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Waste Heat Boiler Deskbook by v. Ganapathy

Library of Congress Cataloging-in-Publication Data Ganapathy, V. Waste heat boiler deskbook / by V. Gartapathy. p. cm. Includes Index. ISBN 0-88173-122-6 1. Waste heat boilers--Handbooks, manuals, etc. I. Title. TJ319.G36

1991

621.1'83--dc20

90-85871

elP

Waste Heat Boiler Deskbook / By V. Ganapathy. ©1991 by The Fairmont Press, Inc. All rights reserved. No part of this publication may be reproduced or transmitted in any form or by any' means, ~lectronic or mechan~cal, including photocopy, r.ecprd~g,. 01: finy information .storage and retneval system, wIthout permIssIon m wntmg from the publisher. Published by The Fairmont Press, Inc. 700 Indian Trail Lilburn, GA 30247 Printed in the United States of America

10 9 8 7 6 5 4 3 2 1

ISBN 0-88173-122-6

FP

ISBN 0-13-950890-2

PH

While every effort is made to provide dependable information, the publisher, authors, and editors cannot be held responsible for any errors or omissions. Distributed by Prentice-Hall, Inc. A Simon & Schuster Company Englewood Cliffs, NJ 07632 Prentice-Hall International (UK) Limited, London Prentice-Hall of Australia Ply. Limited, Sydney Prentice-Hall Canada Inc., Toronto Prentice-Hall Hispanoamericana, S.A., Mexico Prentice-Hall of India Private Limited, New Delhi Prentice-Hall of Japan, Inc., Tokyo Simon & Schuster Asia Pte. Ltd., Singapore Editora Prentice-Hall do Brasil, Ltda., Rio de Janeiro

Contents

Preface ......................................................................................... xi Acknow ledgements ..................................................................... xiii List bf Frequently Used Abbreviations .......................................... xv Introduction .............................................................................. xvii CHAPTER 1: Heat Recovery Systems ............................................. 1 HRSGs for gas turbines ........................................................... 5 Natural versus forced circulation boilers ................................. 6 Auxiliary firing ..................................................................... 9 Computing fuel requirements ................................................. 20 HRSG system efficiency ....................... ;............................... 30 Fresh air firing .................................................................... 31 HRSG design features ........................................................... 33 Finned surfaces and design ................. ;.................................. 33 Steaming economizers ........................................................... 34 Emissions of NOx and CO ..................................................... 35 Methods of reducing pollutants ............................................. 37 Bypass dampers ................................................................... 41 Recent trends ........................................................................ 44 STIG and Cheng cycle systems ............................................... 47 Enhanced on recovery applications ...................................... 51 Reciprocating engine heat recovery ....................................... 52 Hydrogen plant waste heat boilers ....................................... 52 Boilers for sulfuric acid plant ............................................... 57 Incineration and heat recovery ............................................. 62 Solid waste incineration ....................................................... 62 RDF firing ........................................................................... 72 Fluid bed combustors for MSW .............................................. 74 Hazardous waste incineration .............................................. 76 High temperature corrosion .................................................. 77

v

Heat recovery boilers ........................................................... 80 Incineration of wood wastes, tires ......................................... 84 Incineration of liquids, fumes, VOCs ..................................... 87 Air heating applications ...................................................... 91 References ............................................................................ 93 CHAPTER 2: Fire Tube Boilers ...................................................... 97 Guidelines for fire tube boilers .............................................. 99 Design procedure ................................................................. 100 Determination of tube side coefficient .................................. 101 Example of design ""'"'''' ""'" ....................... ""'"'' ............. 103 Effect of tube size on design .................. """ .......................... 108 Simplified approach to design ............................................ 109 Predicting boiler performance .............................................. 117 Simplified approach to predicting performance ................... 119 Checks for fouling ............................................................... 120 Effect of scale on boiler performance .............. " ...... " ............. l21 Hydrogen plant boilers ..................... """"""'''''''' ............... 123 Gas bypass flow calculations ............................................... 124 Determining heat losses from boiler ..................................... 125 References ........................................................................... 126 Nomenclature ..................................................................... 126 CHAPTER 3: Water Tube Boilers ................................................ .131 Guidelines for water tube boilers .......................................... 133 Heat transfer calculations ................................................... 135 Convective heat transfer coefficient .................................... 135 Determination of tube side coefficient .................................. 139 Non-luminous heat transfer coefficient ................................ 140 c~ Gas pressure drop calculations ............................................. 148 In-line versus staggered arrangement ................................... 153 Design of evaporators .......................................................... 155 Performance calculations ..................................................... 162 Selecting designs with low pinch and approach points ............................................................ 165 Comparison of bare versus finned evaporator .......... " ............ 167 Radiant heat transfer ............................. "" ........................ 168 HRSG configuration and circulation""""""" ........ "" ........ " .170 vi

