joints within the assemblies. 9 Beating clearance check on the previously installed beatings as well as the new bearings. Visual and dimensional inspection of the beatings for evidence of rubbing, wiping, and unusual wear. 9 Rotor check balance to correct any unbalance introduced by rotor blade replacement. Rotor tip clearance check to ensure that proper running clearances are established for safe operation. 9 Variable stator vane linkage check to replace any worn bushings or locking/locating pins. 9 Instrumentation check to ensure the operational indicators to be recorded are accurate and available. 9 Coupling inspection for tooth wear on gear couplings and diaphragm check on diaphragm couplings. 9 Axial rotor-to-stator clearance check during reinstallation of the rotor assembly to ensure the proper rotor-to-thrust bearing positioning. 9 Seal clearance check on all shaft and bearing assemblies. The shaft should be inspected on removal from the unit for any signs of seal contact. 9 Inlet filtration system check to ensure the filter elements are clean and secure. In addition, the inlet piping should be inspected from the inside to ensure no loose pieces exist or foreign objects are inside, which could enter the unit and cause damage to the compressor.
SELECTING AN AXIAL COMPRESSOR* (For additional reference material, see also appendices on barrel and isotherm compressors in the previous chapter, as well as environmental factors involved in Environmental Engineering and Management: Sustainable Development for the Power Generation, Oil & Gas and Process Industries, Butterworth-Heinemann 1998). Axial compressors are being used increasingly for applications which not long ago were clearly considered the domain of centrifugal machines. Thanks to their high specific flow capacity, the corresponding low weight, reduced space requirement and particularly their high efficiency, the axial compressors play a major * Source:Sulzer Turbo Ltd, Zurich, Switzerland. Adapted with permission.
488
Process Plant Machinery
role in the reliable and economic operation of modem, large-scale industrial plants. They now form a vital and indispensable integral part of installations like blast furnaces, air separation plants, fluid catalytic cracking units, nitric acid plants, jetengine test facilities, thermocompression, liquefied natural gas (LNG) and synfuel processes. The present-day technique of modem axial compressors is based on decades of experience. Thousands of units are in continuous operation in industrial plants all over the world, to which thousands more serving as combustion air compressors of gas turbines have to be added. The power input ratings of the industrial applications vary between 2000 and 90000 kW. With a view to economical manufacturing and stocking of the major components, the range of compressors has been standardized. The stocking of major components facilitates prompt delivery of machine parts such as rotor blades and stator blades, bearings, joints, etc., for service requirements. The systematic design of components over the whole range of sizes enables the compressor to be adapted optimally to the required operating conditions. Measurements conducted on the blading of various designs and sizes ensure exact conformity of the design data with the operating conditions.
Product Range The axial compressors manufactured by one of the major producers, Sulzer, come in two types: Type A Compressors with fixed stator blades (FIXAX) Type AV Compressors with adjustable stator blades on all or only some stages at the inlet (VARAX) Each type consists of 12 geometrically graduated sizes with rotor diameters extending from 40 to 140 cm. This completely covers a suction volume range of 70 000 to 1250 000 m3/h. The required compressor size and number of stages, together with the corresponding standardized overall length, are selected according to the suction volume and the thermodynamic head.
Axial Flow Compressors
Type A (FIXAX) The A-type FIXAX compressor is generally used whenever the driver is a steam turbine, a split-shaft gas turbine or a variable-frequency high-speed synchronous motor. The required operating points can be attained by speed variation, and there is basically no need for adjustable stator blades. Fixed-blade machines are also selected for installations where only minor flow variations are required, or if the mass flow is adapted by variation of the suction pressure as in aerodynamic test facilities, for example.
Type A V (VARAX) The AV-series with adjustable stator blades permits a large stable operating range at constant speed. It is therefore used for constant-speed electric motor drive. Nevertheless, this type is being increasingly preferred for steam turbines and singleshaft gas turbines as well. In this particular case, the stator blade control either facilitates operation with limited speed control range (increased reliability of operation for certain turbine types) or, in combination with the speed control, provides an additional extension of the operating range and an improvement of the overall efficiency at part load. Furthermore, it offers the advantage of quick adaptation of the compressor to changed operating conditions without acceleration of the s e t - a characteristic which is of great interest for the periodic charging of air heaters in blastfurnace blowing plants. Roughly 50% of the major axial compressors sold since 1960 have been equipped with adjustable stator blades.
Stator Blade Setting with Electric Servomotor For a great number of processes, the reference value of pressure or mass flow is selected at the process control panel and transmitted to the compressor servomotor. In this case this servomotor is of the electric type with the additional possibility of manual adjustment of the stator blades.
Stator Blade Control with Hydraulic Servomotor If the process calls for automatic pressure or mass flow control, the stator blade adjusting mechanism will be operated by a synchronized pair of hydraulic servomotors.
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Process Plant Machinery
With the exception of the stator blades and their adjusting mechanism, the same standardized construction elements are used for both FIXAX and.VARAX types. Performance Data The following diagrams (Figures 12-8 to 12-11) facilitate the selection of Compressor size 9 Nominal diameter 9 Number of stages Power input Speed Discharge temperature Using Capacity Suction pressure Suction temperature Relative humidity of the air or gas Discharge pressure Molecular mass Isentropic exponent Compressibility factor The following factors and symbols are also used for the calculation: Suction volume (actual) Absolute humidity Polytropic efficiency Indices Suction branch Discharge branch Dry Wet Information Information concerning the selection and performance calculation of an axial compressor is provided later in this section.
D (cm) z(-)
P (kW) n (rev/min) T2 (K)/t2 (~ m (kg/s)
pl (bar abs) Tn (K)/h (~ ~o, (%) P2 (bar abs) M (kg/kmol) k (-) z
(-)
V1 (m3/s) x (-) 1 2
t f
Axial Flow Compressors
Important
Type
491
Experienced engineering personnel are always needed to deal with customer inquiries and optimization of equipment layout.
designation
Figures 12-10 and 12-11 are valid for air at suction conditions l bar, 20~ 70% relative humidity. (Source: Sulzer Turbo Ltd., ZUrich, Switzerland) Figures 1212 and 12-13 describe operating characteristics for type A and type AV compressors respectively. Figure 12-14 is a longitudinal cross-sectional view of an AV compressor. Depending on the specific process requirements, such as extremely high turndown, maximum efficiency within a certain range, flat or steep p - V curve, etc.,
FIGURE 12-8 Determination of the absolute humidity x and the molecular mass Mf of the wet gas. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
492
Process Plant Machinery
FIGURE 12-9 Determination of the discharge temperature t2. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland) various parameters of the blading may be changed in order to adapt the characteristics to such special conditions.
Design Features 9 Robust design with cast casing and separate blade carrier 9 Casing supported by means of pendulum supports (minimum expansion forces) or feet with keyways for the smaller frame sizes 9 Solid or hollow rotors, very smooth running characteristics with integrated balancing pistons 9 Blading with optimal aerodynamic characteristics- with high efficiency, large specific capacity, low stressing, favourable control characteristics 9 Great safety against blade vibrations thanks to the careful selection of the blade profiles and the special blade fixation 9 Adjustable stator b l a d e s - of standard design - for optimum flow control
Axial Flow Compressors
493
FIGURE 12-10 Type A (FIXAX). Selection of the compressor size: nominal diameter D (cm) = frame size as a function of the actual wet suction volume flow V f, (m3/s). Determination of the approximate number of stages z, nominal speed n (rev/min) and power input P (MW). (Source: Sulzer Turbo Ltd., Zi~'rich, Switzerland)
FIGURE 12-11 Type AV (VARAX). Selection of the compressor size: nominal diameter D (cm) =frame size as a function of the actual wet suction volume flow Vf, (m3/s). Determination of the approximate number of stages z, nominal speed n (rev/min) and power input P (MW). (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
9 Maintenance-free, oil-free stator blade adjusting mechanism which is also protected against the ingress of contaminants 9 Possibility of fitting various types of bearings 9 Possibility of fitting various shaft seals
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Process Plant Machinery
FIGURE 12-12 Operating characteristics for a compressor, type A, with fixed stator blades and variable speed: NP-Reference point (100%)- Design point; P2/Pl- Pressure ratio. Valid for: constant gas data, constant inlet temperature. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
FIGURE 12-13 Operating characteristics for a compressor, type AV, with adjustable stator blades and constant speed: N P - Reference point (100%)- Design point; P2/Pl - Pressure ratio; a - Angular setting of the stator blades. Valid for: constant gas data, constant inlet temperature. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
Axial Flow Compressors
TABLE 12.2 Selection and performance calculation of an axial compressor
calculation example Air compressor, type A
Given Capacity Suction pressure Suction temperature Relative humidity Discharge pressure Dry molecular mass Isentropic exponent Cp/Cv Compressibility factor
rilt = 146.2 kg/s P l = 1 barabs T1 = 3 0 8 K, tl = 3 5 ~ tpl = 60% P2 = 4.5 bar abs Mt = 28.95 kg/kmol k=l.4 Z=I
Calculation instructions 1. Determination of the absolute humidity x and wet molecular mass Mf (from Tl, Pl, tpl) with Figure 12-8
x =0.021 Mf = 28.6 kg/kmol
2. Calculation of the wet mass flow lilf = riat(l + x)
lilf -- 146.2(1 + 0 . 0 2 1 ) = 149.3 kg/s
3. Determination of the actual suction riaf. T1 [m3/s] volume '~/1 = F * . M f . p l
Vl = 133.7 m3/s
4. Determination of the discharge temperature t2 with Figure 12-9
t2 = 225 ~
5. Selection of the compressor size (nominal diameter D) and number of stages Z, determination of speed n and power input P with Figure 12- l 0
D - - 9 0 cm Z=I2 n -- 3600 rev/min P --- 29 M W
6. Type designation
A 90-12
7. Selection of compressor size AV with Figure 12-1 ! 7.1 Steam turbine drive, i.e. no speed restriction (speed ns = nn" x / Z t h / Z s ) = 3926 .x/10.6/11.0 Type designation
D-90 cm Z--II n = 3854 rev/min P = 29 M W AV 90-11
7.2 Electric motor direct drive 3600 rev/min
D = 100 cm Z=II n = 3600 rev/min P = 29 MW AV 100-11
Type designation 7.3 Electric motor direct drive 3000 rev/min (an) 2 (3533) Zs=Zth" -= 10.6. 9
D = 100 cm Z=I5 n = 3000 rev/min
as
P = 29 MW = 14.7 Type designation nn ns Zth Zs
= = = --
AV 100-15
Nominal speed Selected speed for electric motor direct drive or corrected to arrive at a whole number of stages Theoretical number of stages on the basis of Figures 12-10 and 12-11 Selected number of stages to arrive at a whole number or to account for reduced speed in example 7.3. The reduced rotational speed imposed by the speed of the two-pole motor reduces the tip speed proportionately. This must be compensated by increasing the number of stages to produce the desired pressure ratio. * Factor F, if mass rnt given in Nm3/h = 33505 SCFM (14.7 psia, 60F) = 20844 SCFM (14.7 psia, 70F) = 21245 kg/s = 12.027
495
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Process Plant Machinery
FIGURE 12-14
Longitudinal and cross-sectional view of an A V compressor. (Source: Sulzer Turbo Ltd., Ziirich,
Switzerland)
Casings, Bearing Pedestals Depending on the type of gas and the design pressure, the casings are made of grey cast iron, nodular cast iron or cast steel. The cast design facilitates a rigid construction, effective noise attenuation and aerodynamically favorable layout of the respective ducting. The suction and delivery branches are usually routed vertically downwards. In cases where, due to the composition of the gas and/or the pressure level, steel casings are mandatory, a welded construction can be supplied. The suction branch may then be axial, or both suction and discharge may be routed upwards or downwards. In the vertical central plane, the casing is aligned by two keyways; it is equipped with four supporting feet. It is fixed at one end in the axial direction by one pair of feet. The other pair of feet of large frame sizes rests on pendulum supports with spherical contact surfaces. As a result of this, the casing can take up the thermal expansion in both the axial and lateral direction without difficulty. This feature is particularly advantageous in the case of light steel foundations. On small frame sizes the pendulum supports are replaced by sliding keyways.
Axial Flow Compressors 497
Blade Carrier, Casing Inserts The blade carder inserted in the casing is centered on both the suction and discharge side, and is able to expand freely in the axial direction. The diffusor and the gland inserts are also fitted as separate parts in the casing. The double-casing design with outer casing and blade carder offers various advantages: 9 Rigid casing construction; the clearances in the blade duct are not influenced directly by external pipe forces. 9 Simple fitting of the blades and assembly of the casing parts; the top half of the casing can be raised without dismantling the blade-adjusting mechanism. 9 Possibility of fitting different blade carriers, for adapting the blade duct and thus the compressor characteristics to greatly changed operating conditions. 9 Optimal protection of the adjusting mechanism in the space between the casing and blade carder; the space is kept under suction pressure in order to safeguard the adjusting mechanism against condensation and corrosion attack.
Rotor The rotors are usually of forged solid design. In the case of larger machines or if the moment of inertia must be minimized to limit the power requirement when running up with an electric motor, welded hollow rotors may be used. Integrated balancing pistons at both ends of the rotor facilitate an equalization of the axial thrust. The careful balancing of the rotor at full speed results in highly smooth running characteristics. If necessary, balancing can also be effected in the casing. The labyrinth strips are caulked in the rotor.
Blading Blading with a high degree of reaction, i.e. the increase in pressure takes place exclusively in the rotating components, is employed for the compression of lighter gases such as helium or hydrogen. (See Figures 12-15.) For all other applications, such as the compression of air, blading with a lower percentage reaction is adopted. The increase in pressure is distributed to the rotor and stator blades. This enables the following major advantages to be realized: 9 Higher efficiency with lower aerodynamic loading 9 Widest possible control range with high part-load efficiency at constant speed 9 Largest possible suction volume at given speed 9 Increased reliability of operation thanks to larger radial blade clearances and the omission of guide vane sealings 9 Steeper pressure volume characteristics, especially suitable for capacity control, for the parallel operation of different compressors, for refrigeration processes and for exact adjustment of the blow-off line. The rotor and stator blades are normally made of 13% chrome steel and machined. When handling highly contaminated aggressive air or corrosive gases, alloys with higher chromium and nickel content may be used. The rotor blades have rhomboidal roots and are firmly braced in an exactly defined position in the peripheral grooves of the rotor. This is of particular importance for their vibration-related
498
Process Plant Machinery
FIGURE 12-15 Adjustable stator blade, rotor blade and fixed stator blade with intermediate piece. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland) design. The fixed stator blades are provided with a rectangular foot. The adjustable stator blades are made in one piece with a cylindrical shaft. The latter is seated in a beating bush in the blade cartier. The high damping characteristics of this seating arrangement practically excludes the occurrence of dangerous vibration amplitudes associated with the stator blades.
Stator Blade Adjusting Mechanism for the AV types The adjusting mechanism is located in an annular space between casing and blade carrier, and is thus well protected against contaminants and moisture. It is maintenance-free and does not require any lubrication. Servomotors.
Two types of servomotors are available (see Figures 12-16 and 12-17):
Automatic mass flow, volume or pressure control. The adjusting mechanism is operated by means of two hydraulic servomotors which are affixed laterally to the casing. One of the servomotors is equipped with a positioning transmitter and the second operates hydraulically in parallel. The linear movement of the servomotor piston rods is transmitted directly to the adjusting cylinder by way of two ball and socket joints. Remote setting o f reference value. One single electric servomotor is attached laterally to the bottom half of the casing. Its driving shaft actuates a pivoted fork positioned on either side of the casing in maintenance-free beatings. This fork in turn moves the adjusting cylinder in axial direction. Adjusting cylinder. The adjusting cylinder of welded design can move in the axial direction and is dry-seated. There is no restriction of heat expansion in any direction. U-shaped guide tings are provided on the inner side in which the adjusting levers are engaged. Stator blades. The adjusting levers provided on the end of each stator blade shaft are connected to the guide tings of the adjusting cylinder by means of pivoting sliders. The axial movement of the cylinder is converted into a rotating movement of the stator blades (see Figure 12-18).
Axial Flow Compressors
FIGURE 12-16 Electric servomotor. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
FIGURE 12-17 Hydraulic servomotor. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
The self-lubricating bearing bushes of the blade shafts are seated in the radial holes of the blade carrier. O-ring packings prevent the ingress of contaminants into the stator blade seating.
Shaft Seals Labyrinth seals (Figure 12-19) are used for the standard models. The stainless steel labyrinth strips are caulked on the rotor and are easily replaceable. In case of rubbing due to unbalance, the friction-induced heat is immediately passed to the massive stator, thus avoiding distortion of the rotor. Gas-tight shaft seals and standstill seals can be fitted for special requirements.
FIGURE 12-18 Stator blade adjusting mechanism. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
FIGURE 12-19 Labyrinth shaft seals. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
499
500
Probess Plant Machinery
TABLE 12.3
Material Table, Valid for Air and Similar Gases
Machine component
Material
DIN No.
ASTM No.
Casing Blade carrier Inlet pieces Diffusors Stator blades Adjusting cylinder Journal bearing Rotor Rotor blades
Grey cast iron Nodular cast iron Grey cast iron Grey cast iron Chrome steel Boiler plate Steel and white metal Alloyed steel Chrome steel
GG-25/1691 GGG-42/1693 GG-25/1691 GG-25/1691 X 20 Cr 13/17440 H 1/17155 LgSn 89 28 Ni Cr MoV 85 X 20 Cr 13/17440
Class 35/A48 SA 395 Gr. 60-40-18 Class 35/A48 Class 35/A48 AISI 420 Class A + B/A 285 B 23 A 469-71, Class 6-7 AISI 420
Journal and Axial Bearings Journal bearings. In the normal version, i.e. with the compressor rotor solidly coupled and the rotor thrust transferred to the axial thrust bearing of the prime mover or the gear, the bearing housings are equipped only with journal beatings. Twolobe bearings are provided for the lower speed range; tilting pad journal bearings (Figure 12-20) being generally used for the higher speeds of the smaller frame sizes for reasons of stability. The slight curvature of the adjusting plates allows the bearings to be set accurately on erection. The bearings (Figure 12-21) are firmly held in position by the bearing housing top half. Two-lobe bearings are suitable for both senses of rotation, while tilting pad bearings are essentially for only one direction, although they can tolerate running backwards with a somewhat reduced load capacity. Axial thrust bearings. If a flexible coupling is selected between driver and driven machine, the bearing housing can accommodate the necessary tilting pad thrust bearing. The purpose of this beating is to absorb the remaining thrust of the machine and any significant axial friction thrust of the coupling due to sharp temporary differential expansion between rotor and casing. To provide easy access and reduce the overhang, it is preferably mounted on the free shaft end. The tilting pads are supported on load equalizing segments which allow angularity of the shaft of up to 0.3%. Because the tilting pads are supported eccentrically, thrust bearings are suitable for only one direction, but tolerate a reversed rotation at a somewhat reduced load capacity. Figure 12-22 shows a popular thrust bearing.
Solid coupling It is Sulzer's normal practice to make extensive use of solid couplings allowing the use of only one axial thrust bearing for single- or multiple-casing arrangements. An intermediate shaft, flexible enough to allow for considerable misalignment, is inserted between the two shaft ends of the machines to be coupled together (Figure 12-23). In case of motor-driven units, the normal technique is to use single helical gears provided with thrust collars on the pinion shaft, as shown in Figure 12-24 and 12-25. The thrust collars not only neutralize the axial thrust created by the meshing of the teeth cut at an angle to the axis of the shaft, but also transmit the unbalanced axial thrust of the high-speed rotor train to the thrust bearing on the low-speed wheel shaft. Good gear meshing requires parallelity of gear and pinion shaft and automatically assures parallelity of the contact surfaces of thrust collar and wheel rim. The slight tapering of the thrust collars is responsible for the formation of a wedge-type oil film creating a pressure zone spread out on an enlarged surface with a pressure distribution very similar to that of a standard-oil-lubricated joumal bearing.
Axial Flow Compressors 501
FIGURE 12-20 Multi-segment journal bearing with four tilting pads. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
FIGURE 12-21 Two-lobejournal bearing. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
The relative motion between the two contact surfaces of the thrust collar system is a combination of rolling and sliding and takes place near the pitch circle diameter, resulting in a very small relative velocity. The thrust transmission is therefore effected with almost no mechanical losses. The considerably reduced losses of the single thrust bearing on the low-speed shaft as compared with the high losses of individual thrust bearings on the high-speed train lead to a substantial power saving. Moreover, this low-speed bearing can be more amply dimensioned to provide a much higher overload capacity. For direct turbine-driven compressor trains, the thrust bearing is usually located in the turbine. Also in this case solid couplings with flexible intermediate shafts are much preferred. This coupling arrangement avoids heavy overhung gear couplings which are usually responsible for not clearly defined lower critical speeds and for the phenomena of torque lock leading to additional loading of the axial thrust beating. The resulting axial friction forces can become quite substantial if insufficient attention is given to the cleanliness of the lubricating oil.
Safety Systems and Process Control Two main requirements are to be met by the control system of turbo-compressors. 9 Safety: to prevent the compressor from operating in an unstable range or at other hazardous conditions. 9 Process: to adapt the compressor performance to the demands of the process Sulzer offers the engineering and the supply of complete compressor control and safety systems. This tradition ensures the optimum protection of the compressor and the plant. The control systems may be pneumatic or electronic with hydraulic or electromechanical servomotor. In cases where the compressor control system is engineered or furnished by others, it is Sulzer's practice to review and approve the system in order to ensure the compatibility of all equipment and functions. Thanks to the strict employment of standard signals, the control system can be integrated into other systems without difficulty: it allows remote control, automation of starting and stopping and integration with process computers.
502
Process Plant Machinery
FIGURE 12-22 Kingsbury-type thrust bearing with self-equalizing pads with directed lubrication. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
FIGURE 12-23 The Sulzer solid quill-shaft coupling conforms to API 671 standard and consists of the quill shaft and the two hubs hydraulically fitted onto the shaft ends of the connected machines. On each coupling side, an equal number of tie bolts for axial fixation and tapered dowel pins for torque transmission and centering assure a clearly defined connection. Balancing as a complete assembled unit and correlative marking enable removal and remounting of this intermediate shaft with the connected rotors remaining in place, without affecting the balancing quality and vibration behaviour of the complete string. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
Axial Flow Compressors
..~..~: ............. F:;;.4~,,'., ........................
F,z
/
;"
Cross-section of thrust collar
A-A
F I G U R E 12-24 Method of axial thrust transfer in a single helical gear with thrust collar. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
FIGURE 12-25
1|
._
II
"
Transfer of external forces. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
Safety systems. Antisurge control Axial compressors have a limited stable operating range, regardless of the type of blading and other influencing factors. This range is given by the characteristic curves limited by the surge lines. Surge conditions, occurring on the left side of this surge line, are avoided by an antisurge control system. It measures flow and pressure and can be designed to closely follow the actually measured surge line at a given safety margin. As soon as the operating point approaches the surge line, the controller starts opening the antisurge valve according to the pre-set values (L) (see Figure 12-26).
503
504
Process Plant Machinery
FIGURE 12-26 Characteristics of a turbocompressor p - Discharge pressure; V - Flow rate; C-Compressor characteristic curves; S - Surge line; L - Limit flow (Source: Sulzer Turbo Ltd., Ziirich, Switzerland) For air compressors, the excess capacity is blown off to the atmosphere. For a gas that cannot be wasted to the atmosphere, antisurge control is a bypass control, the unwanted flow being returned to the suction side. A bypass cooler may then be required. The antisurge control system is not a flow or pressure control device, but a safety device which has to act independently of any other control. Vibration, temperature, pressure, power limitation Under certain circumstances external influences may lead to undesired changes of the normal level of vibration, gas and bearing temperatures, pressure and power. A reliable interlock, alarm and shutdown system must protect the compressor and driver from possible damage under such conditions. Auxiliary component control Auxiliary component control assures a safe supply of lube, control and seal oil. Process control. Suction pressure: Constant suction pressure to adapt the compressor flow to an upstream production unit or to maintain a constant evaporation temperature for refrigerating units.
Discharge pressure: Constant discharge pressure in cases where chemical reactions have to take place at a clearly defined pressure, or where the compressor flow has to be adapted to a fluctuating downstream demand. Flow: Constant mass flow control corresponding to a constant plant output.
Process Control System (see Figure 12-27) System for either capacity control and pressure limitation or pressure control with capacity limitation. The capacity is measured by means of a Venturi tube. The computer FY receives input signals from the differential pressure transmitter PDT, pressure transmitter PT as well as the temperature transmitter TT and calculates the actual capacity value for the flow controller FC.
Axial Flow Compressors
The flow controller FC and the pressure controller PC are each equipped with one manual station HIC for the reference value. By way of the minimal-selection relay UY 2, the volume or pressure controller acts upon the positioning controller GC of the stator blade adjusting mechanism. The positioning circuit for the stator blade adjusting mechanism comprises the positioning controller GC, the electrohydraulic converter GY and the position feedback transmitter GT.
Load Limit Control Malfunctioning in the steam or condensate system of the driving turbine may lead in certain circumstances to an undesired drop in speed. In such cases, the stator blades will be closed to such an extent by the load limit controller SC2 that the speed will remain nearly constant. The speed controller SC 2, which is equipped with the reference value station HC, receives its input signal from the speed transmitter ST and has priority over the process controller with minimal selection relay UY2, when necessary.
Antisurge Control System System to maintain stable operation of the compressor, even when the process operating point moves into the unstable range of the compressor performance characteristic. The pressure transmitter PT is used to determine the actual value. A differential pressure transmitter PDT and a function generator UY 3 are employed for setting the reference value. This computer enables the response line (blow-off line) to be
FIGURE 12-27 Antisurge control system and combined discharge pressure~power limit control system: SC 1 - S p e e d controller; E/HGY-Electrohydraulic converter; P D T - Differential pressure transmitter; T T - Temperature transmitter; P T - Pressure transmitter; G T - Position feedback transmitter; S T - Speed transmitter; F Y - Computer; G C - Positioning controller; SC 2 - Load limit controller; UY 3 - Function generator; F C - Volume controller; P C - Pressure controller; B C - Antisurge controller; UY 2 - Minimum-selection relay; H C - Reference value station; HICManual station. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
505
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Process Plant Machinery
well adapted to the surge limit. The output signal of the antisurge controller BC acts by way of the minimum selection relay UY 2 on the positioning controller GC of the blow-off valve. The valve can also be opened by the manual station HIC and override the antisurge controller. The positioning circuit for the blow-off valve comprises the positioning controller GC, the electrohydraulic converter GY and the position feedback transmitter.
Monitoring Logic The monitoring logic supervises various controller reference values, the speed of the set, time-dependent operations such as start up and the positioning of the final control elements. The system facilitates the determination and processing of any disturbancies in the final positioning elements and transmitter signals.
General System Layout The control system of axial compressors is designed for all types of drive and the most diverse process applications. It corresponds to the latest knowledge and experience in the field of electronic control. 9 9 9 9
Transmitter according to the 2-wire system; standard signals 4 - 2 0 mA Isolating amplifier in the case of external measurement data Plausibility supervision for the determination of measurement data faults Output tracking for automatic changeover to manual control made in case of measurement failure 9 Impulse technique of the control units 9 Positioning elements driven by hydraulic servomotors and controlled by electrohydraulic converters 9 Two changeover contacts for each output for alarm and control purposes.
Features of the System 9 9 9 9 9 9 9
Proven building block system for compressor installations Central simulation and control unit Integrated monitoring system acting directly on the final positioning elements Monitoring of important values according to the "20-of-3 system" Three-week tests on load at works under elevated temperature and supply voltage Electronic fuse protection for each individual functional group Optimal adaptation of the system to the individual control parameters, thanks to the modular system 9 Self-contained functional units as plug-type modules 9 Easy extension and modification at any time thanks to the clearly defined wiring plans 9 Simple interface in the case of the use of process computers.
Typical Compressor Plant P & I Diagrams Figure 12-28(a) depicts automatic control by means of adjustable stator blades. Figure 12-28(b) depicts remote setting of the adjustable stator blades.
Axial Flow Compressors
FIGURE 12-28(a) Automatic control by means of adjustable stator blades
List of major components for automatic (Figure 12-28(a)) 1 Driving motor 2 Gear coupling, oil-filled 3 Speed-increasing gear with thrust collars, thrust bearing on the slowrunning shaft 4 Main oil pump, laterally flanged to the gear-box, driven by the slow-running gear shaft 5 Solid coupling 6 Turning gear 7 Axial compressor with adjustable stator blades 8 Labyrinth glands 9 Servomotor of the stator blade adjusting mechanism 10 Two-stage air intake filter consisting of: 11 Inertia dust separator (lst stage) 12 Dust evacuation fan 13 Roll-band filter (2nd stage) with driving motor and automatic forward feed control
control of adjustable stator blades
15 16 17 18 19 20 21 22 23 24 25 26 27
14 Bypass flap with open alarm switch GAO Suction silencer Non-return valve with air-operated closing cylinder and solenoid valve Antisurge valve, hydraulically operated Venturi tube for measuring the air flow Discharge silencer Blow-off silencer Noise-attenuating hood Ventilation fan with silencer Lube oil supply unit Control oil supply unit Electronic control system and monitoring logic Discharge isolating valve Jacking oil pump
507
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Process Plant Machinery
FIGURE 12-28(b)
Remote setting of the adjustable stator blades
List of major components for remote 1 Driving motor 2 Gear coupling, oil-filled 3 Speed-increasing gear with thrust collars, thrust bearing on slowrunning shaft 4 Main oil pump, laterally flanged to the gear-box, driven by the slow-running gear shaft 5 Solid coupling 6 Turning gear 7 Axial compressor with adjustable stator blades 8 Labyrinth glands 9 Electric servomotor and gear for remote setting of stator blade adjusting mechanism by means of controller HIC; gear with local manual intervention facility 10 Two-stage air intake filter consisting of: 11 Inertia dust separator (1 st stage) 12 Dust evacuation fan 13 Roll-band filter (2 nd stage) with driving motor and automatic forward feed control 14 Bypass flap with open alarm switch GAO
setting of adjustable stater blades (Figure 12-28(b)) 15 Suction silencer 16 Non-return valve with air-operated closing cylinder and solenoid valve 17 Antisurge valve, pneumatically operated with air failure to open safety device (pressure accumulator) and solenoid valve. The antisurge valve is controlled via an electric/pneumatic converter E/P by the antisurge controller BIC. Manual control of the valve is possible in the sense "Valve to open" (the antisurge controller thereby being overridden) by operating the manual loader HIC 18 Function generator, computing the set points of the antisurge controller BIC in function of the aspired air flow rate 19 Discharge silencer 20 Blow-off silencer 21 Noise-attenuating hood 22 Ventilation fan with silencer 23 Lube oil supply unit 24 Discharge isolating valve 25 Jacking oil pump
Axial Flow Compressors
FIGURE 1 2 - 2 9
Typical oil supply schematic
Scope of supply and functional description for oil supply components (Figure 12-29) 1 2 3 4 5 6 7
8
9
10
11
12
Lube oil tank with auxiliary equipment Electric oil heaters Oil mist fan Filling sieve and breather Degasifier plates Suction strainers Safety arrangement (4 non-return valves) in the main oil pump suction and discharge lines to prevent backflow of oil if the compressor and hence the main oil pump should accidentally turn in reverse direction Auxiliary oil pump, electric-motordriven; automatic start up of the pump in the event of oil pressure dropping Twin oil cooler, each of the coolers being sized for full flow, with transfer valve for cooler changeover during operation Oil temperature control valve, maintaining constant the oil cooler oil outlet temperature by bypassing more or less the oil around the cooler Oil pressure control valve, maintaining constant the lube oil supply pressure Twin oil filter, each of the filter screens being sized for full flow,
13
14 15 16 17 18 19
20
21
22
with transfer valve for filter changeover during operation and pressure loss indicator as well as alarm PDIAH Testing device for checking the automatic start up of the auxiliary oil pump Control oil tank with auxiliary equipment Electric oil heaters Filling sieve and breather Degasifier plates Suction strainers Control oil pumps, motor-driven; one pump in operation, the other one acting as standby pump, starting automatically if control oil pressure drops Control oil filters, two in parallel each one sized for full flow, with transfer valve for filter changeover during operation and pressure loss alarm PDAH Testing device for checking the automatic start up of the standby pump Bladder-type pressure accumulators assuring that the compressor stator blades are closed properly in the event of control oil pump failure
509
5 lO
Process Plant
Machinery
Symbols for the above schematic diagrams Line symbols
Mechanical
P r o c e s s main G l a n d b a l a n c i n g line G l a n d l e a k a g e line 9,,,,,, L u b e oil s u p p l y ,,,-,,, J a c k i n g oil ,., -- L u b e oil r e t u r n line ,, = -, Oil mist :.'. ..... C o n t r o l oil s u p p l y -,,.. ,+,... C o n t r o l oil r e t u r n line ,,, ,,,, C o o l i n g w a t e r - - - - V e n t line 9 9 - Dust e v a c u a t i o n line 9 9 I n s t r u m e n t air -;- c- E l e c t r i c ~ e l e c t r o n i c s i g n a l Mechanical connection Valve
-r
and apparatus
symbols
Globe/ball/disc valve Gate valve V a l v e with c o n t r o l trim T h r e e - w a y valve GE~ N o n - r e t u r n v a l v e N o n - r e t u r n v a l v e with b o r e IZ] Suction s t r a i n e r Spring-loaded safety valve F l o w orifice Drainplug Double acting piston actuator Piston a c t u a t o r with c o u n t e r spring Diaphragm actuator Solenoid E l e c t r i c motoz Pressure accumulator Electric/pneumatic converter Electric/hydraulic converter ~ ) Low s i g n a l s e l e c t i n g r e l a y
+
m
symbols
~__+0,, F l e x i b l e c o u p l i n g Solid c o u p l i n g ~'-" J o u r n a l b e a r i n g =~ Thrust bearing, tilting-pad type [~] G e a r with thrust c o l l a r
..J"l_
Labyrinth gland chamber Instrument
symbols
M e a s u r e d v a r i a b l e s (first l e t t e r ) B Antisurge F Flow (3 Position H Hand L Level P Preasure PD P r e s s u r e d i f f e r e n c e T Temperature O u t p u t function (following l e t t e r ) A Alarm G S i g h t - g l a s s without m e a s u r e m e n t I Indication S Switch function E x t r e m e v a l u e s (last l e t t e r ) C Closed 0 Open H High L Low S i g n a l processing ~) C o m p r e s s o r d r i v e r s t a r t p e r m i s s i o n " ~ ' R ~ n i n g g e a r start p e r m i s s i o n 4) Compressor driver shutdown C ~ m p r e s s o r driver start up interlock Abbreviations
A
E x h a u s t to a t m o s p h e r e
D
Drain
V
Vent
Axial Flow Compressors
511
FIGURE 12-30 Layout drawing of a compressor, type AV, with driving motor and gear unit. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
Typical plant layout Figure 12-30 shows an example of a blast furnace blower plant. Figures 12-31 to 12-38 are photographs in various field situations.
512
Process Plant Machinery
FIGURE 12-31 Two steam-turbine-driven axial compressors delivering 330000 Nm3/h and 450000 Nm3/h of air at 5.3 bar to the blast-furnace plant of Hoogovens, ljmuiden (NL). Power input 29 000 KW and 40 700 kW respectively. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
FIGURE 12-32 One of the two motor-driven compressor~expander trains in the nitric acid plant of Norsk Hydro at HerCya (N). From left to right: axial tail-gas expander, 123 000 Nm3/h, 7700 kW; centrifugal nitrous gas HP compressor, 140000 Nm3/h, 4330 kW; axial nitrous gas LP compressor, 180000 Nm3/h, 12 150 kW. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
Axial Flow Compressors
513
FIGURE 12-33 Two axial water vapour compressors for a thermocompression/evaporation plant in the Austrian salt works Steinkogel. Capacity 81 t/h or 95 000 m3/h, power input 3520 kW each. The compressors are equipped with a special washing device, periodically injecting demineralized water in order to remove salt deposits which may otherwise accumulate on the critical internal parts within the flow-path. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
FIGURE 12-34 Turbocompressor train for the nitric acid plant of SASOL, Secunda (South Africa), on the test bed From left to right axial tail-gas expander 111 000 Nm3/h, 10950 kW; centrifugal nitrous gas compressor, 122 500 Nm3/h, axial air compressor, 139000 Nm3/h, 11890 kW condensing~extraction steam turbine 6500 kW, 6600 rev/min. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
514
Process Plant Machinery
FIGURE 12-35 one of the three identical two-casing axial compressor sets for the compression of 290000 Nm3/h of a hydrocarbon mixture from 1.42 to 39.2 bar, installed in the natural gas liquefaction plant (LNG) at Skikda (Algeria). Each steam-turbine-driven compressor train requires a power input of 78 MW. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
Completed air compressor module incorporating axial compressor A 71-10, 13-MW steam turbine and condenser. The direct driven compressor supplies 231 000 Nm3/h of catalyst regenerating air at 3.4 bar for a fluid catalytic cracking unit (FCCU) of Statoil at Mongstad, Norway. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland) FIGURE 12-36
Axial Flow Compressors
FIGURE 12-37 The word's first compressed-air energy storage (CAES) installation, operated by the Nordwestdeutsche Kraftwerke AG (NWK) at Huntorf (Germany). The compressor unit, consisting of an axial LP air compressor and a centrifugal HP compressor coupled through a gear to the motor/generator of the BBC gas turbine, pumps 300 000 Nm3/h at 46 to 72 bar into a salt cavern during the charging cycle and absorbs 60 MW. The gas turbine produces 290 MW during the peak-load period. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
FIGURE 12-38 Three of four identical axial compressor sets installed at the US Air Force Aeropropulsion System Test Facility of the Arnold Engineering Development Center, Tullahoma, Tennessee, USA. Each compressor delivers 352 000 Nma/h of air at 3 bar and absorbs 18000 kW. Another two 35-MW compressors of the same frame size can boost the air to 9.6 bar, but both LP and HP machines are operated at different pressure levels for aircraft engine tests at various simulated flight conditions. (Source: Sulzer Turbo Ltd., Ziirich, Switzerland)
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Chapter 1 3 Propeller, Axial, and Centrifugal Fans* Fan applications in modern process plants range from handling fresh air to moving large volumes of corrosive and abrasive gas streams. This chapter deals with the fundamentals of selection and operation of fan systems. Industrial fans can be classified into three basic groups: propeller, axial and centrifugal. PROPELLER FANS
Propeller fans utilize long slender blades twisted in such a manner as to provide some angle of attack on the gas being moved. The blades are fixed to a hub, and the entire assembly rotates in a housing. The housing has little or no effect on controlling the gas flow. Typical applications for propeller fans are wall- and ceiling-mounted exhausters (Figure 13-1) and cooling tower and air-cooled heat exchangers. Pressures generally range from 0 to 1 inch of water column with efficiencies ranging from 10 percent to 35 percent. Tip speeds are often limited so as to minimize noise generation.
AXIAL FANS
Axial fans are basically propeller fans with shorter rigid blades assembled in a tubular housing to provide some degree of controlling and/or streamlining the gas flow. Figure 13-2 shows a small, simple, single-stage unit, while Figure 13-3 depicts a large, heavy-duty industrial version. This type of fan propels gas axially through its housing, which acts as an integral part of the ducting. Axial fans can be further subdivided into two types, namely, vane axial and tube axial. Because of the type of construction employed, these fans lend themselves easily to multistaging if higher pressures are required. Figure 13-4 illustrates a two-stage axial fan. Vane axial fans have vanes installed before and/or after the rotor to direct and streamline the gases for better fan action. The vanes are installed on the stator housing and employ some kind of control mechanism for varying the flow pattern. Vane axial fans are used primarily in clean gas service, handling pressures generally up to 20 inches of water column (single-staging), but higher pressures are available. Multistaged vane axial fans can achieve efficiencies in excess of 85 percent and are generally considerably more efficient than tube axial fans. * Source: Garden City Fan and Blower Company, Niles, MI (except as noted). Adapted by permission. 517
518
Process Plant Machinery
FIGURE 13-1 Wall-mounted exhaust fan (propeller-type). (Source: A CME Engineering & Manufacturing Co., Muskogee, "OK.)
FIGURE 13-2
gee, OK.)
Small axial fan. (Source: ACME Engineering & Manufacturing Co., Musko-
Propeller, Axial, and Centrifugal Fans
t
FIGURE 13-3 Heavy-duty single-stage axial fan. 1 - Access door; 2 - diffuser; 3 - external pitch control lever; 4 - stationary blades; 5 - variable pitch rotating blades; 6 - removable upper fan housing; 7 - main bearing assembly; 8 - coupling; 9 - inlet box; 1 0 - motor; 11 - blade pitch control mechanism; 12 - rotor assembly; 13 - shaft tube; 14 - drive shaft. (Source: C E M A X Fans, Boston, MA.)
FIGURE 13-4 Heavy-duty two-stage axial fan. 1 - Access door; 2 - diffuser; 3 - removable stationary blade; 4 - bolted blade retainers; 5 - removable variable pitch rotating blades; 6 - removable upper fan housing; 7 - coupling; 8 - external pitch control lever; 9 - blade pitch control mechanism; 1 0 - rotor assembly; 11 - main bearing assembly; 12 - drive shaft; 13 - shaft tube; 1 4 - shaft seal. (Source: C E M A X Fans, Boston, MA.) Tube axial fans employ the same basic construction as vane axial fans except for the guide vanes and are generally less efficient than vane axial fans. Recently, variable-pitch rotor blade construction has been offered as a standard package for both tube and vane axial fans. Typical performance characteristics of axial flow fans are shown in Figure 13-5. Note the pronounced "dip," or stall range, in both the static pressure and brake
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Axial flow fan characteristics
horsepower (BHP) curves. This represents the unstable range in which surge phenomena occur and where fan operation should be avoided. It also explains why axial fans operating in parallel are difficult to start up. Although the absorbed horsepower is initially high, its subsequent decline makes it a non-overloading characteristic.
CENTRIFUGAL FANS Centrifugal fans (Figure 13-6) are prevalent in virtually all of the industry. These fans are applied in building and equipment ventilation, hot-gas recirculation, dusthandling systems, and furnace and boiler forced draft/induced draft services. The basic difference between centrifugal fans and the types previously mentioned is the fact that gas enters the impeller axially and leaves the rotating element radially. The casing is designed and installed around the rotating element to direct the discharge gas in a manner suited for the particular application. Centrifugal fans operate with typical efficiencies in excess of 80 percent and an achievable limit of 91 percent with airfoil blade construction. Discharge pressures can be as high as 90 inches of water column or higher and cover a wide operating range. Centrifugal fans can be categorized by three basic construction styles depending on blade orientation with respect to impeller rotation. These styles are radial blade, backward-inclined blade, and forward-inclined blade. Subtle variations of each style are available.
Radial Blade As the name implies, the blades on this type of impeller are radially oriented. This type of impeller construction can be further subdivided into straight radial blade and radial-tip blade styles. The merits of each of these different blades can be observed in Figures 13-7 through 13-10. The straight radial bladed fan illustrated in Figure 13-7 has a best efficiency of approximately 72 percent. The fan has an
Propeller, Axial, and Centrifugal Fans
FIGURE 13-6 Centrifugalfan for typical process plant. (Source: A CME Engineering & Manufacturing Co., Muskogee, OK.)
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FIGURE 13-8
Radial-tip blade fan performance characteristics
overloading horsepower characteristic, but it features stable operation over the entire performance range. Because of low specific speed, straight radial bladed fans have relatively good resistance to abrasion from entrained particles. The radial-tip fan of Figure 13-8 minimizes gas turbulence by curving the blades. It achieves higher efficiencies, with 78 percent to 83 percent being not uncommon. Fan horsepower is nonoverloading in the high flow range. This fan also has low specific speed and medium resistance to abrasion. Figure 13-9 illustrates an open radial bladed fan with a geometry that readily allows the addition of simple wear plates for highly abrasive applications. It is
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Open radial blade fan characteristics
Propeller, Axial, and Centrifugal Fans lowest in efficiency, with 65 percent being typical. This fan has an overloading power characteristic. In summary, radial blade fans are commonly used for induced draft service and are especially suited for contaminated gas streams where airborne particles and fouling could be a problem. Backward-Inclined
Blades
The fan blades on this style of rotor are oriented backwards with respect to the direction of rotation. Three variations of backward-inclined blades are available. Figures 13-10 through 13-12 illustrate both geometry and performance of these variations: backward-inclined fiat plates, curved plates, and airfoil blades, with efficiencies increasing in that order. Typical efficiencies for flat plate blades range from 77 percent to 80 percent and 84 percent to 91 percent for airfoil blades. The flat and curved plate blades can handle fine airborne particles that would normally damage the airfoil design. These fans are commonly used in forced draft services. Modification of the backward-inclined flat plate profile of Figure 13-10 reduces the unstable range to moderate instability in the backward-curved blade (Figure 13-11) and virtually no instability in the backward-inclined airfoil blade (Figure 13-12). Efficiencies show slight increases in the order mentioned; however, the level of tolerance for handling abrasive gas streams decreases as one progresses from backward-inclined to airfoil geometries. All of the backward-inclined blade fans exhibit non-overloading horsepower characteristics. Forward-Inclined
Blades
These fans are styled such that the blades lean forward with respect to the direction of rotation. The two typical variations are the fiat plate and curved plate. Figure 13-13 shows geometry and performance of the forward-inclined curved blade style. This
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type of fan is generally limited to high-volume applications and clean services such as residential heating and air conditioning. Efficiencies range from 72 percent to 76 percent; also, a distinct unstable region exists in the performance curve for this fan type. It has an overloading horsepower characteristic. A summary of the relative characteristics of the principal fan types is given in Table 13.1.
Propeller, Axial and Centrifugal Fans
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Relative Characteristics of Principal Fan Types
Characteristics
Backward Vane
Centrifugal Radial Vane
Forward Lean
Axial
Propeller
High
Medium
Low
High
Low
Good High Low
Good Medium High
Fair Low Low
Good High Medium
Good Medium Medium
Medium
Good
Low
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Good
Low
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Efficiency Stability of operation Tip Speed Noise generation Abrasion resistance Tolerance for polymers F A N
Forward-curved blade characteristics
F U N D A M E N T A L S
A brief review of commonly accepted definitions will assist us in highlighting the fundamental principles involved in fan selection and operation. cfm: Volume flow rate expressed in cubic feet per minute at outlet conditions. S C F M : Volume flow rate expressed in standard cubic feet per minute at a density of 0.075 lb/cu ft. VP: Velocity pressure is the pressure caused by the average outlet velocity of the gas stream expressed in inches of water column. It is measured in the outlet duct by the differential reading between an impact tube facing the air flow and a static reading normal to the air flow. SP: Static pressure is that pressure expressed in inches of water column and measured in a manner to exclude the velocity pressure of the gas stream. Static pressure of a fan is the difference in the reading of a Pitot tube placed in the inlet duct and facing in the direction of gas flow and the static reading obtained in the outlet duct. The pressure is usually referenced with respect to 70 ~ and a density of 0.075 lb/cu ft.
525
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Process Plant Machinery TP: Total pressure is the algebraic sum of the static pressures and the velocity pressures and is basically the rise from inlet to outlet measured by two impact tubes. It is expressed in inches of water column, much as static pressure. B H P : The horsepower required to move a specified volume of gas against the indicated static pressure. This is usually referred to air at standard conditions and is sometimes termed air horsepower. Note that for proper sizing of the driver, this should be calculated with respect to total pressure rather than static pressure and should include the ratio of actual density/air density at standard conditions. BHP =
62.3 x SP x cfm 12 x 33,000 x SE
ME" Mechanical efficiency is the ratio of the horsepower delivered to the input horsepower. SE: Static efficiency is the mechanical efficiency multiplied by the ratio of static pressure to total pressure. SE-
ME x (SP)/(TP)
S N D : Static no delivery or zero flow.
WOV: Wide open volume (zero static pressure). OV: Outlet velocity, cfm/outlet area.
VP:
Gas density - lb/cu ft x (average outlet velocity - ft/min) 2 = 1.203 x 10E6
Note for standard conditions, the density of air - 0.075 lb/cu ft. The unit of pressure is expressed in inches of water column where the density of water is specified as 62.3 lb/cu ft. Specific Speed" The speed in RPM at which a fan would operate if reduced proportionally in size to deliver 1 cu ft/min against a static pressure of one inch of water. Specific speed =
RPM x cfm ~ x (gas density/0.075) 0 75 9 Sp0.75
Fans are governed by certain basic laws known as the fan laws. These laws are reasonably accurate and can be stated as follows: For variation in speed" 1. The volume delivered is proportional to the speed. 2. The static pressure across a fan is proportional to the square of the speed. 3. Brake horsepower is proportional to the cube of the speed. For variation in size: 1. Volume is proportional to the cube of the wheel diameter. 2. Static pressure is proportional to the square of the wheel diameter. 3. Brake horsepower is proportional to the fifth power of the wheel diameter. If both fan size and speed are changed, then the combination factors should be calculated. FAN PERFORMANCE AND SYSTEM EFFECTS
The performance of the centrifugal fan is represented by either a multirating capacity table indicating various points on a family of constant speed performance curves
Propeller, Axial and Centrifugal Fans or by a single graph developed by plotting at a fixed speed the static pressure (SP) in inches of water column versus volume flow (cfm) expressed in cubic feet per minute. The fan develops pressure for various volumes of air flow through it and the proposed system. The characteristic fan performance curves shown earlier are constant speed (RPM) performance curves for operation from wide open volume position (zero static pressure, maximum volume) to static-no-delivery position (zero volume). Each performance curve is plotted for a specific RPM and fan size as well as a fixed air or gas density. A simple system might consist of ductwork connected to either the fan inlet, discharge, or both, in addition to elbows, dust collectors, scrubbers, furnaces, etc., through which air may be forced to flow. The prime mover, of course, is the fan or blower in the duct system, which provides the energy to the airstream to overcome the resistance to flow of its various components. Each fan and blower system is unique, i.e., they have a combined system resistance to a specific flow that varies from system to system and is dependent on the resistance to air flow for each individual component of the system. In a fixed system, the volume flow rate (cfm) will have a corresponding pressure loss (system resistance), generally expressed as static pressure, the resistance to the flow varying directly as the square of the increased or decreased flow rate. The equation defining the relationship between flow and pressure drop in a fixed system is as follows: (cfm-final/cfm-original) 2 = (SP-final/SP-original) Figure 13-14 represents a fixed system resistance curve to air flow and is calculated as a parabolic equation relating flow to pressure as shown in the above expression. If the actual resistance curve is substantially different from the calculated one, the performance will, of course, deviate, and in extreme cases, it will risk operation in the unstable zone. Just as different designs of fans have different characteristic performance curves, different types of systems have different characteristic resistance curves. The more complex the system, the more involved the characteristic curve. However, the complex systems can be separated into their component parts, the individual 160 z 150 o_ m14.0 w o 130
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Process Plant Machinery
characteristics of which are known, and a summation of the resistance of several components of a system will give the composite resistance of the overall system. The terminology resistance to flow is generally accepted as equivalent to the phrase pressure loss for the individual system. The determination of pressure loss may be accomplished by use of handbook values for loss through ducting and branch piping as well as by published or computed values for resistance to flow through the individual pieces or components of the system. It is important to remember that the resistance for the system should be expressed in terms of standard air defined as 0.075 lb/cu ft density. The latter resistance may then be related to fan static pressure requirements that are read directly from fan manufacturers' published performance data based on standard air with a density of 0.075 lb/cu ft. The Air Moving and Control Association (AMCA) has long been aware of the problem in applying fans in systems where inlet and outlet conditions vary from test facilities. The association is publishing a manual indicating guidelines or factors to be used in computing fan static pressure requirements when inlet and outlet connections are less than the desired design. Frequently, the system designer will add a factor of safety to his or her calculations of the system resistance in an attempt to adjust for inaccurate evaluations or to compensate for poor inlet or outlet connections. These safety factors rarely compensate for the "system effect" on the fan performance, and it is necessary either to make adjustments in the system by improving inlet and/or outlet conditions or by increasing the fan speed to obtain the desired fan capacity. However, it will be necessary to establish whether or not the increased fan speed is within the structural limits of the equipment before deciding to increase the speed of the fan. Obviously, poor inlet and outlet conditions should be corrected in order to keep operating horsepower and cost to a minimum. A few of the most common conditions that result in deficient performance of the fan include (1) prerotation of the air into the fan inlet due to a vortex condition that occurs when the fan inlet is placed too close to an obstruction or wall, (2) nonuniform flow due to elbows prior to fan inlet, and (3) omission of an outlet duct on the fan discharge. These conditions will definitely alter the characteristic of the fan performance so that its full volume flow potential is not realized. This system effect in the performance of the fan and blower must be considered in the evaluation of the pressure loss for the total system. If the application engineer or system designer has accurately determined the resistance to flow through a given system, the fan engineer will be in a position to properly evalute and select the fan that will develop the required flow and pressure to meet system requirements. When placed into a system, the fan will operate only where its characteristic performance curve intersects the actual system resistance curve. Figure 13-14 shows the fan characteristic performance curve with the calculated system resistance curve superimposed and the resultant point of intersection that should be the design operating point-(55 percent of wide open volume and 107 percent of design pressure). If the system resistance has been accurately determined and the fan properly applied, the design performance will be at the intersection of the two curves. When the pressure losses have not been accurately determined for the system, the intersection of the fan performance curve and the actual system curve will not coincide with the design point. Occasionally, fans operating in a position other than design point may be subject to unstable or pulsating flow, which may damage the fan and the system components. Should the estimate of a system resistance be less than actually required for the flow rate, the resultant flow will be less than required. This is also illustrated in Figure 13-14. When the actual system resistance is less than the calculated value, the flow will be higher than the design flow rate and will require additional horsepower to drive that particular fan.