Design of superheaters ........................................................ 176 Design procedure ................................................................. 179 Performance calculations ..................................................... 181 NTU method of performance calculations ............................. 181 Metal temperature calculations .../ ...................................... .182 External radiation .............................................................. 186 Flow in parallel streams ..................................................... 189 Minimizing tube wall temperatures ..................................... 190 Steam temperature control. .................................................. 191 Design of economizers ......................................................... .196 Performance of complete HRSG ............................................ 197 References ...........................................................................201 Nomenclature .....................................................................201 CHAPTER 4: Simulation of HRSG Design and Performance ............................................ 205 Importance of HRSG simulation ........................ '" ................ 205 Design and performance calculations .................................... 206 Design temperature profile ..................................................207 Guidelines for selecting pinch and approach points ............................................................209 Example of design ............................................................... 213 Performance calculation procedure ....................................... 216 Software for HRSG simulation - COGEN ............................. 228 Supplementary firing and HRSG efficiency .......................... 231 Improving efficiency of HRSG ............................................. 231 Deaeration steam calculations ............................................. 238 Steam turbine calculations ................................................... 241 Optimizing temperature profiles using COGEN .................... 245 Using field data to simulate HRSG performance ..................248 Multi-pressure HRSG design and performance simulation ................................................. 250 References ...........................................................................254 Nomenclature ..................................................................... 255 CHAPTER 5: Specifying Waste Heat Boilers ............................... 257 Application or system design ............................................~ .. 257 Space and layout guidelines ................................................ 259 vii

~r---------------------------------------------------------

Gas data ............................................................................. 260 Boiler duty .........................................................................263 Auxiliary fuel data .............................................................. 264 Emission data .................................................................... .265 Feed water analysis, blow down .......................................... 265 Surface area, fin configuration ............................................. 270 Cost data for fuel, electricity and steam............................... 271 Drum sizing ......................................................................... 271 References ........................................................................... 274

APPENDIX A: Finned Tubes ........................................................ 275 Heat transfer calculations ................................................... 276 Fin efficiency and effectiveness ........................................... 278 Gas pressure drop ................................................................ 278 Tube wall and fin tip temperature ........................................ 280 Design example .................................................................. .281 Comparison of bare versus finned evaporator ........................ 286 Comparison of in-line versus staggered arrangement ................................................. .287 Fin configuration and performance ....................................... 290 Importance of tube side coefficient ....................................... 291 Effect of fouling factors ........................................................ 292 Surface area and duty .......................................................... 300 Nomenclature .....................................................................304 References ...........................................................................306

APPENDIX B: Low Temperature Corrosion .................................. 307 Causes and cures .................................................................. 307 Methqds of avoiding cold end corrosion ................................312 Condensation on surfaces ......................................................314 Corrosion is stacks, ducts ......................................................315 Heat loss calculations through multi-layer insulation ...................................................320 Hot casing design ................................................................322 Nomenclature .....................................................................326 References ...........................................................................326

viii

APPENDIX C: Heat Transfer Equipment Vibration ........................................................... 327 APPENDIX D: Gas Turbine Data ................................................. 337 APPENDIX E: Gas and Steam Properties ...................................... 349 Specific heat, viscosity, thermal conductivity of gases ........................................349 Enthalpy of gases ............................................................... .351 Estimating flue gas properties .............................................351 Effect of pressure on heat transfer ........................................353 Converting % volume to % weight ........................................ 355 Properties of steam and compressed water ............................355 APPENDIX F: Tube Thickness Calculations ..................................377 Tubes and pipes subject to internal pressure ........................... 377 Designing vessels and tubes subject to external pressure ..................................... 381 APPENDIX G: Conversion Factors ................................................391 INDEX ....................................................................................... 395

ix

-~~~~-~~--

F

~------~---~-----~~~~~,

Preface During the past 20 years I have had the opportunity of engineering a wide variety of industrial boilers and Heat Recovery Steam generators. During the past 7 years at ABeO Industries in particular I have had the pleasure of custom designing over two hundred fire tube and water tube waste heat boilers, each with different gas/steam parameters, which are in operation in the USA and abroad; these units were built for diverse heat recovery applications such as gaseous, liquid, solid waste and hazardous waste incineration systems, gas turbine exhaust, effluents from chemical plants such as sulfuric acid and hydrogen plants, petrochemical plants, cat crackers in refineries and for effluents from clean as well as dirty processes; the gas flow ranged from 2000 to 1.5 million pounds per hour, which implies a wide variety of boiler configurations and design features as you will see in the text. Energy management programs are vital to the economic life of any industry and heat recovery boilers playa dominant role in those projects which otherwise waste energy from hot flue gases. I decided to write this book after reviewing hundreds of specifications for heat recovery boilers prepared by consultants and would be users of the equipment; unfortunately several of them are poorly written without emphasis on the process aspects and optimization of installed plus operating costs, with a result that the end user or the owner gets an equipment which perhaps meets the budget requirement but which could incur significant operating costs in the form of higher gas pressure drop or fuel consumption or lower steam production year after year. The book addresses various aspects of heat recovery boilers, such as engineering, specifying, system design, optimization and performance evaluation. Hence engineers and managers involved in several disciplines of energy management including plant operation will find the book useful and informative. xi

The book is dedicated to professionals involved in any way with energy conservation and heat recovery. As pointed out by one, the earth is not for man, but man is for earth. Hence let us use the limited natural energy resources wisely with the future of mankind and the next generation in mind.

V. Ganapathy

xii

F

I I I

Acknowledgements

I

I would like to thank ABCO Industries for their encouragement and support in the preparation of this book and for the use of several ABCO illustrations and photographs.

I

I would also like to thank the following publications for permitting me to use my articles, which originally appeared in them:

I I

i

Power Power Engineering Chemical Engineering Oil and Gas Journal Hydrocarbon Processing Heating, piping and Air-conditioning Sci-Tech Publications Pennwell Books Marcel Dekker Inc. I would also like to thank ESCOA Corp for permitting me to use their correlations for extended surface heat transfer calculations. Several readers from various continents have been writing to me regarding my publications during the past several years, which has been indeed been motivating and I would like to thank them for their interest.

V. Ganapathy

xiii

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List of Frequently Used Abbreviations ABMA - American Boiler Manufactures Association ASME - American Society of Mechanical Engineers CO - Carbon Monoxide EOR - Enhanced Oil recovery FrB - Fire tube boiler GTE - Gas turbine exhaust HRSG - Heat Recovery Steam Generator MSW - Municipal Solid Waste NIMBY - Not in my back yard NOx - Nitrogen oxides NWL - Normal water level PPB - Parts per billion PPM - Parts per million RDF - Refuse Derived Fuel SCR - Selective Catalyst Reduction System STIG - Steam Injected Gas Turbine TDS - Total dissolved solids WHB - Waste Heat Boiler WTB - Water tube boiler VOC - Volatile Organic Compounds