Propeller, Axial, and Centrifugal Fans Frequently, in the case of direct-connected fans, the design flow and calculated resistance of the system will not fall exactly on the constant speed fan performance curve. In this particular situation, the fan will deliver more or less air, depending on where the system curve intersects the fan curve, and adjustments will be necessary to obtain the required performance. The adjustments may involve the use of a damper to throttle the flow to design requirements or to reduce the speed of the fan until the required flow is achieved. Should the system resistance curve intersect the fan curve at a lower value than design flow, it will be necessary either to reduce the resistance by the use of turning vanes or splitters in the duct elbows, etc., or to speed the fan up to achieve the design flow requirements. This topic will be addressed later. The foregoing comments and figures are based on fan inlet and outlet connections in accordance with good design practice as applied in actual laboratory tests. When fan inlet or outlet conditions vary from good design practice, design performance may not be obtained even though the system resistance has been properly calculated. The system effect on fan performance is dependent on the type of inlet or outlet condition as well as gas properties and velocity. The system effect, therefore, will not be uniform throughout the complete range of the fan performance curve.
Performance Corrections When the fan will be handling a gas at a density other than standard (0.075 lb/cu ft), it may be necessary to use correction factors so that the published tables may be used accurately. Frequently, fan operating specifications are given at temperature, altitude, and/or air density other than standard conditions. The AMCA publishes standards and rating tables for fan equipment. Since fans will operate only where the system resistance curve intersects the fan performance curve, it is important to make certain the fan performance and system resistance are compared on the same basis, i.e., values shown in the same basic units at identical operating conditions.
Example 1" Temperature and Altitude Required: A fan to exhaust 10,000 cfm of 650 ~ air at a static pressure (system resistance) of 4.0 inches at 650 ~ and at an elevation of 2000 feet.
Step 1. The volume of air will be 10,000 cfm at 650 ~ and since no reference is made to weight, this value is correct for selection directly from fan rating tables. Step 2. Static pressure is specified at 650 ~ and 2000 ft elevation. Therefore, it will be necessary to correct this value in order to make use of the performance tables. Step 3. The density ratio obtained from Table 13.2 for gas density at 650~ and 2000 ft elevation show 0.444. To obtain 4.0 inches SP at 650~ we must select a fan at the adjusted static pressure from the standard tables determined as follows" SPc
Pc
SPs
Ps
1 SPs -- SPc x -p'
=4.0x
0.444
= 9.0 in
529
530
Process Plant Machinery TABLE 13.2
Air Density Ratios of Various Altitudes and Air Temperatures*
Altitude in Feet above Sea Level 0
500
1,000
1,500
2,000
3,000
29.92
29.38
28.86
28.33
27.82
26.81
70 100 150 200 250
1.000 0.946 0.869 0.803 0.747
0.981 0.928 0.852 0.788 0.733
0.965 0.913 0.839 0.775 0.721
0.947 0.895 0.823 0.760 0.707
0.930 0.880 0.808 0.747 0.695
300 350 400 450 500 550 600 650 700 750 800 850 900 950 1,000
0.697 0.654 0.616 0.582 0.552 0.525 0.500 0.477 0.457 0.438 0.421 0.404 0.389 0.375 0.363
0.684 0.642 0.604 0.571 0.542 0.515 0.491 0.468 0.448 0.430 0.413 0.396 0.382 0.369 0.356
0.673 0.631 0.594 0.562 0.533 0.507 0.483 0.460 0.441 0.423 0.406 0.390 0.375 0.362 0.350
0.660 0.619 0.583 0.551 0.523 0.497 0.474 0.452 0.433 0.415 0.399 0.383 0.368 0.355 0.344
0.648 0.608 0.573 0.541 0.513 0.488 0.465 0.444 0.425 0.407 0.392 0.376 0.362 0.349 0.338
4,000
5,000
6,000
25.84
24.89
23.98
0.896 0.848 0.779 0.719 0.669
0.864 0.817 0.751 0.694 0.645
0.832 0.787 0.723 0.668 0.622
0.799 0.756 0.694 0.642 0.597
0.625 0.586 0.552 0.521 0.495 0.470 0.448 0.427 0.409 0.392 0.377 0.362 0.349 0.336 0.325
0.602 0.565 0.532 0.503 0.477 0.454 0.432 0.412 0.395 0.378 0.364 0.349 0.336 0.324 0.314
0.580 0.544 0.513 0.484 0.459 0.437 0.416 0.397 0.380 0.364 0.350 0.336 0.324 0.312 0.302
0.557 0.523 0.492 0.465 0.441 0.419 0.400 0.381 0.365 0.350 0.336 0.323 0.311 0.300 0.290
Barometric Pressure in Inches Air Temp. in ~
*Unity basis standard air density of .075 lb. per cu. ft., which at sea level (29.2" barometric pressure) is equivalent to dry air at 70 ~ where SPc = SPs = Pc Ps P'
static pressure at site conditions static pressure at standard conditions(70 ~ and 0.075 lb/cu ft density) -- density of gas at site conditions = density of gas at standard conditions = density ratio Pc/Ps (see Table 13.2)
Select a fan from the manufacturer's published rating tables for 10,000 cfm and static pressure, SP, of 9.0 inches of water column. In selecting the fan from the manufacturer's tables, let us assume we read 1832 R P M and 16.0 BHP. Restating the performance at standard conditions for the selected fan size and type: 10,000 cfm, SP = 9.0 inches, 1832 RPM, 16.0 B H P (standard conditions) . . . .
Step 4. N o w check performance when handling air at the actual operating conditions. Referring to the fan laws for variations in air density: cfm" unchanged -
10,000 cfm
RPM: u n c h a n g e d -
1832 R P M
SPc = SPs • P' = 9.0 • 0.444 - 4.0 in BHP = BHP
x P'
= 16.0 • 0.444 -- 7.1
Propeller, Axial, and Centrifugal Fans where:
BHPc =
brake horsepower at site conditions
B HPs = brake horsepower at standard conditions We have a choice of sizing the motor to drive the fan at start-up (ambient conditions 70 ~ with a 20-HP motor or using a damper and/or a speed control device so as to reduce the motor size to the amount required under operating conditions (7.1 BHP) or a 7.5-HP motor.
Example 2: SCFM- Temperature Correction Select a fan to move 20,000 SCFM at 400 ~ against a static pressure of 2.0 inches at 400 ~ and sea level:
Step 1. Since specifications indicate there will be a requirement for weight flow at 400 ~ it is necessary to correct the SCFM specified to a yield cfm at 400 ~ delivering the same weight. If the SCFM did not indicate a temperature, it would be necessary to either question the specifications or to assume that the SCFM was at 70 ~ Here are the corrections: Step 2.
cfm = (SCFM) Ps/Pc = 20,000 x 1/0.616 = 32,500 cfm
and correction of static pressure SPs = SPc 1/P' = 2.0 x 1/0.616 - 3.25 in The fan unit would have to be selected from the manufacturer's rating tables for 32,500 cfm, 3.25 inches SP. Assuming the selected fan unit requires 700 RPM and 20.0 B HP, at standard conditions, we would determine the horsepower consumed at the actual operating conditions by multiplying by the density ratio factor:
Step 3.
HPc = HPs x P' = 20.0 x 0 . 6 1 6 -
12.32 HP at actual conditions
Example 3: Wheel Diameter Correction We require increased pressure and volume capability, but we are unable to increase the speed. Figure 13-15 illustrates the performance curve of an available 40-inch diameter centrifugal fan operating at 1000 RPM. Table 13.3 shows a functional data form listing operating points taken from the 40-inch diameter fan performance curve of Figure 13-15 and calculated data for plotting the 45 1/8-inch diameter fan performance curve at the s a m e speed. The 40-inch diameter fan curve has been divided into several sections as indicated by a series of conveniently selected points on both the curve and left portion of Table 13.3.
Fan Law Factors
Wheel diameter factor
=
45.125 in 40.000 in
= 1.129
Flow factor
= (1.129)3 _ 1.44
Pressure factor
= (1.129)2 = 1.275
Horsepower factor
= (1.129) 5 = 1.836
531
532
Process Plant Machinery
I.dl:~ ~1./~~1412 13 . . . . . . . . . .~ . l . ~ i 11 ~'
8 7
z-
z-
tu
-' . . . . . . ' ...... ~ . . . . . . . . ~ - - I
i.... ~-. . . . . .
_..23
].
, i I I ' ' -'~ ..--L-. -i---4---~-' -~, ..... , .....
,-7
2 o[
~--!~ ~ i~'!is "
0
!
L-7'P 4
-I- ~-TX~ r~ ;--7
8
16
~
Jr..........
" ~ ' i. . . .j .
I
64
LU
09
-
i -r-"
12
ii "
:, ..../ 20
i
[ i
24
28
rF I.
, ;o
32
(Thousands) FLOW (CFM)
FIGURE 13-15
Constant speed diameter performance curves for different diameter fans
TABLE 13.3
Fan Performance at a Series of Operating Points
Point
CFM
SP
BHP
CFM
SP
BHP
A B C D E F G H I J
0 4,000 6,000 8,000 12,000 14,000 16,000 20,000 22,000 22,500
8.85 9.22 9.21 9.00 7.75 6.70 5.39 2.18 0.40 0
7.50 12.50 15.90 19.50 26.60 30.00 33.00 38.90 41.80 42.50
0 5,760 8,650 11,500 17,300 20,160 23,010 28,800 31,700 32,400
11.28 11.76 11.74 11.48 9.88 8.54 6.87 2.78 0.51 0
13.78 22.95 29.20 35.80 48.80 55.00 60.60 71.50 76.90 78.00
Abbreviations: CFM, cubic feet per minute; SP, static pressure; BHP, brake horsepower. F r o m Table 13.3, we take Point " F " as a sample for applying the calculated fan law factors.
point
cfm
SP
BHP
cfm
SP
BHP
F
14,000
6.7
30
20,160
8.54
55.00
cfm = 1.44 x 14,000 = 20,160 SP = 1.275 x 6.7 = 8.54 BHP = 1.836 x 30 = 55.00 The 45 1/8-inch diameter fan p e r f o r m an ce curve at 1000 R P M has similar characteristics as the 40-inch diameter curve. In rating performance from one size fan to another, the rating process should always be accomplished from a base curve of the smaller fan.
Example 4: Speed Variation A 30-inch radial bladed centrifugal fan handling 3000 cfm of clean air against a system resistance of 15.75 inch water column absorbs 12 B H P when running at
Propeller, Axial, and Centrifugal Fans 1780 RPM. For an uprated volume of 4000 cfm of air, the new speed, BHP and pressure would be calculated as follows" Original Volume
Original speed
New Volume
New Speed
3000
1780
4000
New Speed
New S p e e d -
4000 • 1780
3000 = 2373.3
New Pressure
(2370) 2
15.75
(1780) 2
New Pressure --
(2370) 2 (1780) ~
x 15.75
= 27.92 in Water Column New BHP
(New Speed) 3
Original BHP
(Original Speed) 3
New BHP
(2370) 3
12.0
(1780) 3
New B H P -
--
(2370)3 (1780) 3
x
12.0
= 28.32
CAPACITY CONTROL OF FANS Centrifugal Fans Control of the output capacity of a centrifugal fan is accomplished by adjusting either the fan characteristics or the system characteristics. Fan characteristics may be modified by changing the rotation or speed of the wheel or by altering the rotation or whirl of the inlet air. System characteristics may be altered by either increasing or decreasing resistance somewhere in the system. ,, A number of methods of capacity control are used in current practice. The relative importance assigned to various factors of the particular application governs the choice of method. Basically, the primary design function is to achieve maximum efficiency through control or maximum reduction in operating power to accompany the required reduction in capacity. Selection of a method of capacity control will result from evaluation of such factors as initial cost, operating costs, range of control required, speed of response, simplicity of operation, auxiliary controls, reliability, longevity, maintenance, etc. The effects of the three primary methods of control (inlet vane control, discharge damper control, and speed control) in relation to fan performance are illustrated on the constant speed curves shown in Figures 13-16 through 13-18. The curves are identical in speed, and the performance characteristics are arbitrarily based on the non-overloading airfoil-type fan.
533
534
Process Plant Machinery
AIR
111o 140
OPERATING POINT
13o
i
120 110 100 90 80 70 60 50 40 30 20 10 0
s
RIE31Sl"ANClE ,
!
i
40
~'~
., 60
80
100
120
140
PER(2~ OF MAXIMUMVIX,U~
FIGURE 13-16
Performance curves f o r inlet vane control at constant speed
Inlet Vane Control Inlet vane control consists of a number of vanes or blades located at the inlet to the fan. The varies may be adjusted to various positions so that the entering air is given a change in direction or a spin in the direction of wheel rotation as illustrated in Figure 13-16. The initial spin modifies the basic characteristics of pressure output and power input, resulting in a new and reduced pressure and horsepower characteristic. This is graphically illustrated in the differential between the base curve and percent inlet guide vane closure curves shown in Figure 13-16. Adjustment of the vanes to various positions, thereby changing the extent of the initial spin, gives regulation to any required volumetric flow at only the pressure demanded by the system. A spin in either direction will reduce the capacity of the fan. A spin contrary to wheel rotation will often increase the shaft horsepower. A spin in the same d~rectlon as wheel rotation, as shown, will reduce shaft horsepower. With inlet guide vane control, the fan performance curve is repositioned as the vanes are moved from the wide open to the closed position, with the system resistance curve staying effectively the same. Inlet vane control provides moderate first cost, excellent operating costs, widerange regulation, simplicity in operation and auxiliary controls, low maintenance, and relatively long life. Of course, the vane structure must be sufficiently rigid to withstand significant aerodynamic forces, and the mechanism for moving the vanes should be such that all are positioned equally, with considerable degree of accuracy. Inlet vane control mechanisms are limited to use in relatively low temperature applications (up to 900 ~ and may have limited life in abrasive and/or corrosive atmospheres.
Propeller, Axial, and Centrifugal Fans
It is often believed that inlet guide vane control acts like a multiple-leaf damper. This is quite wrong. A damper is independent of the fan and is not intended to affect fan performance in any way, although if a damper is located near the fan suction, it may do so to the detriment of fan performance. A damper is merely an adjustable resistance in the system intended to waste horsepower. Vane control, on the other hand, acts primarily to modify the fan performance so as to reduce the power required to attain a specific set of conditions. Such resistance as the inlet vane controller interposes is incidental and has no effect on its principal function of fan performance modification.
Control
Damper
A damper consists of one or more pivoted blades acting as a valve in the duct whereby resistance to flow may be varied from wide open to virtually blocked tight. Figure 13-17 illustrates the change in system characteristics achieved with damper control. The same figure also shows two different arrangements of the commonly used multivane damper normally mounted on the fan discharge or in F
~j~
,.r
!'b
i"
! Parallel Vane Operation
160 z 150
' CLO S E ~ I - ~ - - . . .
~
140 o 130
~,.
120 ~ Z -110 n~ Z o_
_
8O
O3
bJ
n~ D_
40 _
~-
20
u~
10
F--
0
."
.,.."
-
... ...:.~ ... 9
.
\
//
.
:" ... .." .... .."
I
.
:
50 3O
/
."
60 _
/~I~.
...
...
." .:
_
,,, 70 -.t n-
:
..
100
',, 9 90
Q_
_
DAMPER POSITION
/
.'"
,/ '/
\
1
/
~ <"
"
\
BHP
\
/ /
/
//
:~'"i
...." 9"
"" /
/
" ~
SYSTEM
..... f
0
'
RESISTANCE
2'0
'
4'o
6'o
~ -
. . . . . 8 '0- - - - ~
100
'. . . .
120
PERCENT OF WIDE OPEN VOLUME
FIGURE 13-17
Effect of damper control on system characteristics
Opposed Vane Operation
535
536
Process Plant Machinery
the duct system downstream of the fan. Each of these arrangements has individual characteristics as to its effect on airflow; but, in general, both provide a variable resistance for the system. The multiple vane damper, often called a louver damper, should be considered as an independent element interposed between the fan and the system. As previously mentioned, the system characteristics and not the fan characteristics are affected by the damper. As shown in Figure 13-17, when the system resistance is altered, the operating point on the fan curve is shifted correspondingly. At reduced volume, the fan develops a higher pressure than the system requires. The curve is shifted to the left as the damper is moved from the wide open to the closed position, with the fan performance curve staying effectively the same. A damper in a duct system controls the flow and dissipates energy by warming the flowing air. A considerable amount of the power input to the fan is converted by warming the air just slightly. Nevertheless, something must be done to control air flow, and the damper is often the most desirable method, even at the expense of wasted horsepower. The damper method of control provides minimal initial cost, moderate operating costs, wide-range regulation, simplicity of operation, and relatively long life. Dampers can be fabricated to withstand extremes in temperature and are capable of satisfactory operation in abrasive and/or corrosive atmospheres. Within design limitations, damper construction can be provided for gas-tight and insulated applications.
Speed Control When a fan is controlled by varying its speed, there is no waste. At each volume, the fan develops a pressure just equal to the system resistance and no more. From the stand-point of the fan, variable speed control is most economically attained. Variable speed control is predicated on the theory that the volume of air flowing is proportional to the fan speed; the pressure developed is proportional to the square of the speed; and the horsepower is proportional to the cube of the speed. Therefore, from a given constant speed performance curve, new curves at other speeds can be easily computed. On reviewing Figure 13-18, it becomes obvious that each fan speed gives a complete and separate performance curve. Since the system resistance has not been physically altered, each reduction in fan speed will provide a new fan performance curve crossing the system resistance curve below the base curve, as illustrated. In like manner, fan speeds in excess of base curve speeds provide performance curves that cross the system resistance curve above the base curve. Many methods of providing variable speed drives are available, each having its own specific desirable and undesirable features. Basically, variable speed control is accomplished by one of two methods: a variable speed prime mover connected directly to the fan shaft; or a constant speed prime mover with interconnecting adjustable means to vary the fan speed. The turbine, direct current (DC) motor and variable frequency alternating current (AC) motor drive fall within the variable speed prime mover classification and may be considered as completely variable. The slip-ring AC motor and the two- and three-speed AC motors also fall within this classification but should be considered as semivariable and stepped. The hydraulic and electrical coupling and the variable pitch drive are interconnecting adjustable mechanisms for variation of fan speed. For small fans, the
Propeller, Axial, and Centrifugal Fans 100 140
,,o,J
1~10
110 -
~
_J~SS SPEED
gO._ IN}
t~os sPz[o
j~
OPEI~TINO POINT
EF'F1CIENCY S "
100
0o /
/
j
/
/
_..j"
40,
p
J
s
10-' 0 -~ -'
eo ~
FIGURE
13-18
IN)
1oo
OF ~
12o
1414)
VtXJAIIs
Variablespeed performance curves
variable pitch sheave and belt drive arrangement is by far the most popular today. Large fans are generally not equipped with belt drives. The ultimate decision as to which control method or methods to use depends on the functional overall system requirements. Comparisons must also take into account differences in first cost, fixed charges, maintenance cost, use factor, environmental conditions, power demands, stability, etc. A combination of two methods of control, such as a two-speed drive with vane control, will frequently provide the most suitable as well as the most economical method of controlling the fan. Also, a variable speed drive and a multivane damper can be combined to permit utilization of a much smaller motor in elevated temperature systems. To ensure proper capacity control of a centrifugal fan, knowledge of the effects of each of the methods described above will be helpful. Figure 13-19 graphically compares the three basic methods.
INLET VANE
DAMPER
SPEED
I ..
"
a.
'il
.:
X
w
of:
.
I I
i
I I
I
I
I
'
iI
v) mr G.
i1' I I
I t
I I
o w ix o -r-
o z
CAPACITY
FIGURE 13-19
CAPACITY
CAPACITY
Comparison of various fan control methods
537
538
ProcessPlant Machinery Axial Fan Control
Although all of the previously mentioned control methods could be used, axial fan control is most often accomplished by adjusting the blade angle. This adjustment can normally be accomplished with the fan either shut down or in operation, and both pneumatic and hydraulic actuating mechanisms are available. The pitch of all blades is changed simultaneously, and the actuating signal can be manual or automatic. Axial fan blade pitch controls have been configured to respond to changes in gas velocity, temperature, pressure, gas composition, and other parameters. Blade angle adjustment allows efficiencies to be optimized under a wide range of pressure and flow conditions. Operation in the stall region of the performance curve can be effectively avoided by this highly flexible control method, which is shown in Figure 13-20. Fan Control Summary
Centrifugal fans can be controlled by the following: 9 Movable inlet guide vanes that vary the pre-rotation angle of the incoming gas. The resulting change produces a shift in the performance curve, which now intersects with the system resistance curve at different pressures and flows, bringing about a different power requirement. 9 Dampers, which throttle the flow either at the fan inlet or outlet and thus cause a shift in the system resistance curve. This shift results in a new point of intersection with the performance curve, and a new pressure-flow relationship is established. Damper control is both less efficient and lower in initial cost than movable inlet guide vane control. 9 Speed changes that produce a parallel shift of the performance curves, which will now intersect the system resistance curve at different pressures and flows. This control method is more energy-efficient than either inlet guide vane or damper controls. Axial fans can be most efficiently controlled by varying the blade pitch angle. Speed, damper, or stationary blade pitch angle controls are feasible, but they are usually less economically attractive. ! 6G
140130 120 110 EFFICIENCY
100 -~
UNES
9080 70 60
~5 4O 30 20 10
,
0 0
120 Pf..qCD(l' Of" WlO( O P ~
FIGURE 13-20
~,.~,~
VOUJI~
Axial flow fan performance with variable pitch blade control
Chapter 14 Reciprocating Compressors Reciprocating compressors are positive-displacement machines that compress and move gases by using a combination of rotational and linear (reciprocating) motion. Reciprocating compressors are used for a variety of industrial services. Their basic function is to raise the pressure level of the gas being compressed. Doing so is desirable for the following reasons: 1. Storage and transmission of energy, e.g., shop air compression. 2. Reduction of volume for the purpose of storage and transport, including gas liquefaction. Typical examples are bottled gases for industrial uses, natural gas storage in underground reservoirs, storage and transport of liquid natural gas (LNG) and liquid petroleum gas (LPG), and compression of natural gas as a fuel for automobiles. 3. Transmission. Examples include pipelines for natural gas, ethylene and other hydrocarbons, ammonia, oxygen, and nitrogen. 4. Process. Chemical reactions that take place at elevated pressures (e.g., ammonia synthesis at 5,000 pounds per square inch absolute [psia]) are illustrative. 5. Energy conversion. Mechanical to thermal energy conversion (refrigeration systems, heat pumps) is an example. Reciprocating compressors have inherent advantages over other compressors in their ability to adapt to a wide range of load, speed, pressure conditions, and pressure ratios (Pdischarge/Psuction). The load may be varied from 0 to 100 percent; the speed may have a wide range, depending on the driver. Pressures may vary from a few inches of mercury absolute suction pressure in the case of a vaccum pump to 50,000 psia or more discharge pressure for process-gas compressors. Pressure ratios may vary from slightly over 1, in the case of natural gas transmission pipeline service, to 8 or more in the case of shop air compressors. Several stages of compression are often used when the overall pressure ratio is high. There are many prime movers suitable for driving reciprocating compressors. These include electric motors, turbines, natural gas engines, diesel engines, and dual-fuel engines. Many electric motors and reciprocating engines have rotative speeds simlar to those of reciprocating compressors and can be directly connected, eliminating speed reduction (or increase) between the prime mover and the compressor. A typical compressor and principal internal construction features are shown in Figures 14-1 and 14-2. * Source: Sulzer-Burckhardt (Winterthur and Basel, Switzerland), unless otherwise noted. Reprinted and adapted by permission. Copyright retained by contributor. 539
540
Process Plant Machinery
FIGURE 14-1 Single-stage dry-running compressor with two horizontally opposed cylinders for ethylene service. This machine is typical of small, skid-mounted equit~ment. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
CYLINDER LINER
~
PISTON ROD CRANK END HEAD \
DRIVE END WIPER SET
1
TOP COVER .~ '~
BREATHER CAP SPACER BAR
/
BEARING CAP
OIL
HANOWHEEL
CYLINDER BODY DISTANCE PIECE HEAD FIXED VOLUME POCKET
VALVE AUXILIARY DRIVE SHAFT
/ PISTON RINGS
' /// CROSSHEAO GUIDE
' V
CRANKSHAFT
DISTANCE PIECE SUPPORT CYLINOER AND PACKING LUBRICATION SYSTEM
BASE
OIL LEVEL SIGHTGLASS
FIGURE 14-2 Major internal parts of a large horizontally opposed reciprocating compressor. (Source: Transamerica De Laval, Trenton, NJ)
Reciprocating Compressors IDEAL COMPRESSOR CYCLE
Positive-displacement compressors are machines in which successive volumes of gas are confined within a closed space and elevated to a higher pressure. The reciprocating compressor is a special type of positive-displacement compressor that elevates the pressure of the trapped gas by decreasing the volume that the trapped gas occupies. A piston moving in a cylinder is used to reduce the volume of the trapped gas (Figure 14-3).
Compression Referring to Figure 14-3, note that the cylinder has filled with gas at the suction pressure with the piston at position a. The piston moves from a toward b, compressing the gas isentropically (with no heat transfer and no turbulent or frictional losses) until the pressure within the cylinder reaches the discharge-line pressure.
Discharge At this point, the discharge valve opens and permits gas to flow from the cylinder into the discharge line until the piston has reached the end of its stroke at point c.
Expansion Since it is impossible to build a compressor with zero clearance volume, gas remains in the cylinder clearance volume at the end of the discharge stroke. The gas remaining expands isentropically to suction pressure as the piston moves from ctod.
b
DS I CHARGE PRESSURE
!~
SUCTION
CLEARANCE VOLUME
VOLUME SWEPT __~ VOLUME [---"1
I I I
DS I CHARGEZ VALVE P,s'roN
SUCTION _~ VALVE
I I I
I i I
I I I
I
'
I I
I I I
I I
I I I
I I L._.J I
0
FIGURE 14-3
Meal compressor cycle. (Source: Transamerica De Laval, Trenton, NJ)
541
542
Process Plant Machinery
Suction When the pressure within the cylinder reaches the suction pressure, the suction valve opens and permits gas at suction pressure to enter as the piston moves from d to a. Since points b and d are determined by the pressures during the cycle, the cycle is described as having a suction stroke (piston moves from c to a) and a discharge stroke (piston moves from a to c).
CLASSIFICATION OF RECIPROCATING COMPRESSORS Reciprocating compressors can be categorized as follows: I. Cylinder lubricated A. Trunk-piston type II. Cylinder nonlubricated A. Dry running piston tings B. Ringless type (labyrinth-piston type) Other criteria for classification include arrangement of cylinder: 9 9 9 9
vertical in-line horizontal (balanced-opposed) V-, L-, or W-arrangement integral with internal combustion engine
and cooling: 9 water-cooled 9 air-cooled Whether to use a compressor with a lubricated or nonlubricated cylinder is, as a rule, dictated by process requirements and the gas to be compressed. Some chemical processes do not permit the use of lubricants in the cylinders, even though the oil can be removed almost entirely by means of separators, filters, and the like, because the slightest traces of lubricant may "poison" the catalyst. In this context, it must be remembered that since most separation methods are based on the difference between the specific weights of the gas and the lubricant, oil separation becomes more difficult with higher gas pressures. Lubricating oil cannot be used in compressor parts making contact with such gases as oxygen and chlorine.
Compressors with Cylinder Lubrication Except for diaphragm compressors, machines with cylinder lubrication are the only reciprocating compressors generally available for pressures in excess of approximately 4300 psia. Small compressors are, as a rule, trunk-piston machines. Their construction details (see Figure 14-4) resemble automotive engines. For power inputs above approximately 100 kilowatts (kW), crosshead-type compressors are used. In contrast to the less costly trunk-piston types, compressors with crossheads permit the use of double-acting pistons. In other words, compression takes place during both forward stroke and reverse stroke of the piston.
Reciprocating Compressors
Mechanical Seal
Oil
FIGURE 14-4 Internal construction of a small trunk-piston reciprocating compressor. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland) Where contamination of the gas by lubricating oil cannot be accepted, compressors with water-lubricated cylinders may be used. The water can then be removed by means of a dryer. These compressors are used mainly for filling high-pressure cylinders, where small volumes of gas have to be compressed to relatively high pressures. Typical usage ranges are suction volumes from 12 to 160 actual cubic feet per minute (acfm), and discharge pressures from 2000 to 3600 psia.
Compressors with Nonlubricated Cylinders Compressors with nonlubricated cylinders are now used wherever possible. The occasional claim that if process permits, it is better to have a lubricated compressor, is not always borne out by the latest experience. Although tremendous progress has been made on nonlubricated compressors, where there is a choice between the oil-lubricated and nonlubricated type, the following factors must be considered: 9 Nonlubricated compressors cost more than oil-lubricated machines. 9 Nonlubricated compressors require more power. 9 With the possible exception of frictionless compressors, nonlubricated compressors require more maintenance. While there may be no ready answer to the question of lubricated versus nonlubricated compressor selection, the user should have little difficulty choosing the most applicable machine. A detailed investigation of service experience and maintenance frequency in a given service will be of great help to the specifying engineer.
543
544
Process Plant Machinery
COMPRESSORARRANGEMENTOVERVIEW As mentioned earlier, reciprocating compressors can also be categorized according to the type and arrangement of their cylinders. Small compressors usually have single-acting, trunk-piston cylinders (see Figure 14-5). Most of the higher horsepower units are double-acting, crosshead units (see Figure 14-6). Double-acting compressors have pistons that compress gas on both ends so that one end is on its suction stroke while the other end is on its discharge stroke. The force resulting from the pressure and area differential across the piston is referred to as the piston-rod load. Reciprocating compressors are rated in terms of their rodload capability rather than by horsepower. Rod-load ratings range up to 175,000 lb. Higher horsepower compressors are built with a basic frame, and a wide range of cylinders that are interchangeable on the frame. The cylinders range from smalldiameter high-pressure cylinders to large-diameter low-pressure cylinders. A line of cylinders is usually designed so that each cylinder matches the rod-load capability of the compressor frame at the maximum working pressure of the cylinder and the expected pressure ratio of the applications for which the cylinder is intended. Double-acting compressors with cylinders of the type described above are designed as vertical in-line, vertical or horizontal straight-line, horizontal balancedopposed, or "V"-or "L"-machines. Other cylinder arrangements are less common. Multi-cylinder compressors with horizontal cylinders are very often designed as balanced-opposed type (see Figure 14-7). The balanced-opposed frame is characterized by adjacent pairs of crank throws 180 ~ out of phase and separated by crank webs only. With this configuration, the inertia forces are balanced if the reciprocating weights of opposing throws are balanced. The balanced-opposed design is a separable frame, thus the basic compressor can be driven by any number of prime movers, including diesel, natural gas and dual-fuel engines, gas and steam turbines, and electric motors. Compressors driven by an internal combustion engine (e.g., diesel engine, gas engine) or by a steam engine can be designed as integral engine-compressor units, (Figure 14-8; see also Chapter 4). The integral compressor has compressor cylinders and power cylinders mounted on the same frame and driven by the same crankshaft. CONNECTINGROD
...r _ r CRANK
~-CYLINDER VOLUME
FIGURE 14-5 Schematic view of a single-acting, Transamerica De Laval, Trenton, NJ)
PISTON ROD7 ROD PACKINGz// CONNECTING--.7 ROD
PISTON 9
/
trunk-piston machine.
Ld2_c __r /
/- CROSSHEAD / z-CRANK-ENDCYLINDERVOLUME Z-HEAD-END CYLINDERVOLUME
(Source:
F-CRANK -"""\
k3~ ' CROSSHEAD GUIDE
FIGURE 14-6 Schematic view of a double-acting compressor design. (Source: Transamerica De Laval, Trenton, NJ)
Reciprocating Compressors
l
PRIME MOVER I
.FLEXIBLECOUPLING
FIGURE 14-7 Schematic view of a balanced-opposed compressor. (Source: Transamerica De Laval, Trenton, N J) POWERCYLINDER
FIGURE 14-8 Schematic view of an integral-type gas engine reciprocating compressor. (Source: Transamerica De Laval, Trenton, NJ)
Some integrals have in-line power cylinders mounted vertically and compressor cylinders extending from one or both sides in the horizontal plane. Others are V engines with compressor cylinders extending from one or both sides. For a more detailed description of gas engines, refer to Chapter 4.
Cylinder Heads and Valves The design of the cylinder heads and valves is dictated by the fact that these parts have to withstand, for years, a pressure that fluctuates considerably, e.g., between 900 and 2500 bar at a frequency of 3 to 4 Hz. Modem methods of investigation led to designs such as that illustrated in Figure 14-9, where inadmissible changes of combined strains could be kept within tolerable limits. With large cylinders, combined suction discharge valves as illustrated in Figure 14-9 are used for very high pressures. This valve is fitted with multiple suction and delivery poppets in order to reduce the moving masses. It is interesting to note that poppet valves, which were used in the last century, have experienced a comeback in compressors that are at the other end of a more than 130-year-old development of reciprocating compressor design.
Capacity Range of Reciprocating Compressors versus Other Compressors The pressure-flow relationship of reciprocating compressors relative to other compressor types is illustrated in Figure 14-10. Application ranges are identified as follows:
545
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Process Plant Machinery
. . . . . . . . . . . . . .
-
-
FIGURE 14-9 Combination suction and discharge poppet valve for a secondary compressor. (Source: Sulzer-Burckhardt, Winterthur and Basel Switzerland) psi
bar 100oo
-m= m
I00(~o
A+G
'X It
A2 =: i
I
10000
)L 1000
I 100
I l0 I I
100 l 10
1000 I 100
10000 1 1000
100000 I 10000
1000000
1 100000
m31hr
cfm
FIGURE 14-10 Simplified capacity versus pressure diagram of the most widely used types of compressors in the CPL Power consumption is approximate and is based on a suction pressure of 1 bar (14.5 psia). (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
9 A l: Reciprocating compressors with lubricated and nonlubricated cylinders 9 A2: Reciprocating compressors for high and very high pressures with lubricated cylinders 9 B: Helical- or spiral-lobe compressors (rotary screw compressors) with dry or oil-flooded rotors 9 C: Liquid ring compressors (also used as vacuum pumps) 9 D : Two-impeller straight-lobe rotary compressors, oil-free (also used as vacuum pumps) 9 E: Centrifugal turbocompressors
Reciprocating Compressors F: Axial turbocompressors 9 G: Diaphragm compressors
9
The most frequently used combinations of two different compressor types are identified in three fields: 9 A-t-G: Oil-free reciprocating compressor followed by a diaphragm compressor 9 E-k-A: Centrifugal turbocompressor followed by an oil-free reciprocating compressor 9 F+E: Axial turbocompressor followed by a centrifugal turbocompressor The above list is never quite complete in either design or application, since compressor innovations enter the marketplace rather frequently. For moderate pressures, rotary screw compressors (Chapter 15) compete with reciprocating compressors. Oil-free screw compressors are built for compression ratios up to 4.5 per stage. Two-stage units with atmospheric intake are available for discharge pressures up to approximately 195 psia and three-stage units are available with pressures up to approximately 300 psia. The oil-flooded screw compressor is available for a compression ratio up to 13 in single-stage and up to 21 in twostage configurations. Booster screw compressors with elevated suction pressure are available for a variety of conditions, with pressure limits in the neighborhood of 260 psia unless special designs and materials are used. Rotor deflection limits the maximum attainable pressure difference. Unlike reciprocating compressors, rotary screw compressors have a fixed built-in pressure ratio. This means that in some cases, higher compression than required by the process occurs inside the compressor, thus causing higher discharge temperatures. With fluctuating intake and discharge pressures frequently occurring in chemical process cycles, it is not always possible to operate a compressor at its builtin compression ratio. Low efficiency and high power consumption may be the result. Sliding-vane Compressors are still widely used. Although they have limited discharge pressure capability (two-stage units are available with pressures up to 130 psia), sliding vane compressors cover roughly the same capacity range as the screw compressors. Since they are not oil-free, they have lost ground to dry-running screw compressors and are not covered in this text. There are no clearly defined limits for the application in series of two or more compressor types. However, in series applications, many factors and alternatives influencing compressor selection must be considered (see later). Reciprocating compressors cover the range from the smallest capacity requirements through 10,000 m3/hour (6000 cfm) and more. The reason why larger suction flows are normally handled by turbocompressors is the relatively high price of very large reciprocating compressors in comparison with turbocompressors. In the chemical processing industry (CPI), reciprocating compressors are typically found with power inputs up to 2000 kW, although standard frames are available with power inputs exceeding 10,000 kW. Most processes in the CPI require discharge pressures below 400 bar (5800 psia). The polymerization of ethylene to produce low-density polyethylene (LDPE) is an exception. This process requires discharge pressures between 1500 and 3500 bar (21,750 to 50,750 psia) and employs so-called secondary booster compressors with suction pressures between 100 and 300 bar (1450 to 4350 psia). Not only are these machines at the upper end of the pressure scale for chemical reactions, but they are also the most powerful reciprocating compressors built thus far. The power consumption of one existing unit is 15,200 kW! Reciprocating compressors for discharge pressures of roughly 8000 bar (116,000 psia) have been used exclusively in high-pressure research (see heading "High Pressure Compressors," later).
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Process Plant Machinery
Oil-free compressors can be built up to about 300 bar (4350 psia) discharge pressure (see heading "Labyrinth-Piston Compressors"). Beyond this limit, high-pressure reciprocating compressors have lubricated cylinders. For oil-free compression to even higher discharge pressures, an oil-free reciprocating compressor followed by a diaphragm compressor may represent the best solution. The diaphragm compressor is discussed at the end of this chapter.
Trunk-Piston Compressors As was mentioned earlier, in trunk-piston machines, the pistons are also the crossheads. Since there are no piston rods, this design is only suited for singleacting pistons, where the gas is compressed on the cylinder head side of the piston only. These machines are used for relatively small power inputs, and ratings rarely exceed 200 kW. A variety of cylinder arrangements is possible; two of these are represented in Figures 14-11 and 14-12. Most small compressors are built along this geometric arrangement. This design is highly suited for air and other noncorrosive gases compatible with the crankcase lubricating oil. Indeed, the gas blowing by the piston tings makes its way under the piston and from there into the crankcase. To prevent a gradual pressurization of the crankcase, the latter has to be connected to a system of adequately low pressure, in most cases the suction line, allowing a recycling of this blow-by gas without letting it escape to the atmosphere. In this case, it is necessary to design the crankcase as a gas-tight pressure vessel with mechanical crankshaft seals. One such machine is illustrated in Figure 14-4. The gas intake and outlet connections of any given cylinder are generally concentrated in one location. This allows the use of combined suction and discharge valves. As shown in Figure 14-13, these concentrically designed valves incorporate circular spring-loaded plates. The design concept represented by Figures 14-14 and 14-15 allows compressors to be built in up to five stages for discharge pressures up to 5000 psia and suction capacities between 3 and 280 acfm. Up to about 80 kW, these units are air-cooled, and above this limit, they are more often water-cooled. Compressors of this size are supplied as compact packaged units on a sturdy steel frame. The assembly essentially includes the compressor, motor, V-belt drive,
FIGURE 14-11 Stepped cylinder arrangement, air-cooled. (Source: Sulzer-Burckhardt,
Winterthur and Basel Switzerland)
Reciprocating Compressors 549
FIGURE 14-12 Stepped cylinder arrangement, water-cooled. (Source: Sulzer-Burckhardt, Winterthur and Basel Switzerland)
FIGURE 14-13 Typical reciprocating compressor valve combining suction and discharge. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland) gas coolers, moisture separators, condensate receiving tank, filter, the piping for gas and, where required, for cooling water. The baseplate typically rests on vibration damping elements. The crankshaft is fitted with counterweights to reduce free unbalance forces to an acceptable minimum. Units of this kind are ready for use when placed on a substantial floor or simple foundation and connected to power, suction, and discharge gas lines and, if water cooled, to water lines.
Crosshead-TypeCompressorsWith Cylinder Lubrication Heavy-duty machines are of the crosshead type with entirely separate and well-controlled cylinder lubrication, water-cooled cylinders, and a relatively low operating speed. Permanently mounted on a good foundation or isolated support system, they can be operated at full load for years with minimum attention.
550
Process Plant Machinery
FIGURE 14-14 Air-cooled trunk-piston compressor for 5000psia service. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
FIGURE 14-15 Water-cooled trunk-piston compressor for 4300psia service. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
While the crankcase is lubricated in the usual manner by a forced-feed lubricating pump driven by the crankshaft or by a separate electric motor, a high-pressure lubricating oil pump, also driven by the crankshaft or by a separate electric motor, feeds the cylinders and piston-rod packings. Each high-pressure lubrication element has an adjustable metering device and visual flow control.
Reciprocating Compressors
551
These machines have a separate crosshead, as shown in Figures 14-16 to 14-18, with a piston-rod connecting crosshead and piston. Crosshead-type machines are the most widely used reciprocating compressors in the CPI. Their power input ranges from around 100 to approximately 10,000 kW. As was mentioned earlier, only secondary compressors for the very high pressures needed in the production of LDPE have a higher power input, i.e., up to 15,000 kW. Heavy-duty reciprocating compressors typically use double-acting pistons and perhaps even step pistons for two, three, or even four compression stages on the same piston rod. Thus the gas forces can be equalized very well. The crankcase side of the piston is sealed off by means of piston-rod packing. With some services, a short single-compartment distance piece is sufficient (Figure 14-16), while for flammable, hazardous, or toxic gases, a two-compartment distance piece as shown in Figure 14-17 is required. Since process compressors are often custom-designed for a given duty, standardization, as a rule, is limited to the frame and running gear. There are two basic factors that guide the designer: 1. The maximum power at a given speed that can be transmitted through the shaft and running gear to the pistons. 2. The load imposed on the piston rod, the so-called pin load.
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FIGURE 14-16 Vertical, four-stage crosshead-type compressor with cylinder lubrication and singlecompartment distance piece f o r nontoxic and nonflammable gases. 1 - C r a n k c a s e ; 2 - f r a m e ; 3 - crankshaft; 4 - bearing; 5 - connecting rod; 6 - crosshead; 7 - cover; 8 - distance piece; 9 - purge chamber; 10 - lubricating pump f o r crankcase; 11 - cylinder lubricator; 12 - cylinder; 13 - cylinder liner; 14 - piston; 1 5 - piston-rod packing; 1 6 - valves; 1 7 - c a p a c i t y control. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
552
Process Plant Machinery
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FIGURE 14-17 Vertical, three-stage crosshead-type compressor with cylinder lubrication and two-compartment distance piece (8 + 9 ) f o r toxic and~or flammable gases. 1 - Crankcase; 2 - f r a m e ; 3 - crankshaft; 4 - bearing; 5 - connecting rod; 6-crosshead; 7 - cover; 8 - distance piece; 9 - purge chamber; 10 - lubricating pump for crankcase; l 1 - c y l i n d e r lubricator; 12-cylinder; 13-cylinder liner; 14-piston; 15-piston-rod packing; 1 6 - valves. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
There are other factors involved, for example, the number of cranks, the distance between them, and the piston stroke. These factors represent the working limits for design purposes. Other limits are set by the cylinders, for example, the maximum allowable pressure. A typical example of a crosshead machine is represented in Figure 14-18. To date, much has been published about the advantages of horizontal versus vertical designs and vice versa. Both designs have their merits, and in some cases, the selection is dictated by the kind of service intended (Figures 14-19 and 14-20). It is claimed that a compressor with horizontal opposed cylinders is cheaper to build than one of the vertical type and that it is better balanced. The pipe connections between cylinders and coolers, being closer to floor level, are simpler and easier to support. This type of compressor can be built with an even number of cylinders only, which means that the very popular three-cylinder configuration is not
Reciprocating Compressors
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16
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FIGURE 14-18 Horizontal, two-stage crosshead-type compressor with cylinder lubrication. In addition to unloaders on all suction valves (17), the second stage cylinder on the right side is fitted with h a n d - o p e r a t e d clearance p o c k e t control (17). 1 - Crankcase; 2 - f r a m e ; 3 - crankshaft; 4 - bearing; 5 - connecting rod; 6 - crosshead; 7 - cover; 8 - packing cartridge; 9 - purge chamber; 10 - lubricating p u m p f o r crankcase; 11 - cylinder lubricator; 12 - cylinder; 13 - cylinder liner; 14 - piston; 15 - piston-rod packing; 16 - valves; 1 7 - capacity control. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
possible. A prominent industry specification, API Standard 618, states that horizontal cylinders are required for handling saturated gases or for gases carrying injected flushing liquids. It also requires that horizontal cylinders have bottom discharge connections. To prevent liquid collecting in the cylinder, arrangements are normally made for delivery valve ports to be on the bottom face of the cylinders, so that liquid is automatically drained by the flow of gas through the cylinder. Horizontal compressors often experience unequal wear on crossheads and their guides, on piston-rod packings, and on piston tings. This may lead to uneven cylinder wear due to force of gravity action in the downward direction. On the other hand, the v e r t i c a l machine is more compact, requires less floor space, and has shorter interconnecting pipes. Also, with a direct-lifting crane over the machine, assembly and dismantling of cylinders and pistons is facilitated. Figure 14-21 demonstrates that a very compact and space-saving arrangement of compressor, gas coolers, and piping is feasible with horizontal compressors as well. Such an arrangement, however, raises the question of accessibility. With this compressor, the remaining free forces are taken up by the springs on which the unit is mounted. This ensures that only a fraction of the forces is transmitted to the substructure; there is no need for a heavy foundation. Vertical compressors in atmospheric suction pressure services might have low-pressure cylinders with extremely large diameters. This might make it necessary to lengthen the distance between the cylinder centerlines, and the compressor frame
554
Process Plant Machinery
FIGURE 14-19 Compact skid-mounted compressor package with a five-stage compressor with four horizontal opposed cylinders. Compression is from atmospheric to 2013 psia pressure, capacity is 1820 acfm, and power input is 690 kW at 705 RPM. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
FIGURE 14-20 Vertical, three-stage compressor with a two-crank frame for sour gas re-injection in an oil field. Compression is from 70 to 2712 psia; the unit is designed in accordance with API Standard 618. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
Reciprocating Compressors
FIGURE 14-21 Three-stage compressor for service on an off-shore platform. Gas compressed is a hydrocarbon mixture containing hydrogen sulfide. Main data: 1680 acfm, 280 psia, 390 kW. Interstage cooling is by gas-to-air coolers. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland) could become too heavy. For atmospheric suction, a horizontal design may thus have advantages over a vertical design, particularly if only two cylinders are required.
Compressors With Water-Lubricated Cylinders Until compressors with dry-running piston rings, labyrinth-piston compressors, and diaphragm compressors became available, compressors with water-lubricated cylinders were the only oil-free reciprocating compressors on the market. Before 1940, virtually all oxygen compressors were water-lubricated. Generally speaking, water-lubricated machines are required whenever oil lubrication of the cylinders is not permitted and other oil-free compressors are not economically feasible. This is usually the case if a relatively small quantity of gas has to be compressed to a high pressure.
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Process Plant M a c h i n e r y
Three-stage compressors, as shown in Figures 14-22 and 14-23, are used for suction capacities to roughly 160 acfm and a maximum discharge pressure of about 3600 psia. While the beatings of the running gear are lubricated in the conventional manner by means of a crankshaft-driven gear-type oil pump, the cylinders are lubricated with demineralized water. The water passes into the suction pipe at about atmospheric pressure and is carried through all three stages. It is recovered in a separator after the third stage, from where it is returned to the lubrication water tank. Since the latter is located above the compressor, the water is fed to the suction pipe by gravity. For compressors with elevated suction pressure, the water lubrication system can be pressurized. The lubrication water tank is equipped with a cloth filter, which prevents solid particles from entering the compressor. Furthermore, it is fitted with a level gauge
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FIGURE 14-22 Section views o f a c o m p r e s s o r with water-lubricated cylinders. 1 - Crankcase; 2 - f r a m e ; 3 - crankshaft; 4 - bearing; 5 - connecting rod; 6 - crosshead; 7 - ratchet drive f o r rotary motion o f piston; 8 - cover; 9 - crankcase lubrication p u m p ; 10 - piston; 11 - cylinder; 12 - cylinder liner; 13 - valves; 14 - gas cooler; 1 5 - c o olin g - w a t e r tank. (Source: Sulzer-Burckhardt, Winterthur a n d Basel, Switzerland)
Reciprocating Compressors
FIGURE 14-23 Three-stage compressor with water-lubricated cylinders. Note the lubrication-water tank on the wall above the compressor. Main data: Capacity at maximum speed, 270 R P M - 160 acfm; maximum discharge pressure - 3640 psia; piston stroke- 250 mm. (Source: Sulzer-Burckhardt, Winterthur and Basel Switzerland) and a level switch and has connections for the feed, condensate return, and vent lines. A solenoid valve in the feed line is electrically interlocked with the driving motor, whereby the lubrication water circuit is automatically opened or shut off. A flow indicator with a needle valve allows the flow of lubrication water to be adjusted and regulated. Should the quantity of lubrication water not be sufficient, the drive motor is immediately switched off by means of a monitoring contact. A ratched mechanism fitted to each crosshead rotates the piston stepwise at each stroke, thus preventing scoring of the mirror-polished running surfaces of the cylinder liners by the piston packing tings. The latter are made of leather in the first- and of fiber in the second- and third-stage cylinder. All cylinders and associated coil-type gas coolers are submerged in a common cooling water tank. This efficient cooling and water lubrication of the piston packings allows compression to take place without any significant temperature rise. These compressors are extremely well suited for compressing oxygen, hydrogen, helium, nitrogen, air, nitrous oxide gas, etc. The moisture content in the gas leaving a water-lubricated compressor is no higher than in an oil-lubricated or nonlubricated compressor handling humid gas, since the gas is saturated at the outlet of the after-cooler, whether the compressor is water-lubricated or not. If moisture in the gas is not permitted, the water can easily be removed by means of an absorption or refrigeration dryer.
557
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Process Plant Machinery
Water-lubricated compressors are low-speed machines. Thanks to their efficient cooling, they allow a relatively high-pressure ratio per stage even in oxygen service. Like diaphragm compressors, they are very well suited for compressing small quantities of very light gases such as hydrogen and helium to relatively high pressures. They are, however, less costly than diaphragm compressors. Both hydrogen and helium are particularly "slippery" gases and have a pronounced tendency to slip past the piston tings during compression in a dry-running ring-type compressor and past the labyrinths in a labyrinth-piston-type compressor. This causes low efficiency and high discharge temperatures. The gas handled by a water-lubricated compressor must be free of dust. Dust causes rapid wear of the cylinder liners and the piston packing rings.
Acetylene Compressors Among the gases for which specially designed compressors with lubricated cylinders are required, acetylene is a typical example. Acetylene is produced from calcium carbide and water. Under certain temperatures and pressures and also through shock, the gas can decompose violently into its elements hydrogen and carbon. If decomposition occurs in a pressure vessel, the gas can reach a pressure about eleven times that which existed prior to decomposition. The higher the pressure, the smaller the initial force required to cause an explosion. Furthermore, acetylene is highly flammable. This dangerous propensity is countered by shipping acetylene dissolved in acetone. One volume of acetone dissolves 300 volumes of acetylene at 175 psia. Acetylene cylinders are partly filled with a porous material that holds the acetone, thus preventing instability of the gas. Full cylinder pressure is 250 psia at 70 ~ Most acetylene compressors are used for filling cylinders. In these machines, the explosion hazard must be minimized by other measures. In Europe and other countries, the design of equipment handling acetylene is governed by the Technical Rules for Acetylene Plants and Calcium Carbide Storage (TRAC) rules. These rules, however, do not apply to acetylene compressors used in a chemical process for which other safety rules exist. Finally, the reader may be interested in reviewing the step-piston construction employed for the acetylene compressor shown in Figure 14-24.
Compressors for the Production of Low-Density Polyethylene A characteristic feature of the high-pressure ethylene polymerization process is that a very large difference in pressure is necessary between the inlet gas entering the reactor and the outlet of the recycled gas. The recirculators, generally called secondary compressors (Figures 14-25 and 14-26), work between two limits, i.e., 100 to 300 bar on the suction side and 1500 to 3500 bar on the delivery side, for most of the existing processes. As the coefficient of reaction lies between 16 and 30 percent, the secondary compressors have to handle three to six times the quantity of ethylene that is polymerized, being thus by far the most powerful machines in the production loop. Their unit capacities, which when the industrial expansion first began were of 4 to 5 ton~our, now lie between 15 and 118 ton/hour, and their power requirement per unit has been increased from 400 kW up to some 15,000 kW. As the whole operating range of these secondary compressors is placed well above the critical point of ethylene, the thermodynamic behavior of the fluid lies somewhere between that of a gas and that of a liquid. This peculiar condition has
Reciprocating Compressors
FIGURE 14-24
Section view of acetylene compressor. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
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Process Plant Machinery
FIGURE 14-25 Large secondary compressor (hyper compressor) (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland) two main effects. The first is a very small reduction of the specific volume with increasing pressure; for instance, at a temperature of 25 ~ the specific volume is 3 dm3/kg at 100 bar, 2 dm3/kg at 700 bar, and 1.5 dm3/kg at 4500 bar. The second effect is a very moderate rise of adiabatic temperature with increasing pressure; for instance, with suction conditions of 200 bar and 20 ~ the delivery temperature will reach only 100 ~ at 2000 bar. These particular thermodynamic conditions greatly influence the design of high-pressure ethylene compressors. Compared with the classical reciprocating compressor, the compression ratio is of little practical significance, the important factor being the final compression temperature, which should not exceed 80 to 120 ~ depending on process, gas purity, catalyst, etc., in order to avoid premature polymerization. The influence of the cylinder head clearance on the volumetric efficiency is slight because of the small reduction of specific volume, and very high compression ratios are therefore possible with quite acceptable efficiency. In addition, the stability of intermediate pressures depends chiefly on the accuracy of temperatures; for instance, in the case of two-stage compression from a suction pressure of 200 bar to a delivery pressure of 2500 bar, a drop in the first-stage suction temperature from 40 to 20 ~ will cause the interstage pressure to rise from 1000 to almost 1600 bar. For these reasons, a secondary compressor is required, which has only one or two stages, despite the very large pressure differences involved. However, this compels the designer to face extremely high mechanical strains due to the high amplitude of pressure fluctuation in the cylinders. Finally, an additional and sometimes disturbing feature of ethylene must be mentioned. If the gas reaches a very high pressure and a high temperature simultaneously (which can easily occur in a
Reciprocating Compressors
561
FIGURE 14-26 Longitudinal sections through secondary compressor. The design of this crank mechanism is based on a moving frame surrounding the crankshaft and connecting the crosshead to the piston rods on either side. Thus the crosshead pin is loaded with the difference between the piston-rod loads only. (Source: Sulzer-Burckhardt, Winterthur and Basel Switzerland)
blocked delivery port owing to the very low compressibility), it will decompose into carbon black and hydrogen in an exothermic reaction of explosive character.
Cylinders and Piston Seals Sealing of the high-pressure compression chamber is a major problem, and this could be solved by avoiding friction between moving and stationary parts. This has been realized for laboratory equipment and small-scale pilot plants by the use of either metallic diaphragms or mercury-based sealants in U-tubes, and such arrangements are still in use for research purposes. In addition, they have the advantage of avoiding any contamination of the compressed gas by lubricants. Unfortunately, chiefly for
562
ProcessPlant Machinery economic reasons, they proved to be impracticable for industrial compressors, at least in the present state of techniques. Thus, as labyrinth seals are out of the question for very high pressures, friction seals have to be accepted; in fact, two solutions are currently u s e d - moving and stationary seals. At first, metallic piston rings were the only sort of moving seals used in the large reciprocating type of compressor for very high pressures. They were generally made in three pieces; two sealing tings, each covering the slots of the other, and an expander ring behind both of them, which also sealed the gaps in the radial direction. The materials used were special grade cast iron, bronze, or a combination of both, with cast iron or steel for the expander. The piston, of built-up design, comprised a series of supporting and intermediate tings with a guide ring on top of them and a through-going bolt (two different designs are shown in Figure 14-27). All parts of the piston were made of high tensile steel and particular care had to be given to the design and to the stress calculation of the central bolt, which was subjected to severe stress fluctuations. The use of piston rings allowed for a simple cylinder design, the main part of which was a liner that had been thermally shrunk to withstand the high variations of the internal pressure (Figure 14-28). The inner sleeve, which had previously been made of nitrided steel, was later produced of massive sintered material like tungsten carbide. The use of this expensive material was justified by two beneficial qualities: it possesses an extremely hard surface and has a high modulus of elasticity. The first considerably improves the conditions of friction and greatly reduces the danger of _+___ m
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FIGURE 14-27 Built-up piston designs for secondary compressors. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
Reciprocating Compressors 563
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FIGURE 14-28 Medium pressure cylinder for secondary compressor (hyper compressor). (Source: Sulzer-Burckhardt, Winterthur and Basel Switzerland) seizure. Owing to the high modulus of elasticity, the amplitude of the "breathing" movement during internal pressure fluctuations is much smaller than with steel, and thus the stress variations in the expanded outer sleeves is appreciably reduced. However, as these sintered materials have a very poor tensile strength, care must be taken to ensure that the inner sleeve is always under compression, even if the temperature increases. This is the main purpose of the external cooling of the liner and not, as is usual, to dissipate the heat of compression. At present, plungers with packings of the self-adjusting type, as shown in Figure 14-29, are most widely used. The packings are usually assembled in pairs, the actual sealing ring tangentially split into three or six pieces being covered by a three-piece radially cut section. Both are usually made of bronze, kept closed by surrounding garter springs, and held in place by locating and supporting steel plates. The sealing rings are pressed against the plunger by gas pressure, which corresponds to the pressure difference across the sealing elements. The supporting steel plates must also be thermally shrunk to resist the high variations in internal pressure. Unfortunately, the use of sintered hard materials is restricted by the fact that the supporting plates are subjected, in the axial direction, to heavy bending and shearing forces, which these materials generally cannot withstand.
FIGURE 14-29 Modem design cylinder with self-adjusting packing for secondary compressor. (Source: Sulzer-Burckhardt, Winterthur and Basel Switzerland)
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Process Plant Machinery
The plungers for medium pressures are made of steel and plated with tungsten carbide. For very high pressures, the use of solid bars of hard metal is the best wear-resistant solution for plungers. The disadvantage of the packed plunger design lies in the much larger joint diameters of the static cylinder parts, which require two to three times higher closing forces than the piston ring design. Large cylinders, such as the one shown in Figure 14-29, need a pretensioning of the cylinder bolts to about ten times the maximum plunger load. This ratio is higher for smaller cylinders. For piston tings and packed plungers, the optimum number of sealing elements appears to be four or five. In both solutions, it is essential that the piston be accurately centered if the seals are to be effective; this is the reason for the guiding ring within or near the cylinder and for the additional guide at the connection between the piston and driving rod. At the base of the cylinder an additional low-pressure gland allows gas leaks to be collected and the plunger to be flushed and cooled. Other separate glands positioned on the rod connecting the piston to the drive (Figure 14-28) prevent the cylinder lubricant from mixing with the crankcase oil, and as the intermediate space is open to atmosphere, it is impossible for gas to enter the working parts. From the point of view of design and maintenance, piston tings would appear to be the most adequate solution, and they are currently used for pressures up to 2000 bar or in some circumstances up to 3000 bar. The choice between them and the packed plungers depends largely on the process and type of lubricant used. One difficulty is that normal mineral oils are dissolved by ethylene under high pressure to such an extent that they no longer have any lubricating power. The glycerine used in earlier machines has been widely replaced by paraffin oil, either pure or with wax additives, which is much less diluted by the gas than other mineral oils. However, it is a rather poor lubricant and is inferior to the various types of new synthetic lubricants, which are generally based on hydrocarbons. Since part of the LDPE is used for the manufacture of plastic films and bags for the foodstuff industry, the ethylene to be compressed must not be contaminated by a toxic lubricant. At present, synthetic oils without toxic ingredients are used almost exclusively. The basic difference between piston rings and plunger packings is that the latter may be lubricated by direct injection, while piston rings are lubricated indirectly. This may be an advantage, since the low polymers carried by the return gas back from the reactor are reasonably good lubricants. However, too large an amount of low polymers causes the tings to stick in their grooves, and some kinds of catalyst carriers also brought back by the gas are excellent solvents for lubricants. Thus, the most convenient solution has to be selected for each specific case. In general, for higher delivery pressures (above 2000 to 2500 bar), better results are obtained with the use of packed plungers. COMPRESSING DIFFICULT GASES
Difficult gases are those that require special attention by the compressor designer due to their specific properties. Very often standard compressors cannot be used for this service and the selection of the proper compressor type must be made very carefully in order to avoid misapplication and other costly consequences. The following- incomplete- list demonstrates for which phenomena the compressor designer must be prepared when difficult gases are to be compressed. These gases may 9 assume the liquid state in intercoolers 9 decompose at high temperature
Reciprocating Compressors 565 9 polymerize 9 dimerize 9 produce the chemical reaction for which they are used already in the compressor cylinder 9 form explosive acetylides in the presence of materials of construction promoting such a reaction 9 attack the materials of construction 9 cause hydrogen embrittlement 9 dissolve materials used for gaskets 9 dilute the lubricating oil and cause oil foam 9 form an explosive and self-igniting mixture with air 9 cause a fire (oxygen) The compressor designer is normally confronted with two or even more of the above phenomena requiting precautions that very often contradict each other. This may explain why difficult gases are also referred to as nasty or unpleasant gases. The answer to all these problems is a purpose-oriented design that is based on previous and occasionally costly experience. While with some gases the problem can be solved by selecting a suitable lubricant, other gases prohibit the use of any lubricant at all and oil-free compressors have to be used. Typical unpleasant gases include the following: 9 9 9 9 9 9 9
acetylene (see previous paragraph) chlorine hydrogen sulfide sour gas (i.e., wet gases containing hydrogen sulfide) hydrogen chloride sulfur dioxide carbon monoxide
While these gases are normally inert to metals and do not attack the commonly used structural metals under normal conditions of use, they will corrode most normal materials of construction in the presence of moisture. Corrosion is accelerated at higher temperatures and pressures. Even if the gas is absolutely dry, corrosion may occur if atmospheric air is not prevented from entering the compressor. This means that the whole system has to be inerted carefully before and after overhaul services. Normally, special materials have to be used for the gas-wetted parts of machines compressing difficult gases, and special precautions have to be taken in order to avoid leakage and to keep the gas temperature below the critical limit fixed in codes and standards for handling corrosive gases.* For chlorine, which is normally compressed to 145 to 175 psia in order to liquefy it, special compressors are available. Since this gas forms hydrochloric acid with oil and water, oil lubrication of the cylinders is out of the question. Compressors with sulfuric acid lubrication used earlier have now been replaced by dry-running machines with piston tings and rod-packing glands of carbon or plastic. Clean and dry chlorine has lubricating qualities, and the wear on the piston rings is astonishingly small. It is, however, essential that the chlorine be pure to prevent sealing tings sticking, and it is highly recommended to wash the chlorine gas with liquid chlorine from the liquefier before suction. Care must also be taken to prevent atmospheric * For wet gases containing hydrogen sulfide (sour gas), see National Association of Corrosion Engineers Standard MR-01-75, Sulfide Stress Cracking Resistant Metallic Material for Oil Field Equipment, published by the National Association of Corrosion Engineers, Houston.
566
Process Plant Machinery
moisture from entering the compressor in order to keep the gas dry and to prevent corrosion. The distance piece of these compressors has to be vented by means of dry air or preferably by dry nitrogen under a pressure slightly above atmospheric. The free length of the rod in the venting chamber is greater than the piston stroke, so that traces of moisture adhering to the rod can never come in contact with chlorine. In order to keep the delivery temperature below approximately 180 ~ chlorine compressors are usually made in three-stage designs for liquefaction at 175 psia. COMPRESSORS WITH DRY-RUNNING PISTON RINGS
In contrast to the compressors described thus far, machines with dry-running piston tings and piston-rod-packing rings utilize no liquid lubricant, neither of a petroleum, synthetic, or other type, nor substitutes (for example, water) within the compression chamber. They belong to the group of nonlubricated compressors. Basically, there are two different types of nonlubricated reciprocating compressors: 9 dry-running compressors with piston rings and rod packing rings, which do not require lubrication 9 frictionless, ringless labyrinth-piston-type compressors, which will be described later The need for nonlubricated reciprocating compressors was first established in the United States by the brewing industry during the immediate postprohibition period. This industry was in the market for machines supplying uncontaminated compressed air. To meet these requirements, it was decided to build a carbon piston ring compressor based on previous experience with carbon brushes on commutator and collector rings in the electrical industry, which indicated that a reasonable life might be expected. It is interesting to note that the first labyrinth-piston-type compressor, built in Switzerland at about the same time, was also designed upon request of a brewery for compressed air service (see later). Since the first appearance of nonlubricated compressors, these machines have gained a considerable market share. There are many reasons why nonlubricated compressors are used, and only the most important ones will be mentioned: 9 Some gases do not permit the use of lubricating oils for safety reasons; oxygen is a typical example. 9 Some gases attack lubricating oil, for example, chlorine. 9 Lubricant contaminates gas stream (for example, instrument air, gases used in the foodstuff industry, and air and carbon dioxide in breweries). 9 Lubricant carry-over fouls heat exchangers. This is an important factor in cryogenic cycles. 9 No suitable lubricant is available for very low and very high temperatures (for example, boil-off compressors for liquid natural gas storage, steam compressors). 9 Lubricant carry-over "poisons" catalyst. Except for plant air, where the presence of trace amounts of oil may occasionally be welcome, there are no compression duties where contamination of the gas by lubricating oil is desired. For this reason, nonlubricated compressors are now used whenever possible and are commercially acceptable. However, when using nonlubricated compressors, one should bear in mind that the gas contains no oil to coat piping, pressure vessels, and heat exchangers. If these components are made of carbon steel, corrosion may occur.
Reciprocating Compressors 567 Carbon used for piston rings and packing rings has been largely superseded in current practice by composition PTFE material, a fluorocarbon resin or plastic together with filler materials, such as glass fibers, bronze, and carbon. PTFE is not a self-lubricated material. The value of this material rests solely in its low coefficient of friction. Although only PTFE is mentioned here, the same comments would apply to future plastics, where improvements seem most likely. The plastics industry has developed low-friction materials for almost every gas. However, currently there is no universal material or compound that gives optimum service under all conditions. Although new engineering plastics, which can withstand relatively high stresses and temperatures, have been developed, there are limits to be observed by the designer when using nonmetallic materials. These limits are set by the following main factors, which may adversely influence the durability of PTFE piston tings: 9 pressure 9 temperature 9 properties of the gas 9 dirt Compressors with piston tings and piston-rod-packing tings of plastic can normally be used for discharge pressures up to some 2900 psia. This, however, is not a fixed limit. With dry gases, excessive wear already occurs at much lower pressures. Some compressor manufacturers claim that they have built dry-running compressors for pressures as high as 4000 psia even for such exacting duties as bone-dry inert gases, e.g., argon and nitrogen. Unfortunately, they often neglect to publish the life or durability of the dry-running parts under these conditions. The author knows of such a compressor compressing ammonia in three stages from 270 to 4000 psia where the piston tings of the third stage had to be replaced every 200 operating hours. For process compressors, where uninterrupted service of 8400 hours (one year with an availability of 96 percent) is required, a discharge pressure of 1500 psia should not be exceeded. For pressure between 1500 and 2900 psia, it is very doubtful whether 8400 hours of uninterrupted service can be achieved, and pressures above 2900 psia are in the experimental range for nonlubricated reciprocating compressors with PTFE piston tings. A large number of nonlubricated compressors with a static suction pressure of around 5000 psia and a delivery pressure roughly 400 to 600 psi higher than the suction pressure- so-called recirculators- have been successfully built. In these compressors, however, carbon tings have been used because PTFE did not withstand the high temperature created by friction heat. With carbon, an average life of 7000 operating hours has been reached. In the meantime, plastic materials have been developed for higher temperatures; however, it seems that these reciprocating recirculators have been phased out by the CPI. Discharge temperature for dry-running compressors with PTFE piston rings should be held to a maximum of 350 ~ in order to achieve acceptable durability of the piston rings. By properly staging a compressor, the temperature can be kept below the critical limit. Other problems can arise from the gas itself. A bone-dry gas, such as nitrogen from a cryogenic air separation plant or boil-off gas from a liquid gas storage vessel, can cause severe ring wear. When compressing argon, a bone-dry inert gas with a specific heat ratio of 1.67, the wear problem is aggravated by the relatively high discharge temperature. For dry-running pistons, dirt is usually the most severe problem. To overcome wear problems, some manufacturers propose so-called minilube cylinders. These machines belong to the category of lubricated compressors.
568
ProcessPlant Machinery The design of nonlubricated compressors is basically the same as for compressors with cylinder lubrication, as described earlier, except for the high-pressure lubricators for cylinders and rod packings, which are not required. In nonlubricated compressors, the compression chamber must receive no lubricant from any source, i.e., not even from the piston rod that normally traverses both crankcase and rod packing. There is a tendency for oil to creep along the piston rod despite the provision of scraper tings and the high pressure in the cylinder. The only way to properly combat this slight oil contamination is to incorporate an extra length distance piece between cylinder and crankcase, longer than the piston stroke, so that the oil-wetted portion of the piston rod does not travel into the rod packing. In addition to this provision, a slinger has to be installed on the portion of the piston rod that passes into neither the cylinder packing nor the frame packing. Difficult conditions- compression of highly explosive, toxic, or extremely flammable g a s e s - require the provision of a two-compartment distance piece with vent and purge connections. Reciprocating compressors with dry-running piston rings and gland-packing tings can be built for power inputs up to several thousands kilowatts.
LABYRINTH-PISTON COMPRESSORS Design Philosophy In contrast to dry-running compressors of conventional design, as described earlier, the distinctive feature of the labyrinth-piston compressor is that no friction occurs in its gas-swept parts. Instead of piston rings, the labyrinth-piston compressor is provided with a large number of grooves producing a labyrinth-sealing effect against the cylinder wall, which is grooved as well. The piston moves with sufficient clearance so that no contact occurs between the latter and the cylinder wall. The same labyrinth-seal principle is used to seal the piston rod, so lubrication of the gland is unnecessary. Figure 14-30 shows a standard machine in cross section, where the oil-free side of the compressor is marked "A," while part "B" is lubricated in the usual manner. This is achieved by using a crankshaft-driven gear-type pump that supplies the main bearings, the connecting-rod bearings, and the crosshead guide with pressurized oil. The piston-rod guide bearings are splash lubricated. Oil scrapers mounted above the piston-rod guide bearings remove the oil from the piston rod, and a slinger on the piston rod (not shown in Figure 14-30) prevents the remaining oil film and oil droplets from entering the piston-rod gland. The distance between the crank gear and the gland has been so selected that the oilwetted portion of the piston rod cannot reach the oil-free gland.
Labyrinth-Piston Compressors versus Compressors With Dry-Running Piston Rings Labyrinth-piston compressors have certain advantages but also disadvantages as compared with the oil-free compressors with dry-running sealing elements. As usual, the advantage of one compressor type are the disadvantages of the other one and vice versa. The successful operation of filled PTFE rings in a dry-running compressor depends on the ability of the PTFE to coat the mating surface of cylinder bore and piston rod. Once the coating process is completed, the PTFE rings ride a film of the same material on the mating surface, minimizing ring wear. Until the PTFE coating
Reciprocating Compressors
569
U_II
FIGURE 14-30 Sections through a single-stage labyrinth-piston compressor. 1 - Crosshead; 2 - guide bearing; 3 - oil scraper; 4 - gland with labyrinth seal; 5 - piston with labyrinth grooves; 6 - cylinder with labyrinth grooves; 7 - cylinder head. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland) is established on the mating surface of piston rod and cylinder bore, however, excessive wear of PTFE rings and packing will occur. Oil carryover by the oil scraper rings will tend to remove the PTFE coating from the piston rod and cylinder bore. This is why periodic inspection of such a compressor is necessary. Another hazard is the possibility that pipe scale or welding beads will be drawn into the inlet. They too remove the PTFE coating. If the compressor is to be idle for some time, condensation can occur in the cylinder because of the temperature difference between the wet gas and the cylinder jacket water, and rust will eventually form on the bore. Even if the cylinder liner is made of stainless steel in order to prevent the formation of rust inside the compressor, in most cases rust can not be avoided in the gas pipes, heat exchangers, and pressure vessels unless these are made of nonrusting materials. Rust and other solid particles are not only a cause for removal of the PTFE coating but have a tendency to embed in the surface of the PTFE rings, which leads to rapid wear of the cylinder liner. Solid particles stem not only from rusty surfaces but in many cases from the catalyst. Below a certain size it is very difficult to remove them by means of suction filters. This problem can be overcome in labyrinth-piston compressors, which have no friction in the gas-swept part. Consequently, they are not particularly sensitive to damage from dust-laden gases. A slight formation of rust in the cylinder can be tolerated, since the proper functioning of the compressor is not dependent on a smooth cylinder surface. This also means that inexpensive materials, such as carbon steel and cast iron, can be used. One of the problems with dry-running reciprocating compressors stems from friction in conjunction with the materials used for piston rings. They set a limit
570
ProcessPlant Machinery to the pressure difference and discharge temperature of each stage. This problem, too, can be overcome with the contactless and frictionless piston, which does not depend on the use of nonmetallic parts. With labyrinth-piston compressors, pressures up to 4300 psia can be attained regardless of the temperature and properties of the gas. Even higher pressures are possible. Until 1988, the highest pressure attained with a labyrinth-piston compressor in industrial service was 6980 psia. The medium compressed was ammonia synthesis gas. Apart from the rod-packing tings, which are made of graphite, no nonmetallic parts, which limit the temperature in the cylinder, are used. Since graphite is a high-temperature-resisting material, relatively high discharge temperatures can be tolerated. The temperature limit is then set by the gas itself, which may decompose, polymerize, or dimerize, or which may become corrosive above a certain temperature. With suitable cylinder materials and a purpose-oriented design, which keeps heat expansion within tolerable limits, very high and very low gas temperatures can be handled. The disadvantages of a labyrinth-piston compressor as compared with a machine with piston rings stem from the labyrinth principle itself. High discharge pressures lead to small pistons in the final compression stage, with a corresponding unfavorable ratio of the ring gap surface area between piston and cylinder-topiston area. Where light gases, such as hydrogen and helium, are being compressed, this ratio imposes a limit on the application of the labyrinth-piston compressor at significantly lower pressures than mentioned earlier. Whereas energy losses in the labyrinth can be negligible for the majority of gases, such losses can be considerable when light gases are being compressed, particularly to higher pressures. However, this does not entirely exclude labyrinth-piston compressors from hydrogen and helium service. A substantial number of these machines have been installed in cryogenic cycles with helium as a refrigerant. Due to the labyrinth losses between piston and cylinder, it is natural that labyrinth-piston compressors cannot be miniaturized at will. Their capacity range is consequently limited, so the lower limit of a single-stage compressor is approximately 10 acfm at atmospheric suction conditions, while that of the multistage machines is considerably higher. For most gases, the minimum suction capacity at atmospheric suction pressure of a four-stage machine for a maximum discharge pressure of 3570 psia is around 100 acfm. These figures are quite different for a light gas, such as helium. A four-stage machine has been built for helium service, and for a suction capacity of 290 acfm at atmospheric pressure, a discharge pressure of 870 psia was attained. These few examples demonstrate what is meant by the quote, "the labyrinthpiston compressor cannot be miniaturized at will." Smaller capacities, however, are not normally required by the CPI. Supercritical gas extraction processes may open in the future a field of application for small high-pressure labyrinth-piston compressors. For these machines, a different design will be required. The upper limit of the capacity range is governed by economic considerations; it is around 6500 acfm. Labyrinth-piston compressors are available as standardized and customdesigned units. More than 30 frame sizes cover a power range from 20 to 2,000 kW. Labyrinth-piston compressors generally have a shorter piston stroke and are designed for higher speeds than compressors with piston rings. The reason why compressors with friction on the reciprocating parts have longer strokes and lower rotational speeds is obvious: the more strokes per minute, the more wear occurs at the points where the reciprocating parts change direction. On the other hand, a high rotational speed has a favorable effect on the labyrinth leakage in compressors with labyrinth pistons.
Reciprocating Compressors
As mentioned earlier, the labyrinth sealing system has its price: labyrinth losses. It has been claimed that labyrinth-piston compressors require more power than nonlubricated compressors with piston rings. This raises the question as to whether the power loss caused by mechanical friction of the piston rings and gland tings is smaller than the power loss caused by labyrinth leakage between piston and cylinder and in the piston-rod gland. Experience gained with both compressor types shows that for medium to large swept volumes per unit time and for gases that are not unusually light, the energy loss due to labyrinth leakage is about equal or even less than that caused by friction. In order to obtain more accurate results, comparative tests on air were made with a vertical single-throw standard crank machine at a rated speed of 750 RPM and two single-stage double-acting cylinders and piston sets. The two cylinders had exactly the same dimensions. The only difference was the surface finish of the cylinder walls. Whereas one piston was of standard labyrinth design, the other was fitted with three plastic piston tings. In both instances, exactly the same compressor valves, pipe work, measuring instruments, and drive elements were used. In this way, the differences in the operating characteristics could be reliably established without interference from side effects. The comparison was made between the two series of tests on the basis of the following efficiencies: p adiabatic r/a d - -
p effective p
p
0is =
isothermal
effective
The test results are plotted in Figure 14-31 as a function of the pressure ratio at a speed of 600 RPM.
Piston-Rod Glands As mentioned previously, the piston-rod glands of labyrinth-piston compressors are also based on the labyrinth sealing principle. The sealing elements are of the floatingring type. These rings have labyrinth grooves inside, facing the reciprocating piston rod. They are made of special grades of graphite. Their good dry-running ability,
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FIGURE 14-31 Efficiency curves f o r a single-stage n o n l u b r i c a t e d piston compressor. ( ) - isothermal; () - adiabatic; o - p i s t o n with three p l a s t i c rings, single-stage, double-acting, 150 m m stroke; A - labyrinth piston; pl - intake pressure o f 14.2 psia; P2 - discharge pressure. (Source: Sulzer-Burckhardt, Winterthur a n d Basel, Switzerland)
571
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Process Plant Machinery
chemical inertness, and low thermal expansion coefficient are almost ideal for most gases to be compressed in the CPI. Frictionless packings cannot "run hot." The gland leakage is normally collected in a ring chamber at the lower end of the gland and vented back to the compressor intake. It does not represent a real loss of process gas (Figures 14-32 and 14-33). In compressors handling air, nitrogen, oxygen, carbon dioxide, and other nontoxic and nonflammable gases, small gas losses to the surrounding atmosphere occurring at over-atmospheric suction pressure may be tolerated. They are vented to the atmosphere. In order to minimize this leakage, the lowest gland ring may be of the rubbing three-segment type with two garter springs around it and a smooth inner surface. Hot running of the lowest gland ring does not occur provided that the differential pressure across it is relatively small. For higher suction pressure, special gland designs are available. Labyrinth piston compressors as shown in Figure 14-32 are used in these services. In compressors handling toxic, flammable, or very expensive gases, such as carbon monoxide, hydrocarbons, hydrogen, helium, and argon, even very slight gas losses to the atmosphere cannot be accepted. For these services, hermetically closed compressors as illustrated in Figure 14-33 are used. The only inevitable opening to the outside, which is the penetration of the crankshaft, is sealed hermetically in both stopped and running condition by means of a double mechanical seal immersed in an internal oil bath. This type of seal is effective even if the crankcase is under vacuum. The crankcase and the distance piece are filled with process gas and are subjected to suction pressure. However, steps are taken to prevent the gas from carrying oil into the process. Depending on the necessity, this may be achieved by an oil separator in the crankcase or by a molecular sieve built into the leakage return line outside the crankcase. Practical results with closed refrigeration cycles have shown that the system remains free of oil even after such a machine has been in operation for years.
FIGURE 14-32 Labyrinth-piston compressor, open type. (Source: Sulzer-Burckhardt, Winterthur and Basel Switzerland)
Reciprocating Compressors 573
FIGURE 14-33 Labyrinth-piston compressor with closed pressurized frame. (Source: Sulzer-Burckhardt, Winterthur and Basel Switzerland) Basically all gases or gas mixtures that are compatible with lubricating oil can be compressed with these gas-tight compressors. They are also used in closed refrigeration cycles, particularly where low evaporation temperatures preclude the use of compressors with cylinder lubrication. Refrigerants used for this purpose are halogenated hydrocarbons, ethane, ethylene, a n d - for cryogenic c y c l e s - helium. In closed cycles, these compressors must be able to withstand standstill pressures occurring when the entire system is warmed up to ambient temperature. To date, compressors for a crankcase pressure of 300 psia have been built. For toxic and corrosive gases that are not compatible with lubricating oil, neither the open-type nor the closed-type compressor can be used. Chlorine belongs to this class of gases. For such hardship cases, compressors with a strict separation between the oil-free cylinder and the lubricated crankcase have to be used. Since leakage to the environment is not allowed, special piston-rod glands have to be applied with vent and purge connections that can be flushed with a sealing gas, for example, dry nitrogen. Piston-rod gland problems are often the decisive factor for the selection of the compressor type. It should be borne in mind that no piston-rod packing is entirely gas tight. As a rule, lubricated glands have less leakage than nonlubricated ones, since oil also acts as a sealant. Entirely encapsulated compressors with mechanical crankshaft seals give best results as far as tightness is concerned, since it is much easier to seal a rotating shaft than reciprocating piston rods.
Typical Applications For Labyrinth-Piston Compressors Labyrinth-piston compressors are often the answer to problems occurring with other types of compressors. Figure 14-34 depicts a three-stage oxygen compressor with a
574
Process Plant Machinery
FIGURE 14-34 Three-stage oxygen compressor, 1700 acfm range. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
FIGURE 14-35 Four-stage labyrinth-piston compressor in oxygen service. This machine compresses 153 acfm to a discharge pressure of 3570 psia. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
Reciprocating Compressors
capacity of 1670 acfm and a discharge pressure of 600 psia. Since rubbing friction represents a source of ignition, frictionless labyrinth-piston compressors have found worldwide application in oxygen service. Packaged plants of the type shown with power ratings as high as 500 kW have been supplied. Figure 14-35 represents a four-stage oxygen compressor with a capacity of 153 acfm and a maximum discharge pressure of 3570 psia. With this compressor, dry oxygen and other dry gases can be compressed. In contrast to water-lubricated compressors, which are still widely used for this service, the gas remains dry and no dryer is required downstream of the machine. Two out of six two-stage propylene compressors located in a U.S. petrochemical plant are shown in Figure 14-36. Each of these compressors has a capacity of 2270 acfm and produces a discharge pressure of 300 psia. The compression ratio exceeds 10. These machines have
FIGURE 14-36 Propylene compressors in a United States Gulf Coast petrochemical plant. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland)
575
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Process Plant Machinery
FIGURE 14-37 Skid-mounted two-stage labyrinth-piston compressor with suction valve unloaders. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland) performed admirably in a fouling service. Two skid-mounted two-stage labyrinthpiston compressors with pressure-tight casing are shown in Figure 14-37. Compressors of this type have found worldwide application as boil-off compressors for the sea transport and storage of liquefied gases, such as hydrocarbons and ammonia. In most cases, the vapor is compressed in order to be reliquefied and then recycled to the transport or storage tank. All suction valves are fitted with unloaders for unloaded start and capacity control. Compressors of this design are also used in refrigeration cycles. DIAPHRAGM COMPRESSORS*
Earlier in this chapter, Figure 14-10 had assigned diaphragm compressors to the high-pressure low-flow range of application. Modem diaphragm compressors are often a combination of two s y s t e m s - a hydraulic system and a gas compression system. A metal diaphragm group is the isolating component between these two systems. In the diaphragm compressor depicted in Figure 14-38, the gas compression system consists of three fiat metal diaphragms, which are clamped between two precisely contoured concave cavities and process gas inlet and outlet check valves. The hydraulic system includes a motor-driven crankshaft that reciprocates a piston in the hydraulic fluid medium. This positive-displacement piston pulses the hydraulic fluid against the lower side of the diaphragm group, causing it to sweep the cavity, displacing the process gas. Except for very small diaphragm compressors, the other components of the hydraulic system are the automatic hydraulic injection pump, the hydraulic fluid check valve, and the hydraulic relief valve. In addition to ensuring that the hydraulic system is always full for the compression cycle, the * Source: PPI Division, the Duriron Company, Inc., Warminster, Pennsylvania. Adapted by permission.
Reciprocating Compressors UPPER (GAS) H E A D
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automatic injection pump provides for fast priming of the hydraulic system after extended compressor shutdowns. On smaller compressors, this same function is sometimes accomplished by a manual priming pump and a gravity oil-feed system. On the compression stroke of the main piston, the hydraulic check valve isolates the hydraulic system from the automatic injection pump or the manual priming pump so that an elevated pressure can be generated in the system against the hydraulic relief valve. A full cycle of operation for the diaphragm compressor begins with the diaphragm group fully deflected to the bottom of the cavity by the gas suction pressure. The hydraulic piston is at bottom dead center, and the hydraulic system has just been filled by a single stroke of the automatic hydraulic fluid injection pump. On the process side of the diaphragm group, the cavity is now filled, at a given suction level, by the process gas that has entered through the inlet check valve. As the crankshaft rotates and the hydraulic piston moves from its bottom dead center position toward top dead center, the hydraulic pressure increases. This occurs because the hydraulic system injection and relief lines are blocked by the hydraulic check valve and relief valve, respectively. As the piston continues toward top dead center, and as the hydraulic pressure reaches the pressure level of the process gas in the cavity, the diaphragm group begins to sweep the cavity toward its top dead center position, thereby compressing the process gas. When the pressure of the process gas in the cavity reaches the pressure level downstream of the discharge check valve, that check valve opens and the process gas is discharged from the cavity. Since the hydraulic system has slightly more displacement capacity than the gas system, the diaphragm group makes metal-to-metal contact with the process head cavity, assuring that all of the gas has been displaced. With the diaphragm group in this fully deflected discharge position, the piston still has a certain amount of travel required to reach its top dead center position. As the piston moves to top dead center, this additional hydraulic fluid volume is "overpumped" through the hydraulic relief valve, which is set at a pressure level above the desired process gas discharge pressure. At this point, the compression portion of the cycle has been completed.
577
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Process Plant Machinery
FIGURE 14-39 Hydraulic system pressures acting over one operating cycle of typical diaphragm compressor. (Source: PPI Division of the Duriron Company, Inc., Warminster, PA)
The hydraulic piston now moves toward bottom dead center. As it does, the diaphragm group is deflected toward the bottom of the cavity by both the expansion of the residual gas contained in the clearance volume and by the additional process gas entering the cavity at suction pressure. During this suction stroke, a synchronized auxiliary eccentric cam on the crankshaft causes the hydraulic injection pump plunger to stroke, injecting an amount of hydraulic fluid equal to that which was "overpumped" through the hydraulic relief valve at the end of the compression portion of the cycle. On small gravity oil-feed units, hydraulic fluid is drawn from the crankcase as the main piston travels to bottom dead center. When the hydraulic piston reaches its bottom dead center position, the hydraulic system is again filled. Since the diaphragm group is now in the bottom-most deflected position, a full gas cavity is assured. At this point, the cycle is complete. The hydraulic pressure versus crank position has been plotted on Figure 14-39.
APPENDIX 14A Performance Calculations
Performance calculations are made to determine the throughput of a reciprocating compressor and the horsepower absorbed by the process. The capacity throughput is determined by the displacement rate of the compressor, i.e., the volume swept by piston or pistons, the actual suction volumetric efficiency, and the specific volume at suction conditions. The volume swept by the piston(s) is determined by the total piston area, the piston stroke, and the piston speed. In the case of a multistage compressor, the calculations must be performed for the first stage only, and the pressure ratio of the first stage has to be used. However, this does not mean that the following formulas cannot be used for each individual as well. The actual volumetric efficiency is defined as follows: k=
actual suction volume volume swept by piston(s)
%
In order to calculate this value, the theoretical volumetric efficiency based on ideal gas conditions and not considering efficiency losses of any kind must first be determined (Figure 14A- 1).
O-- l O 0 - V o
[(/ 1
P2 ~ Pll -- 1 %
where Vo = clearance volume (or space) expressed as percentage of the swept volume P2 = absolute discharge pressure of the first stage p~ -- absolute suction pressure of the first stage k = exponent of the re-expansion curve c - d As a rule, the specific heat ratio is used for the sake of simplicity. The actual reexpansion curve deviates from the isentrope obtained in this way, and the difference has to be included in the coefficient of correction x (see below). This theoretical efficiency has to be corrected by x in order to obtain the actual volumetric
efficiency: X=r/-x% Correction x allows for all influences that decrease the volumetric efficiency, such as pressure losses across the valves increasing the pressure ratio, preheating of the gas in the suction chamber resulting in a decrease of its specific weight, deviation from ideal gas law (compressibility factors at both suction and discharge conditions), and internal gas leakages. Test results determine x. It is normally between 3 and 10 percent. 579
Process Plant Machinery
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%
F I G U R E 14A-1 Theoretical volumetric efficiency versus clearance volume, compression ratio (pressure ratio PI/P2), and K value. Example: K - 1.4; compression ratio- 3; clearance volume- 12%; theoretical volumetric efficiency - 85.5%. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland.)
Figure 14A-2 may be used to determine the theoretical volumetric efficiency. The actual suction volume at intake conditions can then be calculated as follows: V l - Vs x ~. acfm where V s -
volume swept by piston(s) acfm
The throughput expressed as a mass flow is obtained by means of the following formula: V1 W - - - x 60 lb/hr 01 where
O t = specific volume of the gas at suction conditions, which is a function of the pressure, temperature, gas composition, and compressibility factor.
cu/ft lb
Appendix 14A" Performance Calculations
"T
o,} o,j
cap_ C) t~ .C}
q ................ ~--. . . . . . . . . . . . . . . . . .
.Vo
D, v
_
..........
too _O/_o .
.
.
.
.
.
.
.
.
.
.
.
Votume
FIGURE 14A-2 P-V indicator diagram, with shaded area representing work required to compress and deliver the gas. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland.)
POWER REQUIREMENT At a very early stage of each project in chemical engineering and in many cases before compressor manufacturers can be approached for quotations, the chemical engineer is confronted with at least three questions" 1. How big is the compressor I need? 2. What initial investment is required? 3. How much power is required? Since the key to answering the first and second question is the power consumption, it may be useful to have a formula at hand, by means of which it is possible to work out at least an approximate figure for the power absorbed at the compressor shaft. The basic theoretical single-stage horsepower formula is developed for the pv indicator diagram, the net area of which represents the work required to compress and deliver the gas. This is represented by the shaded area in Figure 14A-1. Unlike the diagram of the ideal cycle, the indicator diagram describes the real cycle and includes the pressure losses across the valves. The actual power requirement is related to a theoretical cycle through an overall efficiency that has been determined by test on prior machines. The overall compressor efficiency is the ratio of theoretical to actual power on the compressor shaft. As a rule, reciprocating compressors are compared with the isentropic (adiabatic) cycle. Since this leads to a somewhat complicated formula, a good many engineers still use the isothermal power as a basis, as in the formula below: Pis = R x T l x In P__!2 P~ where R -- gas constant T] = absolute suction temperature Since this is a general formula saying nothing about the type of compressor or the number of stages, it can also be used for multistage compressors provided that P2/P] is the overall pressure ratio and not only the pressure ratio of one stage. Using algebra, the above formula can be converted into the following ones, which are better suited for practical use:
581
582
Process Plant Machinery
Power absorbed in horsepower: P = 10.05 • 10 -3 • p~
x
1 V 1 x log P2 x m = HP Pl r/is
Power absorbed in kilowatts: 1 P = 7.50 x 10 -3 x Pl • V1 • log P~ x = kW Pl /']is
where Pl -" absolute suction pressure in psia P2 = absolute discharge pressure in psia V1 --actual suction volume in acfm r/is - overall efficiency based on isothermal power Overall efficiency based on isothermal power (rlis) is a figure that must be stated by the compressor manufacturer. In most cases, it is between 0.55 and 0.70 (55 to 70 percent). The value of rlis depends on many factors and its magnitude can be influenced only in part by the compressor manufacturer. It is mainly determined by the following: 9 gas to be compressed (specific heat ratio, molecular weight, compressibility factors) 9 size and type of compressor (as a rule, one 100 percent capacity compressor has a better efficiency than two 50 percent capacity machines) 9 speed of the compressor (pressure losses in the gas stream) 9 discharge pressure, pressure ratio 9 number of stages 9 intercooling with multistage compressors, i.e., the better the cooling effect, the better the efficiency CONVERSION OF SUCTION CAPACITY Compressor manufacturers state the actual suction capacity of their machines in acfm or cubic meters per hour, whereas chemical engineers prefer weight rates expressed in pounds per minute or pound moles per minute or flow rates expressed in standard cubic feet per minute (scfm). The actual inlet volume may be calculated from the following formulas by using the proper inlet pressure and temperature and correcting for moisture content at these conditions and deviation from the ideal gas laws (compressibility): From weight flow (W) to acfm (W lb/min, dry) V! = 1 0 . 7 3 •
W
Tl • m xZ1 xC1 = a c f m M p~
From scfm to acfm (Vo scfm at 14.7 psia, 60 ~ Vl--
14.7 520 x V o •
Tl m • Pl
dry) •
From mole flow to acfm (N lb mole/min, dry) VI=
379 • 14.7 520
xNx
T1 m • Pl
xCl=acfm
Appendix 14A: Performance Calculations or T1 V1 = 10.714 x N x m x Z1 x C 1 ~- acfm Pl
where
M T1 Pl Z1 C1
= = -= --
molecular weight absolute inlet temperature absolute inlet pressure compressibility factor, at intake conditions correction factor for the moissture at intake conditions
dimensionless degrees Rankine ~ psia dimensionless dimensionless
(/)
depends on gas, pressure, and temperature. For ideal gases, Z 1 is 1. C 1 depends on moisture content, temperature, and pressure. For dry gases, C1 is 1. For atmospheric pressure, 14.7 psia, 68 ~ and 100% relative humidity, it is 1.024. Use charts for ZI and C1. Z 1
583
APPENDIX 14B Capacity Control
There are two main reasons why compressor capacity regulation is used. The most prevalent one is to adjust the suction flow to match the process demand. The second reason is to save energy. As a rule, capacity control is determined by compressor discharge pressure. In cases where the system upstream of the compressor has to be protected against too low a suction pressure, the control point may be governed by intake pressure. Modem control technology permits using other parameters as control points; temperatures, flows, liquid levels are but a few examples. Where changes in demand are infrequent and slow, the capacity control may be arranged for manual operation, either directly on the compressor or by means of remote control. Modem process plants in which changes in demand are rapid and not always predictable, or where compressors have to be operated without supervision by operating personnel, require automatic control. There are a number of variations that can be grouped under two branches of capacity control, as shown in Figure 14B-1. The optimum capacity control is largely determined by the following parameters: 9 9 9 9 9
the capacity range required how frequently changes in demand occur how long reduced capacity is required size of the compressor type of driver
Not all types of capacity controls can be used with a given compressor model, a specific pressure range, or a given gas. The process engineer involved in specifying the compressor should clearly describe the turndown requirements and work with the machine manufacturer in determining feasible capacity control strategies. It is sometimes necessary to combine two or more types of regulation for best efficiency, flexibility, and reliability. Table 14B.1 lists the main characteristics of the capacity control systems as described in Figure 14B-1. Capacity control by means of an overall bypass can be applied without limitations to all compressors, provided that the gas recycled enters the suction line close to normal suction temperature. This means that an aftercooler or a bypass cooler may have to be used. In addition, a check valve in the discharge pipe is required to prevent the high-pressure gas from flowing back when the compressor is at standstill. Since this regulation is very uneconomical, it should only be used if the compressor has to be operated at reduced capacity for a short time or in combination with an energy-saving type of control. In multistage compressors, a bypass around the first stage or a partial bypass can be used. The absolute power input can be reduced in this way, although the specific input is increased. In fixing the regulation range of a first-stage bypass, it must be remembered that a reduction of the flow to the second stage causes a drop of all interstage pressures, and consequently it can lead to excessively high pressure ratios and discharge temperatures in the following stages. These 584
Appendix 14B: Capacity Control 585 TYPES OF CAPACITY CONTROL MOST WIDELY USED FOR PROCESS RECIPROCATING COMPRESSORS
I,
I
1
CONSTANT-SPEED CONTROL
I
i
I
OVERALL
I 1
I
I
INLET VALVE UNLOADING
I STEP-
I
I
CLEARANCE CONTROL
1
I
TWO-SPEED EL. MOTOR
I
STEPLESSSPEED CONTROL
1
STEPINFINITEINFINITECONTROL STEP CONTROL CONTROL STEP CONTROL
PARTIAL
2
I
I
INTAKE PRESSURE THROTTLING
BYPASS CONTROL
VARIABLE-SPEED CONTROL
3
I
4
I
5
I
6
I
7
8
9
FIGURE 14B-1 Types of capacity control most widely used for process reciprocating compressors. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland.) pressure shifts can cause overload in the last stage. Therefore, the recycled flow has to be restricted by means of an orifice plate in the bypass pipe. The minimum capacity that can be attained depends on the number of compression stages. The more stages used for a given overall compression ratio, the wider the achievable control range. For oxygen service, a partial first-stage bypass is normally the only recommended means of constant-speed capacity control apart from an inefficient overall bypass. With three-stage compressors, the minimum attainable capacity is approximately 60 percent. If a lower turndown capacity is required with such a compressor, the problem can be solved by providing an additional compressor stage or cylinder. Bypass regulation are very practical and easy to apply. Standardized components can be used. Intake pressure throttling will provide an infinite number of steps between full load and reduced load. Capacity changes are achieved by reduction in both gas density and volumetric efficiency, since the latter depends on the pressure ratio. When this method is applied, it is again necessary to investigate beforehand the resulting changes in pressure, discharge temperatures, and load conditions in the upper stages. When compressing flammable gases, care must be taken not to create a vaccuum in the suction pipe in order to avoid the ingress of air. Inlet valve unloading is the most widely used type of constant speed control. This consists simply of holding one or more inlet values open during both suction and discharge strokes, so that the gas taken into the cylinder on the suction stroke is pushed back through the suction valves on the discharge stroke. With one doubleacting cylinder, two s t e p s - 50 and 0 p e r c e n t - are possible. With two first-stage double-acting cylinders running in parallel, 75, 50, 25, and 0 percent capacity can be attained. With multistage compressors and control range down to 0 percent, all suction valves have to be fitted with unloaders. The cylinder shown in Figure 14B-2 is fitted with suction valve unloaders (A). The suction valve plates can be lifted either by means of threaded spindles with handwheels (not shown in the figure) or by means of servomotors, as represented in Figures 14B-3 and 14B-4. These are operated either by oil pressure or, less frequently, by compressed air, or by the process gas.
586
Process Plant Machinery Main Characteristics of Capacity Control Systems
TABLE 1411.1 No.
Type of Control Symbol
Power vs Capacity
Range %
Advantages
Disadvantages
Unloaded Start
I00-0
SIMPLE RESPONSIVE STEPLESS UNEXPENSIVE
WASTE OF POWER
NO
100-,-~,60
STEPLESS INEXPENSIVE
CAN BE APPLIED TO MULTISTAGE COMPRESSORS ONLY INTERSTAGE PRESSURE SHIFT
PARTIAL
DEPENDS ON NUMBER OF STAGES
STEPLESS INEXPENSIVE
INCREASED PRESSURE RATIO HIGHER DISCHARGE TEMPERATURES
PARTIAL
100...0
ECONOMICAL RESPONSIVE ONE STEP PER CYLINDER OR CYLINDER HALF
NOT FOR OXYGEN
YES
100-50 AND 0
ECONOMICAL
NOT FOR OXYGEN NOT SUITED FOR MULTISTAGE COMPRESSORS
YES
100 . . . . 50
ECONOMICAL
NOT FOR OXYGEN CAN BE PROVIDED ON CYLINDER HEAD SIDE ONLY CAUSES DISCHARGE TEMPERATURE TO RISE
PARTIAL
100- "-'50
ECONOMICAL
SAME AS ABOVE
PARTIAL
100 AND 50 (100/67)
ECONOMICAL INEXPENSIVE
FOR ELECTRIC MOTORS BELOW 500KW ONLY
NO
100-'--,50
ECONOMICAL
WITH ELECTRIC MOTORS: EXPENSIVE
NO
p_
v', P
J i
P
/ 1,5
P
V,
P
V,
Appendix 14B: Capacity Control
I
587
I
FIGURE 14B-2 (A) Compressor cylinder with suction valve unloaders and (B) clearance control. (Source-Burckhardt, Winterthur and Basel Switzerland.)
J§
$~ m
FIGURE 14B-3 Suction valve unloader for reciprocating compressors. (A) Servomotor, (B) main spring, (C) oil under pressure, (D) bell, (E) suction valve. (Source: Sulzer-Burckhardt, Winterthur and Basel Switzerland.)
When the servomotor (A) is deactivated, a spring (B)exerts pressure on the spindle, thus keeping the valve plate in the open position. The oil pressure is, as a rule, generated by the crankshaft-driven gear-type lubricating oil pump, which also supplies oil to the compressor beatings. This means that the suction valves are kept open by spring force when the compressor is idle. As a consequence, a compressor equipped with hydraulic suction valve unloaders is started up without load and begins to deliver gas only when the oil pump generates sufficient oil pressure. When the compressor is running, its capacity is reduced in steps according to the number of suction valves kept open. Infinite (stepless) control between 100 and approximately 50 percent of each individual suction valve and hence of each cylinder or cylinder half can be achieved
588
Process Plant Machinery
FIGURE 14B-4 Single-stage labyrinth-piston compressor with two double-acting pistons working in parallel. One cylinder is fitted with suction valve unloaders (A) on both suction valves, while the other one has an unloader on the lower suction valve only. The unloaders are actuated by oil pressure supplied by the crankshaft-driven oil pump (C). The oil flow to and from the servomotors is controlled by three solenoid valves (B) so that capacity steps of 75, 50, and 25 percent can be achieved. (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland.)
by using springs between the servomotor spindle and the bell, which keeps the valve plate open. In this way it is possible to influence the time during which the gas is pushed back to the suction pipe at the discharge stroke. Valves controlled in this way are referred to as timed suction valves. By altering the spring force acting on the bell, the closing of the valve plate can be held up at any point between 0 and approximately 50 percent of the compression stroke, so that the flow can be adjusted continuously between 100 and approximately 50 percent. The spring force depends on the position of the servomotor piston, which in turn depends on the oil pressure below the piston. Below 50 percent capacity, this control system jumps to the no-load position and the suction valve plates are then kept open permanently. This system also permits unloaded start of the compressor. An example of a one-step clearance control is shown in Figure 14B-2 (B) and Figure 14B-5. Such a device is sometimes used in combination with valve unloading in steps. Figure 14B-5 illustrates cylinders of a two-stage labyrinth-piston compressor compressing ethylene from 280 to 1215 psia. The first-stage cylinder (fight) is equipped with a one-step clearance control and with valve unloaders on the lower suction valves. The second-stage high-pressure cylinder has no capacity control. With this combination, steps of 73 and 60 percent can be achieved. Reducing
Appendix 14B: Capacity Control
FIGURE 14B-5 One-step clearance controls on vertical compressor cylinder (right side, top). (Source: Sulzer-Burckhardt, Winterthur and Basel, Switzerland.) the capacity in one cylinder only causes a drop of the interstage pressure, which could be tolerated in this specific case. The actuator for the clearance pocket control is operated by compressed air. The servomotors for the valve unloaders are actuated by oil pressure; however, the three-way valve controlling the oil flow is controlled pneumatically. The system is suited for manual as well as automatic remote control by means of a 60-psig compressed air signal. All reciprocating machines have a clearance space, which is imposed by their design and which, particularly in compressors, is kept as small as possible. At the end of the compression stroke, this space is filled with the compressed gas. When the piston returns, the entrapped gas expands to the suction pressure before the suction valve can open and gas can flow into the cylinder. The larger the clearance space, the greater the loss of suction volume. Capacity control by additional clearance space is based on this fact. By artificially increasing the clearance space, the discharge quantity of the compressor can be regulated with only very slight losses. In its simplest form, this additional space is in the cylinder cover, as shown in Figure 14B-2, or is provided by a clearance bottle. By means of a valve, operated by hand, by air or oil pressure, or by a small electric motor, the additional clearance space can be added to the already existing one. The practical application of this principle today consists in the inclusion of one, two, or more such additional spaces in the design of the cylinder cover, each fitted with suitable valves. With two additional spaces of different sizes, a and b, four stages of regulation- namely full load and the partial loads by connecting a, b, or a + b - can be attained. This method of regulation is reliable and economical, but it permits adjustment to fixed quantities only, so that it sometimes has to be
589
590
ProcessPlant Machinery combined with a bypass or suction valve control in order to bridge the intervals between the regulating steps. Infinite stepless control by clearance space can be obtained if the additional clearance space is designed as a cylinder with a movable piston. Such a design, however, is far from simple and is generally expensive. This method of capacity control is therefore used very rarely, in spite of its technical advantage. Permanent reduction of capacity by increasing clearance space is sometimes necessary if a compressor has to be operated at reduced flow for a relatively long period of time. In this case, capacity control systems are not the optimum solution. Clearance space increases are achieved by lifting cylinder cover and/or compressor valves by means of distance rings. If valves are lifted in this way, the lanterns have to be replaced by shorter ones. The higher the compression ratio, the greater the reduction in capacity that can be achieved. In most cases, a reduction of 10 percent or more is feasible at little cost and with practically no loss in power efficiency. When the full capacity is required again, the distance rings can be removed and the lanterns replaced by the initial ones. VARIABLE-SPEED CONTROL The optimum method of regulation is adapting the flow to process demand by changing the compressor speed. Variable-speed control is used whenever the driver is capable of operating at a speed commensurate with demand. Steam and gas turbines and internal combustion engines are in this class. Most compressors today, however, are driven by electric motors. The simplest and least expensive method is to use a two-speed motor. However, two-speed operation may not be possible with large motors. The problem is aggravated by the fact that in contrast to dynamic compressors, reciprocating compressors have a constant torque over the full speed range. With direct current (DC) motors, a wide speed range can be attained, but DC drive systems can become very costly. Recent developments in thyristor technology and electronic control made it possible to use adjustable frequency drives for large compressors, resulting in excellent overall efficiency. Adjustable frequency drives are very reliable and easier to maintain than other variable-speed electric motors. With adjustable frequency systems, both induction and synchronous motors may be used. A typical adjustable frequency drive system with electronic control consists of the following main equipment: 9 9 9 9 9
a a a a a
circuit breaker or motor starter supply converter DC link reactor motor converter brushless, alternating current (AC) synchronous motor
The use of variable-speed drivers creates some problems that deserve special attention: 9 The flywheel effect decreases in proportion to the cube of the speed. 9 Torsional and other vibrations could coincide with running speeds and cause damage to the machine. 9 Gas pulsations in the pipe system must be kept under control over the full speed range. 9 Poor lubrication of the compressor bearings occurs at minimum speed, particularly if the lubricating oil pump is driven by the crankshaft. This problem can be solved by means of an independently driven pump.
Appendix 14B: Capacity Control 591 9 It may be necessary to add some additional mass to the reciprocating parts of the compressor in order to ensure piston-rod load reversal at the bottom speed. In most cases, this reversal is required to maintain proper lubrication between the crosshead pin and bushing. The above list, which is not complete, demonstrates that in certain cases it is not possible to utilize the entire speed range of a driver.