xv

-----._---------------

.~~-~---------------.-----------

Introduction The book is aimed at engineers, consultants and managers involved in specifying, operating, engineering, marketing and procuring waste heat boilers (WHBs) or heat recovery steam generators (HRSGs). It offers valuable information on not only the heat recovery systems in chemical plants, gas turbine cogeneration and combined cycle plants, solid waste, liquid and gaseous incineration systems and flue gas heat recovery in general, but also provides the characteristics of each system such as gas analysis, fouling and slagging tendencies, high and low temperature corrosion potential and the impact of these on design and performance aspects of HRSGs. During the past 15 years and particularly during the last 7 years at ABCO Industries I have had the opportunity of engineering a wide variety of fire tube and water tube waste heat boilers for different types of applications as mentioned above. Having designed over two hundred boilers with gas flows varying from 2000 to 1.5 million pounds per hour and steam flows varying from 2000 to 250,000 pounds per hour, I feel that custom designing HRSGs is an art as well as a science, as there are numerous configurations possible depending on economics, cleanliness of gas, gas and steam parameters and layout considerations. I have had also the opportunity of authoring four books and over 175 articles on heat recovery boilers and steam plant systems in journals such as Power, Power Engineering, Chemical Engineering, Heating Piping Air-Conditioning, Oil and Gas Journal, Hydrocarbon Processing, Plant Engineering; the feed back from the readers has been very encouraging, which prompted me to bring out this work. Another reason was that in the course of reviewing specifications on HRSGs from various consultants and engineering organizations, I felt that less emphasis was being placed on process and optimization aspects, which is very important in the long run to the owner of the plant. Due to lack of knowledge on HRSGs and their performance

xvii

aspects, several of the specifications are poorly written and do not furnish adequate information to engineer an economically and technically sound design. Many engineers also lack knowledge or do not know how to evaluate alternate design options. For example if you read Appendix A, several examples are given to show that with finned tubes one can have a lower surface area and still transfer more energy by proper choice of fin configuration. Several engineers and purchase managers still purchase HRSGs for critical applications based on surface area and are of the view that more the surface area the better and a design with a lower surface area would not perform. Also, I have come across several specifications which do not place emphasis on HRSG operating costs; during the life time of the HRSG, the cost of moving the gas through the system due to high gas pressure drop or the cost of fuel which is required to generate a desired quantity of steam may be very significant. While the consultant looks at the initial cost alone, to the owner of the plant who has to pay for the fuel and electricity for years to come, the life cycle cost of the HRSG is important. Hence addition of secondary heat recovery surfaces such as condensate h;eater or economizer may have to be looked into, though the initial cost may be slightly more. A few examples on the subject of evaluating operating and life cycle costs are discussed in the book. This book offers useful information on design and off-design performance aspects of Heat recovery systems and components such as superheaters, evaporators and economizers, which are elaborated by over 65 fully worked out examples. You will find quantitative answers to commonly asked questions on heat recovery boilers and systems; some of them are: • • • • • •

How can one improve the efficiency of a HRSG system? What is the effect of auxiliary firing on system efficiency? How to compute the fuel requirements and oxygen consumption for gas turbine exhaust boilers? How to select pinch and approach points? How do they vary with gas inlet conditions? What is the effect of scale on boiler performance and tube wall temperatures and heat flux? How to compute the dew points of hydrochloric acid, sulfuric acid, hydrobromic acid, nitric acid?

xviii

• • • • • • • • • • • •

Which is better arrangement for bare and finned tubes, in-line or staggered? How do boilers with finned tubes compare with bare tube design for the same duty? How to compute the gas temperature at the SCR at off-design conditions. How to avoid high and low temperature corrosion problems? How to use field data to predict off-design performance or fouling of HRSGs? With finned tubes can you transfer more duty with less surface area? What is the effect of fin configuration? How to size and predict off-design performance of fire tube and bare/finned water tube boilers, superheaters, economizers? How to compute tube wall and fin tip temperatures? How to compute thickness of tubes subject to internal or external pressure? How much gas should be bypassed for gas temperature control? What is the effect of gas pressure on heat transfer? How to evaluate HRSGs for possible noise and vibration problems?