APPENDIX 14C
Higher Pressures on Dry-Running Compressors*
In recent years, energy and environment considerations have led to a growing demand for dry running compressed gases in the chemical industry. A new sealing system was developed to enable reciprocating compressors to raise pressures about 100% more than previous limits. Operators of compressor installations now demand at least 8000 hours of continuous running. At an average piston speed of 3.33 m/s the sealing and guiding elements cover a frictional distance of 96 000 km in the course of this-more than twice round the earth. During this time, the elements must neither fail nor permit unacceptable leakage rates. This applies also to compressors that have to run oilfree, as in the food industry, for example. Oilfree ring piston compressors have dry-running sealing systems between the moving and stationary parts, which also guide the piston. Crosshead compressors, which are able to compress gas with the lower piston side additionally, are fitted furthermore with a likewise dry-running piston rod seal too, called a packing. These devices must seal very efficiently, otherwise leakages may result in process gas escaping into the environment. Such losses must be avoided with expensive gases. They are quite unacceptable when inflammable, corrosive or toxic gases are being compressed. Consequently, a great deal of engineering goes into the sealing arrangements. Dry-running friction pairs function by reason of solid lubricants present in one of the friction partners. Very important here is the formation of a transfer film on the partner that does not supply lubricant. Tests with various metal-plastic pairings have shown that the coefficient of sliding friction is reduced from 0.3-0.6 to 0.1-0.3 after the formation of a transfer film. As a result of the self-lubricating action, however, the partner supplying lubricant is gradually worn away, so that sooner or later it becomes necessary to replace the sealing and guiding elements. Filled polymers with a matrix of polytetrafluoroethylene (PTFE) are employed mainly nowadays for sealing and guiding elements in unlubricated reciprocating compressors. Already decades ago, rings of this material ousted the previously predominating rings of carbon/graphite. However, pure PTFE is unsuitable for the purpose in question because its strength diminishes markedly under heating. Fillers matched to the particular application are therefore added to the polymer. Following the trend to unlubricated compression, the performance level of dry-running ring piston compressors has been raised progressively on the heels of the lubricated machines. The application limits are dictated by the properties of the dryrunning materials used.
(PEEK), polyphenylene sulphide (PPS) or polyimide (PI). These likewise are not suitable in the pure form, and require additions of solid lubricants such as carbon/graphite, PTFE or molybdenum sulphide. Rings of various plastic mixtures are now offered by virtually all suppliers in a wide variety of compositions and mixing ratios. Some of these polymer blends have been used with eminent success in areas which were previously dominated solely by PTFE materials containing hard, abrasive fillers such as glass fibres or ceramics. Costly abrasion protection can be omitted as a result. On the one hand, the oscillating motions of the reciprocating compressor are responsible for an inconstant behavior in time of the key variables in the total stress load, like the sliding velocity, surface pressure and operating temperature. On the other hand, owing to the functional principle of friction seals arranged in series there is no clear-cut distribution between the individual sealing elements. To investigate this, MSB operates a test stand based on a modified process gas compressor of the balanced-opposed type. It allows the pressure and temperature distribution in the packing to be measured as well as the gas leakage rate (Figure 14C-1). To enhance the sealing efficiency and reduce wear, the sealing elements are provided in multiple. Nevertheless, trials with various packing ring designs types did not yield the expected pressure distribution pattern. The differential pressure to be sealed was restrained mostly by one or two highly loaded sealing elements. The load was concentrated first on the elements at the edge of the packing, and when these became worn, it migrated inwards from both sides onto the next, previously unloaded elements till the entire packing was leaking. Figure 14C-2 shows the actual pressure reduction and the associated loading of the sealing elements. Extensive trials on the test stand with various types of sealing elements made from the same material revealed that not only the leakage but also the temperature and above all the wear are influenced by the ring section geometry. This shows that the sealing elements are likewise significantly improved by their design. To exploit the potential present here, the "Kaskade" simulation program was developed by MSB, simulating the processes in a dry-running friction ring seal. Assisted by laboratory trials and test stand trials with piston and packing rings, it has been possible to advance the design of sealing elements till the frictional heat generated is minimal - a prime requisite for dry running.
There has recently been a growing tendency to make use of high-temperature polymers like polyether ether ketone
With that it is possible to determine the admissible application limits for dry-running sealing systems according to the operating data of the compressor and to design dependable sealing systems.
* Source: Sulzer-Burckhardt. Adapted by permi~ion.
The efficiency of friction seals can be raised additionally by selective division of the sealing functions.
592
Appendix 14C: Higher Pressures on Dry-Running Compressors 593
Fig. 14C-1 Pressure distribution measured in a dry-running piston rod seal. After the dynamic pressure reduction at the first two rings, there is virtually no further change till the pressure drops from 4MPa to ambient after the 10th ring pair. (Source: SuIzer-Burckhardt, Switzerland.)
Fig. 14C-2 The green line plots the expected pressure reduction in serially arranged friction seals. The measurement (blue line) shows, however, that only the outer sealing elements are highly loaded. (Source: Sulzer-Burckhardt, Switzerland.)
The piston shown sectionally in Figure 14C-3 has been developed specially for compressing very light gases to high pressures. The retained piston rings held at the upper part are provided exclusively for wearfree sealing of the dynamic pressure part, while the gastight twin rings following seal the static pressure component involving wear. By
virtue of the stable separation of functions, with this design the compression of hydrogen to over 20 MPa (200 bar) is possible- completely oilfree. Previously no more than 10 MPA was attained, and this only with heavy gases. The improvement in piston rod sealing means the pressure has been raised from 4 to 10 MPa.
594 Process Plant Machinery
Fig. 14C-3 Dry-running, double-acting ring piston. With a similar piston having the sealing arrangement shown here, hydrogen compression to 20 MPa has been possible for the first time. (Source: Sulzer-Burckhardt, Switzerland.)
Chapter 15 Rotating Positive Displacement Compressors* Rotary screw compressors and rotary piston blowers belong to the machinery group making up rotating positive-displacement compressors. Of these two machines, rotary screw compressors are primarily used in higher pressure air and process gas services, whereas the rotary piston blowers are more typically used in lower pressure, high-volume applications. Both machines can be used as dry or wet fluid movers. Rotating positive-displacement machines offer the same advantage as reciprocating positive-displacement equipment with regard to flow versus pressure relationships, i.e., nearly constant inlet flow volume under varying discharge pressure conditions. Also, positive-displacement machines do not have a surge limitation, which is to say, there is no minimum throughput requirement for these compressors. The rotor tip speeds on rotary screw and rotary piston blowers are low; this allows for liquid injection and handling of contaminated gases. By design, the rotors are self-cleaning during operation, which is a significant advantage in dirty-gas services. ROTARY SCREW COMPRESSORS
Rotary screw compressors are available in oil-free or oil-flooded construction. Fields of application for oil-free machines include all processes that cannot tolerate contamination of the compressed gas or where the lubricating oil would be contaminated by the gas. Oil-flooded machines can achieve slightly higher efficiencies and utilize the oil for cooling as well. Properly designed rotary screw compressors are constructed with no metallic contact whatsoever inside the compression chambers, either between the rotors themselves or between these and the walls of the housing. Although originally intended for air compression, rotary screw compressors are now compressing a large number of process gases in the petrochemical and related industries. These include air separation plants, industrial refrigeration plants, evaporation plants, mining, and metallurgical plants. Practically all gases can be compressed: ammonia, argon, ethylene, acetylene, butadine, chlorine gas, hydrochloric gas, natural gas, flare gas, blast furnace gas, swamp gas, helium, lime-kiln gas, coking-plant gas, carbon monoxide gas, all hydrocarbon combinations, town gas, air/methane gas, propane, propylene, flue gas, crude gas, sulphur dioxide, oxide of nitrogen, nitrogen, styrene gas, vinyl chloride gas, and hydrogen gas can be found on the reference tabulations of experienced manufacturers. * Source: Aerzener Maschinenfabrik, GmbH, D-3258 Aerzen, Germany. Adapted with permission. 595
596
Process Plant Machinery
Application Limits for Rotary Screw Application limits for rotary screw compressors are given by the pressure and temperature ranges and by the maximum allowable speed of the machines. Oil-free rotary screw compressors can be mechanically loaded with pressure differences up to 12 bar, and oil-flooded compressors up to 20 bar. Higher pressure differences are possible in special cases. The maximum allowable compression ratio for one screw compressor stage that will not cause the final compression temperature to rise above the permitted value of 250 ~ will to a very large extent depend on the specific heat ratio Cp/Cv of the gas to be compressed. For example, where the specific heat ratio Cp/Cv equals 1.4, the maximum compression ratio would be approximately 4.5, and where the specific heat ratio Cp/Cv equals 1.2, the maximum compression ratio would be approximately 10 for one oil-free screw compressor stage. Multistage (multicasing) arrangements are not uncommon and can result in pressure ranges from approximately 0.1 bar absolute to 40 bar. Interstage cooling is used in many of these applications. Depending on compressor size, speeds from 2000 to 20,000 RPM can be encountered. The limiting factor is typically the circumferential speed of the male rotor, which typically ranges from 40 to approximately 120 m/sec, and up to a maximum of 150 m/sec for very light gases. Flow volumes up to 60,000 m3/hr can be accommodated in these compressors.
Principal Construction Features Rotary screw compressors (Figure 15-1) are dual-shaft rotary piston machines operating on the principle of positive displacement combined with internal compression. The gaseous medium moves from the suction port to the discharge port, entrapped in progressively decreasing spaces between the convolutions of the two helical rotors, being thus compressed up to the final pressure before it is discharged into the discharge nozzle. Figure 15-2 illustrates this process. On small rotary screw compressors, the housing is vertically parted on the suction side. Cylinder and discharge side plate are frequently combined in one housing. The housings of larger machines are often parted horizontally for easy assembly. Rotors and shafts are milled out of one piece of either forged or stainless steel. Some manufacturers provide rotors with synthetic coatings. Depending on service conditions, this may lead to a rapid drop in compressor efficiency due to loss of coating on the rotor edges. Process gas machines incorporate direction of flow from the top to the bottom, thus facilitating liquid removal from the compression space whenever liquid is injected into the rotor chamber for cooling or cleaning during operation. On-stream cleaning is highly advantageous in services where gases are contaminated or tend to polymerize. The sealing area is equipped with connections for sealing medium supply and relief. In principle, it is possible to apply a cooling medium to the cylinder wall, but uncooled cylinder housings can be used as well. The principal components of a large, two-stage rotary screw compressor are shown in Figure 15-3. Figure 15-4 illustrates typical rotor combinations incorporating an asymmetrical rotor profile. The profile combination 4 + 6 means that the male rotor has four teeth and the female rotor, six. Due to this profile combination, the diameter of
Rotating Positive Displacement Compressors 597
FIGURE 15-1 Modernrotary screw compressor. (Source: Aerzener Maschinenfabrik GmbH, D-3258 Aerzen, Germany.) the rotor core is relatively thick. This allows for operation with large differential pressures. Bearings
Although air machines are often equipped with rolling element bearings, the majority of compressors for process gas applications are furnished with journal beatings and thrust bearings of the type commonly found in centrifugal process gas compressors. The service life of these beatings is practically unlimited as long as proper lubricating and operating procedures are in force.
Seals In many rotary screw compressor applications, it is necessary to provide a sealing barrier between the process gas and the beatings. A number of different seal types are feasible: 9 carbon ring seals 9 barrier water floating ring seals 9 double-acting mechanical seals with stationary spring 9 combined floating ring/mechanical seals At the compressor input shaft, manufacturers often opt for: 9 labyrinth seals, or double-acting mechanical seals with rotating springs
9
598
Process Plant Machinery
FIGURE 15-2 Compression process in rotary screw compressor. (A) Suction intake: Gas enters through the intake aperture and flows into the helical grooves of the rotors, which are open. (B and C) Compression process: As rotation of the rotors proceeds, the air intake aperture closes, the volume diminishes, and the pressure rises. (D) Discharge: The compression process is completed, the final pressure is attained, the discharge commences. (Source: Aerzener Maschinenfabrik GmbH, D-3258 Aerzen, Germany.)
FIGURE 15-3 Principal components of a two-stage rotary screw compressor. 1 - Housing; 2 - male rotor; 3 - f e m a l e rotor; 4 - side plate on intake side; 5 - t i m i n g gears; 6 - graphite ring shaft-seal; 7 - o i l seal; 8 - radial bearing; 9 - a x i a l thrust bearing; 10-torsion shaft; 1 1 - drive shaft; 1 2 - step-up gears; 1 3 - oil pump; 1 4 - coupling; 1 5 - compensating piston. (Source: Aerzener Maschinenfabrik GmbH, D-3258 Aerzen, Germany.) Carbon ring seals with connections for the injection and education of neutral, clean gases are used in cases where leakage gas, even in connection with sealing gas, may enter into the beating areas or into the atmosphere. The gas pressure is relieved across floating carbon tings at the beginning of the seal chamber.
Rotating Positive Displacement Compressors
FIGURE 15-4 Asymmetrical rotor sets for modem Maschinen-fabrik GmbH, D-3258 Aerzen, Germany.)
rotary screw
compressors.
599
(Source: Aerzener
With barrier water floating ring seals, barrier water enters the seal chamber and a small amount of water reaches the compression space. Most of the water is returned to the barrier water system for cooling, filtration, and re-use. Barrier water seals are able to fully prevent gas leakage and can provide valuable cooling and scrubbing duties. Figure 15-5 depicts a flow diagram for a single-stage screw compressor with barrier water seals. A double-acting stationary spring mechanical seal and a combination mechanical and floating ring seal are primarily used for compression with high differential pressures.
600
Process Plant Machinery
o
o
10
!! I1
I
2 7
it 171
16
12
13
FIGURE 15-5 Water sealing flow schematic. 1 - Lateral compensator ss*; 2 - starting strainer; 3 - water injection; 4 - screw compressor; 5 - lateral compensator ds*; 6 - discharge silencer; 7 - non-return valve; 8 - safety relief valve; 9 - gear box; 1 0 - drive motor; 1 1 - c o u p l i n g ; 1 2 - o i i reservoir (base plate); 1 3 - o i l sight glass; 1 4 - l e v e l controller (oil); 15 - non-return valve; 16 - oil heating; 17 - oil drain; 18 - twin oil filter; 1 9 - twin oil cooler; 2 0 - oil pump with motor; 2 1 - gear box oil pump; 2 2 - safety relief valve (oil); 2 3 - barrier water controller; 2 4 - f l o w indicator; 2 5 - barrier water return; 2 6 - valve; 2 7 - slide valve; 28 - manometer. (*ss - suction side; ds - discharge side). (Source: Aerzener Maschinenfabrik GmbH, D-3258 Aerzen, Germany.)
Rotating Positive Displacement Compressors 601
Operating Principles for Oil-Injected Compressors Regardless of whether the screw compressor is executed for dry compression or oil injection, the gas is compressed in chambers progressively decreasing in size that are formed by the intermeshing action of the helical rotors and by the housing wall. However, oil-injected compressors do not incorporate timing gears. Instead, the driven male rotor interacts directly with the female rotor without use of timing gears. Oil injected into the compressor cavity provides intensive lubrication, and a large portion of the compression heat is absorbed. At the same time, the clearances between rotors and cylinder walls are filled with oil. This prevents the reverse flow of compressed gas and increases the overall efficiency of compression. At the compressor discharge flange, gas and oil exit through a check valve to the oil reservoir where most of the oil is separated from the gas. The remaining oil is separated in a downstream separator, and residual oil amounts of typically 5 parts per million (ppm) continue to remain in the gas stream. Oil carryover can be further lowered by downstream cooling and final moisture separation. The oil separation unit has to be properly maintained and the pressure drop across the separator cartridges taken into account to determine the overall performance of the compressor package. It should also be recognized that the efficiency of oil separation depends on the degree of contamination of the separator elements. The principle of oil separation is shown in Figure 15-6, and a typical oilinjected screw compressor is illustrated in Figure 15-7.
Principle
of oil
J
separation
inlet hlter
0
0
oil-cooler oil-filter
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oil r e s e r v o i r / o i l s e p a r a t o r
FIGURE 15-6 Principle of oil separation used with oil-injected rotary screw compressors. (Source: Aerzener Maschinenfabrik GmbH, D-3258 Aerzen, Germany.)
602
Process Plant Machinery
FIGURE 15-7 Oil-injected rotary screw compressor cross section showing slide valve for capacity control. (Source: Aerzener Maschinenfabrik GmbH, D-3258 Aerzen, Germany.)
Compressor Volume Control In principle, it is necessary to consider the problems of volume control for dryrunning and for oil-injection-type screw compressors separately.
Dry Screw Compressors Control by Variable Speed. In consequence of the fact that screw compressors displace the medium positively, the most advantageous method of achieving volume control is that obtained by varying the speed. This may be done in any of the following ways: 9 by variable speed electric motors 9 by use of a torque converter 9 by steam turbine drive Speed may be reduced to about 50 percent of the maximum permissible speed. Induced flow volume and power transmitted through the coupling are in this manner reduced in approximately the same proportion. The allowable turn-down depends on the adequacy of bearing lubrication at low speed and compressor discharge temperature. More than 50 percent reduction is possible in special cases. As mentioned earlier, there is no surge limit for these positive-displacement machines.
Rotating Positive Displacement Compressors 603 Bypass. Using this method, the surplus gas volume is allowed to flow back to the intake side by way of a compressor discharge pressure controller. An intermediate cooler brings the surplus gas volume down to intake temperature. Full-Load~Idling-Speed Governor. As soon as a predetermined final pressure is attained, a pressure controller operates a diaphragm valve that opens a bypass between the discharge and suction sides of the compressor. When this occurs, the compressor idles until pressure in the system drops to a predetermined minimum value. The valve will close once again on receiving an impulse from a pressure sensor. This brings the compressor back to full load. Suction Throttle Control. This method of control is suitable for air compressors only. As in the case of the full-load/idling-speed control method, a predetermined maximum pressure in the system, for example in a compressed air receiver, causes pressure on the discharge side to be relieved down to atmospheric pressure. Simultaneously, the suction side of the system is throttled down to about 0.15 bar absolute pressure. When pressure in the entire system has dropped to the permissible minimum value, full load is once again restored. 5crew Compressors Equipped With Oil Injection Suction Throttle Control. Since the final compression temperature is governed by the injected oil, a greater range of compression ratios, such as may arise when the induced volume is throttled down, can be safely coped with. This permits the main flow volume to be varied within wide limits. Built-in Volume Governor. Large compressors are frequently equipped with an internal volume-regulating device. By operating a slide valve (Figure 15-8) that is shaped to match the contours of the housing and that is built into the lower part of the housing, designed to move in a direction parallel to the rotors, the effective length of the rotors can be shortened. The range of this control mode is typically between about l0 percent and 100 percent. Compared with suction throttling, this type of control offers more efficient operation, as graphically represented in Figure 15-9.
FIGURE 15-8 Slide valve inside housing in partial capacity position. (Source: Aerzener Maschinenfabrik GmbH, D-3258 Aerzen, Germany.)
604
Process Plant Machinery
Capacity point of full load
power absorbed Pk l Olo ] 100 90 tO
I
',
/
~
/
)
/
_\~2/" /
7O 6O
I ~ 0
"
.
50 ~0 10 2o
/ . ' i final preS . . . . . i""
tO o
0
70
20
JO
40
SO
60
l
i
70
aO
90
100 r ' l . ]
capacity
Fall-off in power absorbed Pk expressed as a percentage of the full-load performance when operating at partial capacity and with the constants" theoretical curve curve obtained in actual practice . . . . . . . curve obtained in actual practice by switching over from say 20 % capacity to idling speed (p~ = P2 = 1 bar) . . . . . . . . curve obtained at idling speed (P~ = P2 = I bar). FIGURE 15-9 Capacity versus power curve pertaining to oil-injected rotary screw compressor. (Source: Aerzener Maschinenfabrik GmbH, D-3258 Aerzen, Germany.)
Calculation Procedures As was discussed earlier, the spaces in which the gas is trapped and progressively advanced are those formed between the cylinder walls and the interlocking convolutions of the two helical rotors. The position of the edge of the outlet port determines the so-called "built-in volumetric ratio," vi. The "built-in compression ratio," zri, results from the equation 7/" i = v X . * The compression process is shown in the theoretical p-v-diagram (Figure 15-10). The induced flow volume may be calculated for any compression ratio, provided the data applicable to the particular compressor being considered are known. One revolution of the main helical rotor conveys the unit volume, q0 (liter/rev). This gives us the theoretical induced flow volume, Q0, at n revolutions: n . q0 [m 3/min] Q0 = 166-6 The actual induced flow volume, Ql, is lower by the amount of gas, Qv, flowing back through the very small clearances. Thus QI -- Q0 - Qv [m3/min] * For a description of metric symbols used in this chapter, refer to Table 15.1 (page 610).
Rotating Positive Displacement Compressors 605
I
4--
p Ibar]
/
w
~orking compression ratio > ni
3-built-in compression ratio = n i
working compression ratio <
2-
\ I-
0--
0
0,2
0,6
0,4
018
i,0
V2/%
FIGURE 15-10 Pressure-volume diagram for rotary screw compressor. (Source: Aerzener Maschinenfabrik GmbH, D-3258 Aerzen, Germany.)
The slip loss volume Qv is mainly dependent on the following individual factors" 9 total cross section of clearances 9 density of medium handled 9 compression ratio 9 circumferential speed for rotor built-in volumetric ratio Volumetric efficiency is 0v =
Q!
Q0
=1
Qv
Q0
The theoretical power input required to compress the induced flow volume, Ql, is 10-3 Pth : where
hlad
--
60
" Pl 9Q0" hlad [kW]
the amount of energy required for the adiabatic compression of 1 kg of gas from Pl to P2
The theoretical power input requirement is increased by the dynamic flow loss, Pdyn, and by the mechanical losses, Pv. The latter consist of the losses in bearings, timing gears, and step-up gears. Thus, the power transmitted through the coupling is Pk : Pth + Pdyn -+- Pv [kW] Screw compressors are manufactured to American Petroleum Institute (API) 619 and/or VDI 2045 standards with a permissible tolerance of -t-5 percent in regard to power requirements and induced flow volume. These tolerances result from
606
Process Plant M a c h i n e r y
3 \
2
1
\
16
//
17
/ J
\, ,,\
i _~.
", \
/ I~ t.~tL~, i / I I r1"rl r1"~r--1 . . .a.l.. / ./ . . .~.
11
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11
/
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12 -I ./12
v
8j
2s 13
/
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14
21
FIGURE 15-11 Flow schematic f o r a two-stage rotary compression system. 1 - Slide valve; 2 - lateral compensator; 3 - intake silencer; 4 - starting strainer; 5 - screw compressor 1st stage; 6 - d i s c h a r g e silencer; 7 - g a s cooler; 8 - s e p a r a t o r ; 9 - l a t e r a l compensator; 10 - starting strainer; 11 - screw compressor 2nd stage; 12 - discharge silencer; 13 - non-return valve; 14 - slide valve; 15 - control and shut-off devices; 16 - g e a r box 1st stage; 1 7 - g e a r box 2nd stage; 1 8 - drive motor; 1 9 - n o i s e abatement hood (not shown); 2 0 - noise abatement hood (not shown); 21 - oil system; 22 - sealing water system; 23 - injecting water system; 24 - condensate tank 1; 25 - condensate tank 2; 26 - condensate tank 3. (Source: Aerzener Maschinenfabrik GmbH, D - 3 2 5 8 Aerzen, Germany.)
Rotating Positive Displacement Compressors 607 inaccuracies introduced during the manufacturing process. The final compression temperature is calculated for a dry-type compressor as follows: tZth = tl + Atth [~ Atth = T1
[() ] P2
--7-
--
Pl
-
I
1 - -
fly
[~
When operating on the oil-free, dry-running principle, a screw compressor may attain a maximum final compression temperature of 250 ~ When air is the medium handled, this temperature (isentropic exponent X = 1.4) corresponds to a compression ratio of P2 ~ 4 . 5 Pl On the other hand, gases with a X - 1.2 will permit, within the temperature limits mentioned, a compression ratio as high as P__~2~ 7 Pl In the case of a screw compressor operating on the principle of oil injection, most of the drive energy applied to the machine is removed by the oil in the form of heat. The amount of oil injected is adjusted to ensure that final compression temperatures of approximately 90~ are not exceeded. When taking in air under atmospheric pressure, compression ratios as high as P2/Pl ~ 21 are obtainable.
Typical Flow Schematic Figure 15-11 represents a flow schematic for a two-stage rotary screw compression system.
ROTARY PISTON BLOWERS Rotary piston, or lobe-type, blowers derive from the Roots compressor principle and have been built since 1864. They are used in a large variety of process plant applications, including pneumatic conveying of bulk materials, pressurized aeration of water at treatment plants, creation of vacuum, and gas movement in the petrochemical, pharmaceutical, and metallurgical industries. They range in size from fractional horsepower to literally hundreds of kilowatts.
Design and Construction Rotary piston blowers are twin-shaft rotary machines. The two rotors are axially parallel to one another and located centrally inside the casing. The timing gears ensure that the rotors turn without contact. The rotors are supported on ball and roller bearings. The clearance between the rotors is kept to a minimum and selected for the expected pressure differential and thermal load under working conditions. Smaller rotors are adjusted in such a manner as to enable them to be run in either flow direction. Large blowers and gas blowers of special design are suitable for flow in one direction only, since in this case the clearances between the rotors and the casing must be kept larger on the low-pressure side to compensate for
608
Process Plant Machinery
shaft deflection and bearing clearances. Axial thermal expansion of the rotors is compensated for by larger clearances between the rotors and the end plates at the free bearing end, i.e., at the bearing that is permitted to slide so as to accommodate thermal growth. The shaft diameter is a very important factor in the evaluation of rotary lobe blower quality. It determines the amount of deflection and thus the magnitude of clearance and volumetric efficiency under load. Adequate bearing span provides space for proper sealing components between compression chamber and bearing housings. Needless to say, well-designed seals prevent contamination of gas by the lube oil and vice versa. This extends the life of bearings and gears. The principal construction features of rotary piston blowers include driving and driven rotors, timing gears, beatings, and seals (Figure 15-12).
Method of Operation Two symmetrical rotary pistons revolve in opposite directions timed to one another. The medium to be conveyed flows into the blower housing that surrounds the two rotors. From there, it is moved via positive displacement in the chambers formed between the rotors and the housing toward the discharge side. At the instant when one rotor tip passes the edge of the discharge port, the gas is compressed by the back flow from the discharge port as can be visualized from Figure 15-13. The final pressure adjusts itself to the pressure loss in the piping and in the plant equipment downstream from the blower. The capacity of a given blower can be calculated for all types of gases and for every possible load condition. Each revolution of the rotors causes four separate volumes q0/4 (liters/revolution) to be conveyed and compressed. The power transmitted through the blower coupling is P = Pth --b Pv(kW) Dnvenrotor
....
Cylinder(conveyingchamber)
,,. ~t ..
.t Unfixedbearing
Fixedbeanng Timinggears
~ ~ l ~
~
splaP;sl~ r:ngai
Oil thrower Gearcasing
__ ~ideplate
;~i;ing ;r~ Drivingrotor
FIGURE 15-12 Principal parts of a rotary piston blower. (Source: Aerzener Maschinenfabrik GmbH, D-3258 Aerzen, Germany.)
Rotating Positive Displacement Compressors
rl
FIGURF 15-13 Operating principle for rotary piston blower. (Source: Aerzener Maschinenfabrik GmbH, D-3258 Aerzen, Germany.) The main component, the theoretical power for compression, is thus independent of the type of gas involved and directly proportional to the working pressure differential and to the blower speed. Since no compression occurs internally, power absorbed when running under no-load conditions is nearly equal to the power loss, Pv. This will be approximately 3 to 5 percent of the power transmitted by the coupling when running on full load. Rotary piston blowers are typically manufactured with a tolerance band of 4-5 percent, referred to the power consumption and the intake volume. These tolerances are composed of the sum of all manufacturing tolerances. This results in the theoretical capacity n ' q 0 (m3/min) Qo = i66-6 The actual capacity is obtained by taking the theoretical capacity and reducing it by the amount of gas, Qv, flowing back through the clearances: Ql = Q0 - Qv (m3/min) The clearance losses depend on the specific density of the gas at intake, Ap, and on the total area, F, of the clearances. Volumetric efficiency is Ql 0v=~=l
Q0
Qv
Q0
Since the rotor clearances are extremely small, efficiency under working conditions is high. The capacity varies very little with changing loads (Figure 15-14). The amount of power needed to compress the capacity, Ql, is theoretically Pth -
Qo" Ap (kW) 600
This power is in actual fact further increased by mechanical friction in beatings, timing gears, and sealing elements, and also by the dynamic losses occurring within the blower parts and in the compression chamber.
609
610
Process Plant Machinery lOOO
/
E . ~a~_,400 ~ o
o~ 200 ,
\
0""
5
10 15 20 25 Intake volume flow Q~ [mVmin]
30
FIGURE 15-14 Volumetric characteristics of a small rotary piston blower. (Source: Aerzener Maschinenfabrik GmbH, D-3258 Aerzen, Germany.) TABLE 15.1 Nomenclature Typically Used for Rotating Positive Displacement Calculations
Symbol h X n p~ P2 Pe
--Pe Pth Pv
Pdyn Pk qo Qo Qv Ql t~ t2
Atth TI vi 0v zr~
Unit
Meaning
J/kg min-~ bar bar bar bar kW kW kW kW 1/U m 3/min m3/min m3/min ~ ~
Specific adiabatic work of compression Ratio of specific heats Speed of rotation Suction pressure Discharge pressure Compression gauge pressure Negative vacuum pressure Theoretical power input Mechanical losses Dynamic losses Power transmitted through coupling Unit volume Theoretical induced volume Slip loss volume Actual induced volume Inlet temperature Discharge temperature Theoretical increase in temperature Absolute inlet temperature Volumetric ratio Density Volumetric efficiency Compression ratio
~ K kg/m 3 -
Capacity Control The capacity of rotary piston blowers can be controlled by either varying the speed or by bypassing flow from discharge back to suction. Speed variation is, of course, the more efficient method, and turndown capacities of 70 percent can be reached in some cases. Power demand and pressure rise are almost directly proportional to blower speed.
APPENDIX 15A
Water-Flooded Single Screw Process Compressor
Technology*
The origin of single screw process compressor technology lies in two diverse fields- naval shipboard air compression applications and natural gas applications. It is the naval shipboard air application which fostered the water flooded single screw technology. The combination of water injection and water lubricated bearings has made possible totally oil free gas compression. Since 1992, significant opportunities for the application of this technology have been recognized in the process gas market. Certain gas compression applications exist in which oil contamination of the gas is unacceptable. Many of these applications also require that compressor components exhibit a high degree of corrosion resistance. These technological challenges have been met with the development of a family of water flooded single screw process compressors manufactured entirely from austenitic stainless steel. Other materials are available which can be applied to a wide range of operating environments, from relatively inert to highly corrosive. In concert, the inherent flexibility of the single screw compressor design, the range of materials of construction and the prospect of totally oil free gas compression provides an opportunity to pursue new applications in the process gas market. This compressor class represents an improvement in the technology available to the markets served.
Fig. 15A-1
THEORY OF OPERATION The single screw process compressor is a positive displacement, high speed, clearance critical, rotary machine. There have been four different variations of the single screw machine. The "PP" design utilizes a planar screw and planar gate-rotors. The compression grooves are located on the front and back faces of the main screw. Two gate-rotors are used for each side (4 in all) and the shafts for these gate-rotors are all parallel each other. The "PC" design has a similar planar main screw but utilizes 2 cylindrical gate-rotors. Compression occurs on the front face of the screw. The gate-rotor shafts are oriented at opposing angles from the axial face of the main screw. The third design "CC" has both cylindrical gate-rotors and a cylindrical screw. Compression from the two gate-rotors occurs on the top and bottom face of the main screw. As with the "PC" design, the gate-rotor shafts are oriented at opposing angles. The final design, and the one discussed in this appendix and referred to as the single screw process compressor, is the "CP" design. This compressor has a cylindrical main screw which meshes with two planar gate-rotors (Figure 15A-1 and 2). The gate-rotor shafts are parallel to each other with axes perpendicular to the axis of the main screw. Compression occurs in the volume created when the gate-rotors mesh with the main screw. This volume, along with tight screw to housing clearances, forms a compression chamber. The cycle begins as suction gas is drawn through the front of the compressor
Single screw process compressor - driveline section. (Source: Dresser-Rand, Olesn, NY.)
* Source: Dresser-Rand, Olean, NY (Paper by F. Heidrich). Adapted with permission.
612 Process Plant Machinery
I
m
Flg. 15A-2
Slngle screw process compressor - gate-rotor sectlon. (Source: Dresser-Rand, Olean, NY.)
and fills the exposed grooves of the main screw (Figure 15A3). As the screw rotates, the gas becomes trapped within the three sides of the screw, the cylindrical casing and the gaterotor. Further rotation causes the groove volume to diminish which increases the internal gas pressure. Near the end of the compression cycle, a machined port in the housing is exposed and the gas is discharged into a plenum. As the groove discharges on one side of the gate-rotor, the same groove is filled with suction gas on the opposite side and the process is repeated. By means of a viscoseal on the high pressure end of the screw and equalization holes in the housing, the main screw is subjected to suction pressure on both ends. Since the front and rear faces of the main screw are nearly equal in surface area, axial thrust loading is substantially reduced. In the absence of this thrust load, Llo life for the main drive bearings increases dramatically. This feature extends the continuous running life of the compressor and enhances its overall reliability. During the compression process, water is injected into the compression groove. Water enters the compression groove through fixed port holes drilled into the housing. The water acts as a sealant to increase efficiency and a coolant to remove the heat of compression. The injected water is directed through the discharge port with the gas mixture. A liquid separator is used to remove the gas from the water. The compressed saturated gas is delivered to the customer's process line while the separated water is cooled, filtered, and recycled through the compressor again. This closed-loop injection water system is typical for all compressor sizes. The injection water system not only supplies both sides of the main screw but also lubricates the compressor bearings.
DESIGN CONSIDERATIONS Gate-Rotor Support Loading The planar gate-rotor rides on the horizontal face of a gate-rotor support. This gate-rotor support rotates about a stationary shaft. The gate-rotor support is subjected to
load fluctuations throughout the load cycle. For example, at any time, suction pressure may be acting on one gaterotor tooth, an intermediate pressure may be acting on an adjacent tooth, and discharge pressure can be acting on its adjacent tooth. This combined loading must be supported by the gate-rotor bearings. A computer program has been developed to determine this combined loading. The program first determines the relevant geometric properties of the screw, including volumes and areas at each rotational step. Once this information is stored, the program determines the pressure-volume profile of the compression process. Leakage and water injection effects are incorporated into the model. Next, the program determines the number of teeth exposed to the gas at any given time during the compression. Then the pressures stored for that time are multiplied by the exposed area for the tooth. These forces are totaled for the load step and the program continues. Once the cycle is complete, a diagram depicting the loads is generated (Figure 15A-4). With this information, loading on the gate-rotor support can be determined. The average load is converted into radial and thrust loads acting on the bearings (Figure 15A-5). For the single screw process compressor, two fixed profile bearings are used to support the gate-rotor support as it rotates about the stationary gate-rotor shaft. The top bearing is referred to as the free end bearing while the bottom bearing is the thrust bearing.
Water Lubrication The gate-rotor bearings are lubricated by the system's closed loop water. Using water as a lubricant has a number of benefits. A costly sealing mechanism is not required to seal the injectant from the lubricant. The bearings have relatively cool running temperatures due to the high specific heat capacity of water. However, water lubrication presents a number of design challenges also. The two most significant are the low viscosity of water and its relatively poor properties in the boundary lubrication regime. 1 Laboratory tests have been
Appendix 15A: Water-Flooded Single Screw Process Compressor Technology 613
Fig. 15A-3 Single screw compressor compression cycle - CP design. (Source: Dresser-Rand, Olean, NY.) performed to determine the optimal combination of bearing and journal materials that can withstand the water environment. The single screw process compressor utilizes a hardened stainless steel journal with a graphite grade bearing material. This material has proven to be resistant to the localized water surface velocities and soft enough to interact as a mating surface for the journal. Alternate grades of material have been evaluated in terms of corrosion resistance and could be used if the conditions of the application warranted. From a design point of view, standard design calculations are used for water lubricated journal bearings. It is important to note however, that the bearing lubricant is the same water used in the process. Therefore, it is imperative to fully determine the fluid and thermodynamic properties of the gas-liquid solution. Once this is known, standard design practice 2.3 can be followed.
Sealing The single screw process compressor utilizes a number of different sealing mechanisms to improve efficiency and prevent process gas leakage to the environment. Three primary seals are employed; the first one located on the highpressure end of the main screw and the other two located behind the screw on the drive shaft. Seals are not required for the gate-rotor supports or the suction end of the screw. The first seal is a spiral groove seal, or viscoseal. The primary function of the viscoseal is to increase efficiency by reducing circumferential and axial leakage from the compression grooves. Thus it is located on the high-pressure end of the main screw. Very often, this seal is confused with a labyrinth seal. In operation, these two types of seals are very different. The labyrinth seal is a widely used non-contacting shaft seal. Labyrinth seals are based on positive, finite running
614 Process Plant Machinery
Flg. 15A-4 Typlcal gate-rotor loadlng. (Source: Dresser-Rand, Olean, NY.)
Flg. 15A-5 Gate-rotor loadlng. (Source: Dresser-Rand, Olean, NY.) clearances which are sufficiently large to preclude the possibility of contact between the seal and shaft. This type of seal is effective by reason of generation of eddies within the cavities created by the grooves. This results in a highly turbulent flow path and inefficient flow process which produces low leakage. Similar to a labyrinth seal, the viscoseal relies on tight running clearances, but it does not depend upon pressure drops across its grooves. A viscoseal is a dynamic seal and is comprised of continuous helical groove cut at a prescribed pitch. Relative motion between the main screw, on which the viscoseal is machined, and the housing induces viscous, or Couette drag, flow in the helical grooves of the seal. 4 A viscoseal can have three possible operating modes; (a) a zero-leakage seal, (b) a leaking seal and (c) a pumping seal. A zero-leakage seal is classified as having sufficient pumping
action to retard the forcing flow. This does not, however, indicate 100% sealing. Liquid droplets will leak from the seal due to instabilities at the interface of the two opposing flows. A leaking seal is classified as a seal which can not overcome the forcing flow. This may be due to excessive clearances or a relatively short seal length. The last type is a pumping seal. This seal occurs when the pumping action caused by the viscoseal is much greater than the forcing flow. It is in this case where reverse flow may occur. Due to the high reliability of the mechanical face seal located directly behind the viscoseal, the single screw process compressor utilizes a minimum flow leaking viscoseal. There are many methods available to model the behavior of a viscoseal. Typically, the designer will need to determine the
Appendix 15A: Water-Flooded Single Screw Process Compressor Technology 615 magnitude of the two opposing flows based on the physical dimensions of the seal and housing. For the forcing flow, or the liquid that is being sealed, bushing-flow equations can be used: = 7rDh3
(~ushing----~(~-~-~)Gzr (see p. 618 for Nomenclature) To model the drag flow induced by the viscoseal;
evs= ~Oh3r (~"~-~)[ (1Gzr F3tan2 ~Gxg ] /.i. - c~) + - 0.57rDvh, r tan 0(F - 1 )
As evidenced by the equations, viscoseal flow is highly dependent on the geometry of the seal. Other factors such as eccentricity, Reynold's number and rotative speed are considered during the design of the seal. An effective viscoseal is determined when the two equations are equal. 4 The next seal within the single screw process compressor is a mechanical face seal on the main shaft behind the screw. The purpose of this seal is to prevent gas and/or fluid leakage from migrating along the shaft and into the environment. The seal utilizes two mating surfaces that form a tight seal during operation. As with the bearings, a significant amount of research went into material compatibility issues. A typical seal will have one relatively soft surface riding on a substantially harder material. For the single screw process compressor lineup, both materials are impervious to corrosive gases and process fluids. The third line of defense is a vent/purge cavity to which the customer can attach a line to safely remove and leaking gas from the cavity or purge the seal cavity with an inert gas. This cavity will come into play if there is a failure of one of the previous seals. Finally, directly behind the vent/purge cavity is one additional seal. This seal protects the bearings from becoming contaminated with any leaking gas. For most processes in which this compressor may be applied, it is imperative the process fluid and/or gas does not enter the oil flooded drive bearings.
Dlscharge Temperatures An appealing feature of single screw compressors is the ability to compress large ratios in a single stage. Typical applications range from 2 - 1 6 ratios, yet higher ratios are achievable. The liquid injection in the compression process enhances the machines ability to reach these pressure ratios. For a reciprocating compressor, the discharge temperature is dependent upon a number of different factors; i.e. lubrication, friction, valve losses, piston leakage, suction preheat and cylinder cooling. However, a theoretical temperature can be estimated using the isentropic relationship:
m2=m11~11)ly-~-yll Using this estimator, reciprocating compressors reach fairly high (>149~ discharge temperatures after 3.5-4.0 pressure ratios. This heat has detrimental effects on compressor efficiency, as well as, negatively affecting the overall life of the compressor components. The discharge temperature for a single screw is determined differently. Since water is constantly injected into the compression chamber, a significant amount of heat is removed from the gas by the water. The relationship can be modeled by equating the enthalpies of the gas and liquid. Using the energy equation, the indicated horsepower, or the power consumed to compress the gas, is a direct function of the change
in enthalpy of the gas plus the change in enthalpy of the fluid. In equation form, Power =
mgasCpgas A Tgas +
mcoolantCpcoolant~, Tcoolant
For water, the measured discharge temperature is 3 - 5 ~ greater than the water injection temperature. In fact, with the water-flooded single screw process compressor, the discharge temperature of the gas can be less than the suction temperature. This eliminates the need for an expensive gas aftercooler.
Efficlencles The water-flooded single screw process compressor utilizes a fixed discharge port design. The groove volume is reduced to the same value for all applications. Despite this, efficiencies for the water-flooded compressor have shown to be quite good. Unlike conventional compressors, the heat of compression is not a factor in the overall efficiency of the unit. The single screw process compressor has volumetric efficiencies ranging from 95% at low ratios (2-6) to 80% at higher ratios (> 18). Discharge port location is determined through the use of a computer program. The single screw process compressor has been designed to offer the most efficient compressor for the widest range of applications and gases. As with most fixed port screw compressors, there are slight inefficiencies when the discharge pressure is reached before and after the discharge port is exposed. Testing has shown however, that the overall machine efficiency is still very good. The nearly isothermal compression process minimizes volumetric losses noted in other positive displacement compressors. In addition, the single screw has the added benefit of not having compressed gas trapped in clearance pockets during suction. The flow through a reciprocating cylinder is reduced until the trapped pressurized gas is re-expanded. This explains why the performance of the single screw process compressor is not heavily dependent on compression ratio.
Deslgn Program Throughout this segment, references have been made to algorithms used to determine machine geometrics and thermodynamic properties. A single screw process compressor design program has been written to handle these calculations. The user inputs the machine type, coolant and gas mixture. From this information, the program first determines the area and volumes and at all steps during the compression cycle. This information is then stored. Next, a separate gas analysis is performed to determine all relevant thermodynamic properties, i.e. critical temperature, critical pressure, volumetric exponent, as well as, enthalpies and entropies for each state. Next, a routine utilizes the stored machine geometry, coupled with appropriate flow equations to determine the coolant requirement for the application. Next, the pressure-volume trace is calculated and stored. Once this is known, bearing loads and required bearing flowrates are determined. Finally, the program converts the numbers into the required units and reports the results. A variation of this program has been written to be used as a compressor sizing tool. The variant utilizes the same routines but is geared towards sizing a known compressor based upon a customer's requirements.
SAMPLE APPLICATION One application in which water-flooded single screw process compressors have made inroads is the compression of oxygen/ozone for medium consistency pulp bleaching systems. 5 In this system, the main function of the compressor is to draw gas from an ozone generator, reduce its volume to a desired phase (or volume) ratio and deliver the compressed gas to a mixer. Typical pressures are 0.2-0.7 barg delivered from the generator compressed to 7.0-12 berg. The unique
616 Process Plant Machinery characteristics of ozone make near isothermal compression a requirement. The traditional compressor for ozone is the liquid ring compressor. In this machine, a curved impeller, rotating within a housing, causes a water and gas mixture to move towards the center of the rotating ring. As this happens, the movement of the water compresses the gas. On the inside ring portion of the casing, a port allows the compressed gas to be discharged. The liquid ring can generally compress only 4 ratios per stage, therefore two or more stages may be required in an application where a single stage single screw process compressor may be specified. The liquid ring is used for ozone compressor because of its reduced discharge temperatures and low maintenance. The single screw process compressor exceeds these requirements by offering reduced discharge temperatures, higher isentropic efficiency (lower specific horsepower), a compact design and the ability to compress up to 20 ratios in a single stage. Furthermore, virtually no degradation of the ozone gas occurs across the compressor. The water-flooded single screw compressor has been shown to be an efficient rotary, positive displacement compressor. Its pressure balanced compression cycle helps maintain smooth operation and extends component life. The single screw process compressor comes as a total package complete with separator, motor and filter. Installation requires only process piping and electrical connections. Due to its high resistivity to corrosive gases and unfriendly environments, the single screw process compressor is an excellent candidate to fill many petrochemical compression requirements. NOMENCLATURE
Cpcoolant
Cpgas
D
Gzr Gxg hr
specific heat of coolant specific heat of gas bearing diameter flow correction factor flow correction factor radial land clearance
L mcoolant mgas P P1 P2 QDushJng
Qvs
T1 T2 A Tcoolant A Tgas F y # v
bearing length mass flow rate of the coolant mass flow rate of the gas bearing unit load inlet gas pressure discharge gas pressure volumetric flow through a bushing volumetric flow through a viscoseal inlet gas temperature discharge gas temperature change in coolant temperature change in gas temperature ratio of groove to land widths helix angle ratio of groove to land clearance volumetric exponent fluid dynamic viscosity fluid kinematic viscosity
REFERENCES
1. Smith, D.M., Journal Bearings in Turbomachinery. Chapman and Hall, (1969): p. 147. 2. Hersey, M.D., Theory and Research in Lubrication. John Wiley & Sons (1966): p. 190. 3. Wilcock, D.F. and Booser E.R., Bearing Design and Appfication. McGraw-Hill, 1957. 4. General Electric Co., Viscoseal (Spiral Groove Seal), Seals
Design Guide, Study of Dynamic and Static Seals for Liquid Rocket Engines, July 1968- September 1969.
5. Rounsaville, J., Breed, D., Nakamoto, S., Griggs, J., High
Concentration Ozone-System Design & Economics.
Ingersoll-Rand Co., Sumitomo Precision Products Co., Ltd, Dresser-Rand Co., 1994 TAPPI Pulping Conference AIChE Session.
Chapter 1 6 Mixers and Agitators Many process operations involve fluid mixing. Table 16.1 lists basic classifications, including liquid-solid mixing, gas-liquid mixing, liquid-liquid mixing, blending of miscible liquids, and fluid motion. In practice, most mixing operations involve several of these. As shown in Table 16.1, process performance can be judged by physical uniformity, determined by taking samples and calculating the degree of uniformity produced, or by the time it takes to achieve a certain degree of uniformity. When choosing mixing equipment, the viscosity of the fluids and mixtures is an important consideration, as is the density of materials and the resulting density of the mixture of materials involved. Mechanical mixing equipment can be categorized according to (1) the kinds of impellers used to suit specific process requirements; (2) the types of mechanical drives required to accommodate power, speed, and shaft length; (3)the necessity for sealing the tanks against high pressures imposed by certain processes. Another consideration is the requirement of stabilizing devices on the impellers as well as steady bearings in the tank. However, we will first take a look at some fluid mechanics involved in the mixing tank and some of the parameters that must be measured and o b s e r v e d - the mixers and agitators. IMPELLER FLUID MECHANICS
A prime consideration is the power drawn by the mixer. In the average plant it is often necessary to measure horsepower, usually by means of electrical measurements for electric motor-driven equipment. Other types of drives are possible; these would include hydraulic motors, air motors, and steam or gas turbines, but electric motors are by far the most common. The most reliable way of measuring mixer power is by means of a recording wattmeter. Clamp-on ampmeters can be used to give an approximation, but they usually must be ratioed to full-load nameplate amperage and may not take into account voltage fluctuations that could exist in plant operation. When taking a wattmeter reading, the electrical efficiency of the motor must be considered. Alternatively, a motor curve from the motor manufacturer would be helpful in relating amperage readings to the actual output horsepower. Motor manufacturers and mixer suppliers normally have these curves available for different horsepower ratings, different types of enclosures, and varying starting characteristics for motors. The curves would indicate average values expected for particular motor types. If a more accurate reading is desired, an individual motor curve should be considered. Most motors of 100 HP and over have had these electrical characteristics recorded, and motor manufacturers can often produce a motor curve for a particular motor on a given installation. * Source: LIGHTNIN, Rochester, NY. Adapted by permission 617
618
ProcessPlant Machinery TABLE 16.1 Mixing Processes
Physical Processing
Application Classes
Chemical Processing
Suspensions Dispersions Emulsions Blending Pumping
Liquid-solid Liquid-gas Immiscible liquids Miscible liquids Fluid motion
Dissolving Absorption Extraction Reaction Heat transfer
Source: LIGHTNIN, Rochester, NY. In any event, a subtraction must now be made for mechanical efficiency in the gear. Typically, gear reducer efficiencies for spiral bevel or helical gears are about two percent of rated horsepower per reduction. Another subtraction must be made if there is a stuffing box or mechanical seal involved in the equipment. This estimate usually must be obtained from mixer or seal manufacturers. The resulting quantity is the shaft horsepower, and all this horsepower is transformed into heat in the mixing vessel. It is not recommended that no-load readings be taken as a subtractor for wattmeter average readings. Motor efficiencies are often very low at no-load readings, and these give errors that are too large for accurate calculation of shaft horsepower. All the power produced in the fluid is proportional to the pumping capacity (Q) and impeller velocity head (H). This can be expressed as P c x QH. However, since the typical mixing system does not incorporate a casing around the impeller, the definite pumping capacity must be through some arbitrarily chosen discharge area of the impeller. Considerable variation exists in practice in this regard, and a particular definition must be obtained when observing pumping capacity data from mixer manufacturers. In addition, the velocity head is not easily measured, so this too is a quantity that cannot be readily related to the shaft horsepower. In concept, these two quantities are extremely important. The pumping capacity circulates fluid throughout the tank, and the process entrains other fluids with it; so the total flow of the tank may be anywhere from a few percent to as much as ten times higher than the actual flow from the impeller. The velocity head is related to the fluid shear rate of the system. The fluid shear rate is a velocity gradient and is the only means by which particles get together in the mixing system. If it were not for shear rate, the particles would never meet each other; each one would go around in its own velocity pattern, which is identical to the pattern of every other particle. Shear rate is an important factor in many processes, and process engineers must carefully consider the shear rate required and make sure that it is compatible with the process result desired. When low-viscosity media are mixed in tanks that are unbaffled, they tend to swirl and will produce a vortex, as shown in Figure 16-1. Sometimes this action is desirable, but usually it is not. Hence, four wall baffles are often used to produce the flow pattern shown in Figure 16-2. These baffles are usually 1/12 of the tank diameter in width. In calculating the strength for supporting the baffles, the mixing tank must resist all be applied by the mixer. Torque is the power divided by the impeller speed or the torque consistent units. Torque(in_lb ) =
HP(63,000) N (RPM)
The baffles must resist this torque, allowing us to divide the torque by the number of baffles, and then by the tank radius to the center of the baffle. This gives the actual torque acting on each of the baffles. Although this torque is distributed
Mixers and Agitators
BOTTOM
SIDE FIGURE 16-1
Typical swirling condition of tank without baffles. (Source: LIGHTNIN
Rochester, NY.)
:Ji
,r -
,!
.-BAFFLES
l |
l
tl /
BOTTOM VIEW
SI DE VIEW MAXIMUM SHEAR RATE
FIGURE 16-2 Typical flow pattern with axial flow impeller tank with wall baffles. (Source: LIGHTNIN, Rochester, NY.) over the length and width of the baffle, it is acceptable to assume that the torque operates in equal amounts at the location of the various impellers in the system. This concentrated load at the midpoint can be used to calculate the baffle thickness and the strength of the support arms required.
Tank Shape Figure 16-3 depicts the nomenclature used in explaining mixer layout and design. Most tanks are cylindrical and typically have a liquid height over tank diameter (Z/T) ratio of about 1.0. On occasion, tanks have very low Z/T ratios of 0.2 to 0.4, which is typical of the large storage tanks in the petroleum industry. On the other hand, tanks can be very tall and slender. Multiple impellers must be used on large, tall tanks, while single impellers are used on tanks with Z/T ratios of less than approximately 0.8. While the least power is usually required for tanks that have a Z/T ratio of 0.6, the specifying engineer must investigate some additional parameters in an effort to optimize the entire installation. He or she should be aware that tanks with a Z/T ratio of 0.6 are more costly to build than tanks having a Z/T ratio of 1.0, and this may offset the power savings. Square tanks are often used, as are rectangular tanks. Up to a power level of about 1 HP/1000 gallons, these tanks are self-baffling and do not require additional
619
620
Process Plant Machinery
k..,A.A.A.A.A.AJ
k.A.A.A.X.A.AJ
Z --I I
/
DI C
T
-
T
FIGURE 16-3 Nomenclature employed in typical mixer applications. (Source: LIGHTNIN Rochester, NY.) baffles. Above that level, baffles are required, as shown in the drawing. Four baffles are typically used in square tanks and two baffles in rectangular tanks. There are also tanks with very complicated shapes, including elliptical heads, spherical heads, spherical tanks, and horizontal cylindrical tanks.
Impellers
There are two basic kinds of impellers, radial flow and axial flow. Figure 16-4 shows a spectrum of flow and shear rates for different impeller types. It will be noted that axial flow impellers are typically high-flow, low-shear devices, compared with the radial flow impellers. The axial flow device shown in Figure 16-5 has historically been the most common impeller used for axial flow situations. Axial flow impellers incorporate blade angles of about 45 ~ to the horizontal, although this angle can vary between 5 ~ and 60 ~ An impeller diameter over tank diameter (D/T) ratio of 0.3 to
RAKES. GATES SPIR A LS.ANCHOR.P ADDLE
""
HP
"
F
PROPELLER A X I A L FLOW TURBINES FLAT BLADE TURBINE BAR TURBINE.BLADELESS IMPELLER.IMPELLER & STATOR (CLOSE CLEARANCE! COLLOID MILLS.HOMOGENIZER
/
/
HEAD
FIGURE 16-4 Chart showing difference between flow and shear. (Source: LIGHTNIN Rochester, NE )
Mixers and Agitators
0.5 is quite common, and these ratios are particularly applied to areas where high flow is needed, such as blending and solids suspension. All impellers have a Reynolds number power curve, as shown in Figure 16-6. This allows the calculation of the Reynolds number when viscosity is known, the
FIGURE 16-5 100
Typical axial flow turbine. (Source: LIGHTNIN Rochester, NY.) I
I
i
I
I
Pg pN3D 5 10
.,,..
INE BAFFLED
1.0
-PROPELLER SQUARE PITCH0.1
I
10
10 2
10 3
C = I D D2N_.__._ppZ - T D IMPELLER DIAMETER N IMPELLER ROTATIONAL SPEED p9 LIQUID DENSITY
10 4
D/T
10 5
1 0 6/
= I/3
,U, LIQUID VISCOSITY P POWER g GRAVITY CONSTANT
FIGURE 16-6 Reynolds number-power number curve for two different impeller types, radial low and axial flow. (Source: LIGHTNIN, Rochester, NY.)
621
622
Process Plant Machinery
FIGURE 16-7
Typical flat blade turbine. (Source: LIGHTNIN, Rochester, NY.)
FIGURE 16-8 Typical high-speed disc-type radial flow high-shear impeller. (Source: LIGHTNIN, Rochester, NY. )
Mixers and Agitators
Reynolds number being the product of the impeller speed, the impeller diameter squared, and fluid density divided by the fluid viscosity. This has to be in consistent units. Having established the Reynolds number, Figure 16-6 can be used to obtain the power number, which is the power times g divided by density, speed cubed, and diameter to the fifth power. From this, the power can be calculated. These curves must be available for the particular geometry of the impeller, baffles and tank configuration. The fiat portion of the curve depicts the turbulent region, which is commonly for low-viscosity materials. There is also a viscous region in which the slope is minus 1, with a transition area in between. Radial flow turbines are normally used where higher shear rates are required and where lower pumping capacity is needed. Radial flow turbines include the fiatblade turbine shown in Figure 16-7 and high-speed disc turbines for high-speed applications (Figure 16-8). An axial flow airfoil impeller (Figure 16-9) has higher pumping capacity and lower shear rates than does the axial flow turbine and is particularly suited to blending and solids suspension. It is also particularly suited for large tanks, where often very poor blending conditions exist compared with flow patterns in small tanks in the pilot plant. Axial flow airfoils are not particularly desirable for piolt plant operation, since blend times can appear to be adequate whereas full-scale counterparts may not be. For viscous materials, normally with Reynolds numbers of less than 10, the helical impellers shown in Figure 6-10 and anchor impellers (Figure 6-11) are typical. Both are very effective in providing visual blending throughout large-scale systems in viscous fluids and typically operate at speeds as low as 5 to 15 RPM. This is considerably lower than the speed of radial mixing impellers. During the 1980s highly flow efficient impellers (Figure 16-10) were originally designed to maximize the outlet flow in terms of pumping capacity and axial flow direction. This was done to optimize their performance capability in processes such as low viscosity blending and solids suspension. Now there is a whole family of high efficiency impellers" one for gas-liquid applications, another for the paper stock industry, and a third for draft tube circulators. Recently, additional advances have been made in fluidfoil impellers to extend the range of these impellers beyond the turbulent fluid regime to higher viscosities. Measurements of fluid velocities performed using a dual channel laser doppler velocimeter (LDV) and blending studies have demonstrated equal blend times at one-half of the power consumption compared to a pitched-blade turbine which has historically been used in these flow regimes. The results show that this axial impeller can be effectively used in the transitional flow regimes. A structural composite mixer with hydrodynamic design and proplet tips (Figure 16-11) represents another dramatic change in fluid mixing technology. This vinyl ester resin axial impeller reinforced with graphite and glass fibers, is highly resistant to corrosives and abrasion. Its lightweight composition makes possible the use of longer shafts and increases its effectiveness in deep tank applications.
Portable Mixers Portable mixers are often provided with suitable clamping devices that permit mounting to the side of an open tank. Portable mixers range in horsepower up to about 3. Constant speed versions are usually either direct driven at 1150, 1450, or 1750 RPM or operate at either 280 or 350 RPM with a gear drive. The geardriven portable mixer has a larger impeller and operates at slower speed that its direct-driven counterpart. It therefore develops more flow and less shear rate than the comparable direct-drive unit, which has a smaller impeller and higher speed.
623
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Process Plant Machinery
FIGURE 16-9 Typical fluid foil impeller used to provide maximum flow and minimum shear rates. (Source: LIGHTNIN, Rochester, NY.)
FIGURE 16-10 A320 high flow efficiency impeller with laser doppler velocimeter measurements. (Source: LIGHTNIN, Rochester, NY. )
Mixers and Agitators
FIGURE 16-11 Fluid foil impeller made from structural composite materials; also shown are laser measurements of flow characteristics. (Source: LIGHTNIN, Rochester, NE)
FIGURE 16-12 Illustration of proper angle to achieve top-to-bottom turnover without wall baffles with portable mixers. (Source: LIGHTNIN, Rochester, NY.) Using a ball and socket joint for a clamp allows the impeller to be skewed, as shown in Figure 16-13. Good flow patterns can thus be obtained without using baffles. These mixers can also be provided with a permanent angular mounting for open tanks. Portable mixers typically use the airfoil impeller (Figure 16-9), although other impellers may be used if called for by special fluid conditions. The gear-driven portable mixer can be more expensive, since it requires a gear drive; this follows the general principle that mixers for low speed and high flow normally require higher capital outlays than mixers producing high shear and low flow. There are many other options available, such as hydraulic motors, air motors, and various kinds of variable-speed motors. Variable frequency, adjustable speed drive mechanisms represent another important option.
Heavy-Duty, Top-Entering Mixers Industrial fluid mixers for large tanks usually incorporate a two- or three-stage gear speed reduction mechanism. These mixers go up from 2 HP to several thousand HP. The torque provided by the gear box can reach one million in-lb. These mixers are normally installed on the tank centerline. Wall baffles are typically used, and
625
626
Process Plant Machinery
•)~176 I I
400 -=
/
I
/ I
/ /// / /z / /
I
/ /
/ / /
IO
/-
/ / 2 / / S /."/_
,~
/
/
/
/
/
/
/
/ !
I
SPEED
I IIIII I000
ioo RPM
FIGURE 16-13 Typical chart showing power versus output speed for various sizes of speed reducers. Most speed reducers are rated essentially at constant torque. (Source: LIGHTNIN, Rochester, NY. )
impellers may be either radial flow turbine, axial turbine, or airfoil type. Most gear drives are rated at constant torque. Figure 16-13 illustrates the typical range of gear boxes; it indicates that for any given size, the mixer output is proportional to the output speed. Also shown on this curve is a typical process curve for a blending or solids suspension process. This confirms that for flow-controlled applications, there is a trade-off between operating costs and capital expenditures. For example, the choice of mixers can be between a 75-HP mixer, size five casing, or a 190-HP mixer, size four casing, which would cost less in capital dollars but would require more power. For high levels of power and high fluid shear rates, the radial flow flat blade turbine offers good mechanical stability. It has the ability to improve almost any power level that is required for the process. The axial flow turbine, since it has a higher pumping capacity, is limited to lower power levels. These usually range from 0.5 to 5 HP/1000 gallons. Axial flow turbines are used in applications where circulating capacity is needed with a moderate amount of fluid shear rates. When the only process requirement is pumping capacity and minimum horsepower is needed, the airfoil impellers use power levels on the order of 0.1 and 2 HP/1000 gallons and are very adaptable to blending in solids suspension applications. Typical D/T ratios for all three impeller types are usually between 0.3 and 0.5. When the D/T ratio gets much beyond 0.5, there is a tendency for additional entrainment of fluid to become very minimal. The advantage of large-diameter, slow-speed units is to give more total pumping capacity is not realized. One of the key design elements of top-entering mixers is that they must run below the first natural frequency of the mixer shaft. In designing a mixer of this type, the operating speed is selected for the particular diameter and power level desired for the process result. The operating speed is divided by the ratio of operating speed and critical speed desired, which most manufacturers set somewhere in the range of 0.7 to 0.8. The critical speed of the shaft can then be calculated.
Mixers and Agitators
The radius of the forces acting on the impeller must be calculated from the fluid mechanics and the fluid forces involved in these systems along with a suitable stress level for the material being used. The shaft diameter required at the maximum bending moment point is calculated. In addition, the diameter of the shaft is calculated based on torque requirements. The largest of these three diameters is the shaft required to satisfy all foreseeable conditions. As a general rule, there is some shaft diameter that will allow any length of the shaft to be run overhung with no steady bearing, operating at a suitable distance from the natural critical frequency. Using a steady bearing at the bottom of the tank allows the shaft diameter to be reduced. The cost of maintenance of the steady beating must be balanced against the extra cost of a large diameter shaft without a steady beating.
Light-Duty,Top-EnteringMixers There is a gap between the range where typical portable applications stop and where the heavy-duty mixers described earlier must be used. This undefined range is usually in the vicinity of 5 to 15 HP with speeds on the order of 68 to 125 RPM. A variety of mixers are available in this area, all of which have somewhat differing basic design characteristics and which yield satisfactory performance at reasonable cost.
Side-EnteringMixers Side-entering mixers must be properly positioned in the tank (Figure 16-14). Positioning about 10 ~ off the tank diameter gives effective flow patterns for many types of blending applications. These mixers are quite commonly used in paperstock suspension blending, or in gasoline and crude oil tanks, as well as to prevent settling
"
~
~ ' -~ /
~
CORRECT
l
o
,
10o
INCORRECT
}
FIGURE 16-14 Illustration of correct flow patterns for installation of side-entering mixers and large petroleum storage tanks. (Source: LIGHTNIN, Rochester, NE )
627
628
Process Plant Machinery
of solids in the sediment and water component of many oil pipelines. Side-entering mixers often produce quiescent zones 90 ~ from the entry location. These zones can cause difficulty with solids settling if caution is not exercised in the process design. The major advantage of side-entering mixers is the low initial cost compared with that of top-entering mixers. Top-entering mixers must have a support structure, which adds considerably to the total cost of the installation. Side-entering mixers will always require higher horsepower for the same application, since they run at higher speeds, typically 150 to 400 RPM. They have smaller diameter impellers than corresponding top-entering mixers and thus require more horsepower for the required circulating capacity. A relativity short shaft and cost savings on the gear box favor the economic evaluation. However, since the mixer shaft must penetrate the tank wall, a mechanical seal is needed. Materials that are very abrasive or corrosive can cause severe problems unless component configurations and materials of construction are carefully chosen and properly engineered.
Bottom-Entering Mixers Bottom-entering mixers, as shown in Figure 16-15, are not as frequently used as top-entering mixers. They are chosen for certain reasons - one of them may be the desire to eliminate the shaft passing through the liquid surface, thus being subject to fouling due to collection of process material on the shaft. While it may be desirable to perform drive maintenance at floor level rather than on an elevated structure, it must be recognized that bottom-entering mixers require a mechanical seal at the bottom of the tank. Seal failure can cause the tank contents to discharge. Bottom-entering mixers allow the head space to be used for other types of piping connections. One disadvantage is that the high Z/T ratio is not readily handled, since the second and third impellers above the lower impeller are positioned on a shaft that is unsupported and may complicate the design.
Shaft Sealing Devices There are two types of shaft sealing devices for mixers and agitators. One is the somewhat unsophisticated stuffing box that utilizes soft packing pressed against the shaft by an adjustable gland plate (Figure 16-16), upper half). The other sealing method employs mechanical face seals, as illustrated in the lower half of Figure 16-16. Mechanical face seals are an indispensable component of high-pressure, toxic, flammable, or high-value mixer applications. Unlike soft packing, which must be adjusted so as to allow a small amount of leakage in order to provide a lubricating fluid film between shaft and packing, mechanical face seals can be arranged for virtually zero leakage. Moreover, mechanical seal housings can be designed to include stabilizing beatings (Figures 16-17 and 16-18) or emergency shutdown features (Figure 16-19). Figures 16-17 and 16-18 illustrate double mechanical seals, which are typically surrounded by a buffer fluid that is maintained at about 10 to 25 psi over tank pressure. Should seal leakage occur, the buffer fluid captured between the two seals would either leak into the tank or toward the atmospheric (bearing housing) side. However, leakage rates are either extremely small and will be replenished from a makeup container, or else loss of buffer fluid can be detected by appropriate instrumentation and a safe shutdown be initiated without delay.
Mixers and Agitators
FIGURE 16-15
Bottom-entering mixer. (Source: LIGHTNIN, Rochester, N.Y.)
FIGURE 16-16 Shaft sealing area for mixer-agitators. Soft-packing is shown above the centerline of shaft; a double mechanical seal is shown below the shaft centerline for comparison. LIGHTNIN Co., Inc., Rochester, N.Y.)
629
630
Process Plant Machinery
flushing
t
throttle nng
v)
l)uffer INLET
fluid
T
drainage
most e f f e c t i v e c o o l i n g flange 3 0 m m fhermal shatt m o v e m e n t . - -I~.
"o o
Q,
\
FIGURE 16-17 Double mechanical seal for a modern mixer. (Source: Burgmann Seals America, Houston, TX.)
FIGURE 16-18 Double mechanical seal for vertical mixer. Radial and thrust bearings are incorporated in the seal housing. (Source: Burgmann Seals America, Houston, TX.)
Mixers and Agitators
STD
k
Shut-down seal Product side shut-down seal with rubber elastic sealing element (item 1), pneumatically or hydraulically operated.
1
Seal can be changed with the tank filled Suitable for top and bottom drive.
Flange with connection dimensions to DIN 28138 can be supplied.
..L
.....
,__.
Sealing element not operated ...
____L
... operated
Shut-down Seal with shaft centering Product side shut-down seal for agitator with bottom drive.
product side
Seal can be changed with the tank filled. Item Ir 1 O-ring 2 centering cone 3 thrust ring
j l
shut-down seal not operated ...
Emergency Seal
...
operated
Emergency sealing of the atmospheric side of an agitator seal is possible by pneumatic or hydraulic operated O-rings (item 1). -.---4
With this version a seat cannot be changed during operation, but aggressive products are prevented from escaping to atmosphere if the seal fails. In order to change the seal the agitator must be stopped.
.
.
.
.
.
\
/
~\',~
\ \
\ \
inoperative
O-rings sealing
FIGURE 16-19 Shutdown or emergency seals for mixers and agitators. (Source: Burgmann Seals America, Houston, TX.) Figure 16-19 shows three of many shutdown or emergency seals for mixers and agitators. Upon loss of a main seal, the mixer will be automatically shut down and the emergency seal activated to contain the product in a safe fashion. It is almost always a serious mistake to purchase the least expensive seal for a mixer or agitator application, as the resulting frequent seal failures will burden the user with high maintenance expenses. On the other hand, properly designed mechanical seals have extended the run time of modern mixers by orders of magnitude. The downtime and repair cost avoidance achieved with superior mechanical seals will often pay for the seals in the first few weeks of operation.
631
This Page Intentionally Left Blank
Chapter 1 7 Separators Centrifugal separators are used in a variety of applications, some of them in conventional technology, such as food processing, for instance dairy creamers and chemical conversions. More recently, the design demands placed on this technology have increased with mineral beneficiation applications, such as extracting synthetic crude from oil sands. When the prototypes of this application were first tried in 1977, the centrifuge nozzles had a life of a few hours. The material had to be changed to one that better withstood the highly abrasive sand involved. The nozzle angle also had to be varied.
INTRODUCTION* Frequently, mixtures of solids and liquids must be separated into their components in order to be effectively utilized. The mixtures may be of different solids or the liquid fraction may contain dissolved solids that are to be removed. Such situations occur in food processing, mineral beneficiation and chemical conversions. When the solids and the density difference is small and the flow volume is large, disk nozzle centrifuges are often the best means to accomplish the purification. The separation which takes place within the rotor of a disk nozzle centrifuge is effected by the G force, the "rising rate" of the liquid (related to the feed flow) and the separation area provided by a set of conical, close-spaced disks as well as the process factors of fluid viscosity, particle size, shape, and density. In addition, the design of the equipment must allow for the quantity of solids to be handled, the flow characteristics of the slurry and other practical engineering considerations. Whereas disk nozzle centrifuges have been in use for concentration purposes for a long time, they are now employed as purifiers; in which instance a large flow of "upflowing" liquid greatly enhances the purity of the products. A benchtop illustration makes it easier to see how this elutriating stream functions. Using this concept we can deduce the beneficial action of displacement washing versus dilution washing and how the improved flow pattern enhances the "classification" of particles. Three examples are presented which show typical processes. Disk nozzle centrifuges have been used for over 60 years for the concentration of fine solids in a stream of slurry feed. Such centrifuges are now in common use around the world for handling food products, chemicals, minerals, biological materials and waxes. These centrifuges are made in a variety of materials and sizes and in many different countries by various manufacturers with differing design concepts. * Source: Dorr-Oliver, Milford, CT. Adapted with permission 633
634
Process Plant Machinery
However, the significant principles are well established and many publications show the relationships of the operative factors. In addition to simple sedimentation, where the objectives are to obtain a clarified effluent or thickened solids-loaded fraction or the separation of two liquid phases, it is also possible to simultaneously introduce a stream of "wash" into the centrifuge. There may be several purposes served. First, the discharging solids may exit in the "wash" fluid rather than in the mother liquid, or upflow action of the wash stream may flush out a smaller size solids fraction from the larger size solids fraction. In the first case we have purification by washing (solubles removal) and in the second case we have purification by classification (slower-settling solids removal). Figure 17-1 is a photo of an intermediate size disk nozzle centrifuge. This machine is fitted to operate under elevated temperature and pressure conditions and for the purification of terephthalic acid crystals. Note the electric motor, the overhead V-belts and the flexibly mounted beating assembly. These power the pendulumsuspended rotor which has somewhat of a double cone shape, with the nozzles being located at the largest periphery. The housing material is Hastelloy for extra corrosion resistance and the rotor is similarly special to withstand the severe mechanical and chemical conditions. Figure 17-2 is a cutaway view of a disk nozzle centrifuge that shows the flow pattern. It is easy to follow the path of the feed slurry as it flows continuously down
FIGURE 17-1
ford CT.)
Shop photo of Merco | PCH-30 centrifuge. (Source: Dorr-Oliver Inc, Mill-
Separators
FIGURE 17-2 Cutaway view of a disk nozzle centrifuge and its flows. (Source: Dorr-Oliver Inc, Millford CT.) into a central rotating feed distributor and laterally into the main separating chamber. Here the high sedimenting force (of perhaps 5000 g) acts to draw the heavier solids outward where they discharge from the rotor through backwardly reacting nozzles. This slurry is then gathered in a collecting volute and recycles (by means of its velocity head) back to a re-injection port in the bottom of the stationary housing and jets back into the rotor hub where it is re-accelerated. A major portion of the underflow can be drawn off through a valve located appropriately in the return loop. Meanwhile, the surplus flow (the feed minus the draw-off) moves inwardly through the separating disks, where fine solids are removed, and it discharges from the top of the rotor as clarified overflow. The method of feeding into the disk stack through a set of vertically punched holes and the arrangement of spaces on the disks are well known. They are sized and located in a specific fashion appropriate to the application. Similarly the recycled flow is directed through special tubes back towards the nozzle region. Figure 17-3 is a cutaway view of the centrifuge with a special wash inlet system inserted at the bottom of the housing. This system makes it possible to inject large volumes of wash at the interior of the rotor where the flow must travel inwardly and counter-currently to the outward motion of the solids. This action can accomplish a great increase in the washing capability of a single stage or it can significantly enhance the sharpness of the separation between two classes of solids. In the process of washing, the most important thing is to remove the contaminants as completely as possible. Thus, the use of large quantitites of wash are generally of benefit. However, we often want to conserve the wash fluid for economic reasons. Accordingly, a balance is struck and the degree of efficiency of washing becomes important. In the process of classification, the most important thing is to remove all of the slow settling solids but to not remove the other solids. Thus, the appropriate quantity of wash has to be sought by testing.
635
636
Process Plant Machinery
FIGURE 17-3 Cutaway view of Merco | centrifuge with a special wash inlet. (Source: Dorr-Oliver Inc, Millford CT.) DISK NOZZLE CENTRIFUGES
The solid-liquid separation in a disk nozzle centrifuge is involved with the physical properties of the solids and the fluid as well as the fluid dynamic phenomena in the disks. Since the flow pattern inside the rotor is very complicated, much effort has been devoted to the understanding of the separation mechanism and to improving the separation efficiency. The process of sedimentation of solid particles in a tank can be contemplated in order to understand the mechanism of centrifugal clarification. Figure 17-4 shows the continuous clarification of solids in a sedimentation tank. A solids suspension with different particle sizes is fed to the tank and flows out at a velocity w. The suspended solids settle at a velocity v to the bottom of the tank. The processing capacity of the sedimentation tank can be calculated from the terminal velocity of the particles with a critical diameter (dc). The critical diameter is the minimum diameter of the particles that can be completely separated in the sedimentation tank. Stokes' law predicts the sedimentation velocity (v) of particles under gravity as follows: d 2(Ps - Pl ) v= g (17.1) 18/.t where d is the diameter of the particles; Ps the density of solids in the suspension; Pl the density of the liquid phase; /z the viscosity of the liquid phase; and g the acceleration due to gravity. By inserting the critical diameter, densities and viscosity in the above equation, we can readily obtain the sedimentation velocity (Vg) of particles with the critical diameter. The retention time of the process stream in the sedimentation tank can
Separators ENTRANCE
J 0
9
o
9
EXIT
e
O0 0
9
e
0 o
h 41,
0
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9
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00 9 9 *
co oo8"o~
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L FIGURE 17-4 Continuous clarification of solids in a sedimentation tank. (Source: Dorr-Oliver Inc, Millford CT.) then be calculated by dividing the volume of the tank by the flow rate (Q), thus, V t = --
Q
(17.2)
where V = bhl and b is the width of the tank, h the height of the tank and l the length of the tank. In the same period of time, the particles settle to the bottom of the tank. The longest time that particles take to do this is for particles with the critical diameter, so h t = --
Vg
(17.3)
By equating the above two equations, the capacity of the sedimentation tank can be obtained as follows: Q = blvg = Avg (17.4) where A is the bottom area of the tank. The above equation shows that the height of the sedimentation tank does not influence the process capacity and that the capacity is linearly proportional to the bottom area of the tank. Accordingly, more horizontal plates might be inserted into the tank as shown in Figure 17-5 in order to increase the capacity of the tank. Thus, the total capacity of the tank is Q = NAvg (17.5) where N is the number of horizontal plates. With a continuous flow of a slurry stream in the above mentioned tank, the separation path can be easily clogged by the settled particles after some period of operation. If the horizontal plates were inclined at an angle, then the sedimented particles will slide down to the bottom of tank during the sedimentation process. In this case, the area term in equation (17.5) is the projected area of the inclined plates. Figure 17-6 shows the concept.
637
638
Process Plant Machinery ENTRANCE
1
0
e
v
0
O0 0
0
9
0
~
0
0
~ v
0 o
0
o 0
o 0
0
O0
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0 o
~
~
o~
o o~
o 0
0 0
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~
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FIGURE
9
(,"~ ~
9
0
fk~
0
et"~
(,',)
0
r
0
oo
0 O o
v
Sedimentation tank with horizontal plates. (Source: Dorr-Oliver Inc, Mill-
17-5
ford CT.) ENTRANCE
,ll
o.
.
000
o .oo..o O
ee
.
9
9
9
. (
.o
9
.
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17-6
o
0
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FIGURE
0
-
0
9
9 0 d~
o ~
0 d~
0 "
-
0
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eO
9 ~
~
0
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~
o( ~
dPmVJ%d
Sedimentation tank with inclined plates.
(Source: Dorr-Oliver Inc,
Millford CT.)
If the sedimentation tank with inclined plates is transformed into the configuration shown in Figure 17-7 to allow the use of centrifugal force to enhance the separation power, we have a prototype of the disk centrifuge. Within the centrifuge, the traveling distance of the solids is confined in the space between two disks. Usually, rectangular metal bars are placed on the top surface of disks, and their thickness determines the disk spacing. The disk spacing is adjusted according to the solids load and the properties of the process slurry.
Separators ENTRANCE CLARIFIED FLOW
I
FIGURE 17-7
Prototypeof a disk centrifuge. (Source: Dorr-Oliver Inc, Millford CT.)
Since the rear surface of disks is critical for solids traveling down away from the separating zone during centrifuging, the surface is manufactured to be as smooth as possible. The metal bars also separate the disks into many sectors in which separation takes place. The segregation by the metal bars also limits the influence of the Coriolis effect which can reduce the separating efficiency to a great extent. During continuous operation a steady state is usually maintained so that the underflow concentration is held constant, which is very beneficial to the downstream process. In this case, the underflow flowrate should be able to be adjusted so that a desired concentration can be achieved. The discharge from the centrifuge underflow may not always provide the desired concentration so a recycle device is needed. Dorr-Oliver's Merco centrifuges are equipped with a recycle bend, and this permits more functions for different applications. In addition to clarification and concentration, operations such as washing, classification and soluble recovery can be done with centrifuges with an underflow regulating valve and a return bend. The motion of a particle in a sedimenting centrifuge may be regarded as strictly analogous to that of a particle settling under gravity, with the difference that the gravitational acceleration is replaced by centrifugal acceleration, w2r. At a given value of acceleration, a particle will accelerate until it reaches a constant terminal velocity at which the centrifugal force is balanced by an opposing drag force due to the friction between the surface of the particle and the fluid through which the particle is traveling. The terminal velocity subject to a centrifugal force may be expressed as follows.
-
7)1
---
(w~r)
d2(ps Pl)(w2r) = Vg ~ 18/z
(17.6)
where w is the angular velocity of the particle in the settling zone and r is the radius at which settling velocity is determined. The term (w2r/g) refers to the G force. In order to justify the validity of Stokes' Law used for settling particles in the centrifugal bowl, the Reynolds number (Re) of the particles has to be evaluated. The
639
640
Process Plant Machinery
TABLE 17.1 The Reynolds Numbers of Spheres (SG = 2.8) at Different Accelerating Forces Particle diameter
Reynolds number of particles
(microns)
at gravity
at 2000 G
at 4000 G
at 8000 G
0.1 1 10 100
1 x 10 -9 1 x 10 -6 1 x 10 -3 1
2 x 10 -6 2 x 10 -3 2 2 x 103
4 x 10 -6 4 x 10 -3 4 4 x 103
8 x 10 -6 8 x 10 -3 8 8 x 103
much higher accelerating force generated in a centrifuge raises the possibility that a different flow pattern might result. If spherical particles having specific gravity of 2.8 settling in water is taken as an example, the Reynolds number can be calculated for differing sizes of particles under certain gravitational forces. Table 17.1 indicates the results. It provides reasonable confirmation for the general assumption of Stokes' law for usual industrial combinations of particle size and gravity. Whereas the sedimenting surface area linearly affects the capacity of the sedimentation tank as discussed above, the "Sigma" concept proposed by Ambler has been widely used in centrifugal studies, in which the capacity of the centrifuge is expressed in terms of the equivalent area of a gravity settler performing the same duty. Similar to equation (17.4), the throughput of a centrifuge can be expressed as follows: V a = ~3t-~(17.7) where V is volume of liquid held in the system and S is the thickness of the liquid layer. Comparison of equations (17.1), (17.4), (17.6) and (17.7) shows the following equality: a --
oj2rV gS
(17.8)
This is termed the Sigma factor, E, for a given centrifuge, since it is dependent only upon the characteristics of the centrifuge and not upon the solid/liquid system which it is handling. Sigma theory shows that Q = 2VgE
(17.9)
if the concept of a 50% cut point is taken into consideration. The 50% cut point is the diameter of particles half of which will be removed in passage through the centrifuge bowl, and half of which will not be removed. WASHING
Washing processes are commonly employed in the food, chemical and pharmaceutical industries. The main objective of washing is purification, which is to reduce the concentration of soluble materials (impurities) in the mother liquor. There exist two modes of washing, dilution washing and displacement washing. Taking a bath and taking a shower are examples showing the difference between these two modes. Usually displacement washing gives better efficiency than does dilution washing. Typical washing efficiency curves for these two washing modes are shown in Figure 17-8. Figure 17-9 shows the conventional system of washing by dilution. In one stage the feed mixture is diluted, and then the insoluble solids settle since their specific gravity is higher than that of the liquid. The soluble impurities in the mixture are
Separators 100 ~
_
displac,~
~
~
~
641
~
~'~'-dilution ~
, '
1
/ I/
: 0
1
2
3
4
5
Vol. of wash added/Vol, of underflow liquid drawn off FIGURE
ford CT.)
FIGURE 17-9
17-8
Advantage of Merco centrifuge washing. (Source: Dorr-Oliver Inc, Mill-
Dilution washing mode for 50% mother liquid removal. (Source: Dorr-Oliver Inc, Millford CT.)
reduced by the ratio of the quantity of washing liquor to the quantity of mother liquor. For example, if a ratio of 1:1 is applied, the soluble impurities concentration is reduced by 50%. In an actual operation, a series of stages may be set up to operate countercurrently. Figure 17-10 shows a process in which the process material mixes with the overflow stream from the second stage before going to the first stage. The overflow of the first stage goes back to the process, while the underflow is concentrated in the first stage and fed to a mixing tank to mix with the overflow of the third stage. The circuit of the streams extends to as many stages as needed in such a pattern. A governing equation for describing the counter-current dilution washing scheme is as follows: S, = r, Sn+l + 1 (17.10) where S,, is the soluble ratio at stage n and r,, is the wash ratio (overflow rate over underflow rate) at stage n. This equation states that the soluble ratio for any stage
642
Process Plant Machinery
FEED F'I:tOM~JI9
~OIN TOF-~
CX~NCENTF~IA'rlE
!9
FIGURE 17-10
-
ODISlIC:E:J~ITI~TE
p
Countercurrent dilution washing circuit. (Source: Dorr-Oliver Inc, Mill-
ford CT.)
is the wash ratio times the solubles ratio for the succeeding stage plus one. Since, in the final stage, the soluble ratio is unity and if the underflow rate is constant, then the soluble reduction ratio in n stages in a series is as follows: r n+l
S --
--
1
(17.11)
r-1
The above equation cannot be used for the case when the wash ratio is one. When the wash ratio is one, the reduction ratio for n stages can be calculated as follows" S-
n + 1
(17.12)
Figure 17-11 describes the impurity reduction ratio for different numbers of stages at differing wash ratios. The dilution washing mode can be found in processes where such solid-liquid equipment as hydrocyclones, screens and decanters are operated.
stages 13
10030(~ 11
O
o.,,
O
o,.,
"1=
E
0
0.5
1 1.5 2 2.5 Overflow liquid/Underflow liquid
3
FIGURE 17-11 Impurity reduction ratio by dilution washing mode at various wash quantities and stages. (Source: Dorr-Oliver Inc, Millford CT.)
Separators
643
1 Dart washing liquid
[ mother liquid
r ~. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
70% mother liquid 30% washing liquid displaced liquid
interface zone .~:j~::(-~i:i:~;i~.~i:i:i~:i:i:i:i:i:i:!:i:i:i:i::::ii!
~::~iii:~i~2,:'~i:i:i~i::ii:i:!ii! 30% mother liquid 70% washing liquid
displaced liquid
washed liquid FIGURE 17-12
Mechanism of displacement washing mode. (Source: Dorr-Oliver Inc, Millford CT.)
The mechanism of displacement washing is shown in Figure 17-12 where the washing liquid enters from below and pushes the mother liquor in the container upward and replaces the mother liquor. In this way the washing efficiency can be greatly increased. Seventy percent efficiency is assumed in Figure 17-12 so that the displaced liquid contains 70% mother liquid and 30% washing liquid, whereas the wash liquid with all of the insoluble solids contains 30% mother liquid and 70% washing liquid. The efficiency depends on the stability of the interface as well as the degree of mixing. When a liquid is displaced by another immiscible fluid of lower viscosity, the interface between the phases is frequently unstable. The phenomenon of "viscous fingering" will result. The degree of mixing is caused by the turbulence in the eddies of the flow patterns. In the displacement washing mode, the concentrations of overflow and underflow streams are different. The concentration of soluble impurities in the underflow can be separated into two parts; a certain fraction, R, being from the feed, and the remainder, I-R, from the wash stream. An equation can be derived for the displacement washing mode as follows: S,, = m,,Sn+l + an
(17.13)
where mn -
1 R
D
~
an
and (l-R) an =
R
W. Un
If the underflow rate and wash rate are constant in the series of counter-current washing stages, the values of a and rn being the same for every stage, a general
644 ProcessPlant Machinery equation describing the soluble reduction ratio as a function of displacement ratio (wash rate over underflow rate) is as follows:
S=mn+a(
mn-1)m-1
(17.14)
When the displacement ratio (W/U) is 1, the above equation cannot be used, instead, the following equation should be employed:
S = na + 1
(17.15)
This is a much more effective procedure than dilution washing. In fact, in most cases only one stage need be used but, of course, there are cases where the impurities must be eliminated to such an extent that several stages are employed and these are arranged to be counter-current as shown in Figure 17-13. Figure 17-14 describes the impurity reduction ratio in a series of stages at different displacement ratios (W/U). From Figures 17-11 and 17-14, we can easily observe that fewer stages are needed for the displacement washing mode than for dilution washing. For example, to achieve an impurity reduction ratio of 100, four stages are needed for displacement washing at the displacement ratio of 1.5, while nine stages are needed for dilution washing at the wash ratio of 1.5. The displacement washing mode may occur in rotary vacuum drum filters, and in some centrifuges such as the Merco (disk nozzle with return bend) and Mercone centrifuge (screen with wash injection). Classification
A different aspect of purification involves the removal of solid impurities from a slurry of mixed solids based upon their difference in settling rates. This is called "classification". The solids may be different in nature (protein and starch, for example) or just in size (kaolin clay particles, for example). If two kinds of particles are to be separated and one particle has a specific gravity lower than that F EED ~
O V ~ TO PRCX~ESS
,
T
UN~R.CI~
OVERFLOW
DISPLACEMENT
UNI~.RR.CI~ ~P
WASH
,d
DISPLACEMENT
~
ll
!
1 FIGURE 17-13
Millford CT.)
Counter-current displacement washing circuit. (Source: Dorr-Oliver Inc,
Separators
stages
10000000
/_.~
1000000 0
"= 0
=9~ ~
"~
/ "
f
,,.,-
.'..
~
~
5
...._ .-.
.
1000, 100
6
,,~"
t
/
,;-
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--"
.. lip
~
7
j...-
100000 10000
645
I
O
B
,,
4
'
3
u
9
.,-
-
-,
2
=
10:
,= = I
0
0.5
1 1.5 2 Wash liquid~nderflow liquid
2.5
3
FIGURE 17-14 Impurity reduction ratio by displacement washing mode at various wash quantities and stages. (Source: Dorr-Oliver Inc, Millford CT.)
of the medium and the other particles have a higher specific gravity than that of the medium, then it is not difficult to separate them, as one would float and the other sink. If both kinds of particles have higher specific gravity than that of the medium, sharp and complete separation of both groups of particles becomes challenging. In this case of classification, the wash stream will be employed and provides another force to be manipulated so that the solids can be classified as desired. Figure 17-15 shows the conventional classification method for the separation of small and large spheres as a time sequence. The large spheres have a diameter two times that of the small spheres. According to Stokes' law as stated in equation (17.1), the settling velocity of the large spheres is four times that of the small spheres, assuming both have the same specific gravity. At time 0 the mixture is suspended uniformly and then the solids are allowed to settle. The large spheres move downward more quickly. If the mixture is decanted at time interval 2, and the mixture is split into two equal parts, all of the large spheres will be recovered whereas a portion of the small spheres will be present with the large spheres in the underflow stream. If the time interval is increased for example, at time interval 4, more small spheres will settle with the large spheres and make the separation worse. Thus, the operation of classification following the conventional method will be very touchy and the set-up of the machine specifications becomes tricky. An improved method is proposed for classification as shown in Figure 17-16. An additional stream is fed as wash and is adjusted to a rate such that the velocity of the wash stream is greater than that of the small spheres and smaller than that of the large spheres. In this case, the velocity of the wash is two times that of the small spheres. Accordingly, the large spheres settle down into the underflow, whereas the small spheres cannot settle down and are carried over to the effluent (overflow). Therefore, sharp and complete separation of spheres can be effected by proper adjustment of the wash flow.
646
Process Plant Machinery
Current Classification
i!.
Method
By difference in settling velocity
I0
0
Ox
O 0
O 0
0
0~.~ 0>r 0 '>< i
A
B
Time-0
Time- 1
Time-2
Time-3
Time-4
Upper half
4A+4B
3A ~2B
3A+0B
2A+08
2A+OB
Lower haft
4A+4B
4A+6B
5A+BB
5A+88
6A+BB
FIGURE 17-15 Principles of the conventional classification method. (Source: Dorr-Oliver Inc, Millford CT.)
Improved Classification
Method
By displacement method
1
Time..0 Overltow
i"i"l
Time.,1 1A
ITi
~
Time-2
ITI
Time-3
2A
3A
Time.4 4A
Upper half
4A§
5A.18
6A.0B
7A.0B
8At0B
Lower half
4A.4B
3A,,6B
2AtSB
tA+SB
0A.SB
FIGURE 17-16
Millford CT.)
Principles of the improved classification method. (Source: Dorr-Oliver Inc,
Separators 647 The fine particles in the overflow can be thickened in another centrifuge and the clear liquid fraction can be recycled back to serve as the wash stream. Obviously, the upflow rate of the wash liquid is critical and specific to the situation. If its velocity is too high, all of the spheres may be lost and if the velocity is too low, then some contaminating small spheres may remain with the large spheres. One restriction to this operation is that the size range of the small particles cannot overlap the size range of the large particles. In the overlapping case, a sharp separation cannot be achieved. Merco centrifuges have washing capabilities which provide a unique feature allowing the operation of classification of particles simply by following the principles illustrated. APPLICATIONS
Three kinds of processes, i.e. classification of kaolin clay particles, washing of terephthalic acid crystals, and dewaxing of lube oil, are employed for the purpose of discussing the application of the disk nozzle centrifuge in the purification of solids mixtures occurring in the chemical industry. During the processing of kaolin clay, the raw material is mined and slurried and the coarse sand and grit is removed. Several stages of centrifugation are employed in order to separate the slurry into various fractions for other uses. Such equipment includes decanter centrifuges and batteries of hydrocyclones. High-speed disk nozzle centrifuges are used in several positions in the flowsheet but the most critical station is for "fractionation." The slurry is a mixture of clay solids suspended in water and the solid material is all of the same density and character except for size. The most valuable clay fraction is less than 2 microns but larger than 0.5 microns in size and is used to make paper whiter and improve the appearance of print. Here is where we "cut out" the slimes and leave the bulk of the supply as product. A typical feed would have 30-40% DS and the"underflow" product would be 5 0 - 6 0 % DS and the "overflow" product would be 15-20% DS depending upon the product requirement. The disk nozzle centrifuge would typically operate at high speed and with a large number of large size rotor nozzles so as to cope with the flow requirements. The feed rate to the centrifuge and the underflow compositon (% solids) are controlled by valves so as to render the desired "cut" (the overflow). Operation in this way is quite different than when a centrifuge is in a "clarification" mode. Both fractions are opaquely white and the sharpness of the "cut" requires that the particle size distribution of each fraction be determined. The analyses are done using special instruments. By using the elutriation wash principle, as described earlier, it is possible to remove the slimes more completely and to simultaneously recover more of the valuable material. The action takes place exactly as explained. The wash is introduced into the centrifuge rotor at the bottom and it enters the slurry bed at the "bottom" (the inboard periphery near the discharge nozzles). The water rises through the suspended bed carrying slimes with it and the flow passes on through the separating disks and discharges as "overflow". Accordingly, the underflow has less slimes remaining in it and the overflow has more. There is some dilution of the overflow due to the additional water. When this scheme is practiced, a series of runs must first be conducted at various wash water addition rates. The results will show how much benefit each increment of water provides. One can then calculate the optimum combination of feed and wash rates to perform the purification. According to the test results, normal classification by differential settling of kaolin particles gives 75% recovery of the
648
Process Plant Machinery
most valued fraction, whereas classification enhanced by elutriating wash results in 85% recovery. The processing of terephthalic acid affords a rather novel example of the benefit of displacement washing when using a disk nozzle centrifuge. In this case, the product (PTA crystals) is washed while it is suspended in a hot corrosive fluid under elevated pressure and temperature conditions. Terephthalic acid is made in many countries around the world and by a variety of processes. Some of its uses are fiber, film, and plastic bottles. In most processes centrifuges are employed to separate the solid crystals (less than 200 microns in size) from the mother liquor and a counter-current washing scheme is used to assure the production of purified crystals. The solids content of the feed may be of the order of 35% DS and the underflow may be exactly the same or at a higher concentration. The large size of the crystals (peaking around 70 microns and none smaller than 20 microns) makes it possible to operate with high upflows of the wash fluid. Most of the Merco centrifuges operate at 10 bar, or less, pressure but some models operate at 20 bar. Higher pressures and temperatures improve the performance by reducing the fluid viscosity. High-speed disk nozzle centrifuges are used in series where the displacement wash efficiency can be over 90% per stage if high wash ratios are used, such as W / U of 2.2. These particular centrifuges have outstanding performance records for on-stream time and have required very little mechanical maintenance. A third example involves the processing of petroleum wax which is removed from lube oil by chilling the feed stock together with the introduction of certain solvents which enhance crystal formation. In this process the conventional arrangement is to use rotary vacuum filters in series in order to completely dewax the oil. The procedure is practiced around the world in similar fashion but on differing flow scales. Merco disk nozzle centrifuges are uniquely able to perform in this process even though the density difference between the wax crystals and some mixtures of solvents and oil are very small as well as having the handicap of higher-thannormal viscosities. The centrifuges are completely sealed and may be pressurized, as required. In this example, the percentage of wax crystals is less than in the previous cases and this value will vary with the temperature. Even a small rise in temperature will liquefy some w a x - there is no fixed melting point. Both the wax and oil vary according to the source of the crude oil and specific testing has to be done in order to define the performance in each case. Fortunately, this can be done easily. It is also important not to have emissions. This is for environmental pollution reasons as well as the hazard of fire. In summary, several criteria must be met in order to use a disk nozzle centrifuge for the purification of a solids-liquid slurry, as follows: 1. A continuous supply of a large volume stream of a solids-liquid mixture. 2. An impurity being present that can be separated away by virtue of a differential settling rate to a solids fraction (includes classification and washing). 3. The solids to be small (below 1 micron up to 200 microns) and difficult to settle (otherwise a simple gravity device would suffice). 4. The product to have a substantial value. 5. A complex set of operating conditions can be overcome by virtue of having diverse experience already (corrosion, temperature, pressure, solvents, hazards). 6. High solids concentrations can be reached and high wash rates can be used as well as using multi-stage counter-current systems.
Separators 649 Dilution and displacement are the two washing modes used in process industries. Displacement washing is the more effective of the two, as with substantially less wash, more liquor can be displaced. Disk nozzle centrifuges with internal recirculation are uniquely equipped for displacement washing, because the wash is introduced into a preconcentrated underflow slurry, of which most of the mother liquor had been previously displaced within the separation chamber. Using equal amounts of wash versus feed liquor, displacement washing achieves a 70% removal while dilution washing is limited to 50%. Increasing the wash liquor to 2:1 improves the washing efficiency to 90% versus 67% for dilution washing. Washing of starch to remove protein, washing of TPA to remove soluble constituents and the removal of wax fractions from petroleum are just some of the more common and proven applications. The same principle of upflow used in displacement washing is used for classification purposes to classify very fine from coarse particles. A typical application is in kaolin where 0.5 micron or smaller particles are separated from 2 micron material. Centrifuges have the added advantage of being totally enclosed, contain a minimum of (process liquid) inventory, operate in a continuous mode, can be easily temperature controlled and are available in corrosion resistant materials.
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Chapter 1 8 Internal Mixers: Single- and Twin-Screw Extruders* INTERNAL MIXERS Internal (Banburyt-type) mixers have found widespread application in the rubber industry, where high-horsepower mixing is required for masticating and compounding; the technology has been successfully applied to plastics and chemicals for high- and low-viscosity systems:
Rubber Applications
Plastic Compounds
Other Applications
Tires, tubes
Polyvinylchloride (PVC)s (flexible, semirigid)
Adhesives, sealants Carbon electrodes Ceramics Chewing gum
Packing, sealing, roofing
PVC scrap reclaim ABS (molding) Polyethylene, polypropylene, PVC, ABS color concentrates Phenolic
Flooring (sheet)
Thermoplastic/rubber blends
Molded articles, profiles Hoses, gaskets, seals Shoe soles, heels Foam, sponge rubber
Dewatering, devolatilizing Pharmaceuticals
The common goal is to mix solid and/or liquid additives into a rubber or plastictype matrix. Additives (agglomerated particles or droplets bound by surface tension) must be separated, reduced in size, and uniformly distributed within the matrix: Two types of mixing phenomena are involved: extensive mixing and intensive mixing. Extensive (also known as distributive) mixing is responsible for spatial distribution of the individual particles within the polymeric matrix. Intensive (also known as dispersive) mixing is responsible for separating and reducing the particle size of the additives.
Principle of Operation Internal mixers are designed to provide intensive and extensive mixing. Intensive mixing occurs in the narrow gap formed between the rotor tip and the mixing chamber wall. The mixture is repeatedly passed through this high shear field where fluid mechanical stresses separate and rupture agglomerates; dispersive forces are * Source: Werner & Pfleiderer Corporation, Ramsey, NJ. Adapted with permission. t Banbury is a registered trademark identifying internal batch mixers made by Farrel Company and its predecessor companies since 1916. 651
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Process Plant Machinery
similar to those in a two-roll mill. Extensive mixing takes place between the rotors: the mixture is circulated from side to side and from one end of the mixer to the other after passing through the shear zone. Energy dissipation from intensive mixing results in heating of the mixture. This heat is removed through the walls of the mixing chamber, the rotor bodies, and other contact parts (ram, discharge door, etc.) through cooling channels. The effective heat transfer of internal mixers can be the limiting factor to intensive work, since discharge temperatures cannot exceed the critical temperature of the mixture, which is determined, for example, by thermal breakdown of organic phase, onset of undesirable reactions (e.g., cross-linking), or the decrease of continuous-phase viscosity to a point where dispersion cannot proceed. Poor heat transfer will cause these temperature limits to be reached before dispersion is complete.
Design Features of Internal Mixers Optimum dispersion is achieved through proper selection of machine type and process parameters. Internal mixers are available today with intermeshing and tangential (nonintermeshing) rotors. Rotor design is a critical factor in mixer performance, and some manufacturers offer various rotor configurations. (This topic is covered in detail later). Figure 18-1 shows the major components of internal mixers. A completely enclosed mixing chamber houses the spiral-shaped rotors, which rotate in opposite directions and at the same speed (intermeshing design) or at different speeds (tangential design) to keep the material circulating. The gap between the rotor tips and the chamber wall produces intensive shearing of the mixture. A hopper allows for loading ingredients, and an air-operated ram in the feeding neck confines the batch within the mixing chamber. A discharge door allows for quick and efficient unloading of the mixture at the end of the mixing cycle. The mixing chamber, rotors, and discharge door are all temperature controlled with steam and/or water.
Mixing Chamber The mixing chamber body is usually of two-piece construction, split vertically. This allows for removal on-site without dismantling the entire mixer. The internal body halves are lined with wear-resistant materials to maintain rotor tip-to-wall clearances; the exterior is heavily reinforced for mechanical loading. Temperature control of the mixing chamber is accomplished with steam or circulating water. The body halves are fitted with one of three types of chamber sides: cored, spray-type, or drilled. Cored sides have passages arranged in a serpentine pattern running along the length of the sides of the chamber. These passages are usually formed in the casting, through which water or steam is circulated. Spray-type sides use nozzles to spray water onto the sides for cooling effect; flood-type design is also available, without nozzles. Drilled sides are the most common type of temperature control for the mixing chamber. Holes are drilled laterally in the body halves and provide a serpentine path for the flow of water. These holes are smaller and greater in number than the cored passages, as well as being closer in proximity to the chamber wall.
End Frames The ends of the mixing chamber provide support for the mixer body. End frames carry the rotor bearing assemblies and dust-stop seals.
Internal Mixers: Single- and Twin-Screw Extruders
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Self-aligning roller beatings are standard on most internal mixers. Double-acting axial thrust bearings can be used on large internal mixers to increase lifetime and efficiency of the rotor seals. Dust stops are used to seal dust (carbon black, pigments, etc., used in the mix) within the mixing chamber. There are several designs available; one type relies on a hydraulically actuated yoke to apply pressure on the sealing rings. Self-sealing dust stops use the mechanical pressure of the mixture to maintain seals. Powders or dust that pass into the dust stops are flushed with process oil.
Discharge Doors The discharge door of internal mixers is designed to provide quick and efficient dumping of the mixture. The door top must be hard-surfaced to withstand the mixing environment; some designs include a removable/replacement door top section. Most discharge doors are provided with passages for heating or cooling media; the door is usually cooled to prevent the mixture from sticking. Two types are available: drop door and sliding door. The drop door type is pivoted on a shaft running through bearings located in the end frames. The door swings downward (135 ~ to 180 ~) and away from the mixing chamber, providing a clear path for the material discharge. The sliding door is mounted on an air-operated cylinder. Guides are necessary to provide clearances for door operation. Also available are mixing chambers that tilt over (up to 140 ~) to discharge the mixture through the top opening. Tilting mechanisms are electrically or hydraulically driven.
Feed Hopper The feeding of materials into an internal mixer can actually take longer than the mixing cycle. Thus, efficient designs are utilized on the hopper assembly of internal mixers to facilitate quick charging of ingredients. The feed opening itself must be large to allow venting of air (and dust) as the batch is charged. The feed opening and throat on larger mixers are sized to accommodate rubber bales intact. Air-operated doors are provided in the hopper assembly to allow for loading of material into the mixer. Openings on the sides or in the rear of the hopper for charging of fillers, accelerators, curatives, and other dry powder components are normally connected to weigh hoppers or other means of feeding.
Ram Cylinder Assembly A ram is used to confine the batch within the mixing chamber. Air pressure (10 to 120 psi) is applied to the cylinder, forcing the ram down into the mixing chamber. The bottom of the ram is usually shaped to conform to the gap between the rotors (V-bottom). The ram is fitted with a height indicator, used to gauge the state of the mix. Cavity cooling is used for temperature control of the ram bottom. The ram cylinder assembly is usually air-operated, but it can also be a hydraulic device. Ram pressure can have a significant impact on dispersion quality, and it is sometimes used as an operating parameter.
Drive Train There are three mixer drive arrangements available: "standard," semi-unidrive, and unidrive. Standard drive trains use reduction gears mounted on the mixer base to drive the rotors; one rotor shaft is longer than the other, functioning as pinion shaft and reduction gear. Semi-unidrive systems use a separate reduction gearbox prior to the rotor shaft. One rotor shaft is longer than the other, carrying the pinion gear. Unidrive systems have speed reduction and dual output shafts within a
Internal Mixers: Single- and Twin-Screw Extruders single gearbox. Rotor shafts of equal length are coupled to the pinion shafts from the gearbox. Rotor speed directly influences mixing quality, mixing time, and batch temperature. Optimum rotor speeds are chosen to process materials at their highest acceptable temperature within the shortest cycle time. A single-speed drive limits the number of formulations that can be processed optimally. Two-speed motors increase internal mixer flexibility. Several other alternatives are also available: 9 constant speed motor with integral (or separate) two-or four-speed gearbox 9 variable-speed DC or variable frequency AC motor (ultimate flexibility)
Bed Plate The bed plate is the base frame that anchors the mixing chamber (and possibly gear-box). It is strengthened to withstand torque transfer and vibration of the rotors, as well as to evenly distribute the load of the mixer onto a foundation.
Auxiliary Systems Operation of internal mixers requires several dedicated systems. The tempering system supplies constant-temperature fluid to heat or cool the various mixer components. Steam/water, pressurized water, or heat-transfer oils can be used. Several separate circuits may be needed for efficient operation (e.g., ram and mixing chamber at one temperature, discharge door at another temperature, and rotors at yet a third temperature). The lubricating oil system ensures adequate supply of lube oil to critical components in the drive train, rotor bearings, dust stops, etc. The process oil injection system injects oil into the mixing chamber. The injection nozzle should be a self-sealing type to prevent fouling from mixture. The hydraulic or pneumatic system operates discharge door (or tilting mechanism), ram cylinder, feed chute door, etc. It is controlled by means of solenoids. The temperature sensor, strategically located, indicates batch temperature. It may be located in the feeding ram, in the discharge door, or through the end frames. Sensors should have intimate contact with the mixture to provide accurate readings.
Instrumentation and Controls The degree of sophistication of internal mixer control systems can vary. They can 1. 2. 3. 4.
mix batch until desired temperature is reached mix batch for predetermined time period mix batch until predetermined energy is consumed use various combinations of the above
Efficient operation of internal mixers is attained with automation of the mixing process. Increased productivity and consistency in product quality can be realized with computerized control of the mixing line from batch weighing through downstream processing. Manual Control. Standard internal mixer control systems provide interlocks and data acquisition for manual operation. Interlocks are installed on ram cylinders, feed hoppers, and discharge doors for operator safety and on drive components for overload protection. Operating data are usually recorded on strip-charts housed in a control panel for production monitoring of discharge temperature, power
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consumption, and cycle times. Stop/start push buttons are provided for drive motor and auxiliary equipment. Automated Mixer Control Systems. Application of a process computer system to an internal mixer significantly improves batch quality as well as quality consistency from batch to batch. Flow of raw materials, mixing control, downstream equipment, and production planning can all be integrated into a supervisory computer system. Optimum mixer control is achieved with logic controls and/or combinations of mixing time, energy input, rotor speed, stock temperature, and chemical reaction parameters. Adaptive process control systems allow the mixing process to follow predefined energy and temperature curves that are stored in memory with each formulation. Rotors
Mixing is achieved in internal mixers with the rotors, rotating toward each other at the same speed (intermeshing design) or at different speeds (tangential design). The rotors are designed to interact with each other via the rotor blades, called wings. Short wings and long wings are used in combination on each rotor. Mixers are available with either two-wing or four-wing rotors. Two-wing rotors have one short and one long wing, while four-wing rotors have two short and two long wings. The rotors are arranged in the mixer such that the long wing of one rotor interacts with the short wing of the other rotor (Figure 18-2). The edge of the blade is called the wing tip, which forms the shearing gap between the rotor blade and the chamber wall. The rotors are temperature controlled with steam and/or water, which flows through the center of each shaft. Spray-type cooling or forced circulation is available. Rotors and shafts can be manufactured as a one-piece steel casting or as two pieces: forged steel rotor shafts with rotor bodies shrunk onto the shafts. The entire rotor bodies are hard-surfaced for wear protection or are chrome plated with hardfacing only on the tips.
FIGURE 18-2 Plan view of tangential four-wing rotors, rotor bearings, and dust stops. (Source: Werner & Pfleiderer Corporation, Ramsey, NJ. )
Internal Mixers: Single- and Twin-Screw Extruders
Rotor Design Intensive mixing occurs where the material is compressed between the rotor wing tip and the mixing chamber wall. The width of the tip, the clearance between the wing tip and the chamber wall, and the leading/trailing angle affect dispersion. Wing length and helix angle influence the distribute mixing from rotor to rotor and from one end of the mixer to the other.
Tangential Rotors Conventional internal mixer designs use tangential rotors with two or four wings. The rotor diameter is equivalent to the center distance between the rotors. Tangential rotor mixers are mainly used for large-volume mixing, as in the tire industry. Two-wing rotors are characterized by two flow regions due to different radial clearances between the rotors. Mixing intensity is directly influenced by ram pressure with a two-wing rotor system. Four-wing rotors provide constant radial clearances between the rotors. Total wing length is greater than the two-wing system, producing higher specific energy input and better homogeneity. The torque capacity of internal mixers supplied with four-wing rotors is approximately 30 percent greater than with the two-wing rotor system. Modified four-wing rotor designs have been developed with enhanced longitudinal (distributive) mixing.
Intermeshing Rotors Manufacturers of internal mixers have developed intermeshing rotor systems in response to quality and productivity requirements of compounders. The diameter of intermeshing rotors is greater than the center distance between the rotors (Figure 18-3). A calendar effect is created by the intermeshing geometry of the wings, resulting in improved dispersion. Intermeshing rotor systems are capable of higher rates of energy input and better heat transfer than tangential designs. The number of mixing steps, as well as mixing time for each step, may be reduced by changing from tangential to intermeshing rotors. Internal mixers with intermeshing rotors are used mainly for high-quality mixing.
Operation of Internal Mixers Several mixing techniques are commonly practiced using internal mixers: singlestage, masterbatch, and multistage mixing. Single-stage mixing is used for materials that can tolerate relatively long mixing time at low rotor speed; temperature rise is the limiting factor. All ingredients can be charged at the same time or added sequentially while mixing. The sequence of addition of plasticizers and oils is critical to dispersion quality. Masterbatching is an implied two-stage mixing process. Viscous components and fillers are mixed first. This stage can tolerate higher rotor speed and material temperature. Thus, short mix cycles produce complete dispersion. The batch is discharged and allowed to cool. A fraction of the first stage (masterbatch) is then loaded with the balance of ingredients that make up the total formulation. The predispersed masterbatch mixes efficiently and can produce higher quality dispersions in less total mixing time than a comparable single-stage mixing process. Multistage mixing can include several masterbatching steps, a remilling stage to disperse the masterbatch, and final mixing.
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FIGURE 18-3 Tangential (top) and intermeshing (bottom) rotor geometries. (Source: Werner & Pfleiderer Corporation, Ramsey, NJ.)
Operating Parameters Optimization of internal mixers requires knowledge of how operational parameters affect mixing quality. Generalized statements can be made as to the influence of fill factor, mixing sequence, ram pressure, rotor speed, mixing time, and temperature control. Fill factor is an indication of the working volume of the mixer. Chamber volume is used to specify internal mixing capacity (in liters or cubic inches); empty volume is measured with the rotors installed. The working volume is the space occupied by the mixture (function of batch weight, specific gravity). The ratio of working to empty volume gives the fill factor. Internal mixers are available with chamber volumes up to 650 liters (1), which can handle a 560-kilogram (kg) batch; laboratory-scale internal mixers used for research and development have net volumes as low as 0.41, requiring less than 0.3 kg per batch. Fill factors vary for different mixer geometries (two-wing versus four-wing; intermeshing versus tangential) and different mixing tasks. Intermeshing rotor mixers generally use lower fill factors than tangential rotor systems, but they can achieve the same level of dispersion in a shorter mixing time. A compound that runs on a tangential mixer at 70 to 75 percent fill factor (four-wing rotor) would run between 60 and 65 percent fill factor on an intermeshing mixer.
Internal Mixers: Single- and Twin-Screw Extruders
Fill factor directly influences dispersion quality, specific energy input, and discharge temperature by providing empty volume for the mixture to circulate within. Mixing sequence has a significant effect on dispersion quality, specific energy input, and discharge temperature when large amounts of oil are being processed. Oil or plasticizers are typically added later in the mixing cycle, after fillers have been dispersed. Ram pressure is used as a process variable to influence the specific energy input and discharge temperature. Tangential rotor mixers are more sensitive to ram pressure than intermeshing rotor mixers. Increased ram pressure can shorten the mixing cycle by compressing the batch within the chamber to intensify mixing. Discharge temperature can be reduced by decreasing ram pressure. Rotor speed is used to control the rate of specific energy input. Higher rotor speeds can reduce mixing time by dissipating more energy in a shorter time period. Rotor speed and ram pressure can be independently varied to achieve a target energy input, dispersion quality, or material temperature. Internal mixers supplied with a variable-speed motor have more flexibility in optimizing rotor speed for a given formulation than a single- or two-speed mixer. Large internal mixers operate at rotor speeds of 10 to 60 RPM; laboratory-scale internal mixers operate at speeds up to 110 RPM or higher. Mixing time in conjunction with fill factor (batch weight) determines the throughput capacity of the mixer. Decreasing the mixing time (by increasing rotor speed or ram pressure) then increases mixer output. Mixing times vary widely depending on formulation and quality of mix. For example, single-stage mixing of one particular rubber formulation and oil, which takes eight minutes on a tangential mixer, can be processed in only five minutes on an intermeshing mixer. A typical tire formulation running on a tangential rotor system takes about three minutes for the masterbatch stage, three minutes for the remilling stage, and two minutes for final mixing. Temperature control of mixer components (chamber body, rotors, ram, and discharge door) has several effects: adherance of material to rotors (rotor temperature), discharge temperature (cooling effect through chamber body), and efficiency of discharge (discharge door temperature). The mixer is usually started up hot to prevent slippage, and then cooling is applied when the material is in a fluxed state. Accurate temperature control of each part of the internal mixer helps maintain batch-to-batch uniformity.
Downstream Equipment An internal mixer usually discharges directly into some type of shaping/forming equipment or onto conveyors. Downstream equipment is sized to provide a continuous process from the batch mixer; two mixers can also be arranged with alternating discharge.
Mixing Mill Roll mills are used for cooling and shaping as well as for after-homogenizing and mixing. Cross-linking chemicals that cannot be added in the internal mixer due to temperature limitations in single-stage mixing can be added in the mixing mill. Mixing mills are built with fixed friction ratios (constant speed on rolls), or variable friction can be achieved with variable-speed control of each roll. Adjustment of the roll gap can be carded out under load with hydraulics.
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Extruder Single-screw extruders are used to produce pellets or sheet from the internal mixer. Pelletizing extruders are fitted with screens or strainers and die-face pelletizers or underwater pelletizers. Force-feeding devices (screws or rams) can also be installed. Sheet stock is produced from extruders equipped with a roller-die.
Maintenance of Internal Mixers Routine maintenance of internal mixers requires attention to the various subsystems responsible for smooth operation (e.g., hydraulic, pneumatic, and temperaturecontrol systems). Mixer manufacturers provide recommendations for preventive maintenance on rotor bearings and seals, lubricating oil changes, gearbox overhauls, etc. High on-stream factors are achieved when maintenance records are kept and factory recommendations are followed. Major overhauls of internal mixers are required when the product specifications (dispersion quality, discharge temperature, etc.) are no longer acceptable. Abrasive wear on rotors, chamber body, and ram becomes evident in mixer performance: dispersion quality cannot be maintained as clearances increase between rotor tip and chamber wall; discharge temperature increases as a result of increased clearances and subsequent reduction of heat transfer. Manufacturers of internal mixers can provide a field inspection service to periodically document wear. Critical components (rotors, bearings, dust stops, etc.) are usually kept in stock for emergency delivery. Rebuilding of internal mixers is a service performed by the manufacturers. Worn rotors and chamber bodies can be rebuilt to factory specifications. New duststop parts, door tops, and end frames are installed. Older mixer designs can be converted from cored or spray-type sides to drilled sides, from two-wing rotors to four-wing rotors, from standard gear to unidrive, etc.
SINGLE- AND TWIN-SCREW EXTRUDERS Screw extruders were designed as continuous mixers for dispersing additives into a molten polymer. Intensive or extensive mixing takes place in the extruder as a function of screw geometry and operating conditions. Mixing time in continuous mixers cannot be arbitrarily chosen as in batch mixers, but is determined by operating conditions, screw geometry, and extruder length (residence time). Evaluation of single- and twin-screw extruder designs for various process tasks requires a basic knowledge of key components: gearbox, drive motor, screw shafts, screws, and barrel (Figure 18-4).
Mechanical Description
Gearbox The gearbox performs speed reduction, torque transfer, and thrust loading for the extruder. These functions can be handled separately with individual components or combined in a single unit. Drive motor speed is reduced to operable extruder screw speed in the gearbox. Various gear ratios are available to provide different output speed ranges from standard motor input speeds. Torque is transferred from the drive motor to a single output shaft (single-screw) or must be equally split for dual output
Internal Mixers: Single- and Twin-Screw Extruders
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FIGURE 18-4 Extruder components. 1 - T h r u s t bearing assembly; 2 - m o t o r with overload safety clutch; 3 - gear box; 4 - f e e d barrel; 5-electrical resistance heater shells; 6 - kneading elements; 7 - thermocouple for stock temperature; 8 - venting section; 9 - screw elements; 1 0 - stock pressure gauge; 11 - thermocouple for barrel temperature; 1 2 - start-up valve; 13-screen pack changer; 14-pellet/water discharge; 15-UG(under-ground pelletizer); 1 6 - w a t e r inlet. (Source: Werner & Pfleiderer Corporation, Ramsey, NJ.)
shafts (twin-screw). Axial thrust loading (back-pressure) from the screw shaft(s) is taken up by thrust bearings. Single-screw extruders are able to utilize large thrust bearings. Twin-screw extruders, however, must use alternative methods, since the close proximity of the screw shafts precludes the use of large bearings. Twin-screw extruder gearbox designs are inherently more complex than single-screw designs.
Drive Motor Small single-screw (<8 in) and twin-screw (<130 mm) extruders use constanttorque variable-speed DC drive motors to provide process flexibility. Extruder systems can also use constant speed AC motors with mechanical variable-speed gearboxes, variable-frequency AC motors, or even hydraulic drives to achieve the same flexibility. Very large extruders are usually equipped with fixed-or two-speed drives, with an auxiliary drive for start up. Drive motors can be direct-coupled to the gearbox or belt-driven. Mechanical slip clutches, positive disengagement couplings, or shear-pin devices are installed to provide overtorque protection. Typical installed power for twin-screw extruders is shown in Figure 18-5.
Screw Shafts The screw shaft(s) transfer the motor torque to the screw(s). Splines, keys, and other methods of torque transfer are used to mount the screws to the shafts. One end of the screw shaft is coupled to the output shaft of the gearbox; on some extruders the screw shaft is an extension of the pinion shaft and is not removable. The screws are fixed to the screw shaft with a screw tip that compresses the screw bushings together on the shaft. Screw shafts can be cored for flow of heat transfer fluids that can enter through the gearbox end or the discharge end.
Extruder Screws The extruder screw(s) transfer the motor power into the material via viscous dissipation. Single screws are built as one piece or may be assembled from modular
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Installed power for twin-screw extruders. (Source: Werner & Pfleiderer Corporation, Ramsey, NJ.)
sections. Some twin screws are manufactured as one piece, but most are of modular design. The discrete screw elements offer two significant advantages" 1. Process flexibility to alter the screw geometry, thereby influencing shear, mixing, and residence time distribution. 2. As screw elements wear, only those pieces that are subject to abrasion need to be replaced, not the whole screw. Special materials of construction can be used for those particular screw pieces subject to abrasion or corrosion instead of for the entire screw. Single screws typically have one or two flights; twin screws can have one, two, three, or more flights together on the same shaft (Figures 18-6 and 18-7). Screw elements can have various pitch angles, lengths, and number of flights, and can be forward or reverse conveying. Co-rotating twin-screw extruders additionally use staggered screw discs (called kneading discs, mixing paddles, etc.) to provide additional mixing. The screw discs have various widths and stagger angles, as well as being forward or reverse conveying to tailor the amount and type of mixing required.
Extruder Barrel As with screws, barrels can be built as one piece or of modular construction. The modular design provides the same advantages as for screws. Barrel sections are open for feeding and venting or closed for conveying and pressure build-up. The extruder barrel can have several openings for venting or feeding solids downstream. Access ports are provided to inject liquids or for installing melt thermocouples or pressure transducers. Temperature control of the barrel can be accomplished in several ways: electric heating (with water or air cooling), steam heating, or fluid heating/cooling.
Internal Mixers: Single- and Twin-Screw Extruders
FIGURE 18-6
Cross section of counter-rotating screws. (Source: Werner & Pfleiderer Corporation, Ramsey, NJ.)
FIGURE 18-7 Cross section of co-rotating screws: single-flight (top), two-flight (center), three-flight (bottom). (Source: Werner & Pfleiderer Corporation, Ramsey, NJ.) Barrels either have drilled passages or are equipped with external jackets for flow of heating/cooling fluids. Maximum operating temperatures are usually limited by the materials of construction of the extruder barrel, screws, etc. Electric heating is provided with bolt-on resistance or induction heaters. Water or air is used to trim temperature. Several temperature zones can be installed, each with a dedicated controller. Steam heating requires mixing valves to control temperatures
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Process Plant Machinery
or flow control valves to control flow rate. Pressure ratings of barrel jackets or cored passages may limit steam pressure for high-temperature applications. Fluid heating/cooling provides very stable temperature control; commercial heat-transfer oils operate up to 650 ~ while glycol solutions may be used for intensive cooling. Extruder barrels are designed to provide access to the screws for cleaning, maintenance, etc. Clamshell designs are available, which open horizontally (top and bottom) or vertically (side-to-side). In other cases, the screw is pulled from the barrel, or the barrel housing can be pulled, leaving the screw attached to the gearbox. Barrels can be manufactured with a replaceable liner for wear- or corrosionresistance. The liners are press-fit or a split-barrel designed to bolt around the liner. Only those barrel sections where wear or corrosion is anticipated need to be built of special materials or replaced when worn.
Miscellaneous A feeding device, located near the feed opening of the extruder, is a starting point common to single- and twin-screw extruders. Its purpose is to provide a continuous source of raw materials from a storage bin or mixer/surge system to the extruder. Volumetric or gravimetric metering equipment is used to regulate throughput with twin-screw extruders; a simple gravity bin with bottom discharge can be used with single-screw extruders, since the extruder operates at 100 percent degree-of-fill in the feed section (flood feeding). Starve-fed extruders are very sensitive to feeding; disturbances or fluctuations may not be dampened out within the extruder. Separate feeding of several components to a twin-screw compounding extruder requires a gravimetric feeding system for accurate metering. Force-feeding devices are used in conjunction with single- and twin-screw extruders for improved feeding of low-bulk-density materials. These are available in single-or twin-screw configurations for side feeding or vertical feeding (Figure 18-8). Shear control mechanisms are designed to alter the shear intensity of the screw(s) while the machine is running. Several designs are available; most operate on the principle of a variable orifice restricting flow in the screw channels or at the screw tips. Downstream equipment is used in conjunction with single- and twin-screw extruders to form, pelletize, and filter. Extrusion dies are available to produce strands, sheet, film, fibers, and profiles. Heat- and shear-sensitive materials can be processed in a two-stage extrusion process: a second extruder or gear pump is used to pressurize the die, relieving the compounding extruder of this task. Pelletizing equipment depends largely on the properties of the material. Molten strands can be pulled through a water bath (if the material has enough strength) and fed into a strand cutter. Underwater pelletizers are used if the material tends to smear when cut in air. Water-ring pelletizers and other types of die-face cutting are possible if the material can be cut in air and cooled in an air or water conveying system. Screens are used to filter molten polymers for film and fiber extrusion. Fixed screens, screen-pack changers, or deep filter elements are installed between the extruder and the die assembly.
Process Task In most thermoplastic extrusion operations, the basic function of the extruder is to accept a product, melt it through a combination of mechanical and thermal energy,
Internal Mixers: Single- and Twin-Screw Extruders
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FIGURE 18-8 Downstream feeding of abrasive resin extenders and additives reduces extruder wear and maintenance costs. (Source: Werner & Pfleiderer Corporation, Ramsey, NJ.)
and generate sufficient pressure to pass the molten polymer through a die. Sometimes the product is fed to the extruder in the form of a melt. In these cases, the task of the extruder is limited to providing additional energy input for homogenizing the melt with other ingredients and then to generate sufficient pressure to pass the molten material through a die. Other process tasks include 9 9 9 9
devolatilization of residual monomers, solvents, water, etc. chemical reactions such as polymerization, polycondensation, polyaddition polymer alloying and blending incorporation of fibers (glass, carbon, etc.) for reinforcing
Screw extruders are applied in many areas outside of the plastics industry, for example: 9 9 9 9 9
food processing (extrusion cooking of starch-based raw materials) energetic materials (mixing and forming of propellants) adhesives and sealants (compounding and feeding of coating equipment) chemicals and pharmaceuticals (reactive processing) hazardous waste treatment (encapsulating, solidification)
Design Feature- Single-Screw Extruders A helical screw that rotates inside a cylindrical barrel housing can be referred to as a single-screw extruder. The screw consists of three sections: feeding, compression, and metering.
665
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Process Plant Machinery The feed section consists of deep screw flights designed to accommodate feed intake of low-bulk-density powders or pellets. Single-screw extruders usually operate with the feed section filled 100 percent (flood feeding); less than 100 percent fill in the feed section is called starve-feeding. The compression section (sometimes referred to as the transition zone) is designed to input mechanical energy for melting, cooking, etc. Compression ratio is the term used to describe the ratio of flight depth or pitch in the feed zone to flight depth or pitch in the metering zone; typical compression ratios are from 1:1 to 5:1. There are two basic design approaches to achieve compression: 1. increase root diameter (decreasing flight depth) 2. decrease pitch, constant root diameter (constant flight depth) The metering section is located before the die restriction. Most of the mechanical energy dissipation occurs in the metering section, producing the highest temperature in the extruder, high shear rates, and mixing. Enhanced mixing is achieved in the metering section with interrupted flights, mixing pins, screw barriers, etc. The barrel housing may be grooved on the inner surface to improve frictional characteristics in the feed zone. Small grooves are cut into the barrel wall, either straight (axially down the barrel length) or spiral (helical grooves cut opposite to the conveying direction of the screw). Several hybrid single-screw extruder designs are available; one incorporates a single screw that reciprocates as it rotates. Fixed pins on the barrel fit between interruptions in screw flights to promote mixing within the screw channel with each revolution. Some designs provide blades instead of pins, which can be adjusted to more or less restrict flow, depending on the angular position of the blade; shear intensity can be varied to suit the process application.
Design Features- Twin-Screw Extruders Twin-screw extruders can be classified according to the direction of screw rotation and to the amount the screws intermesh with each other (Figure 18-9). A screw system that is lengthwise open (down channel) has a passage from the inlet to the outlet of the machine. Material exchange can take place lengthwise in the screw channel. In a lengthwise closed design, the screw flights in the longitudinal direction are closed at intervals. Crosswise open channels allow material exchange from one flight to another; in crosswise closed channels no material transfer is possible (neglecting mechanical clearances). Whether the screws are open lengthwise or crosswise or have a closed geometry has a direct effect on conveying characteristics, mixing, and pressure-generating capacity of the screw system. Nonintermeshing twin-screws extruders are closest in principle to single-screw extruders. Two screws are arranged next to each other with the center line distance a little greater than the screw diameter; the screws can rotate at the same or different speeds. The interaction of the screws is limited to a random passage of material from one screw to the other (nonintermeshing screws are open both lengthwise and crosswise). Nonintermeshing systems are available with co-rotating or counterrotating screws. Fully intermeshing twin-screw systems can be co-rotating (both screws turn in the same direction) or counter-rotating (screws turn opposite, either toward or away from each other at the tip of the extruder). Fully intermeshing, counter-rotating twinscrew extruders are characterized as lengthwise and crosswise closed, similar to a screw pump. Fully intermeshing, co-rotating twin-screw extruders are lengthwise open and can be crosswise closed with conveying screws or crosswise open when
Internal Mixers: Single- and Twin-Screw Extruders SCREW ENGAGEMENT
SYSTEM
COUNTER-ROTATING
CO-ROTATING
LENGTHWISE AND CROSSWISE CLOSED
THEORETICALLY NOT POSSIBLE
LENGTHWISE OPEN AND CROSSWISE CLOSED LENGTHWISE AND CROSSWISE OPEN
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LENGTHWISE OPEN AND CROSSWISE CLOSED LENGTHWISE AND CROSSWISE OPEN
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Screw meshes and configurations. (Source: Werner & Pfleiderer Corporation, Ramsey, NJ.)
staggered screw discs are used. Fully intermeshing twin-screws are also known as self-wiping. Partially intermeshing twin-screw extruders are lengthwise open and can be crosswise closed (counter rotating) or crosswise open (co-rotating or counterrotating). The degree of interaction between the screws is more pronounced than the nonintermeshing systems. Twin-screw machines are now available that can operate both as a co-rotating and counter-rotating extruder using the same gearbox; a lever device is mounted on the gearbox to change rotation, and the appropriate screws are installed. Another twin-screw system that is not shown in Figure 17-9 is a fully intermeshing, counterrotating conical screw system. The screw diameters decrease from the feed toward the discharge. This system provides a large free volume for feeding low-bulk-density materials and a large heat transfer surface area.
Single-Screw versus Co-Rotating versus Counter-Rotating
Feeding and Conveying of Solids When fed with a solid, conveying is the first extrusion operation. Like any screw conveyor, an extruder operates on the principle of frictional relationships between a solid and the surrounding barrel walls. In order to analyze solids conveying, a force balance has to be made. The balance of the frictional forces determines the solid plug transport. The driving force is the friction between the barrel surface and the plug and is proportional to the pressure exerted by the plug on the barrel. The retarding forces are friction forces on the flights and the root of the screw.
668
ProcessPlant Machinery An increased coefficient of friction between the plug and the barrel wall (such as barrel grooves) and a lower coefficient of friction at the flights and the root of the screw (smooth screw) will result in better solids conveying. This conveying mechanism applies both to single- and twin-screw systems. Single-screw extruders have deep-cut flights in the feed section to maximize volumetric capacity. Throughput is influenced by screw speed, since flood-feeding is commonly used. Barrel grooves improve frictional relationship for solids conveying. Counter-rotating twin-screw extruders turn outward at the top, inward at the bottom (if the screws turned inward on top, the material would have to be pulled in by the calendar gap formed between the screws). The material is conveyed to the lower wedge where it is partially compressed and conveyed as a unit volume. In order to utilize the largest possible free volume, multiflighted deep-cut screws are used. Since the conveying principle of the counter-rotating screws forms closed chamber unit volumes, this machine is ideal for feeding solids. The entire free volume can be utilized, assuming optimum pitch angle. Co-rotating twin-screw extruders convey material from one screw toward the lower wedge, where the material is compressed and then picked up by the other screw and conveyed further. The flow path is then a figure-eight shape, material being continuously transferred from one screw to the other around the periphery as it moves downstream.
Melting Process The melting process in all screw machines can be divided into two sections: 1. Compacting of the solids to eliminate air pockets. The air in most cases escapes through the feed opening. Simultaneously, a melt film is formed on the barrel walls; the material fuses due to heat and pressure. 2. Generation of melt through sheafing of the material and mixing of the unmolten particles with the already formed melt. As a result, a large heat exchange surface is generated between the solid particles and the melt. Single-screw extruders achieve compaction in the compression section of the screw where air is forced back to the feed opening. A melt film on the barrel wall is scraped off by the pushing side of the screw flank. The melt pool collects in front of the pushing flank. Fully intermeshing, counter-rotating twin-screw extruders with tight clearances cannot efficiently compress the air out of the closed chambers toward the feed opening. Loose clearances are intentionally created for this reason. Solid particles are drawn into the wedge area and softened; the plasticized mass is taken in by the calendar gap and collected at the trailing flank. Solids compression in co-rotating twin-screw extruders is produced through flow restriction. Reverse pitch screw elements or staggered screw discs are used to convey in the opposite direction, causing a pressure build-up upstream of the reverse pitch section. Air is allowed to flow back toward the feed, since the system is lengthwise open. Melting results from the backup length generated by the reverse pitch section.
Venting Removal of volatiles, moisture, or entrapped air is commonly practiced in singleand twin-screw extruders. Screw geometry, vent port design, and process parameters will vary for venting on single-screw, co-rotating, and counter-rotating twin-screw extruders. Vent ports can be operated atmospherically or under vacuum, depending on the amount of volatiles to be removed.
Internal Mixers: Single- and Twin-Screw Extruders
Single-screw extruders use reduction of pitch or flight depth (increase in root diameter) prior to vent ports to produce a pressure drop for venting. The degree of fill under the vent is less than 100 percent, preventing material from being forced out of the screw. Downstream of the vent(s), the screw must have another compression zone and metering zone to overcome die pressure. Counter-rotating twin-screw extruders have a reduced pitch prior to the vent opening to create a pressure drop under the vent where pitch may be increased again. Non-intermeshing (tangential) designs turn into one another at the top and drag material away from the vent opening. Lengthwise and crosswise closed channels on fully intermeshing counter-rotating extruders create melt sealing for vacuum venting. Fully intermeshing counter-rotating screws turning away from each other also prevent material from being forced out of the screw channels. Co-rotating twin-screw extruders use reverse pitch screw elements or staggered screw discs to form a restriction in the screw. Conveying screws upstream of the restriction must overcome this resistance, resulting in a pressure drop in the vent area. Large pitch screws are installed in the vent section to provide a low degree of fill. Several types of vent ports are used to keep the up-turning side of the screw(s) covered, preventing material from being forced out of the vent opening. Mechanical vent "stuffers" are available to keep material in the screw flights while under vacuum.
Pressure-GeneratingCapacity The pressure-generating capacity of an extruder is a function of the screw geometry. Single-screw extruders have lengthwise open channels; pressure is generated by the metering section. Single-screw machines generally use the entire screw length to overcome die resistance (100 percent degree of fill throughout extruder). Fully intermeshing, counter-rotating twin-screw extruders provide positive displacement through tight clearances. Closed chamber volumes intermittently releasing may lead to pressure fluctuations; enlarged clearances are sometimes used to overcome these pulsations, which also decreases positive-displacement capabilities. Fully intermeshing, co-rotating twin-screw systems provide partial positive conveying through the wedge resistance. Material viscosity plays an important role in pressure build-up where a lengthwise open channel exists. Due to downstream restrictions, such as dies, the melt accumulates along the screw channel upstream, over several turns of the screw. Back-up length (zone of 100 percent degree of fill) is a function of material viscosity, screw geometry, and die resistance.
Process Variables Process optimization on screw extruders requires special attention to the interaction of process variables and their influence on process parameters. Length/diameter (L/D) ratio is used to characterize extruder size. Diameter refers to outside screw diameter, length refers to the effective barrel length. Singleand twin-screw extruders are available from L/D = 3 to L/D = 48. The L/D ratio is kept constant for geometric scale-up. Residence time is a function of extruder L/D ratio. Screw geometry is a key factor in extruder performance. Screw configuration will influence residence time, mixing quality, specific energy input, discharge temperature, and degree of fill.
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Process Plant Machinery
Single screw compression ratio, length of metering zone, and other geometric variables are well documented and predictable for thermoplastic processing. Barrier-type screws permit melt recirculation to improve melting and mixing performance. Counter-rotating twin-screw extruders require selection of screw pitch, clearances between screws and between screw and barrel, and flight width. Interrupted flights are also used to enhance mixing. Co-rotating twin-screw extruders offer different pitches and conveying directions (fight or left hand); additional shear/mixing is provided with staggered screw discs (kneading discs). These mixing elements have various widths and stagger angles to influence mixing intensity and residence time distribution (Figure 18-10). Screw speed has a direct influence on mixing quality, specific energy input, residence time, and degree of fill. Single-screw and co-rotating twin-screw extruders can operate at 300 to 500 RPM; counter-rotating twin-screw extruders and very large co-rotating twin-screw extruders run much slower due to mechanical limitations (Figure 18-11). Throughput is used as an independent variable on twin-screw extruders; throughput is dependent on screw speed in single-screw extruders that are flood-fed. Throughput affects degree of fill, residence time, specific energy input, and mixing quality. Common throughputs for twin-screw machines are shown in Figure 18-12. Barrel temperature has an effect on specific energy input and discharge temperature. The temperature profile can influence material viscosity, effectively thinning or building viscosity in the extruder screw. Extruders processing high-temperature polymers operate with barrel temperatures at 400 to 500 ~ extruders used for intensive cooling have glycol solution circulating through the barrel jackets and screw shafts at - 15 ~ Die pressure can have a significant impact on extruder performance. Open channel screws will have difficulties in generating high pressure with a very low viscosity product, where a closed screw channel will have no problem. Die pressure can be used as a variable in this way to influence degree of fill (called backup length), residence time, and specific energy input (open channel machines). Extrusion pressures for plastics processing are in the range of 500 to 3500 psi. Screw extruder designs are available for continuous service ratings of 9000 psi.
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FIGURE 18-10 Different screw arrangements in the plasticizing zone of a twin-screw, co-rotating extruder: (A) left-hand screw; (B) right-hand kneading block, left-hand screw; (C) right-hand kneading block; (D) large pitch left-hand screw. (Source: Werner & Pfleiderer Corporation, Ramsey, NJ. )
Internal Mixers: Single- and Twin-Screw Extruders 550
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Commercially common screw speeds for twin-screw machines. (Source: Werner & Pfleiderer Corporation, Ramsey, NJ.)
FIGURE 18-11
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FIGURE 18-12 Common throughputs for twin-screw machines. (Source: Werner & Pfleiderer Corporation, Ramsey, NJ.)
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Process Plant Machinery
Process Parameters The combination of extruder process variables produces a unique set of process parameters (residence time distribution, specific energy input, heat transfer, mass transfer, and degree of fill) responsible for a given product quality. The objective in scale-up is to reproduce those unique process parameters at a higher rate with larger equipment.
Residence Time Residence time distribution (RTD) in a screw extruder is a compromise between an ideal tubular reactor (plug flow) and a cascade of ideally mixed stirred tank reactors. The RTD of a screw extruder is more plug-like than a pipe of comparable L/D. Single screw, non-intermeshing and partially intermeshing twin-screw extruders are not self-cleaning and exhibit long residence time tails. Fully intermeshing twinscrew extruders are self-cleaning; residence time distribution is sharply defined. Axial mixing in a longitudinal direction is shown in the width and length of the residence time distribution curves. Average residence times can vary from less than ten seconds (short extruder, high speed) to more than ten minutes (long extruder, slow speed, low throughput).
Specific Energy Specific energy input is defined as kilowatt-hours per kilogram (or horsepowerhours per pound) and describes the mechanical energy introduced into the product by the screw (specific mechanical energy) and thermal energy through the barrel jacket (specific thermal energy). Specific energy is a function of degree of fill, screw design, extruder length (L/D), and temperature.
Heat Transfer Heat transfer in screw extruders occurs between the material in the screw channel and the barrel wall (or through the screw shaft for additional heat transfer surface area). Screw geometry and extruder design play a decisive role in achieving good heat transfer. Fully intermeshing co-rotating and counter-rotating twin-screw extruders provide constant renewal of material layers in the screw channel; multiflighted screw systems provide shallow flight depths for better conductive heat transfer. Measurements with polymer melts on fully intermeshing, co-rotating twinscrew extruders have given heat transfer values of 200 to 500 kcal/m 2 hr ~
Mass Transfer Devolatilization of solvents, monomers, reaction products, or water is accomplished on screw extruders with the use of vent port openings in the barrel. Volatiles can be removed atmospherically or under vacuum, depending on the process requirements. Limiting factors for mass transfer in screw extruders are equilibrium volatile content, residence time in the vent, vapor velocity (vent open area), and film thickness in the screw channel. Devolatilizing extruders usually have several vents to provide staged vacuum strength, increased residence time, and maximum open area. Single-screw extruders use shallow flight depths, while twin-screw extruders use large screw pitch under the vents to produce thin film thickness (low degree of fill) for enhanced mass transfer.
Degree of Fill The degree of fill in screw extruders is influenced by a combination of process variables. Single-screw extruders are typically 100 percent filled throughout the screw length (except for venting extruders), while twin-screw extruders operate
Internal Mixers: Single- and Twin-Screw Extruders
at anywhere from 20 to 100 percent fill along the screw length. Degree of fill is manipulated by screw geometry (forward or reverse conveying, screw pitch), throughput, and screw speed. Low degree of fill is required for venting or downstream feeding and high degree of fill is necessary for pressure build-up.
Extruder Instrumentation and Controls Extruders are equipped to varying degrees with instrumentation for obtaining process information. Analog or digital meters and strip-chart recorders are usually provided with the appropriate sensors: 9 Torque indicates percent of available power that is transferred to the process
9 9 9
9 9
material. Torque may be displayed as percent or motor amps used to indicate load (torque instrumentation must be calibrated to 100 percent of available power). Kilowatts may be displayed to provide an absolute value of energy consumption. The signal is the product of torque and speed signals. Screw speed is taken from a tach generator on the drive motor or a speed pickup in the gearbox. Die pressure is displayed as an absolute value (psi) from a pressure transducer placed in the extruder die, or locally indicated with pressure gauges. Thrust bearings in the extruder gearbox may be fitted with pressure transducers or strain gauges for pressure measurement. Barrel temperatures indicate the actual steel temperature detected by thermocouples placed along the extruder barrel. Product temperature is taken from a thermocouple placed in the barrel (or die) that actually contacts the material in the extruder.
Additionally, thermal (heater) power may be indicated with ampmeters or totalizers, flow rates of components from gravimetric feeders, pelletizer speed, and so on. Start/stop push buttons are provided to energize main drive, auxiliary pumps, feeders, and potentiometers used to control screw speed, pelletizer speed, etc. Screw extruders are operated with open-loop (manual) or closed-loop (automatic) control systems. Microprocessor-based control systems can operate any number of extruders, including upstream and downstream equipment.
Open-Loop (Manual) Controls Open-loop control is used to describe manual setting of feed rate, screw speed, barrel temperature, etc. Steady state conditions are reached when material temperature, torque, die pressure, and barrel temperatures are at equilibrium ("lined-out"). Thermal equilibrium takes the longest to achieve, since the mass of material within the screw is very small compared with the mass of metal in the barrel. Several control system interlocks are typically used to prevent mechanical damage from occurring: 9 Underload protection prevents the extruder screw(s) from running dry in the
event of a feeder malfunction or loss of feed material. Typically, extruder drive shuts down when motor torque remains below a preset value for a specified time period. 9 Overload protection prevents damage to screws, shaft(s), and gearbox. High motor torque shuts down the extruder main drive or causes the coupling to disengage. 9 High pressure interlock for protecting the thrust bearings in the gearbox will shut down the main drive.
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Process Plant Machinery 9 Lube oil pressure and temperature alarms will shut down main drive for low lube
oil pressure or high oil temperature to protect the gearbox from damage. 9 Large extruders may have strategic bearings in the gearbox instrumented for shock pulse measurements (SPM) to warn of possible bearing failure. The SPM system is usually programmed to shut down the main drive when the signals reach preset values.
Closed-Loop (Automatic) Controls Closed loop control of screw extruders provides the same interlocks as in an openloop control system, and additionally may include automatic start-up and shutdown, data acquisition, and control of one or several process parameters. Automatic start-up and shutdown is optimized for each product to reduce the amount of scrap material produced, resulting in increased productivity. Start-up and shutdown is reproducible; ramp functions control screw speed, feeders, etc. Data acquisition requires constant monitoring of process variables. Production reports are generated from process data. Automatic control of screw extruders provides a method for on-line quality measurement and control. Several control strategies are possible: 1. Constant melt temperature can be controlled by adjustable shear-control devices. 2. Constant pressure at the screw tips can be controlled by adjustable throttle devices. 3. Constant specific energy can be maintained by adjusting feed rate or screw speed. 4. Constant melt viscosity can be controlled by metering viscosity-controlling additive(s) (e.g., free radical chain scission reactants) or process variables (e.g., screw speed). On-line methods of measuring various aspects of product quality (e.g., viscosity, color, composition, dispersion quality) are being developed to further increase productivity of screw extruders.
Scale-Up Scaling up is an extrusion operation that requires monitoring of all process variables on laboratory-scale or pilot plant equipment. Scale-up factors are increasing as larger extrusion equipment becomes available; e.g., polyolefin compounding extruders are capable of 20,000 kg/hr, and larger capacity extruders are being designed. Determination of the limiting factor in scale-up places emphasis on accuracy of measurements and reproducibility of pilot plant experiments.
Scale-Up of Feeding and Conveying A step-by-step analysis of extrusion technology can be accomplished by breaking the process down into its elementary unit operations (feeding, melting, mixing, conveying, venting, and pressure generation). Process parameters that influence scale-up can be isolated, including the effect of basic extruder design (free volume). Small single screw extruders (3-in to 6-in screw diameter) are relatively easily controlled in terms of power input, feed rate, and conveying. In larger extruders that have much deeper screw channels, the frictional relationship between the material and barrel is very difficult to control. Figure 18-13 shows a typical solids-fed singlescrew extruder with one vent port. In the feed section, the solids feed must be compacted to create sufficient friction on the inside barrel walls to assure efficient intake; careful temperature control of the feed barrel area is required. Axial or spiral
Internal Mixers: Single- and Twin-Screw Extruders
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CONVEYING
ENERGY CONVERSION
FIGURE 18-13 Basic conveying functions of single-screw extruder and need for addition and withdrawal of energy. (Source: Werner & Pfleiderer Corporation, Ramsey, NJ.)
grooves in the barrel may be necessary in addition to precise temperature control, especially when materials with low coefficients of friction are fed (e.g., polymers with slip additives or lubricants). The basic conveying capabilities of single-screw extruders depend on a number of interrelated factors, some of which are not easily controlled. To achieve proper conveyance in the solids, melt, mixing, and pressure build-up areas, screw design and screw operating conditions are all closely interdependent. This interrelationship becomes more critical with larger screw diameters and larger screw channels. Twin-screw extruders have improved feeding and conveying characteristics due to the screw geometry. Feed intake and conveying of slippery or wet materials is predictable, dependent solely on free volume (function of third power of screw diameter) and screw speed. Twin-screw compounding extruders usually run faster and provide higher free volume/length than comparable single-screw machines and, therefore, have improved volumetric intake capability.
Scale-Up of Melting and Mixing An extruder is an energy conversion device. In order to maintain a proper energy balance, it is generally necessary to add heat by conduction and to withdraw heat by cooling. The poor heat transfer characteristics of polymer melts are well known; problems associated with heat transfer become more severe with increasing screw channel depth and increased thickness of the polymer layer in the screw channel. The heat transfer area in an extruder increases with the second power of screw diameter, but the volume of material with a given residence time increases with the third power of screw diameter (assuming a constant L/D ratio and comparable screw geometry). Heat transfer scale-up in single-screw extruders becomes increasingly difficult with larger screw diameters; the ratio of inner barrel surface area to volume in twin-screw pilot plant extruders does not differ much from production size units. Laminar flow patterns in the screw channel of single-screw extruders are disrupted with sufficient statistical frequency by leak-flow over the screw flights leading to changes of shear direction; this is essential for good mixing. As screw diameter increases, the statistical frequency (per volume of material) of shear direction changes decreases, often to a point where adequate mixing cannot be achieved. Additional shear and mixing devices are used as modifications of basic single-screw concepts (e.g., barrier-type screws, mixing pins, etc.) with good results in operation, but are difficult to scale-up.
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Process Plant Machinery
Twin-screw extruders provide rearrangement of material layers with every rotation as material is transferred from one screw to the other. The direction of shear relative to the original shear planes can be changed with selection of screw pitch (counter-rotating) or use of staggered screw discs (co-rotating).
Scale-Up of Venting Operations In removing large quantities of air or volatiles, the problems of foaming and entrainment exist. In the case of removal of volatiles in low concentrations, the problems are primarily caused by very slow diffusion speed of the volatiles in the molten polymer. Residence time, film thickness, and renewal of surfaces become a significant factor in venting. The thickness of the material layer controls diffusion rate; increasing screw channel depths on large-diameter screw extruders makes diffusionlimited scale-up difficult for single- and twin-screw machines. Vent open area and exposed screw surface area are scaled to produce similar vapor velocities in the vent ports.
Scale-Up of Pressure Generation Screw extruders dissipate energy in generating sufficient pressure to pump molten material through a die. This energy results in temperature rise of the material being pumped. Controlling discharge temperature becomes increasingly difficult with deeper screw flights due to poor heat transfer. The pressure drop for a given die geometry can be calculated, given the rheological behavior of the material. The scale-up task is to predict the temperature rise in screw extruders as a function of pressure. The "wetted" section of screw used to build up pressure is called the backup length (zone of 100 percent degree of fill). In single-screw extruders, the entire screw is 100 percent filled (except for vent areas); the length of the metering zone with fully molten material is used to determine pressure-generating capability and temperature rise. The backup length in twin-screw extruders might only be a few turns of the screw. Because of improved pumping efficiency, viscous energy dissipation of intermeshing twin-screw extruders is greatly reduced. Scale-up models can predict temperature rise in generating a specific pressure with a given screw geometry (single- or twin-screw) and rheological properties of the material; backup length (twin-screw) or length of the metering zone (single-screw) necessary to develop this pressure can also be calculated.
Materials of Construction Screw extruders experience abrasion and/or corrosion depending on the process application; materials of construction for screw and barrels are selected for optimum service life. Most screw extruder components are manufactured from carbon steel and may be nitrided or through-hardened for wear protection. Modular designs of screw elements and barrels have distinct advantages for corrosion and wear protection. New machines only have to be equipped with wear-protected components in the area where wear or corrosion is expected. In a corrosive environment, the generation of a galvanic element by pairing of different materials has to be avoided. With varying barrel wear along the processing section, barrels can be exchanged and relocated so that all barrels are worn down uniformly to the point where a replacement becomes necessary. The same philosophy applies to screw elements. Besides adhesive wear, which is experienced in some extruders due to contact between screw and barrel, the following wear mechanisms may be observed:
Internal Mixers: Single- and Twin-Screw Extruders 1. abrasive wear due to the material being processed or fillers contained in the
material (e.g., glass-fiber-reinforced thermoplastics, mineral fillers, etc.) 2. corrosion due to aggressive acidic or basic components in the processed material
(e.g., fluoropolymers, chemical reactions, etc.)
Designs for Wear-Protected Barrels Several designs are available for manufacture of barrels for corrosion- and wearprotection. Exchangeable liners provide the most flexible and economic solution. Liners can be press-fit into barrels or split-barrels designed to bolt around the liner. The space between the liner and split-barrel is usually filled with heat-transfer cement to eliminate air gaps. Barrel liners are typically through-hardened, manufactured from castings, or machined from forgings. High chrome, high carbon alloys provide excellent wear resistance due to a high content of chromium carbides. Bimetallic liners may also be utilized, with two bimetallic tubes welded together for twin-screw applications. The bimetallic liner has high adhesive wear resistance (NiCrB or NiCrMo alloys with or without carbide reinforcement) with a ductile backing material. Barrels that do not have liners may have surfaces welded with hardfacing materials and machined to original bore dimensions. Unlined barrels may also be machined to accept a liner. Corrosion-resistant barrels or liners may be manufactured from hardenable chromium steel or CoMoCr alloys for extremely high corrosion resistance in addition to very good wear resistance.
Designs for Wear-Resistant Screws Extruder screws are subject to the same corrosive and/or abrasive environment as barrels and are manufactured from material suited to the application. The crests of screw flights can be welded with hardened materials (e.g., stellite), screws can be through hardened (e.g., tool steel), spray-coated (e.g., carbide coatings), or chrome plated; corrosion-resistant materials (e.g., lnconel | -type, hardenable stainless steel) are also available. Extruder manufacturers can provide recommendations for materials of construction based on field experience.
Extruder Maintenance Routine maintenance of screw extruders can be divided into two sections: gearbox/drive motor and process section. Extruder manufacturers provide recommendations for preventive maintenance intervals based on experience in normal operations; modifications to these intervals will depend on the particular process application.
Gearbox~Drive Motor Maintenance Scheduled maintenance on the extruder gearbox involves checking lubricating oil level, temperature, and pressure on a regular basis. Lube oil specifications are provided by the extruder manufacturer. Service life of gear oil depends on gearbox loading and environmental factors, but it typically should not exceed 12,000 operating hours or three years. Oil should be inspected every 2,000 operating hours (sample analysis). Radial bearings have an expected lifetime as a function of average screw shaft speed and average load. This information is provided by the extruder manufacturer to schedule radial bearing replacement in the gearbox.
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Process Plant Machinery
Thrust bearings also have an expected lifetime as a function of average screw shaft speed, average material pressure, and average load. As with the radial bearings, expected lifetimes are calculated for normal operating conditions; operation at low screw shaft speed, low material pressure, or low motor loading increases lifetime. In all cases, bearing overhaul is typically recommended within 64,000 operating hours. Shock pulse metering (SPM) or acceleration-spike energy systems are available to monitor extruder bearing conditions, giving early warning of possible failure. Gearbox overhauls are usually performed on site with or without supervision from the extruder manufacturer, or gearboxes can be returned to the manufacturer for overhaul. Torque-limiting clutches that are not positive-disengagement-type should be checked if clutch slipping occurs; a clutch will disengage at a different setting as pads or shoes wear (recalibration may be necessary). Maintenance of drive motors is generally performed under recommendations from the motor manufacturer.
Process Section Maintenance Maintenance of extruder screws and barrels becomes necessary when machine performance deteriorates due to abrasive or corrosive wear; routine inspection of process section components is recommended. Worn barrels and screws can be reconditioned to factory specifications depending on their condition and material of construction. New crests can be welded on worn screw flights and precision-ground to original dimensions. Barrel cooling channels may become fouled or have scale deposits that restrict the flow of heating/cooling fluids. Plugged or fouled cooling bores are often repaired by the extruder manufacturer.
SUGGESTED READING Internal Mixers Funt J.M.: Principles of Mixing and Measurement of Dispersion. Rubber World, February 1986, pp. 21 - 32. Giffin H.: Interlocking Rotor Developments. European Rubber Journal, May 1984, pp. 25- 29. Melotto M.A.: Rotor Design and Mixing Efficiency. Rubber World, February 1986, pp. 37- 39. Nevett R.F.: Mixing Practice. Rubber World, February 1986, pp. 33-36. Schmid H.M: Quality and Productivity Improvements Using Intermeshing Rotor System for Internal Mixers.
Extruders Eise K., Herrmann H., Wemer H., Burkhardt U.: An Analysis of Twin-Screw Extruder Mechanisms. Advances in Plastics Technology, Vol. 1, No. 2, April 1981, pp. 18-39. Accuracy in Extruder Scale-Up. Polymer Processing News, Vol. 5, No. 1, Wemer & Pfleiderer Corp. An Analysis of Conveyance in Twin-Screw, Co-Rotating Extruders, Polymer Processing News, Vol. 11, No. 2, Werner & Pfleiderer Corp. Twin-Screw vs. Single-Screw Extrusion. Polymer Processing News, Vol. 4, No. 2, Werner & Pfleiderer Corp.
Chapter 1 9 Scraped Surface Crystallizers, Dewaxers and Chillers* Crystallization is a very low energy separation method for many organic mixtures. Organic chemicals are most frequently separated by cooling crystallization which often means relatively low temperatures are involved, minimizing corrosion problems. Typical fluids include fatty acids and chlorobenzenes. High purity is normally possible in one or two steps.
CONTINUOUS CRYSTALLIZERS FOR ORGANIC CHEMICALS Solubility Thermodynamics The Ideal Case There are a number of frequently encountered cases where the ideal liquid mixture assumptions are applicable, and in such cases the solubility, and therefore the ease of separation can be easily calculated. Many such systems are reaction mixtures which do not lend themselves to conventional separation methods. Some frequently encountered examples are: mixed xylenes mixed chlorobenzenes paraffins many multi-substituted benzenes
Non-ideal Case There are a number of cases where these assumptions are not true, these include: 9 Polar solutes in polar solvents such as fatty acids in acetone 9 Polarizable solutes in polar solvents, such as naphthalene in methanol 9 Dimerization or hydrogen bonding, such as many organic acids
Prediction of Solubility Under those conditions where the assumptions of ideality can be considered to hold, the solubility relationship is quite simple. In the ideal case, the solubility curves and eutectics can be fairly accurately predicted by the Van T'Hoff equation:
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Process Plant Machinery
is the mole fraction in solution of component a is the molar heat of fusion of component a is the gas component is the melting temperature of component a at the system pressure is the system temperature Thus given the melting point of a substance and its molar heat of fusion, it is possible to predict its solubility in an ideal mixture and by judicious use of these results, predict the eutectic temperature and composition. Figure 19-1 shows a direct plot of the Van T'Hoff equation, relating the mole fraction of both ortho and paradichlorobenzene in solution in an ideal mixture at Xa
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Scraped Surface Crystallizers, Dewaxers and Chillers the temperatures shown. This means that for the case of an ideal mixture of para in a solvent, the composition of the saturated liquid phase is as indicated by the para curve. For instance this ideal liquid mixture would contain about 26% para at 0 degrees centigrade, regard less of solvent, so long as the system is ideal. The ortho curve has the same significance. A mixture of ortho and para isomers is a frequently encountered case, and closely approximates the ideal predictions. In this case, mole and weight percents are identical as the two components are isomers. In this binary case, the solubility of para is read from its curve, and the ortho is the balance needed to sum to 100 percent. The balance, which is the ortho percent, must be less than or equal to the percents indicated by the ortho curve, otherwise an impossible case could arise of more than 100 percent. At the point where the sum of the ortho and para curves equal 100 percent, the eutectic occurs, and further change of composition is impossible for that system. There is a unique temperature where this will occur. At this point the system shows equilibrium between the liquid phase mixture of the two isomers, and solid pure para and also pure ortho as solid. The phase rule indicates this is a determinant point, and if the temperature is further lowered, solidification will take place, producing the homogeneous eutectic. As nearly as the plot can be read, this turns out to be at - 2 4 ~ with a composition of 14% para, and 86% ortho. The International Critical Tables show the eutectic to be at - 2 3 . 4 ~ with a composition of 86.7% ortho, and 13.3% para, which is certainly excellent agreement. Plotting the figure in another way, solubility expressed as grams para per 100 grams ortho will give the results most often used to calculate the recoveries, as in Figure 19-2. Here it is abundantly clear that the way to separate a mixture of these two components is to cool the mixture, driving the para out of solution. As the cooling approaches the eutectic point nearly total recovery is attained. A simple mass balance will show that a mixture of 65% para, 35% ortho can be reduced to the eutectic point with a theoretical recovery of around 90% of the para. At that point the mixture, ortho rich mother liquor can be sold, or perhaps the eutectic split using solvents to produce pure substances. While the calculations shown are theoretical in nature, the results are almost indistinguishable from published solubility data. Such calculations can be quite useful for basic feasibility work to evaluate whether crystallization is a desirable alternative to more conventional separation methods. If this is the case, verifications of the predictions of the theory are usually advisable, and easy to carry out in simple laboratory equipment.
Low Energy Separation of Organic Chemicals Many organic mixtures may be separated by cooling crystallization, as indicated in the solubility thermodynamics part of this chapter. In simplest terms a cooling crystallization means that a mixture of organic chemicals is reduced in temperature, without removal (by evaporation) of any of the components. In this simple cooling operation, the heat removal consists of sensible heat, which is generally low for organics, and also the latent heat of fusion, which is also quite low. Crystallization is a "one way" process; the heat is removed, crystals are formed, and the mixture of crystals and solids formed are then separated. This must be compared to distillation which is inherently a refluxing operation, where products are repeatedly evaporated and re-condensed.
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FIGURE 19-2 Theoretical Solubility of para in ortho dichlorobenzene 3. (Source: Armstrong Engineering Associates, Inc West Chester, PA) The latent heat of fusion is generally much lower than the latent heat of vaporization. Since the latent heat of fusion must be removed only once, instead of many times as in distillation, the energy requirements are drastically lower for crystallization. Most distillations take place at elevated temperatures, which means the materials being processed must be heated up, and then cooled back down again normally with energy losses each way. Also many distillations are run under vacuum to effect a better separation, however, most vacuum systems are also energy consumers on a large scale. Many crystallizations take place at around ambient temperatures so there is little heat up or cool down required to get to the right conditions for the separation to start.
Scraped Surface Crystallizers, Dewaxers and Chillers 683
FIGURE 19-3 Large carbon steel unit for crystallizing petroleum wax. Refrigeration is by boiling refrigerant. (Source: Armstrong Engineering Associates, Inc West Chester, PA) In the great majority of cases, the crystals which form are 100% pure material, as opposed to something only slightly richer than the feed material, thus it is not necessary to repeatedly melt and refreeze to obtain quite high purities. Certainly the pure crystals will have some impure mother liquor on the surfaces and sometimes contained within the crystals as occlusions. However, the purity increase is extremely rapid and normally one or perhaps two crystallizations can give very high purities. Many of the materials which separate most easily on crystallization are in any event very difficult to separate by distillation. Examples include ortho and para disubstituted benzenes, fatty acids, and many others of commercial significance. In these cases extremely large distillation columns are required, with exceedingly high reflux rates, which are very energy intensive. In many cases it is possible to arrange for energy recovery in crystallization systems by using the cold separated product to carry out part of the cooling required to obtain the crystals, thus reducing the energy required even more. With rising energy costs, it makes sense to take a critical look at many separations to see if crystallization may be the best altemative.
Advantages of Crystallization Utility requirements are low, as discussed in the previous section. Low temperature operation, which means low corrosion rates, and often use of less costly alloys compared to evaporation based separations. The low temperatures
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Process Plant Machinery
FIGURE 19-4 Scraped surface chiller with two cooling circuits of boiling refrigerants. One of several in the installation, surface 168Oft2 (156 m 2) (Source: Armstrong Engineering Associates, Inc West Chester, PA) utilized also mean little or no product degradation which for heat sensitive materials may be extremely important. There is no formation of "tars" which represent a yield loss, and increasingly a severe waste disposal problem, and normally require additional separation equipment (and energy) for their removal, to give desired product color, etc. Enclosed systems, with little chance of leakage of dangerous or noxious fluids. The systems are normally simple and require few pieces of equipment and relatively little instrumentation. Favorable equilibrium; often freezing points are widely separated, allowing for relatively easy separation by crystallization, where distillations may be exceedingly difficult. High purity; the crystals which form in the great majority of cases are 100% pure material. While impurities may adhere to crystal surfaces, or be included inside the crystal, recrystallization can normally achieve very high purities, with relative ease. 95 to 99.5% are normal product purities, although higher figures are also often reached. Crystallization can easily be made continuous; in general, the only reason to work with a batch crystallization is very low design capacity. If a high enough design capacity is desired (normally above 500,000 pounds of product annually), it is easy to make a fully continuous crystallization, with consequent savings in energy and manpower.
Scraped Surface Crystallizers, Dewaxers and Chillers
Crystallizer for Paraxylene Continuous scraped surface crystallizers for paraxylene purification may be designed to use either boiling refrigerant, or brine, depending on temperature level, and the flow rates involved. Brine cooling has a number of advantages in the higher purity, high temperature ranges, while boiling refrigerants offer advantages at lower temperatures. There is no practical upper limit on capacity, as some of the largest plants in the world are currently using this equipment. Rugged construction and field proven design assure long run times even in very tough duties. Process design consultation is normally involved with such facilities to assure troublefree operations. Design and construction is available in most international pressure vessel codes, such as ASME, Stoomwezen, TUV, CODAP, ISPESL, Swedish and Australian Pressure Vessel Code, etc. Crystallization is a very powerful separation method for obtaining high product purities, such as those demanded by current paraxylene customers. The crystals which form are inherently pure (i.e. 100% paraxylene). The only impurities are mother liquors sticking to the surface, or included in the crystal matrix, where washing can remove them. Moderate temperatures are required in high-purity paraxylene crystallizations, so special materials, and high refrigeration loads are not involved. Skimming is a cost-effective method to recover paraxylene from various recycle streams which are relatively rich in this isomer. Current practice often mixes such streams with other, much leaner streams, losing the relative advantages of the purity already achieved. Crystallization facilities may be expanded on an incremental basis which allows for easy de-bottlenecking should this become attractive in the future.
Working Principles of a Surface Continuous Crystallizer A scraped surface crystallizer is built as a double pipe element, generally with 6, 8, 10, or 12 inch inner pipe, and an outer pipe suited to the duty. The area between the inner and outer pipe is filled with cooling fluid. The inner pipe contains a rotating scraper element, shown schematically in Figure 19-5 which mixes the process fluid flowing through the inner pipe, and removes deposits which may form on the wall as cooling occurs. In effect, the scraped surface crystallizer is a heat exchanger but quite an unusual one. It generally performs as a cooling crystallizer, since the solubility thermodynamics portion of this folder indicates the high potential for cooling crystallization for organic chemicals. Heat transfer occurs across the wall in the inner pipe, with cold fluid outside, and process fluid inside. As this cooling occurs, crystals tend to form, and foul, on the inner pipe wall. The scraper blades there rotate and remove crystalline deposits which would inhibit heat transfer. In general, the majority of the crystallization takes place in the bulk of the fluid, as opposed to the wall, thus allowing growth of easily separable crystals. Scraper blades rotate at moderate speeds, generally 15 to 30 rpm. Higher speeds will give higher heat transfer, but the high shear which accompanies it will normally produce much smaller crystals. An extension of this is the margarine or ice cream
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Process Plant Machinery
//
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FIGURE 19-5 Cross-section of scraped surface crystallizer. (Source: Armstrong Engineering Associates, Inc West Chester, PA) freezer, which are designed specifically to crystallize such materials, but produce tiny, inseparable crystals. This is accomplished by use of extremely high rpm, high shear, and high heat transfer. The ice cream machine is often confused with a crystallizer, since the principle is similar, but the result very different. Crystallizers are designed specifically for adequate heat transfer, consistent with good crystal growth. A normal commercial sized installation will consist of several double pipe elements, each with a length of 20 to 40 feet, connected in series. This gives a long thin flow path which promotes a close approach to plug flow, which is very important in many crystallizations. Cooling is carried out by whatever coolant is selected, and run through the annulus between inner and outer pipe. Coolants may be cooling tower water, chilled brines, evaporating refrigerants, mother liquor stripped of crystals, or combinations of these. Fabrication of the equipment is available in alloys normally found in chemical plants, such as carbon and stainless steels, nickel alloys such as inconel, incoloy. monel, etc. Construction is available to international pressure vessel codes such as ASME, BS5500, TI~IV, Stoomwezen, Swedish Pressure Vessel Code, Australian DLI, etc. Many organic chemical crystallizations are done using batch cooling in stainless steel or glass-lined kettles. By and large this represents continued growth from specialty chemical to commodity, without a great deal of engineering attention to the crystallization part of the process. Continuous crystallization offers a number of advantages over batch crystallization: (a) Smaller equipment which generally means less expensive installations and small volumes of sometimes hazardous or expensive material in process. (b) Less floor space. (c) Less operator labor.
Scraped Surface Crystall&ers, Dewaxers and Chillers 687 (d) No duplication of instrumentation, piping, etc. (e) Less upsets, or peak utility demand. (f) Better process control.
Advantages of Scraped Surface Crystallizers Many continuous crystallizations are done in evaporative or cooling crystallizers based on designs typically used for inorganic chemicals where there is normally a very fiat solubility curve, i.e., a change in mixture temperature produces relatively few crystals. Other are sometimes done by cooling and partially crystallizing in shell and tube exchangers. Scraped surface continuous crystallizers offer advantages as follows: (a) May be run for extended periods between hot washing where many shell and tube exchangers would plug up in minutes. (b) May be run at much higher temperature differences between process fluid and coolant than may ever be attempted with shell and tube equipment, without serious fouling, or plugging. (c) May be used over an extremely wide temperature range, from - 7 5 ~ to 4-100 ~ It is normally very difficult to run vacuum crystallization equipment over a broad range of temperatures. (d) Can be used with a high percentage of solids. Vacuum crystallizers are normally limited to about 25% by weight or less solids, while we have worked into the range of 65% by weight solids as slurry. (e) High viscosities are not a major problem, with several crystallizations being carried out from mother liquors with viscosities of 10,000 cp or higher. (f) Flow pattern in once-through operation closely resembles plug flow so conversion of batch to continuous processes is easy, and virtually any time/temperature pattern desired is possible. (g) In small throughput cases, a scraped surface crystallizer will be very inexpensive, even where in much larger installations a vacuum crystallization may be most attractive. (h) Modular construction allows for very easy expansion with growth in demand. (i) Simple, self-contained construction, with minimum instrumentation, and auxilliaries such as condensers, vacuum systems, etc.
Applications For scraped surface cooling crystallizations
Good Solubility Curve Cooling crystallizations are obviously most suitable where the solubility curve will give good yields with simple cooling of the mixture. This is true of a wide variety of organic mixtures.
Low Temperature Crystallizations Scraped surface crystallizers offer the best approach to low temperature crystallization as, for example, the separation of meta and para xylenes, or oleic and linoleic acids.
Products with High Boiling Point Rise Some mixtures of inorganic chemicals in water show very high boiling point rises as concentration proceeds, reducing the vapor pressure, and much increasing the
688
ProcessPlant Machinery vacuum requirement. Many such mixes produce abundant crystal growth on cooling. Often a scraped surface unit may be used in conjunction with a vacuum unit, with a vacuum unit doing the high temperature part of the crystallization and a scraped surface unit doing the low temperature part.
Products with Similar Vapor Pressures Many aromatic chemicals, particularly isomers, have nearly identical vapor pressure characteristics, making distillation very difficult. These same mixes often have widely varying freezing points, making crystallization simple, and effective.
High Viscosity Fluids High viscosity, due either to high mother liquor viscosity or a high percentage of solids do not present problems to the scraped surface crystallizer but may make other types of crystallizers totally inoperable.
Severe Fouling The fouling tendencies of many slurries are opposed, and deposits on the heat transfer surfaces continuously removed.
Partial of List Products Crystallized anthracene anthraquinone
naphthalene nitrochlorobenzene
benzene hexachloride benzoic acid bisphenol A butyl cresol butyric acid
oligomers
caffeine calcium nitrate caprolactam cyanoacetamide dibutyl cresol diglycerides dimethyl hydantoin dimethyl terephthalate fatty acids lactose laurolactam levulinic acid menthol methionine monoglycerides
palm/palm kernel fats paracresol paradichlorobenzene paraxylene pentaerythritol potassium chloride potassium nitrate sebacic acid silver nitrate sodium carbonate sodium lauryl sulfate sodium sulfate sorbic acid sterols tall oil fatty acids tallow fatty acids tetrachlorobenzene tetramethyl benzene vitamins waxes
Separation of Chiorobenzenes Para and ortho di chlorobenzene, which were used in the example on solubility thermodynamics, represent two important chemical products which lend themselves
Scraped Surface Crystallizers, Dewaxers and Chillers
689
to separation by cooling crystallization. The para isomer crystallizes at temperatures far above the point where either ortho crystals, or the eutectic is reached. Para dichlorobenzene forms extremely tough crystals, which adhere readily to any cooled surface, requiring vigorous scraping to remove them. The tough crystals can stand a certain amount of abuse without degradation in size. The mixture normally worked with produces a very thick slurry, and great care must be exercised handling it. The extremely steep solubility curve presents many opportunities for crystal growth, as well as dangers of uncontrolled crystallization, and must be handled carefully or the whole unit may freeze solid. Strong equipment, and ingenious slurry handling, often with staged operations, are the basics of this process, and similar separations of xylene isomers, cresols, and other di-substituted benzenes.
Separation of Fatty Materials Fatty acids from tallow or tall oil, mono-, di-, triglycerides, fatty alcohols, and related compounds all may be separated by crystallization where other separation means will not work. These separations all must take into account the extremely delicate nature of the crystal, and the sensitivity to shear which can rapidly produce an inseparable crystal.
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FIGURE 19-6 Secondary recovery by a scraped surface crystallizer. (Source: Armstrong Engineering Associates, Inc West Chester, PA)
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ProcessPlant Machinery The time/temperature relationship is of extreme importance as well, sometimes requiring sophisticated cooling arrangements on the shell sides of the equipment. Solvents are sometimes used to obtain optimal separations, although solvent free separations using detergents to separate saturated and unsaturated compounds have also frequently been used. Typically, crystal growth is relatively slow and care must be exercised to allow time to grow a decent crystal, which may be easily separated. Reducing shear is more important than producing a rugged machine for these delicate materials.
Secondary Recovery An opportunity exists to recover additional product from waste streams from vacuum or other crystallizers due to the better cooling capability of scraped surface crystallizers. In many cases, either due to boiling point rise or temperature characteristics, vacuum crystallizers can only cool so far, for example to say 35 to 40 ~ as the steam consumption begins to get extremely high for the amount of vapor removed. With high steam consumption goes high cooling water requirements as well, plus large capital equipment needs. Scraped surface crystallizers can be refrigerated easily and can cool as low as minus 100 ~ without great difficulty and therefore can easily cool the effluent from the vacuum crystallizer to effect additional recovery of product beyond that which can economically be obtained in the vacuum crystallizer. It is possible the recovered product can quickly pay for the additional equipment, also it is an easy way to increase output without otherwise increasing capital equipment. The solubility chart (Figure 19-7) illustrates a typical solution. The normal feed to the vacuum crystallizer enters at (1) and is cooled to (2). The temperature at that point is the economic limit, either due to being a low temperature, or being the result of high boiling point rise. Therefore it is either impossible or highly uneconomical to carry the process further with vacuum equipment. There are many such situations in existence today. The scraped surface crystallizer can cool from point (2) down to point (3), effecting appreciable further recovery of crystals because it can easily reduce the temperature, and consequently lower the solubility.
Other Applications of Scraped Surface Exchangers While the vast majority of scraped surface exchangers are in crystallization, the equipment is sufficiently versatile that a wide range of process problems have been successfully solved by use of scraped surface exchangers.
Crystallization of Inorganics Many inorganic chemicals are crystallized from aqueous solutions on a massive scale. Typically, however, this is done by evaporative crystallization as the solubility curve is extremely fiat, so cooling crystallization will not produce many crystals. There are, however, some important exceptions. Sodium sulphate, sodium carbonate, potassium nitrate are all often encountered in quantities where cooling crystallization has merit, and have been successfully handled in scraped surface crystallizers.
Cooling Viscous Liquids (to 100,000 cp) There are many instances where heat exchange is difficult because of the viscosity of a process stream. In such cases, Armstrong scraped surface heat exchangers have shown their merit, handling materials with up to 100,000 centipoises viscosity with
Scraped Surface Crystallizers, Dewaxers and Chillers 691 INLET PRIMARY FEED
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SECONDARY RECOVERY BY SCRAPED SURFACE CRYSTALLIZER
LIMIT OF ECONOMIC OPERATION OF STEAM BOOSTER ON VACUUM CRYSTALLIZER
LOWER TEMPERATURE EASILY OBTAINED IN SCRAPED SURFACE CRYSTALLIZER TEMPERATURE
FIGURE 19-7
Solubilitydiagram of Secondary Recovery. (Source: Armstrong Engineering Associates, Inc West
Chester, PA) high heat transfer. Added benefits of scraped surface exchangers in such duties are the inherently low pressure drops of this type of equipment, and the ability to remove any sticky foulant which might from such viscous fluids.
Scraped Reactor Coolers Scraped surface exchangers have demonstrated their ability to handle high viscosity, fouling fluids, producing good heat transfer, and near plug flow. These unique characteristics make it a very effective reactor for a wide variety of exothermic reactions. Ease of temperature control adds further important flexibility.
Sublimation Condensing Condensation of a vapor as a solid, or desubliming, is another area where scraped surface exchangers have proven their value. In such cases it is necessary to continuously remove a fouling film and allow the solid condensate to be carried away from the unit, while permitting the process stream to pass through. There are a
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Process Plant Machinery
wide number of organic chemical reactions where an off gas contains a valuable component which can be recovered by the use of a scraped surface condenser, but where a normal condenser would rapidly foul with solid condensate. Such condensers also find use in purification operations where a flash distillation separates components and purifies the condensate, which condenses as a solid.
SCRAPED SURFACEDEWAXING EXCHANGERSAND CHILLERS In many cases MEK dewaxing units are in basically good condition with the exception of the scraped surface dewaxing exchangers and chillers. In today's world, it is often difficult to justify putting up a new grass roots lube facility, however, rebuilding existing plants offers a number of advantages. Far lower capital costs are involved, timing is very short, and there may be no environmental approvals required. At the same time, substantial upgrades of equipment are possible, improving process and mechanical performance. As in many other fields, the scraped surface dewaxing equipment designed today is considerably better than that made 30 or 50 years ago. Fortunately, retrofits are easily possible for such important parts as the scraper internals, shear pin sprockets, drive ends, and in some cases, the double pipes themselves. This allows the reuse of some of the existing equipment, but upgrading of the wear parts to current practice. The scraper internal is the heart of the scraped surface exchanger or chiller. It must run for a number of years without maintenance, or even the possibility of
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FIGURE 19-8 Replacement drive-all bolted, no hot work required. (Source: Armstrong Engineering Associates, Inc West Chester, PA)
Scraped Surface Crystallizers, Dewaxers and Chillers examination. As moving parts, there will be wear on the scraper internals. Although installed in a lubricating oil plant, the solvents cut any lubricating qualities of the charge stock so the internals work in an environment which is likely to cause rapid wear. To counter this, only the best materials of construction, fabrication tolerances, and design approaches can be depended on to give maximum life. As wear parts, the scraper internals must be replaced from time to time. This gives the opportunity to retrofit state of the art intemals on a normal maintenance basis. Next to the scraper internals, the drive mechanism is typically the part of the scraped surface exchanger which causes most maintenance headaches. Armstrong have for many years been involved in the business of replacing the originally furnished drive mechanism with drive trains as illustrated above. In such a replacement, a return bend is frequently removed and replaced. With a fabricated crossover which would bolt to the inner pipe. The drive end itself is in turn bolted to this. The drive mechanism is designed to be extremely compact, but rugged. In operation it has proven to be very reliable, and easy for mechanics to understand. While an overhung drive shaft is used, the thickness of the shaft and short dimensions have produced a reliable drive which has caused virtually no problems in operation. In changing over to this type of drive, it is normally necessary to send an engineer to the site to measure exactly what is required, as often old drawings are lacking in sufficient detail to make such a changeover. While this is expensive, it does save problems of things not fitting properly which can cause expensive last minute modifications which have led to mistakes being made in a number of cases, where local machine shops have been asked to do similar work. It is necessary to understand the function of the drive to design alterations.
Mechanical Seals Environmental constraints have increasingly indicated that less leakage from stuffing boxes is highly desirable. Some installations have opted for the use of mechanical seals to minimize leakage at the entry point for the drive shaft. This is not inexpensive, but it has cut leakage drastically which has paid off in reduced solvent losses, solvent attack of the drive, and of course reduced air pollution. Such drives are available as replacements to existing stuffing box drive ends. Some instances of overhaul of old equipment have involved rebuilding the double pipe elements. There have been many variations on just how this has been carried out. In some cases the "carcass" of the double pipes alone has been fabricated and shipped out, to be fit on location with drive and intemals, etc. In other cases the whole exchanger has been removed pulled apart and the inner pipe, drive and internals replaced or variations made as needed. In a number of cases it has been decided to merely scrap the existing equipment and purchase totally new exchangers and chillers, but with the provision that they fit on the existing foundations, and meet existing piping. Most scraped surface exchangers or chillers are fit with a shear pin to protect the scraper internals in the event of something becoming lodged in them. Traditionally these have been made with a sleeve bearing between the halves of the sprocket. When the shear pin breaks, the inner part remains stationary while the outer part rotates with the driving chain. This causes some wear on the sleeve bearing. The wear allows a slight wobble which can cause premature failure of the shear pin, in turn causing more sleeve bearing wear, etc. This problem has been eliminated by the Armstrong shear pin sprocket. An antifriction needle bearing has been fitted between the halves of the shear pin sprocket. Upon failure of the shear pin the outer part rotates but there is no bearing wear. This is true for extended running periods - once nine months before the shear
693
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ProcessPlant Machinery
SHEAR PIN SPROCKET
GREASE SEA
SHEAR PIN ,HARDENED SHEAR PIN BUSHINGS SHEAR PIN HUB GREASE SEAL NEEDLE BEARING
SNAP RING GREASE FITTING
FIGURE 19-9 Chester, PA)
Schematic drawing of shear pin sprocket. (Source: Armstrong Engineering Associates, Inc West
pin could be reinstalled (after removing a rock from inside the scraped surface unit!). After such a long time, it was remarkable that simply replacing the shear pin was sufficient to reactivate the sprocket and have it work properly. Shear pin sprockets also wear out due to wear between the teeth and the drive chain. Armstrong shear pin sprockets thus may be used as replacement items when the older type wear out. These sprockets have been used frequently as replacement parts on scraped surface exchangers and chillers originally made by other firms.
Chapter 20 Conveyor-Based Processing Systems* With numerous process plants employing conveyors of one type or another, it was felt that this text should give at least an introduction to this type of machinery by focusing on one of the more sophisticated executions: steel-belt conveyors. The use of steel-belt conveyors has spread throughout the processing industries. Applications of steel-belt conveyors include cooling/solidification, drying, pressing, freezing, baking, and materials conveying. The steel belt is made from fiat strip steel from a rolling mill, prepared through special techniques that straighten, flatten, and make the ordinary strip suitable for welding into endless bands continuously running around two terminals. The conveyors based on this specialized technology are designed for the processing industries according to the needs of the product and the special needs of the steel belt. Table 20.1 summarizes a wide variety of steel-belt applications and the important steel-belt properties that make the applications successful. The general categories that are shown in Table 20.1 are material handling, food processing, industrial processing, and presses for particle boards, plastics, and rubber. Table 20.1 also indicates the four major steel-belt grades that are in common use. The following discussions describe the applications and process for which steelbelt conveyors have been selected as the best of competing alternatives, including the types of materials used for conveyor belts. Throughout this chapter, the reader will find examples of industries and products where steel belts are being used.
BELT GRADES There are numerous grades of steel that can be rolled and processed to be suitable for use as a conveyor belt. The selection of belt material is governed by the product to be conveyed and the process conditions. For the large majority of applications, four grades of steel conveyor belts are used:
Austenitic Stainless Steels American Iron and Steel Institute (AISI) type 301 is available in widths up to 61 inches. For conveyor belt applications, it is cold rolled to a hardness in the range of Vickers Hardness (HV) 380 resulting in a high tensile strength steel in * Source: Sandvik Process System, Inc., Totowa, NJ. Adapted with permission. 695
696
Process Plant Machinery
TABLE 20.1
Areas of Steel Belt Application Possible Steel Belt
Important Steel Belt Properties
I
Steel Belt Installation Type
I
~
~
~
""+
r~
r~
~
Conveyors, General 9 Material Sorting systems 9 Handling Work tables, general Meat cutting tables Food Industry Bottle handling 9 Bake ovens (9 Contact freezers Belt coolers/ food Belt coolers/ chemicals Industrial Process Steel belt dryers Applications Flow-through belt unit Double belt presses 9 Belt skimmers Single Opening Particle presses Board Industry Multi opening presses Rotation presses 9 Rubber Rotation presses 9 Plastics Rotation presses 9
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X
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X
X
X
X
X X
X
(o)
X
X
(o)
X
X
(e) = property sometimes of importance 9= property always of importance Source: Sandvik Process System, Inc., Totowa, NJ.
the range of 160,000 psi. This steel has good corrosion resistance and remains stable at temperatures ranging from - 2 0 0 ~ to + 4 0 0 ~ Due to physical properties (relatively low thermal conductivity and high thermal expansion of austenic stainless steels), the temperature must be kept uniform over the width to avoid thermal distortion, which could affect the product or process. This material has good welding properties; however, the weld must be work hardened to restore hardness and strength in the weld area.
Conveyor-Based Processing Systems 697 AISI type 316 has a higher nickel content and is alloyed with molybdenum to obtain better corrosion resistance. This material is also cold rolled to increase strength, with maximum hardness in the range of HV 320. Tensile strength is about 130,000 psi at 167~ As type 316 is more difficult to produce and has lower mechanical properties, it generally is used only when 301 is unsuitable.
Hardened and Tempered Carbon Steel This grade approximates AISI 1065 and is available in widths up to 48 inches. After cold rolling it is hardened and tempered through heat treatment to obtain a hardness of about HV 400 and tensile strength up to 180,000 psi at 120 ~ The hard springy nature of this steel makes it suitable for almost any application where the risk of corrosion is low. It has excellent heat transfer properties and a low coefficient of expansion, making it ideal for heating or baking ovens and many material handling applications where a hard, smooth, wear-resistant belt surface is desirable. Due to the carbon and manganese content and heat treatment of this steel, the welding procedure is rather complicated; however, with proper temperature control and heat treatment of the weld, very good results can be obtained.
Low Carbon Martensitic Stainless This is a grade characterized by high strength with hardness of HV 350 and tensile strength of about 155,000 psi, with good corrosion resistance and excellent welding properties. This material is furnished in widths up to 61 inches in heat-treated condition. This material has a thermal conductivity slightly better than austenitic grades and a thermal expansion closer to carbon steel, making it more stable against thermal distortion than the austenitic grades. Strength of the weld is very close to the parent strength.
Precipitation-Hardened Stainless Steel This material is characterized by extra-high tensile strength, high fatigue strength, good corrosion resistance, excellent weld strength, and good repairability. It is available in widths of up to 61 inches and has a hardness of HV 500 with tensile strength up to 240,000 psi. This grade is not suitable for use at low temperatures. As stated earlier, there are numerous factors to consider in selecting the best belt grade for any given process or product. Table 20.2 gives some of the physical properties of steel-belt materials. All of the grades mentioned can be longitudinally welded to obtain wider belts. Depending on the material, the longitudinal welds are made by the same procedure as transverse welds, with work hardening or heat treatment requirements identical. With existing equipment, it is possible to make up to three longitudinal welds with a maximum width of 180 inches. The length of any belt segment is limited to approximately 300 feet; however, these segments may be transverse welded to produce almost any length. Other specialized treatments of steel belts include perforation, which is usually done to allow air to flow through the belt, but is also used in belt washers and for baking. Polished or matter finishes may be supplied for special product surface effects. A variety of edging or vee-rope guides may be attached to belts for guiding or holding a liquid on the belts.
698
Process Plant Machinery TABLE 20.2
Selected Physical Properties of Steel Belt Materials
i
Yield Strength Tensile 0.2% offset Strength Vickers (ksi) Vickers (ksi)
Hardness (HV)
Thermal Thermal Conductivity Expansion Parent Weld Parent Weld Parent Weld (BTU/hr-ft OF) (10-6/~ Austenitic stainless steel AISI 301 AISI 316 Carbon steel Low carbon martensitic Precipitation hardened stainless steel
142 132 174
91 73 128
160 141 186
128 102 144
380 320 400
300 250 350
9.5 9.3 22
9.8 8.9 6.2
145
138
157
152
350
350
11.8
6.0
236
217
239
232
500
500
9.2
6.1
,,
Notes: Yield strength and tensile strength at room temperature; thermal conductivity at 212 ~ thermal expansion over range 68- 212 ~ Source: Sandvik Process System, Inc., Totowa, NJ.
COOLER/FLAKERS The most prevalent application of processing on steel belts is for cooling and solidifying molten chemicals. Some of the products presently being processed on steel-belt coolers are listed in Table 20.3. Steel-belt coolers are also known as flakers. In the case of sulphur solidification, the steel-belt cooler is called a slater, for the 1/4- to l/2-inch thick fiat sulphur product known as slates. The competitive device for solidifying chemicals is the drum flaker, a circular, internally cooled drum that has become relatively expensive compared with steel-belt flakers and therefore is not presently used very often for new installations. Manufacturers of standard steel belt coolers have designs for a variety of belt widths (i.e., 500, 800, 1200, and 1500 mm). Zones or modules in the United States are eight, twelve, or sixteen feet long in the direction of the conveyor. Any number of these zones (within reason) may be combined to give the desired process and the desired capacity.
Cooling a Product on a Steel Belt In the majority of steel-belt cooler applications, a coolant liquid is sprayed against the underside of the top strand of the conveyor. Temperature of the coolant may be TABLE 20.3 Products Cooled or Flaked on Steel Belt Conveyors Anti-oxidants Hot melt adhesives Resins of all types Aluminum sulfate Sulfur Agar-agar Maleic anhydride Animal fat
Phthalic anhydride Asphalt and pitch Chocolate Pesticides Surfactants Atactic polypropylene Stearic acid Waxes
Source: Sandvik Process System, Inc., Totowa, NJ.
Conveyor-Based Processing Systems
specially selected depending on the nature of the product being carded on the belt, or in some cases, it may be simply whatever is available. Selection of the optimum coolant program is discussed below. A typical cross section showing coolant impinging on the underside of the belt is pictured in Figure 20-1. Volumetric flow rafts are varied from 0.5 to 1.5 gpm per square foot of cooled belt, depending on temperature level, heat transfer from the hot product, and limitations on heating the coolant. The coolant film is scraped off as it passes over periodic cross bars, being immediately restored by the next set of spray nozzles. Film coefficients of a variety of spray nozzle types were measured for this configuration as approximately 40 BTU/hr-ft 2 ~ for design flow rates. For thin, thermally conductive products, the coolant film coefficient is relatively important for optimum flaker performance. Good spray nozzle operation must be maintained. On the other hand, thick insulating products cannot be cooled more rapidly by improvements in coolant film coefficient. The temperature of the product as carried down the conveyor is as shown in Figure 20-2. As a rule of thumb for typical belt cooler/flakers, one-half the heat load from the product is absorbed in the first third of the cooling length. One of the advantages of processing on a conveyor is the ability to put a number of cooling zones or tanks in series. Each zone can be treated differently than the others, if desired. If a product is placed on the belt in a thick layer and does not immediately "set up," retaining side strips on the belt may be required. Some products, placed .
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,
,
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,~
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FIGURE 20-1 Cross section showing conveyor cooling water being sprayed on bottom of belt. (Source." Sandvik Process System, Inc., Totowa, NJ.)
T E M P
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FIGURE 20-2 Product temperature versus time for a cooling conveyor. (Source: Sandvik Process System, Inc., Totowa, NJ.)
699
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ProcessPlant Machinery
molten on the cooling belt, do not remain as fiat solids but tend to "curl" up at the edges or take a characteristic gull-wing shape. This tendency is probably due to nonuniform changes in the coefficient of expansion as the product is solidified and cooled. When one of these shapes occurs, the product pulls away from the belt, conductive heat transfer from produ~'t to belt ceases, and cooling of the product takes much longer. In many cases, curling of the product can be minimized by tempering the coolant water, i.e., by using high-temperature coolant in the initial zone and gradually reducing the coolant temperature in successive zones. Often, three different temperature levels are required. It is postulated that in the cases where tempering is effective, the product cools uniformly throughout its thickness, thus minimizing any differential contraction/expansion effects. The opposite condition from a curling product is when the product sticks too rigidly to the belt. Some products tend to form a good adhesive bond to the belt when the belt is clean, dry, or heated. To avoid sticking, (1) the belt could be cooled before the product is placed on it, (2) a release agent such as silicone could be used, or (3) moisture on the belt could also act as a release agent. The challenge to the conveyor process engineer is to find the set of conditions whereby the product will initially adhere to the belt, thus promoting good heat transfer, but on reaching the discharge terminal will separate cleanly from the belt. A surprisingly large fraction of the products run on steel belts have these characteristics. Without mechanical cooling from above, the product, cooling to the ambient air, still represents 10 to 15 percent of the total heat load. With properly designed fans and nozzles directed against the product, 50 percent of the heat load could be transferred above the belt. For thin, brittle products, high-velocity air may tend to blow the product off the belt and so must be used with appropriate caution.
Supercooling Materials Supercooling materials do not solidify as do "well-behaved" materials, but instead cool below their melting points without a transition to a hard, solid state. These supercooling materials may remain as cold, mushy pastes unsuitable for packaging or further processing. Typical examples of supercoolers are antioxidants for the rubber industry and aluminum sulphate. When supercooling occurs, the techniques used for obtaining solidification include careful temperature control, addition of solid, granulated seed material, and energy addition through use of a scraped-surface heat exchanger before the product is deposited on the conveyor/cooler. After being formed into a sheet on the conveyor, cooling must be carefully controlled for solidification to a hard flake to take place. An additional curing conveyor is sometimes used to provide additional time for product hardening. Figure 20-3 shows the complete process requirements for aluminum sulphate.
Steel-Belt Accessories Various feeding devices are used to form continuous sheets, strips, or droplets. The sheets or slabs of product are typically one to three millimeters thick, although there are certainly examples outside this range. Sheets are formed by (1) extruding through a die, (2) liquid flowing over a weir, and (3) a variety of special devices suited to the particular properties of the product. Hot-melt adhesives, in particular, are formed in strips, cooled, and then cut into slats or chicklets by a multibladed cutter at the discharge end of the steel-belt
Conveyor-Based Processing Systems
i
FIGURE 20-3 Aluminum sulphate process with steel-belt solidification. 1 - S e e d i n g material; 2 - vibrator; 3 - agitator; 4 - boiler; 5 - f a n ; 6 - s t e e l belt cooler; 7 - c o a r s e crusher; 8 - accumulation and recooling belt; 9 -fans; 1 0 - belt conveyor; 11 - mill; 12-elevator; 13-vibratory screen; 14-silos. (Source: Sandvik Process System, Inc., Totowa, NJ.)
cooler. Strip forming of brittle materials such as caprolactam is used to produce a more uniform particle size than if conventional flaking is used. Maximum uniformity of solidified product is obtained with drop-forming or pastillating devices. In the past, these devices were reciprocating, required relatively slow belt speeds, and required a lot of maintenance for optimum product quality. Now, a rotary system is available that overcomes the previous objections. Producers are interested in forming drops that solidify to become an almost completely dustfree product in order to provide a marketing advantage over their competitors and because there is virtually no loss in production capacity for a given area of cooling belt. Some of the products presently being formed in pastille shape are shown in Table 20.4. A photograph of a rotating drop-forming device on a steel belt is shown in Figure 20-4. For producers not having a need for uniform, dust-free products, the steel-belt conveyor manufacturers provide rotating-bar breakers and crushers that reduce flake size for easy handling and bagging.
TABLE 20.4 Products Being Drop-Formed on Steel Belts Hydrocarbon resin Bitumen mix Antioxidants Pitch Cobalt naphthenate Cobalt stearate Terpene resins Rosin resins Vegetable fat
Wax Acrylic resin Nickel catalyst Petroleum resin Hot melt adhesive Sulphur Epoxy resin Food sauce Polyester resin
Source: Sandvik Process System, Inc., Totowa, NJ.
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Process Plant Machinery
FIGURE 20-4 Rotoformer formation of pastilles on a steel belt. (Source: Sandvik Process System, Inc., Totowa, NJ.)
Double-Belt Cooler In some specialized cases, an effective method of cooling is to place a second steel belt above the product, as shown in Figure 20-5. The product is loaded onto the upper surface of the unit's lower belt (1), which carries it into the cooling zone, where the pressure of the upper belt (2) ensures contact with both cooling surfaces. The upper belt is sprayed with cooling water (3) from above, and the lower belt is sprayed from below. Loading (4) and discharge methods (5) are simply arranged to suit individual products and can incorporate breaker equipment (6) at the discharge station if required. The double-belt cooler is used for 9 curling products that cannot be controlled otherwise 9 minimizing oxidation of products while hot 4
2
/3 ~
5
l
/
9 1 L/.r
FIGURE 20-5
I
i
Schematic of a double-belt cooler. (Source: Sandvik Process System, Inc., Totowa, NJ.)
Conveyor-Based Processing Systems 703 9 increased capacity where space is a limitation 9 more uniform discharge temperatures across width to eliminate downstream grinding problems
HEATING/DRYING APPLICATIONS There are several prominent applications of steel-belt conveyors in which the product is heated, dried, or baked while on the belt. In each case, the unique configuration of the wide, fiat, moving surface contributes substantially to the process. The types of applications for which steel belts provide the best solutions to the needs of manufacturers are 9 film or sheet forming 9 hot air drying or baking where sanitation or sequenced zones are important 9 high temperature heating on a smooth belt Examples of these applications will be discussed. Film formation or casting a liquid on a steel belt uses the belt as part of the forming process and then uses the belt as the carrier during the drying or evaporation stage. Often the belt is polished. Starting with the standard 2B mill finish, belt polishing can be carried out to any degree up to a mirror finish. Product film properties will reflect the belt surface. Photographic film, filter membranes, polyvinyl alcohol film, and modified tobacco sheet are common examples of steelbelt processed products. Dried tobacco sheet prepared to an approximate thickness of 0.022 inches and a moisture level of 11% is shown in Figure 20-6 being doctored off the discharge drum of a steel-belt conveyor moving at approximately 150 feet per minute. This sheet is formed from a high-moisture-containing slurry and requires careful atmospheric control of temperature and humidity to maintain the integrity of the drying sheet. For another group of applications, direct heating of the belt by gas ribbon burners and/or indirect convective heating is used for baking cookies, drying cereals, and drying fruit such as apples. For the cookies, the carbon steel belt provides a surface on which a forming roll drops individual pieces. By passing through the continuous baking oven, the belt provides the proper thermal properties for this particular baking process. In drying some products such as apples, initial drying of the wet product is performed with a relatively thin bed and high temperature. When the surface water is evaporated, the temperature in the next zone in the dryer is reduced and the bed depth is increased. Zones may be used in several ways in belt dryers to provide the optimum conditions required by the product. With a single belt, each module along the length of the belt may have a different temperature and humidity. With multiple belt systems, even more flexibility is possible. Cereal preparation on a steel belt provides a sanitary surface that permits cutting and stamping operations. Sugar application is practical because the belt can be continuously washed if desired. Many rubber profiles are cured on steel belts running in high temperature ovens. The belt provides a smooth, clean surface with belt temperature capability to 750 ~ DOUBLE BELT PRESS
The double belt press is based on the principle of combining and compressing products between two solid steel conveyor belts that are welded endless and are
704
Process Plant Machinery
FIGURE 20-6 Tobacco sheet on a steel-belt conveyor dryer. (Source: Sandvik Process System, Inc., Totowa, NJ.) passing continuously through the press. As heating and/or cooling are generally involved in most press applications, the moving belts and belt support systems must have good heat transfer properties and must be capable of continuous operation at elevated temperatures and pressure. A smooth, uninterrupted surface is also required for a good surface finish on the product. The smooth, fiat, high-strength endless steel belt meets all of these requirements. Because the press can be built up in modules, it is possible to satisfy a great variety of process requirements with various combinations of pressing, heating, and cooling zones. The double belt press is normally designed to operate in a pressure range of one to ten bar. If the belts were placed directly in contact with the platens, the sliding friction created between the belt and press platen would be too high for pulling the belts through the press. Therefore, a unique traveling roller support system between the belts and press platens is used to eliminate sliding friction, allowing the belts to glide easily through the pressing zones (Figure 20-7 and 20-8). Figure 20-8 shows the overall press configuration with two terminal sections, a heating press zone and a heating or cooling press zone. Each belt is continuous and the product may be produced in any length. The roller chain assemblies called out in Figure 20-7 are detailed in Figure 20-8. The rollers are on an endless loop and travel at half the linear speed of the belt. A large number of closely spaced moving nip pressures pass over the surface of the product through the steel belt, even though there is no relative motion between
Conveyor-Based Processing Systems
Infeed Equipment To Suit
" Hydraulic Cylinders
Roller Chain Assemblies
DischargeEQuipment to Suit
"l'ension-Feed E
~-- Pre Heat - ~
FIGURE 20-7 Totowa, NJ.)
II
Heat,o g o, Cooling ~- Heat,ng PressZone --~ ]<-- PressZone ~
Drive"lerm,nal
Double-belt press general assembly. (Source: Sandvik Process System, Inc.,
TOPPRESSPLATE
PRODUCT ' O 11
Q
~BOTTOM
STEELBELT ROLLER
BOTTOMPRESSPLATE ~ HEATIhJG I:'LUID
FIGURE 20-8 Product is sandwiched between steel belts that are backed up by a traveling chain of rollers. (Source: Sandvik Process System, Inc., Totowa, NJ.) the product and the steel belt. This action tends to average out any tolerance variations, resulting in very uniform thickness control. Original designs were for belt widths up to 1550 mm, the maximum belt width previously available. With the development of new belt grades, such as precipitation hardened stainless steel and longitudinal techniques, it is now possible to increase width, but at a somewhat lower maximum pressure due to structural deflections and machine tolerances. As seen in Figure 20-8, heat transfer or cooling fluid is circulated through the upper and lower press platens. The heating or cooling is transmitted through the rollers to the steel belts and to the product. By controlling oil temperature and press pressure, and by using various length zones, almost any time, temperature, and pressure profile can be furnished to suit the product requirements. It should be noted that a product can be passed directly from a heating zone to a cooling zone without releasing pressure and with minimum loss of energy, since only the relatively thin steel belts are repeatedly heated and cooled. For low pressure, up to 0.33 bar, double belt presses can be designed with a closely spaced, stationary roller support system for the steel belts. In this case, the belts can be heated or cooled in zones by direct air impingement on the back side of the belts. This approach is not practical for pressures higher than 0.33 bar due to belt deflections between rollers (and therefore high stress) and roll deflection. Maintaining thickness uniformity in the product becomes more and more difficult.
705
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Process Plant Machinery
TABLE 20.5 Conveyors
Products Frozen on Steel Belt
Fish fillets Hamburgers Shrimp Coffee concentrate
Whole herring Chicken pieces Food sauces
Source: Sandvik Process System, Inc., Totowa, NJ. The press can be operated as a fixed pressure machine by allowing the upper platens to "float" at controlled hydraulic pressure on the product. The thickness control in this case is a result of the volume and compressibility of the material entering the press. The press can also be operated as a fixed-gap machine by pulling the top platens down against fixed stops. In this case, the thickness is held uniform; however, final density is controlled by the volume of material entering the press. In either case, a pressure release system is built in to protect against overloading and exceeding the design limits. Some typical applications for the double belt press are as follows: 9 glass-reinforced plastic laminate (PP, PET, PE, etc.) 9 Ski laminates (fiberglass rovings with epoxy resin) 9 printed circuit board (copper clad) 9 fiberglass-reinforced laminate 9 rubber mats 9 PVC sheet form-fused granules 9 synthetic leather FREEZING CONVEYOR
The stainless steel conveyor belt forms the basis of a freezing system for a variety of foods. Ease of maintenance and excellent hygiene properties are important qualities. Most products are frozen hard and readily break loose from the belt as the belt curves around the discharge pulley. Products frozen on steel belts are given in Table 20.5. Typical freezing time for a raw fish fillet, one-half inch thick, is ten minutes. Thinner products freeze even faster. Yield from this process is virtually 100 percent, since moisture losses can usually be eliminated and product breakage is minimal because of the smooth belt surface. The coolant used under the stainless steel belt is usually calcium chloride or propylene glycol. Upper belt coils may be arranged for ammonia or freon, and defrosting can be achieved by electric, water, or hot gas methods. Belt freezers have also been designed with refrigerated air both above and below the belt. New products may be evaluated in laboratory tests and production equipment sized accordingly.
Chapter 21 Filtration Systems Filtration is the process of separating suspended solids from liquid by means of a porous medium. Solid materials can be grouped as follows: 1. 2. 3. 4.
A A A A
soluble phase in which the material is soluble in the liquid as a true solution soluble or insoluble colloidal phase having a diameter of 0.001 to 1.0 lxm supracolloidal phase of particles 1.0 to 100 lam in diameter coarse-particle phase which settles or floats
Filtration is concerned with Groups 2 and 3, for Group 1 requires chemical separation and Group 4 is usually treated by settling or flotation. The mechanism of separation within all filter media is not wholly understood but is generally accepted to be any combination of the following: straining, sedimentation, agglomeration, adsorption, molecular cohesion, streaming potential, zeta potential, and ionization. When the filter medium loses porosity because of retention of solids to the point at which further filtration is ineffective or uneconomical, it must be cleaned or replaced. CLASSIFICATION OF FILTERS
Filters may be classified by combinations of: 1. Method of liquid movement through the filter, e.g., a. Pressure b. Gravity c. Vacuum 2. Method of medium handling, e.g., a. Permanent b. Periodic replacement c. Terminal (end-of-cycle) replacement 3. Type of medium, e.g., a. Screen b. Cartridge c. Cloth or paper d. Coarse granular e. Filter aid
Simple Screen Filters Screen filters have fixed pore sizes generally equal in size and not of great depth. Therefore, all solids above the fixed pore size are prevented from passing * Source:
De Laval Hardbook.
707
708
Process Plant Machinery
through. Screen filters may be made of wire mesh, perforated metal screens, porous membranes, or the edges of a series of disks or spiral-wound wire. Screens are generally cleaned and reused.
Cartridge Filters Cartridges are removable filter elements, preassembled or molded into the proper shape. They depend on depth filtration, removing solids within the labyrinth path of flow. They are made of fibrous material, metal wool, porous stone or sintered metal, filter aids, or combinations of these items. Cartridges are generally thrown away but may be cleaned or occasionally washed in place.
Cloth Filters Filter cloths are made of paper, textiles, plastics, rubber, and various other materials or combinations thereof. Although filter cloths are used in cartridge filters, they are more generally used in filter presses or fiat-bed indexing filters. Filter presses are a series of chambers formed by recessed filter plates or alternate frames and plates, forced together by a closing device. Filter cloths are placed over the filter plates and removed to be discarded or cleaned at the end of each cycle Flat-bed indexing filters use a roll of cloth which is indexed or pulled across the filtering area at the end of the cycle or on a continuous basis.
Coarse Granular Filters Granular filters (see Figure 21-l) all depend to a varying degree on depth filtration. The granules should be of uniform size and density to get the most effective filtration.
INLET DISTRIBUTOR AND BACKWASHTROUGH
~VE'L AIR
EFFLUENT INFLUENT
Iol
BACKWASH IN
BACKWASH OUT
FIGURE 21-1
Flow schematic: sand filter (Source: De Laval, Trenton, NJ)
Filtration Systems* Some filters use more than one filter medium, the size being successively reduced so as to give series filtration. In this case, different densities must be used to allow restratification after backwashing. Flow is normally downward but may be upward. Some filters have a moving bed, which is externally cleaned and returned continuously. Good filtration depends on even distribution of flow across the filter area. This can be a function of the underdrain system which supports and retains the filter medium while allowing the liquid to flow through. The underdrain system and the overflow or backwash troughs must evenly distribute the backwash for complete cleaning. Slow sand filters operate at very low flow rates of about 0.5 gal/ft 2. min and are not backwashed but scraped on the surface. Rapid sand filters produce high clarity up to about 8 gal/ft 2. min by using multiple media. High-rate filters operate up to 20 gal/ft 2 9min or higher and normally produce a lesser degree of clarity.
Precoat Filters A very fine, porous material such as diatomaceous earth, cellulose fibers, or perlite is used as a filter aid or medium in precoat filters (see Figures 21-2 and 21-3). The filter aid is distributed on and held by a septum or filter element. A very thin coating (about 1/16 in) is used initially and forms a precoat. The septa may be tubular or flat but generally are vertical and present a large amount of surface area within a small volume. The filter aids have the property of bridging across openings much larger than their own particle size. The cake is distributed and held on by the velocity of the liquid through it. Therefore, flow must never be stopped, and when filtration ceases temporarily, a recirculating or holding flow is established. The filter aid itself forms the filter medium, and use of varying sizes or grades of particles will change the effluent clarity. Most removal is accomplished on the surface, but some is due to the depth filtration of the many labyrinth passages. Most filter aids are also added continuously to the incoming stream as body feed. This BUMP AIR
VENT
CLEAN BACKWASH
EFFLUENT
t
O
I RECIRCULATION I PRECOATOR HOLDING
FILTER
INFLUENT FILTER PUMP BODYFEED
r"l
~~~~KWASH
DRAIN
PRECOAT
FILTER AID SLURRY TANK
FIGURE 21-2
Flow schematic: precoat filter (Source: De Laval, Trenton, NJ)
709
710
Process Plant Machinery
FIGURE 21-3
Pressure precoat filter (Source: De Laval, Trenton, NJ)
FIGURE 21-4 3500-gal/min automatic precoat filter for municipal water treatment (Source: De Laval, Trenton, NJ)
Filtration Systems*
FIGURE 21-5
Typical water-filter control panel (Source: De Laval, Trenton, NJ)
allows the cake to grow, entrapping the removed contaminants and maintaining porosity to increase the cycle length. Figure 21-4 shows the installation of a precoat filter for municipal-water treatment, and Figure 12-5 shows the control panel for such a system. FILTRATION WITH OTHERWISE ACTIVE MEDIA
Sometimes filtration is accomplished as an adjunct to another type of removal such as ion exchange in high-rate demineralizers or powdered-resin filters. Because of the expense of the other process, filtration is usually treated as secondary and should not be the controlling design factor. BACKWASHING
Backwashing is the term applied to cleaning or renewing the filter medium by using liquid and/or air. For permanent media, backwashing dislodges the entrapped contaminants and transports them to waste. With replaceable media, the medium and the contaminants are both removed. Liquid is used both as a washing and as a transporting medium. Generally a substantial velocity is required to perform these functions, although care must be exercised to prevent the loss of any permanent medium. Air is used either for agitation and scouting or to create a pressure release and impart high initial velocity to break the contaminants free. Backwashing is one of the most important aspects of filtration. The ability to renew full filtration capability after each cycle contributes to sustained effluent quality, reduced maintenance, and economical operation.
711
712
Process Plant Machinery
PRETREATMENT
Pretreatment is used with filtration either to condition the particulate matter by agglomeration so as to simplify or improve filtration or to precipitate soluble compound into filterable particulate material. One of the most common pretreatments is with coagulants, such as alum or ferric sulfate, which form a charged floc to attract small particles electrostatically. FILTER SELECTION
The selection of a filter to suit a specific process requires evaluation of many factors that are generally interrelated: 1. 2. 3. 4. 5. 6. 7.
Required clarity of effluent Maximum size of particle in effluent Variation of inlet flow and/or solids concentration Requirements for medium replacement or backwashing Disposal of removed solids and backwash liquid Use of pretreatment Operating economics
Since a greater degree of filtration requires higher operating costs, it is most important that care be exercised in determining this requirement. If a cutoff particle size is absolutely necessary, a screen or fixed-pore-size filter will be needed since depth filtration can give only a nominal rating based on average percentage removal for various sizes. Filters work best under constant conditions. Variations affect effluent, equipment sizing, and economics. Also of major importance is tying the filter application with the backwashing of equipment. The external requirements for backwashing and the disposal of the materials all affect the selection. The effect of any pretreatment on the operation of a filter must be considered. Finally, when filters are capable of performing the requirements, a full evaluation of economics must be undertaken. ECONOMICS
The costs of filtration are interrelated. They cover: 1. 2. 3. 4. 5. 6. 7.
Power to overcome friction and filter pressure drop and to perform backwash Backwash liquid Medium replacement Pretreatment costs, if used Amortization of investment Maintenance Labor
Power is utilized by pumps, compressors, agitators, controls, etc. An increase in maximum pressure drop will increase power consumption but is offset by less backwash cost and medium replacement because of longer runs and by a reduction of first cost. Backwash liquid is an expense if it costs to produce or handle it. Sometimes it is expressed in net output. Backwash air is reflected in power and investment cost.
Filtration Systems* Medium replacement is reduced by increasing the length of run with higher terminal pressure drop or lower filtration rates. However, these measures involve other offsetting costs. With filter aid, precoat usage is a function of surface area and therefore of filtration rate as well as length of run. An increase in body-feed usage will increase length of run and can result in reduced power usage or increased filtration rate with an associated reduction in first cost. When pretreatment is used for precipitation or treatment as an adjunct to filtration, it is an independent cost. When presettling or conditioning of particles is involved, costs must be weighed against the resultant reduction in other filtration costs. For true evaluation, first cost should be amortized as a cost of the final product. This should reflect interest rate on money, required time of depreciation, and equipment usage rate. Maintenance includes not only equipment upkeep but also replacement of permanent filter media. Most filtration systems can be automated, which may sizably reduce labor costs.
713
This Page Intentionally Left Blank
APPENDIX Conversion of Units (US and SI)
CONVERSION TABLES: METRIC (SI) SYSTEM OF UNITS* The modernized metric system is called the international system of units (with the international abbreviation SI). This system of metric units was established for international use in 1954 by the General Conference of Weights and Measures on the International System of Units, and the name international system was adopted by the same body in 1960. The SI units are classified into three categories: basic units, derived units, and supplementary units. The basic SI units, seven in number, are listed in Table A.2. The SI system was established by international agreement to provide a logical and interconnected framework for all measurements in science, industry, and commerce. The system is built upon the base units listed below. All other SI units are derived from these base units. Multiples and submultiples are expressed in a decimal system as in Table A.3. The seven basic units of the SI listed below are relatively easy to use and to convert to or from U.S. customary units. It is in the area of derived units, i.e., those units of measure which are derived from one or more of the basic units listed above, that readers unfamiliar with the SI metric system might meet difficulties. Further, certain units are expressed only in terms of other derived units. One SI metric unit which initially might cause difficulty is the unit of force, or the newton. This unit enters a variety of calculations in civil, mechanical, and electrical engineering and is defined as the force which produces an acceleration of one meter per second squared when applied to a mass of one kilogram. Table A.4 includes the more common derived units used in SI.
CONVERSION FACTORS AS EXACT NUMERICAL MULTIPLES OF SI UNITS Table A.5 expresses the definitions of various units of measure as exact numerical multiples of coherent SI units and provides multiplying factors for converting numbers and miscellaneous units to corresponding 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 " - 0 2 2.54*" expresses the fact that 1 inch = 2.54 x 10 -2 meter, exactly, by definition. Most of the definitions are extracted from National Bureau of Standards documents. Numbers not followed by an asterisk are only approximate representations of definitions or are the results of physical measurements.
* Condensed from McGraw-Hill Metrication Manual, New York, 1972. 715
716
Process Plant Machinery TABLE A.1
To convert
Conversion Tables: US Customary System
Multiply by
To obtain
To convert
Multiply by
To obtain
36 3 0.000568 63,360 5280 1760 6076
ft mi in ft yd ft
Units of length in
in in ft ft ft ft
0.0833 0.0278 0.0000158 12 0.3333 0.000189 0.000165
~ yd mi in yd mi mi(nautical)
yd yd yd mi mi mi mi(nautical)
in
Units of area in 2
in 2 ft 2 ft 2 ft 2
0.00694 0.000772 144 0.1111 0.00002296
ft 2 yd 2 in 2 yd / acres
9 0.0002066 43,560 4840 1296
ft 2
ft 3 ft 3 yd 3 yd 3 yd 3
0.0370 0.0000230 46,656 202 27
yd 3 acre.ft in 3 gal
acre.ft bbl (oil) imperial gal acre.ft
325,800 42 1.2 43,560
gal gal gal
yd / yd 2 acres acres yd 2
acres ft 2 yd 2 in 2
Units of volume in 3
in 3 in 3 gal gal gal gal gal gal ft 3 ft 3
0.00433 0.000579 0.0000214 231 0.1337 0.00495 0.00000307 0.0238 0.8327 1728 7.48
gal ft 3 yd 3 in 3 ft 3
yd 3 acre.ft bbl (oil) imperial gal in 3 gal
ft 3
ft 3
Units of weight gr
gr gr OZ oz OZ
gal lb
0.00229 0.0001429 0.0(O0(K)0714 438 0.0625 0.00003125 8.3322 0.12002
oz lb tons gr lb tons lb* gal*
lb lb lb tons tons tons lb/h* gal/min
7000 16 0.0005
14,000,000
gr OZ
tons gr
32,000 2000 0.0020003 499.925
oz
0.0002778 0.01667 0.0001894 1.689 1.467 88 5280
ft/s ft/min mi/h ft/s ft/s ft/min ft/h
24 0.0329
Hours Months
lb gal/min Ib/h*
Units of velocity ft/s ft/s ft/s ft/s ft/min ft/min ft/min
60 3600 0.682 0.592 0.01667 60 0.01136
ft/min ft/h mi/h kn ft/s ft/h mi/h
ft/h ft/h ft/h kn mi/h mi/h mi/h Units of time
Seconds Seconds
0.01667 0.0002778
Minutes Hours
Days Days
Appendix: Conversion of Units (US and SI) TABLE A.1
Conversion Tables: US Customary System (Continued) \
i
To convert
Multiply by
To obtain
To convert
Multiply by
To obtain
Seconds Minutes Minutes Minutes Hours Hours Hours Hours Hours Days Days
0.00001157 60 0.01667 0.000694 3600 60 0.0417 0.001370 0.0001142 86,400 1440'
Days Seconds Hours Days Seconds Minutes Days Months Years Seconds Minutes
Days Months Months Months Months Months Years Years Years Years Years
0.00274 2,628,000 43,800 730 30.42 0.0833 31,536,000 525,600 8760 365 12
Years Seconds Minutes Hours Days Years Seconds Minutes Hours Days Months
Btu Btu Btu Btu in.lb in.lb in.lb in.lb ft.lb ft.lb
9340 778.3 0.000293 0.000393 0.000107 0.0833 0.00000003138 0.0000000421 0.001285 12
in.lb ft.lb kWh hp.h Btu ft.lb kWh hp.h Btu in.lb
kW kW kW kW kW kW hp hp hp hp hp hp ft.lb/s ft.lb/s ft.lb/s ft.lb/s
1.341 738 44,260 0.948 56.9 3413 0.7455 550 33,000 0.707 42.41 2545 0.001356 0.001818 60 0.001285
hp ft.lb/s ft.lb/min Btu/s Btu/min Btu/h kW ft. Ibis ft.lb/min Btu/s Btu/min Btu/h kW hp ft.lb/min Btu/s
Units of work, energy, and heat ft.lb ft.lb kWh kWh kWh kWh hp.h hp.h hp.h hp.h
0.0000003766 0.000000505 3413 31,872,000 2,656,000 1.341 2545 23,760,000 1,980,000 0.7455
kWh hp.h Btu in.lb ft-lb hp.h Btu in-lb ft.lb kWh
0.0771 0.00002260 0.0000303 0.01667 0.00002141 0.001285 1.055 1.415 778.3 46,700 60 0.01758 0.02357 12.97 778.3 0.01667
Btu/min kW hp ft.lb/s Btu/s Btu/min kW hp ft.lb/s ft.lb/min Btu/min kW hp ft.lb/s ft.lb/min Btu/s
Units of power ft. Ibis ft-lb/min ft.lb/min ft.lb/min ft.lb/min ft.lb/min Btu/s Btu/s Btu/s Btu/s Btu/s Btu/min Btu/min Btu/min Btu/min Btu/min
Units of pressure (water, mercury at 68 ~ in water in water in water in water in water ft water ft water ft water
0.0833 0.0736 82.98 0.03602 5.1869 12 0.8832 995.8
ft water in mercury oz/ft 2 lb/in 2 (psi) lb/ft 2 in water in mercury oz/ft 2
oz/ft 2 oz/ft 2 oz/ft 2 oz/ft 2 oz/ft 2 lb/in 2 (psi) lb/in 2 (psi) lb/in 2 (psi)
0.01205 0.001004 0.000887 0.000434 0.0625 27.762 2.314 2.314/sp gr
in water ft water in mercury lb/in 2 (psi) lb/ft 2 in water ft water ft (any liquid)
717
718
Process Plant Machinery TABLE A.1
Conversion Tables: US Customary System (Continued)
To convert
Multiply by
To obtain
ft water ft water ft (any liquid) in mercury in mercury in mercury in mercury in mercury in mercury atm
0.4322 62.24 0.4322 x sp gr 13.57 1.131 1128 0.4894 70.47 0.03342 29.92
Ibis Ibis Ibis lb/min lb/min lb/min
60 3600 86,400 0.01667 60 1440
ft3/s
60 3600 7.48 448.8 26,930 646,317 1.983 0.01667 60 0.1247 7.48 448.8 0.0002778 0.01667 0.002078 0.1247 7.48 0.1337
lb/in 2 (psi) lb/in 2 (psi) lb/ft 2 lb/in 2 (psi) lb/in 2 (psi) lb/in 2 (psi) in water lb/in 2 (psi) ft water lb/ft 2 oz/ft 2 lb/ft 2 lb/in 2 (psi) lb/ft 2 lb/ft 2 lb/ft 2 atm lb/ft 2 in mercury atm Weight- time rates lb/min lb/h* lb/h lb/h lb/day lb/h Ibis lb/h lb/h lb/day lb/day lb/day lb/day Volume- flow rates ft3/min gal/s ft 3/h gal/s gal/s gal/s gal/min gal/s gal/h gal/min gal/day gal/min acre.ft/day gal/min ft3/s gal/min ft3/h gal/min gal/s gal/min gal/min gal/h gal/h gal/h ft 3Is gal/h ft3/min gal/h gal/s gal/h gal/min bbl/min (oil) gal/h bbl/day (oil) ft3/s acre.ft/day
ft3/s ft3/s ft3/s ft3/s ft3/s ft3/s ft3/min ft3/min ft3/min ft3/min ft3/min
ft3/h ft3/h ft3/h ft3/h ft3/h gal/s
To convert
Multiply by
To obtain
2.043 230.4 144 0.06802 0.1928 0.01607 0.01419 16 0.00694 14.7
in mercury oz/ft 2 lb/ft 2 atm in water ft water in mercury oz/ft 2 lb/in 2 (psi) lb/in 2 (psi)
0.0020003 0.0002778 0.01667 24 0.00001157 0.000694 0.04 17
gal/min Ibis lb/min lb/day Ibis lb/min lb/h
8.022 481.3 60 3600 0.00223 0.1337 8.022 0.01667 60 499.925 0.0000371 0.00223 0.1337 0.0002778 0.01667 42 0.0292 0.5042
ft3/min ft 3/h gal/min gal/h ft3/s ft3/min ft3/h gal/s gal/h lb/h* ft3/s ft3/min ft 3/h gal/s gal/min gal/min gal/min ft3/s
*Water at 68 ~F. TABLE A.2
Basic SI Units
Quantity
Unit
Length Mass Time Electric current Temperature (thermodynamic) Amount of substance Luminous intensity
meter (m) kilogram (kg) second (s) ampere (A) kelvin (K) mole (mol) candela (cd)
Appendix: Conversion of Units (US and SI) TABLE A.3
Prefixes for SI Units
Multiple and submultiple 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 =
TABLE A.4
1012 109 106 103 102 10 10 -1 10 -2 10 -3 10 -6 10 -9 10 -12 10 -15 10 -18
Prefix
Symbol
tera giga mega kilo hecto deka deci centi milli micro nano pico femto atto
T G M k h da d c m /z n p f a
D e r i v e d Units o f the International S y s t e m
Quantity
Name of unit
Area Volume Frequency Density Velocity Angular velocity Acceleration Angular acceleration Volumetric flow rate Force Surface tension
square meter cubic meter hertz, cycle per second* kilogram per cubic meter meter per second radian per second meter per second squared radian per second squared cubic meter per second newton newton per meter, joule per square meter newton per square meter, pascal* newton-second per square meter, poiseuillet meter squared per second joule, newton-meter, watt-second
Pressure Viscosity, dynamic Viscosity, kinematic Work, torque, energy, quantity of heat Power, heat flux Heat flux density Volumetric heat release rate Heat transfer coefficient Heat capacity (specific) Capacity rate Thermal conductivity
watt, joule per second watt per square meter watt per cubic meter watt per square meter kelvin joule per kilogram degree watt per kelvin watt per meter kelvin
Unit symbol or abbreviation, where differing from basic form
nz
N N/m,J/m 2 N/m 2, Pa*
Unit expressed in terms of basic or supplementary units* m2 m3 s -1 kg/m 3 m/s rad/s m/s 2 rad/s 2 m3/s kg. m/s 2 kg/s 2
N. s/m 2, PIt
kg/m. s 2 kg/m. s
J,N 9m, W . s
m2/s k g . m2/s 2
J/kg. K W/K
k g . m2/s 3 kg/s 3 k g / m . s3 kg/s 3 . K m 2/s 2 9K k g . m2/s 3 9K
Jm W/m. K, .mm-rTK
k g . rrds 3. K
W, MS
WIm 2 W/m 3
Wlm 2. K
*Supplementary units are plane angle, radian (rad); solid angle, steradian (sr). *Not used in all countries.
719
720
Process Plant Machinery "rABI.E A.5
Listing by Physical Quantity
To convert from
To
Multiply by
Acceleration foot/second 2 free fall, standard gal (galileo) inch/second 2 acre are barn circular mil foot 2 hectare inch 2 mile 2 (U.S. statute) section township yard 2
meter/second2 meter/second2 meter/second2 meter/second2 Area meter2 meter2 meter2 meter2 meter2 meter2 meter2 meter 2 meter2 meter2 meter2
-01 +00 -02 -02
3.048* 9.806 65* 1.00" 2.54*
+03 +02 -28 -10 -02 +04 -04 +06 +06 +07 -01
4.046 856 422 4* 1.00" 1.00" 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*
+03 +04 +01 +02
1.00" 2.767 990 5 1.601 846 3 5.153 79
Density gram/centimeter~ lbm/inch 3 lbm/foot 3 slug/foot 3 British thermal unit (ISO/TC 12) British thermal unit (International Steam Table) British thermal unit (mean) British thermal unit (thermochemical) British thermal unit (39 ~ British thermal unit (60 ~ calorie (International Steam Table) calorie (mean) calorie (thermochemical) calorie (15 ~ calorie (20 oC) calorie (kilogram, International Steam Table) calorie (kilogram, mean) calorie (kilogram, thermochemical) electron volt erg foot-lbf foot-poundal joule (international of 1948) kilocalorie (International Steam Table) kilocalorie (mean) kilocalorie (thermochemical) kilowatthour kilowatthour (international of 1948) ton (nuclear equivalent of TNT) watthour
kilogram/meter~ kilogram/meter~ kilogram/meter3 kilogram/mete? Energy joule joule
-+-03 1.055 06 +03 1.055 04
joule joule joule joule joule joule joule joule joule joule
+03 1.055 87 -t-03 1.054 350 264 488 +03 1. 059 67 +03 1.054 68 +00 4.1868 +00 4.190 02 +00 4.184" +00 4.185 80 +00 4.181 90 +03 4.1868
joule joule joule joule joule joule joule joule
+03 +03 -19 -07 +00 -02 +00 +03
4.190 02 4.184* 1.602 10 1.00" 1.355 817 9 4.214 011 0 1.000 165 4.1868
joule joule joule joule joule joule
+03 +03 +06 +06 +09 +03
4.190 02 4.184" 3.60* 3.600 59 4.20 3.60*
Appendix: Conversion of Units (US and SI) TABLE A.5
Listing by Physical Quantity (continued)
To convert from
To
Multiply by
Energy/area time Btu (thermochemical)/foot2-second Btu (thermochemical)/foot2-minute Btu (thermochemical)/foot2-hour Btu (thermochemical)/inch2-second calorie (thermochemical)/cmE-minute erg/centimeterE-second watt/centimeter2 dyne kilogram-force (kgf) kilopond kip lbf (pound-force, avoirdupois) ounce-force (avoirdupois) pound-force, lbf (avoirdupois) poundal angstrom astronomical unit cable caliber chain (surveyor's or Gunter's) chain (engineer's) cubit fathom fermi (femtometer) foot foot (U.S. survey) foot (U.S. survey) furlong hand inch league (U.K. nautical) league (international nautical) league (statute) light-year link (engineer' s) link (surveyor' s or Gunter' s) meter micrometer mil mile (U.S. statute) mile (U.K. nautical) mile (international nautical) mile (U.S. nautical) nautical mile (U.K.) nautical mile (international) nautical mile (U.S.)
watt/meter2 watt/meter2 watt/meter2 watt/meter2
+04 +02 +00 +06
watt/meter2
+02 6.973 333 3
watt/meter2 watt/meter2
- 0 3 1.00" +04 1.00"
Force newton newton newton newton newton newton newton newton Length meter meter meter meter meter meter meter meter meter meter meter meter meter meter meter meter meter meter meter meter meter wavelengths Kr 86 meter meter meter meter meter meter meter meter meter
-05 +00 +00 +03 +00 -01 +00 -01
1.134 893 1.891 488 3.152 480 1.634 246
1.00" 9806 65* 9.806 65* 4.448 221 4.448 221 2.780 138 4.448 221 1.382 549
1 5 8 2
615 615 5 615 543
- 1 0 1.00" +11 1.4959789 +02 2.194 56* - 0 4 2.54* +01 2.011 68* +01 3.048* - 0 1 4.572* +00 1.8288" - 1 5 1.00" -01 3.048* +00 1200/3937" - 0 1 3.048 006 096 -+-02 2.011 68* - 0 1 1.016" - 0 2 2.54* +03 5.559 552* +03 5.556* +03 4.828 032* +15 9.460 55 - 0 1 3.048* - 0 1 2.011 68* -I-06 1.650 763 73* - 0 6 1.00" - 0 5 2.54* +03 1.609 344* +03 1.853 184" +03 1.852" +03 1.852" +03 1.853 184" +03 1.852" +03 1.852"
260 5* 260 5* 260 5* 76*
721
722
ProcessPlant Machinery TABLE A.5
Listing by Physical Quantity (continued)
To convert from pace parsec perch pica (printer's) point (printer's) pole rod skein span statute mile (U.S.) yard carat (metric) dram (avoirdupois) dram (troy or apothecaries') grain gram hundredweight (long) hundredweight (short) kgf-second2/meter (mass) kilogram mass Ibm (pound mass, avoirdupois) ounce mass (avoirdupois) ounce mass (troy or apothecaries') pennyweight pound mass, Ibm (avoirdupois) pound mass (troy or apothecaries') scruple (apothecaries') slug ton (assay) ton (long) ton (metric) ton (short, 2000 pounds) tonne Btu (thermochemical)/second Btu (thermochemical)/minute calorie (thermochemical)/second calorie (thermochemical)/minute foot-lbf/hour foot-lbf/minute foot-lbf/second horsepower (550 foot-lbf/second) horsepower (boiler) horsepower (electric) horsepower (metric) horsepower (U.K.) horsepower (water) kilocalorie (thermochemical)/minute kilocalorie (thermochemical)/second watt (international of 1948)
To meter meter meter meter meter meter meter meter meter meter meter Mass kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram kilogram Power watt watt watt watt watt watt watt watt watt watt watt
watt watt watt watt watt
Multiply by -01 7.62* +16 3.083 74 -bOO 5.0292* - 0 3 4.217 517 6* - 0 4 3.514 598* +00 5.0292* +00 5.0292* -b02 1.097 28* -01 2.286* +03 1.609 344" -01 9.144" - 0 4 2.00* - 0 3 1.771 845 195 312 5* - 0 3 3.887 934 6* - 0 5 6.479 891" - 0 3 1.00" -b01 5.080 234 544* -b01 4.535 923 7* -bOO 9.806 65* +00 1.00" -01 4.535 923 7* - 0 2 2.834 952 312 5* - 0 2 3.110 347 68* - 0 3 1.555 173 84* -01 4.535 923 7* -01 3.732 417 216" - 0 3 1.295 978 2* -b01 1.459 390 29 - 0 2 2.916 666 6 +03 1.016 046 908 8* +03 1.00" +02 9.071 847 4* +03 1.00" +03 1.054 350 -b01 1.757 250 -bOO 4.184" - 0 2 6.973 333 - 0 4 3.766 161 - 0 2 2.259 696 +00 1.355 817 +02 7.456 998 +03 9.809 50 +02 7.46* +02 7.354 99 +02 7.457 +02 7.460 43 +01 6.973 333 +03 4.184" +00 1.000 165
264 488 4 3 0 6 0 7
3
Appendix: Conversion o f Units (US and SI) TABLE A.5
Listing by Physical Quantity (continued)
To convert from
To
Multiply by
Pressure atmosphere bar barye centimeter of mercury (0 ~ centimeter of water (4 ~ dyne/centimeter ~ foot of water (39.2 ~ inch of mercury (32 ~ inch of mercury (60 ~ inch of water (39.2 ~ inch of water (60 ~ kgf/centimeter 2 kgf/meter 2 lbf/foot 2 lbf/inch 2 (psi) millibar millimeter of mercury (0 ~ pascal psi (lbf/inch 2) torr (0 ~ foot/hour foot/minute foot/second inch/second kilometer/hour knot (international) mile/hour (U.S. statute) mile/minute (U.S. statute) mile/second (U.S. statute) Celsius Fahrenheit Fahrenheit Rankine day (mean solar) day (sidereal) hour (mean solar) hour (sidereal) minute (mean solar) minute (sidereal) month (mean calendar) second (ephemeris) second (mean solar)
newton/meter ~ newton/meter 2 newton/meter 2 newton/meter 2 newton/meter 2 newton/meter 2 newton/meter 2 newton/meter 2 newton/meter 2 newton/meter 2 newton/meter 2 newton/meter 2 newton/meter 2 newton/meter 2 newton/meter 2 newton/meter 2 newton/meter 2 newton/meter ~ newton/meter 2 newton/meter 2 Speed meter/second meter/second meter/second meter/second meter/second meter/second meter/second meter/second meter/second Temperature kelvin kelvin Celsius Kelvin Time second (mean solar) second (mean solar) second (mean solar) second (mean solar) second (mean solar) second (mean solar) second (mean solar) second second (ephemeris)
+05 1.013 25* +05 1.00" - 0 1 1.00" +03 1.333 22 +01 9.806 38 - 0 1 1.00" +03 2.988 98 +03 3.386 389 +03 3.376 85 +02 2.490 82 -I-02 2.4884 +04 9.806 65* +00 9.806 65* +01 4.788 025 8 +03 6.894 757 2 +02 1.00" +02 1.333 224 +00 1.00" +03 6.894 757 2 +02 1.333 22
-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
tx = tc + 273.15 tK = (5/9)(tF + 459.67) tc -- (5/9)(tF -- 32)
tK -- (5/9)te +04 8.64* +04 8.616 409 0 +03 3.60* +03 3.590 170 4 +01 6.00* +01 5.983 617 4 +06 2.628* +00 1.000 000 000 Consult American Ephemeris and Nautical Almanac
second (sidereal) year (calendar)
second (mean solar) second (mean solar)
- 0 1 9.972 695 7 +07 3.1536"
723
724
ProcessPlant Machinery TABLE A.5
Listing by Physical Quantity (continued)
To convert from year (sidereal) year (tropical) year 1900, tropical, January, day 0, hour 12 year 1990, tropical, January, day 0, hour 12 centistoke stoke foot2/second centipoise Ibm/foot-second lbf-second/foot 2 poise poundal-second/foot 2 slug/foot-second rhe acre-foot barrel (petroleum, 42 gallons) board foot bushel (U.S.) cord cup dram (U.S. fluid) fluid ounce (U.S.) foot 2 gallon (U.K. liquid) gallon (U.S. dry) gallon (U.S. liquid) gill (U.K.) gill (U.S.) hogshead (U.S.) inch 3 liter ounce (U.S. fluid) peck (U.S.) pint (U.S. dry) pint (U.S. liquid) quart (U.S. dry) quart (U.S. liquid) stere tablespoon teaspoon ton (register) yard 3 *Exact by definition.
Multiply by
To second (mean solar) second (mean solar) second (ephemeris)
+07 3.155 815 0 +07 3.155 692 6 +07 3.155 692 597 47*
second
+07 3.155 692 597 47
Viscosity meter2/second meter2/second meter2/second newton-second/meter2 newton-second/meter2 newton-second/meterz newton-second/meter2 newton-second/meter 2 newton-second/meter2 metexa/newton-second Volume
-06 -04 -02 -03 +00 +01 -01 +00 +01 +01
meted meted
+03 1.233 481 9 -01 1.589 873
mete? meted meted meted meter3 meted meted meted meted meted meted meted meted meted meted meted meted meted meted meted meted meted meted meted meted meted
-03 -02 +00 -04 -06 -05 -02 -03 -03 -03 -04 -04 -01 -05 -03 -05 -03 -04 -04 -03 -04 +00 -05 -06 +00 -01
1.00" 1.00" 9.290 1.00" 1.488 4.788 1.00" 1.488 4.788 1.00"
304* 163 9 025 8 163 9 025 8
2.359 737 216" 3.523 907 016 688* 3.624 566 3 2.365 882 365* 3.696 691 195 312 5* 2.957 352 956 25* 2.831 684 659 2* 4.546 087 4.404 883 770 86* 3.785 411 784* 1.420 652 1.182 941 2 2.384 809 423 92* 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 1.00" 1.478 676 478 125" 4.928 921 593 75* 2.831 684 659 2* 7.645 548 579 84*
Index Acceleration head 329 Acetylene production 558 Adjustable frequency drives 20-23, 590 Aeropropulsion Test Facility, compressors 515 Affinity laws, pump performance 255, 277 Agitators s e e Mixers Air density ratios table 530 Air horsepower 526 Air Moving and Control Association (AMCA): design guidelines 528 rating tables 529 Air separation plants, compressors 463, 477 American Gear Manufacturers Association (AGMA), speed reducer standards 358 American National Standards Institute (ANSI): dimensional standards for process pumps 212 pressure containers 410 American Petroleum Institute (API): compressor standard 553 heavy duty pump standard 217, 219 screw compressor standards 603 Ammonia plant, compressors 466 Antisurge control, compressors 384, 461,503, 505 ASME Standard, pressure containers 410 Atomization, in liquid fuel burners 65 Autoclave cover construction 414 Avon gas generators 70-2 Avon turbine s e e Industrial Avon turbine Axial fans 517- 20 control 538 Axial flow compressors 47, 475 applications 475, 477 bearing pedestals 496 bearings 500 blade carrier 497 blading 497-8 casings 496 control systems 501,504-6 corrective maintenance 487 design fundamentals 478-82 environmental considerations 487 for gas turbines 47 inspection 485 instability conditions 482-4 internal cleaning 486 oil supply system 509 operational limitations 482-4 P & I diagrams 506-10 performance capabilities 477-8 process control systems 501-6 product range 488 rotating blades 480-2 rotor 497
safety systems 503-4 selection 487- 8, 490- 95 shaft seals 499 solid couplings 500-1 standard maintenance 484-5 stationary blades 478-80 Type A (FIXAX)489-90 Type AV (VARAX)489-90 Axial flow impellers 620-3 Axial flow pumps 231 Back-pressure turbines 182 Backwashing, filters 711 BACT (Best Available Control Technology) 94 Banbury-type mixers s e e Mixers, internal Barrel compressors 407 applications 409, 422-3 casing design pressures 412-13 construction materials 422-3 couplings selection 418- 20 design 410-11 RB Series design features 414-15 selection 407-9 shaft seal applications 417 shaft seal functions 416-17 shaft string layout 421 - 2 Bearings: for electric motors 12 for gas compressors 375- 7 for gas turbines 74 for turbocompressors 435 for vertical turbine pumps 276 Bendix starter drive 132- 3 Best Available Control Technology (BACT) standards 94 Bevel gears 39-40 Blade frequency analysis 167-9 Blast furnace plant, compressors/blowers 477, 511-12 Bleed turbines 188-9 Borescope inspection: gas generators 90 gas turbines 117 Bottom-entering mixers 628 Brake horsepower (bhp) 251 Burner design, for turbines 65 Butadiene plants, axial flow compressors 477 Bypass regulation, compressor 584-5 Cam and piston pump s e e Vacuum pumps, rotary oil-sealed Campbell diagram 167-9, 482 Capacity, pipeline 251 725
726
Index
Capacity control systems: compressors 584- 91 rotary piston blowers 610 Carbon dioxide emissions 94 Carbon monoxide emissions 94 Carbon ring seals 598 Cartridge filters 708 Casing, barrel compressor: design pressures 412-13 functions 410-12 Cavitation 252- 3 Centrifugal compressors 364-5 aerodynamic design 393 auxiliary equipment 382 bearings 375-7 casings 381 construction 371,373 control equipment 383-4 design diagrams 394-6 drive components 382 environmental considerations 392 for gas turbines 47 high-speed 398-406 horizontally split 365-6 impellers 373-4 lube oil system 385-7 monitoring equipment 391 - 2 operating characteristics 393 process coolers 383 rotor 374-5 seal liquid system 387-90 sealing elements 377-81 sidestream 366-9 stationary components 382 thermodynamic design 393 train arrangement 370 use of intercooling 397 vertically split 369 Centrifugal fans 520 backward-inclined blade type 523 capacity control 533 - 8 cooling towers 354-5 forward-inclined blade type 523-4 radial blade type 520-3 Centrifugal pumps 207 affinity laws 255 canned motor 237-9 capacity 251 cavitation 252- 3 characteristic curves 253-4 environmental considerations 207 net positive suction head 252 parallel operation 258 peripheral velocity 249 power and efficiency 251 pressure head 249- 51 sealless magnet-drive 237, 239 specific speed 251-2 system curves 255-8 types and applications 208-12 variable vane 207 volute 207 see also Conventional process pumps; High-speed pumps
Centrifuges disk nozzle 633-5 applications 647-9 calculations 636-40 Characteristic curves, pump 253-4 Chemicals pumping, construction materials table 260-3 Chlorine compressors 565-6 Chlorobenzenes, separation process 688-9 Choke control, axial compressors 484 Choked flow instability, axial flow compressors 484 Circuit breaker 5 Circulator pumps 231 Clarification process 639 Classification process 639, 644-7 conventional 645 improved 645, 647 Cleaning, compressor internals 486 Clearance space control, compressors 588-90 Closed-loop control, extruders 673 Cloth filters 708 Coal gasification plant, oxygen compressors 466 Cogeneration principle 67 Combustion inspections, gas turbine 116-17 Compressed air energy storage plant (CAES) 515 Compressor cycle, ideal see Ideal compressor cycle Compressor support systems, two-stroke gas engines 125- 7 Compressors: bypass capacity control 584-5 capacity control systems 584-91 for chlorine 565-6 cleaning 93 for difficult gases 564-6 dry-running 567- 9, 592-4 gas seal design 467-74 for gas turbines 46-8 oil-injected 601 rotating positive-displacement 595 selection 581 - 2 variable-speed control 590-1 volume control methods 602-7 see also under specific type names
Concentration cell corrosion 260 Concentration process 639 Condensing turbines 182-4 Conduit box, electric motor 13 Control enclosures, electric motors 5 - 6 Conventional process pumps: ANSI type 212 API type 217, 219 axial flow 231 between-bearing type 213 circulator 231 close coupled 214 dual volute 215 frame mounted 214 high-capacity 219 high-speed 214 in-line 215-17 low flow high head centrifugal type 213 multistage centrifugal 213 multistage horizontal split case type 213
Process Plant Machinery
for paper stock 219 Roto-jet 222, 303-7 self priming 214 for slurry 222-7 for solids handling 214 submersible wastewater type 214 vertical turbine 229- 31 Conversion Table: English-Metric 173 SI units 715-24 Conveyors see Steel-belt conveyors Cooling: double-belt 702 on steel-belt 698-700 Cooling fans, electric motor 13 Cooling system, two-stroke gas engines 129-32 Cooling towers: airflow 347-9 construction 349 counterflow 347 crossflow 347-9 double-flow 348 dry type 350 fan drive shafts 36 fan safety 361 fan speed reducers 357-9 fan-assisted natural draft 346 fans 352-7 hybrid draft 346 induced draft 345- 6 mechanical draft 345-6 mechanical equipment 351 plume abatement 350 propeller type fans 353-4 shapes 349 spray-filled 348-9 water conservation 350 Cooling water supply systems 345-6 fans 352-61 heat transfer methods 350 mechanical components 351 see also Cooling towers Coriolis effect 639 Corrective capacitance 19 Corrosion 259-60 from difficult gases 565 Corrosive chemicals table, pumping materials selection 260- 3 Couplings, gas compressors 418- 20 Crevice corrosion 260 Critical speed, pump rotor/system 288 Crosshead-type compressors 549-55 Cryogenic turboexpander plant 200-3 Crystallization process 679, 681 - 2 advantages 683-4 batch 686 continuous 686- 7 for paraxylene 685 Crystallizer, scraped surface continuous 685-7 advantages 687 common applications 687-90 secondary recovery 690 special applications 690-1 Curtis staging, steam turbines 137, 173, 176 Cut impellers 278
Cycling (of electric motors) 7 Cyclone separators 271 D-R Gas Seal 467, 472 Damper control, fans 535-6 Degree of fill, extruders 672-3 Demand charges 19 Dewaxing exchangers/chillers: mechanical seals 693 scraped surface 692-3 shear pin sprockets 693-4 Dewaxing lube oil 648 Diaphragm compressors 548, 576-8 Diaphragm pumps 318- 22 Differential gear set 39-40 Direct action brake motor 9 Discharge, from mixers 654 Discharge pressure control, gas compressors 384 Disk nozzle centrifuges 636-40 principles 633-5 Displacement ratio 644 Distillation process 681 - 2 Double cone burner see EV-burner Double reduction gears 34 Double-belt coolers 702 Double-belt press 703-6 Double-extraction turbines 188 Double-helical gear 36 Drip-proof motor 7 Drip-proof motors 2 Drive shafts, cooling tower fans 360 Drop-forming process 701 Dry Low-NOx burner 100 Dry screw compressors, volume control 602-603 Dry running compressors: higher pressures 592-4 piston ring 568- 71 use 567-9 Dual fuel burner see EV-burner Dynamic discharge head 250 Dynamic head 250 Dynamic suction head 249 Dynamic suction lift 249 Dynamic system head 257 Efficiency performance, electric motor 13-16 Electric motors 1- 2 adjustable frequency controllers 20-23 controls 5 - 6 cycling 7 efficiency performance 13-16 energy-saving designs 9 explosion-proof 2, 25 in hazardous areas 9 heaters 4 high-efficiency design 3 horizontal 1 insulation design 3 large heavy-duty AC 8 load matching 17-18 maintenance 18 motor service factor 3 - 4 multispeed 9 open 2
727
728
Index
Electric motors 1-2 (continued) operating costs 16-18 selection 13, 16, 23- 5 service factor 3 - 4 special designs 8 - 9 starter controls 5 submersible 9 three-phase design 2 - 3 torque design 4 two-speed design 3 ventilation 2 vertical 1, 9 wiring system design 6 see also under specific type names
Elliptical gears 29 End shields, electric motor 13 Energy demand see Demand charges Engine-compressor units 544 Environmental considerations: axial flow compressors 487 centrifugal pumps 207 gas turbines 46, 94 seal performance 271 steam turbines 135 Epicyclic gears 30, 32 Equalizer gears 29 Erosion-corrosion process 260 Ethylene production facilities compressors 347, 540, 558-60 power generation 107- 8 EV-burner 94- 5 annular combustion chambers 100 applications 100 operating principle 95-101 silo combustors 100 Expanders: design and construction 200-3 for gas separation 203-4 maintenance 205 operation 204- 5 for power generation 199- 200 for power recovery 199 for refrigeration 199 Expansion factor, liquid 333-4 Explosion-proof motors 8 External gear pump 321,323 Extraction steam control 149-53 Extraction turbines 177- 9, 188 Extruders 660 automatic (closed loop) controls 674 barrel construction 662-4 construction materials 676-7 dies 664 double-screw 660-4, 666-7 instrumentation 673 maintenance 677-8 manual (open-loop) controls 673-4 mechanical description 660-4 melting process 668, 675 mixing process 657, 675 pressure-generating capacity 669, 676 process parameters 672 process task 664-5 process variables 669- 71 scaling up factors 674-6 single-screw 660-6
type comparisons 667-9 venting 668, 676 Fan law factors 526, 531- 3 Fans: capacity control 533-8 classification 517 definitions 525-6 fan law factors 531- 3 performance 526- 9 performance corrections 529- 31 system resistance effects 527-9 see also under specific type names
Fatty materials, separation process 689-90 Feed hoppers 654 Feeding devices, extruders 664 Field testing, pumps 271-4 Fill factor 658 Filters: backwashing 711 classification 707-10 economics 712-13 extruder 664 with other methods 711 selection 712 FIXAX compressors see Axial flow compressors, Type A Flakers see Steel-belt conveyors, for cooling Floating ring seal 379-80 Fluid catalytic cracking plant, compressor module 514 Fluid mechanics, impeller power requirements 617-23 Flywheel effect, motor startup 6 Formulae listing, for pump design 259 Frame, electric motor 12 Freezing conveyor 706 Friction head 249 Fuels: contaminants 113 effect on parts life 113 Fusible safety switch 5 Galvanic corrosion 259-60 Gas compression machinery, overview 363-4 Gas engines: basics 121 reciprocating 121 see also Two-stroke gas engines Gas film stiffness 469-70 Gas generators: borescope inspection 90 compressor cleaning 93 condition monitoring 88-90 dual fuelled 81 fuel control systems 79, 81 hydraulic oil supply 82-3 lubrication 82 mechanical monitoring 90 modular construction 73 performance characteristics 84 performance monitoring 90 Phase I model 76, 79, 82 Phase II model 77-9, 82-3 power turbine matching 85 routine servicing 90-3
Process Plant Machinery
Gas injection compressor train 425 Gas seals 381 gas film stiffness 469-70 groove pattern 468-9 hang-up syndrome 470-2 L sleeve design 473-4 pressure profile 468-9 principle 467-8 Gas transmission project, compressor 427-8 Gas turbines 45-6, 103 air systems 63-5 balancing 62 bearings 74 burner design 65 closed cycle 103 cogeneration 67 combined-cycle 106 combustion chamber design 48, 51-3 combustion process 48-50 compressor design 46-8 condition monitoring 88-90 environmental factors 94-101 fuel supply system 51, 79-83 generator and power matching 85 heat recovery cycles 105-108 heavy-duty industrial 109-11 hydraulic oil supply 82-3 life cycle usage 112-14 load duty 114 lubrication 82- 3 maintenance inspections 115-19 major inspections 119 nominal power output 67 operational track record 67-8 performance characteristics 84-8 power losses 57 regenerative-cycle 107 routine servicing 90-3 selection 67-9 simple-cycle 103- 5 starting system 75-9 thermal cycling effects 113 thermal efficiency 67 turbine balancing 62 turbine construction 58-60 turbine design 54-7 turbine materials 60-1 see also EV-burner Gases, compressing difficult 564-6 Gear speed reducers, fans 358 Gears 27 bevel 39-40 helical 33-7 hypoid 40-1 special applications 42 spur 27- 33 worm 37-8 Glands: piston-rod 571 - 3 quench type 271 GMVH gas engine see Two-stroke gas engines Goodman diagram, modified 167, 169 Goodman factor 169 Granular filters, coarse 708-9 Guarded motors 2
Heat recovery steam generators (HRSG) 105-6 combined-cycle system 106- 7 Heating/drying applications, steel-belt conveyors 703 Helical gears 33-7 Herringbone gear 36 High-speed centrifugal compressors 398-406 High-speed pumps: applications 243-4 basic concept 239-40 design advantages 240-1 hydraulic capabilities 241 - 2 maintenance 247-8 mounting arragements 242-3 system controls 244-6 Hot gas path parts (gas turbines) 112-15 inspections 118-19 Hot-melt adhesives, forming 700 Hydraulic diaphragm pumps 318- 22 Hydraulic horsepower (whp) 251 Hydrogen sulfide (sour gas), compression difficulties 565 Hydrogen-rich gas compression 369 Hypoid gears 40-1 Ice cream/margarine machine 685-6 Ideal compressor cycle 541-2 Impellers: axial flow airfoil 623 cut 278 fluid foil 623 for gas compressors 373-4 high efficiency design 623 low- and high-capacity 280-1 for mixing equipment 620-3 radial flow 623 vane tip filing 279 varying vane numbers 281,284 Impulse turbines 56, 136, 138, 155-6 casing 140- 3 configuration overview 137, 139 journal bearings 147- 8 pressure-compounded (Rateau) staging 137, 174-6 thrust bearings 147 velocity-compounded (Curtis) staging 137, 173, 176 Impulse-reaction turbines 56 Impurity reduction ratio 642 Induction motors: components 11-13 standard design 7 - 8 Industrial Avon turbine: development 71 gas generator options 76-7 Infinite stepless control, compressors 587-8 Inlet valve unloading, compressor control 585 Inlet vane control, fans 534-5 Inorganic chemicals, crystallization advantages 690 Inspection patterns: for axial flow compressors 485 for gas turbines 115-19 for high-speed pumps 247 Intake pressure throttling, compressor control 585
729
730
Index
Intercooling, gas compressor 397 Intergranular corrosion 260 Interference diagram s e e Campbell diagram Intermittent gears 30 Internal gear pump 323-4 International Critical Tables 681 International Standards Organisation (ISO), fuel burning conditions 111 International system of units 715 Isotherm turbocompressors: applications 463-6 ARI design features 440-8 ARI technical data 459 axial-radial concept 440 blade carder 444 casing 436, 444 common features 430-34 control system and oil supply 461 coolers 438, 448 design features 448 inlet guide vanes 438 journal and axial bearings 435 P & I diagram 462 power oil supply 447 RIK design features 435-8, 449 RIO design features 439, 453 rotor and impellers 438, 447 seals 436, 438, 451 skid-mounted single-lift 435 solid couplings 434 stator blades 444-536 typical plant layout 460 Isothermal power 581 - 2 Journal beatings, steam turbines 147-8 Kaolin clay particle classification 647 KPV-type pump s e e Vertical mixed flow variable pitch vane pump Labyrinth seals: for expanders 202 for gas compressors 377- 9, 613 for steam turbines 148, 152 Labyrinth-piston compressors: applications 573-6 design 568 disadvantages/advantages 570-1 operation 568- 71 piston-rod glands 571 - 3 Lean premix burner 100-1 Liquid jet vacuum pump 341 Liquid ring vacuum pump 338-40 air ejector 341 booster 341 Liquid sealing systems 416 Load matching, motor 17 Load-torque requirements, electric motors 20-3 Lobe-type blowers s e e Rotary piston blowers Lobe-type pumps 325 Low NOx burners 94-5 applications 100 development history 100 operating principle 95, 99 Low-density polyethylene production 558, 564
Lube oil dewaxing 648 Lubricating/lube oil systems gas compressor 385 gas engine 125- 7 Lubrication, by water s e e Water lubrication Magnetic chip detection 92 Magnetic flow meter 272 Maintenance: axial flow compressors 484-5, 487 electric motors 18 expanders 205 extruders 677-8 gas turbines 115-19 high-speed pumps 247-8 mixers 660 seals 248 Manometer 272 Mass transfer, extruders 672 Masterbatching (mixing process) 657 Materials selection, pump construction 260 Maximum/minimum flow limit control 244-6 Mechanical contact seal 380 Mechanical diaphragm pumps 318, 322 Mechanical seals 268-9 contact 380 corrosion 271 environmental controls 271 for exchangers/chillers 693 tandem 270-1 temperature control 271 types 270 Metering pumps 312-14, 322 hydraulic diaphragm type 318- 21 mechanical diaphragm type 318 packed piston type 314-15 packed plunger type 317-18 Metric units, conversion tables 715- 24 Mining plant, compressors 465 Mixers: bottom-entering 628 heavy-duty top-entering 625-7 light-duty top-entering 627 portable 623, 625 shaft sealing 628- 31 side-entering 627-8 Mixers, internal: auxiliary systems 655 control systems 655-6 design features 652-6 downstream equipment 659-60 heat transfer 652, 659 maintenance 660 operating parameters 658-9 operating techniques 657 principles 651 - 2 rotors 656-7 Mixing equipment 617 fluid shear rate 618 power requirements 617-19 pumping capacity 618 tank baffles 618-19 velocity head 618 Mixing mills 659 Mixing sequence and time 659 Modular construction, Rolls-Royce turbines 73
Process Plant Machinery
Mollier diagram 171 Motor torque 4, 20-6 Motors, electric see Electric motors Multiscrew pump 325-6 Multistage mixing 657 National Aeronautics and Space Administration (NASA), aerofoil sections 480 National Electrical Code: explosion-proof motors 8 hazardous location requirements 2 National Electrical Manufacturers Association (NEMA): AC motor classification 13-14 control enclosures 5 - 6 motor rating 16 motor standards 2, 8, 13-14 Natural gas, supply and distribution 135 Natural gas liquefaction plant, compressors 514 Net Positive Suction Head (NPSH) 252 Net Positive Suction Head Required (NPSHR) 331 Newton (unit) 715 Nitric acid plant: axial flow compressors 477 compresser/expander trains 512 isotherm air compressor 464 turbocompressor train 513 Nitrogen compressor train 464 Nitrogen oxide emission reduction 94, 100 Nonfused disconnect switch 5 Nozzle, for capacity measurement 272 Nozzle guide vanes 58, 60 Nozzle passing freguencies 167-8 Nozzle passing frequencies 167-8 OEM (original equipment manufacturer) 46, 75 Offshore locations: compressor modules 426-7, 555 gas generator modules 73-4 Oil field, sour gas compressor 554 Oil and gas production, compressors 422, 429 Oil separation, oil-injected compressors 601 Oil-free compressors 548, 555 Oil-injected compressors 601 On-line cleaning, axial flow compressors 486 One-step clearance control, compressors 588-90 Open-loop control, extruders 673-4 Operating Deflection Shape (ODS), vertical pump 286 Organic chemicals: continuous crystallizers 685-94 crystallization advantages 683-4 low energy separation 681 - 3 solubility thermodynmics 679- 81 Orifice: for pump discharge capacity measurement 272 for pump discharge capacity reduction 284 Overfiling, vane tips 279 Overload protection 6 Overspeed devices, steam turbines 155 Overspeed trip devices 153 Oxygen compressors 422, 574-5 Oxygen plant 465-6
Packed piston pumps 314-15, 322 Packed plunger pumps 317, 322 Packets, turbine blades 168 Paper stock pumps 219 Parallel operation pumps 258 Paraxylene, crystallization process 685 Pastillating devices 701 Peaking duty, gas turbines 112 Pellitizing equipment 664 Peristaltic pumps 326-7 Petrochemical plant, compressors 575 Petroleum refineries: axial flow compressors 475 power generation 107 Petroleum wax, removal from lube oil 648 pH values, corrosion indicator 260 Pipeline capacity calculation 251 Piping system design, effect on pump operation 263-7 Piston rings 562-4 dry-running 566- 71 packed plungers 563-4 PTFE 567-9, 592 Piston-rod glands 571 - 3 Piston-rod load 329, 544 Pitot tube 273 Pitting corrosion 260 Planetary gears 30, 32 Plastics, for dry running components 592 Plastics plant, power generation 109 Pollutant reduction 94 Portable mixers 623, 625 Positive displacement pumps: peristaltic 326- 7 reciprocating see Reciprocating displacement pumps rotating see Rotating displacement pumps Power and efficiency, pump 251 Power factor 18-19 Power pumps, metering 312- 21 Precoat filters 709, 711 Pressure head 249 Process control, turbocompressor 461 Process coolers 383 Process pumps, conventional see Conventional process pumps Propeller-type fans 517 cooling tower 353-4 operating laws 355-7 Propylene compressor 575 Protective devices, electric motors 5 - 6 PTFE, for dry-running use 567-9, 592 Pulsation Control Tools (PCTs) 329 Pump capacity: characteristic curves 253 measurement 272-4 reciprocating pump 329 Pump efficiency 251 Pump formulae 249-52 Pumps: construction materials 260-3 field testing 271-4 varying performance 277-84 see also under specific type/use names
Purification process see Washing processes Purifiers 633
731
732
Index
Rack and pinion gears 29 Radial bearings 375-6 Radial expansion turbines s e e Expanders Radial flow impellers 620-23 Ram cylinders, internal mixers 654 Ram pressure, mixing process 659 Rateau staging, steam turbines 137, 174-6 RB211 s e e Rolls-Royce RB211 turbines Reaction turbines 56, 136, 155-6 beatings 163 blade tip sealing 161 casing supports 162 casings 157 guide blade carders 157-9 rotor and blading 161 shaft glands 161 valves and governing system 163-5 Reactive KVAR 18 Reactor coolers, scraped 691 Real power (KW) 18 Reciprocating compressors 539 arrangement types 544 balanced-opposed 544 capacity range 545-7 classification 542- 3 crosshead 544, 549-53 cylinder heads and valves 545 with cylinder lubrication 542-3, 549-53 double-acting 544 with dry-running piston rings 566-8 horizontal opposed cylinders 552-3, 555 ideal cycle 541 - 2 labyrinth-piston 568-76 with nonlubricated cylinders 543, 566 performance calculations 579-80 piston rings 562-4 plungers 563-4 power requirements 581 - 2 seals 561 suction capacity conversion 582-3 vertical 552-3 with water-lubricated cylinders 555-8 see a l s o Diaphragm compressors Reciprocating gas engines 121 Reciprocating pumps 309 operating principles 329-35 power type 311 - 21 steam type 310 Reed Frequency 285, 287-8 Refineries, barrel compressors 422 Refrigeration processes, gas compressors 366-9 Residence time, extruders 672 Resistance temperature detectors (RTDs) 248 Restriction orifices, pump discharge 284 Reynolds number 621,623, 639-40 Rod load s e e Piston-rod load Rod-load ratings, compressor 544 Roll mills 659 Rolls-Royce RB211 turbines: materials specification 72- 3 modular construction 73 starting system 7 5 - 9 Rotary piston blowers 607 capacity control 610
design and construction 607-8 operation 608-10 Rotary screw compressors 547, 595 application limits 596 beatings 597 calculation procedures 604 construction 596 oil-injected 601 seals 597-9 volume control 602- 3 Rotating case design pumps 303-7 Rotating positive displacement pumps 321, 323 - 4 calculation nomenclature 610 external gear 321,323 internal gear 323-4 lobe type 325 multiscrew 325-6 single-screw 325 vane type 324 Rotating-bar breakers 701 Roto-jet pumps 222, 303-307 Rotor blades, compressor 480 Rotors: electric motor 12 gas compressor 374-5 internal mixers 656-7, 659 Running inspections, gas turbines 115-16 Safety shutdown controls, gas engine 134 Safety systems, turbocompressor 461 Salt works, compressors 513 Sanitary codes requirements 9 Scaling up, extruders 674-6 Seal liquid system 387-90 Seals: axial-flow compressor 499 barrel compressor 416-17 dry 417 gastight 418 maintenance 248 piston-ring 416 rotary compressor 597- 9 shutdown/emergency 631 stuffing box 267-8 turbocompressor 436, 438, 451 wet 417-18 s e e a l s o Gas seals; Labyrinth seals; Mechanical seals Secondary compressors (recirculators) 558 Secondary recovery, crystalllizer 690 Sedimentation principles 636-40 Selective catalytic reduction (SCR) 94 Self-priming pumps 214 Separators: basic principles 633-5 s e e a l s o Centrifuges Service factor, electric motor 3 - 4 Shaft sealing 416 for mixers and agitators 628-31 Shaft string layout, barrel compressor 421-2 Shear rate, mixing plant 618 SI system of units 715- 24 Side-entering mixers 627-8 Sigma factor (centrifugal studies) 640
Process Plant Machinery
Simple screen filters 707-8 Single screw process compressors, water-flooded 611 applications 615-16 design principles 612-15 design program 615 discharge temperatures 615 efficiencies 615 lubrication 612-13 nomenclature 616 operation theory 611 - 12 sealing 613-15 Single-screw pump 325 Single-stage mixing 657 Slaters see Steel-belt conveyors, for cooling Sliding vane compressors 547 Slip (in electric motors) 1 Slip loss volume, compressor 605 Slip (reciprocating pump) 331 Slurry, definition 224-5 Slurry pumps 222-7 Solid couplings, axial flow compressors 500-1 Solid materials, suspended 707 Solids-handling pumps 214 Solids-liquid slurry, use of disk nozzle centrifuge 648 Solubility thermodynmics: ideal case 679 non-ideal case 679 predictions 679-81 Soluble recovery process 639 Soluble reduction ratio 644 Specific diameter 398 Specific energy 672 Specific speed 251,398, 526 Speed control: by frequency adjustment 20-3 cooling tower fans 357-9 fans 536-7 gas compressors 383-4 Speed-torque characteristics 20-3 Spiral bevel gears 39-40 Splash-proof motors 2 Spur gears 27-33 Squeeze film, bearings 74 Squirrel-cage induction motors 1 Stability limits, barrel compressors 409 Stall instability, axial flow compressors 482 Standby inspections, gas turbines 115 Starter controls, electric motor 5 Stator blade adjustment, axial flow compressors 506-8 Steam balance, process plant 179-81 Steam turbines: building block approach 185- 7 construction materials 139 control valves 189-92 design factors 172-6 diaphragms 145-6 environmental factors 135 extraction steam control 149-53 for gas turbine heat recovery 105-7 labyrinth seals 148, 152 mechanical drive applications 136 nozzle rings 143-5 overspeed devices 155
rotor categories 141 - 3 selecting and sizing 172-6, 182-8 steam consumption 171 theoretical steam rates 171 trip devices 153 turbine components 140 see also under specific type names
Steel-belt conveyors 695 accessories 700-1 belt grades 695-7 for cooling 698-700 double-belt 702 double-belt press 703-6 freezing applications 706 heating/drying applications 703 Steelworks plant, compressors 466 Stepped gears 32-3 Stokes' law 636, 639-40, 645 Stop motion gears 30 Strainers 271 Stress corrosion 260 Strip-forming 700-1 Stuffing box sealing 267-8 Sublimation condensing 691 - 2 Submersible wastewater pumps 214 Suction capacity 582-3 Suction head 249 Suction lift 249 Suction piping design, centrifugal pumps 263-7 Suction pressure limit control 246 Suction volume 579-80 Sulphur solidification 698 Supercooling materials 700 Surge control, axial flow compressors 482-3 Symbols listing 259 Synchronous motors 1 Syngas compressor unit 425, 428 System curves, pump 255-8 Tank design, mixing process 619-20 Technical Rules for Acetylene Plants and Calcium Carbide Storage (TRAC) 558 Temperature sensors 248 Terephthalic acid processing 648 Theoretical Steam Rates (TSR) 171 Throttling, compressor plant 384 Thrust bearings 376-7 axial flow compressors 500 steam turbines 147 Thrust collar gears 434 Timed suction valves 588 Top-entering mixers heavy duty 625-7 light-duty 627 Total pump head, testing 271- 2 Totally enclosed fan-cooled (TEFC) motor 7 Trip devices 153 steam turbines 153 Trunk-piston compressors 548-9 Tube axial fans 519 Turbine blade design 58-9, 61 Turbine disks 58, 60-1 Turbine pumps 207 Turbines see Extraction turbines; Gas turbines; Steam turbines Turbocharger, two-stroke gas engines 133-4
733
734
Index
Turbocompressors: control system 5 0 1 - 6 see also Axial flow compressors; Isotherm turbocompressors Turboexpanders see Expanders Two-stroke gas engines 122 components 123- 5 compressor support systems 125- 7 control system 134 cooling system 129-32 fuel system 127- 9 power train 132- 3 turbocharger 133-4 Underfiling, vane tips 279 Underwriters Laboratory (UL), motors approval scheme 2, 9 V-belt drives, cooling tower fans 359 Vacuum formula 249 Vacuum pumps 337 air ejector 341 liquid jet type 341 mechanical booster 341 multistage combination units 342-3 rotary oil-sealed 3 4 3 - 4 single-stage liquid ring type 338-41 van T'Hoff equation 679-81 Vane axial fans 517 Vane type pump 324 VARAX compressors see Axial flow compressors, Type AV Variable-speed control, compressors 590-1 Velocity head 249 Venturi meter 272 Vertical mixed flow variable pitch vane pump 299- 302
Vertical turbine pumps 228-31 axial thrust calculation 2 7 5 - 6 bearing materials 276 motor changes 289 pumphead changes 289-90 turbine nomenclature 274-5 vibration nomenclature 295- 8 vibration reduction 285-95 see also Vertical mixed flow variable pitch vane pump Vibration monitoring, gas compressor 392 Viscoseal 613-15 Viscous fingering 643 Viscous liquids, cooling 690-91 Voltage unbalance 16 Volumetric efficiency: compressor 579 in hydrocarbon pumping 333 reciprocating pump 329, 332-3 Volute tip, trimming 279 Vortexing, pump suction 264 Wash ratio 641 - 2 Washing processes 634-5, 639-40 counter current dilution 6 4 1 - 2 counter current displacement 644 dilution 6 4 0 - 4 1 , 6 4 9 displacement 643-4, 649 Water compressibility table 332-3 Water-lubricated compressors 555 - 8 Wear mechanisms, extruders 6 7 6 - 7 Weather protected motors 2 Weir, for capacity measurement 272 Wiring system design 6 Woodward hydraulic relay 134 Worm gears 3 7 - 8