The first chapter deals with heat recovery systems. HRSGs are used in various applications such as gas turbine exhaust, incineration systems, chemical plants and refineries to mention a few. In order to design a HRSG for any application, the characteristics of the gas stream are important. For example, auxiliary firing in gas turbine HRSGs is discussed in depth with examples on computing fuel requirements, oxygen consumption, impact on system efficiency and emissions. Features of boilers such as natural or forced circulation, single or multiple gas pass design, insulated casing or fully water cooled membrane wall construction are discussed, along with methods of minimizing steaming concerns in economizers. Various aspects of WHBs in Municipal Solid Waste (MSW) applications, Refuse Derived Fuel (RDF) fired units and other incineration systems are discussed with emphasis on type of boilers, whether fire tube or water tube, fouling and slagging concerns and high and low temperature corrosion potential. Methods of minimizing these concerns through boiler design and selection of steam parameters are addressed. The second and third chapters deal with Design and off-design performance calculation procedures for fire tube and water tube

xix

boilers with bare and extended surfaces. Plant engineers can use the simplified procedures described in the text for instance to check for fouling, estimate bypass flow for gas temperature control or estimate the gas temperature at the Selective Catalytic Reduction system SCR) at different load conditions. Effect of tube size and arrangement whether in-line or staggered on design and performance is elaborated quantitatively. Examples are also given on how to compute the tube wall temperatures, including the effect of scale. Arrangement of headers on flow mal-distribution in superheaters and the effect of tube configuration on direct radiation to tubes are also discussed. Circulation aspects are also discussed along with various configur- . ations available for superheaters, evaporators and economizers. Simulation of single or multi-pressure unfired or fired HRSGs can be performed using the methodology described in chapter 4. Guidelines on selecting pinch and approach points are discussed. One can predict the performance of complex unfired and fired multipressure HRSGs under different load conditions without actually designing the unit. Such studies would be helpful to consultants in simulating the entire combined cycle or cogeneration plant behavior and economics. Methods of improving the efficiency of HRSG systems through addition of condensate heater, deaerator coil or heat exchanger are addressed. Examples illustrate how one can also optimize the temperature profiles b;y rearranging the heating surfaces. Methods of computing deaeration steam requirements and power output from steam turbines are discussed with examples. The software COGEN which is used in HRSG evaluations is recommended to those involved in engineering combined cycle and cogeneration projects, as on can simulate complex HRSG systems without actually designing the plant, saving a lot of engineering time Chapter 5 shows how one should specify waste heat boilers from the process view point. Adhering to the guidelines will save a lot of time for both the boiler designer and the purchasing manager responsible for evaluating alternate bids. Advantages of extended surface over bare tube is discussed in detail with examples in Appendix A. Effect of arrangement of tubes i.e. in-line versus staggered and the selection of fin configuration are elaborated. Examples also show how one can transfer more energy with less surface area with finned tubes. The effect of tube side xx

coefficient and tube and gas side fouling factors on tube wall and fin tip temperatures are illustrated with examples. Appendix B cites the causes of low temperature corrosion and suggests methods of minimizing the problems. Dew points of hydrochloric, sulfuric and hydrobromic acid may be computed using the correlations given. Heat losses thrqugh casing may be evaluated using the program described. One can evaluate an HRSG design for possible noise and vibration due to vortex shedding using the methods discussed in Appendix C. Gas turbine based HRSGs are widely used in cogeneration and combined cycle plants. Appendix D gives the exhaust gas data for several widely used machines. Gas and steam properties are provided in Appendix E with correlations for saturated and superheated steam. Example illustrates how one can compute gas mixture properties. The effect of gas pressure on heat transfer inside and outside tubes is also addressed. Appendix F shows the method of computation of thickness of tubes subject to internal and external pressures according to recent ASME code procedures. In sum, over sixty five examples from real life situations are worked out covering design and off-design performance aspects of various types of waste heat boilers and systemsi in addition, elaborate matter of fact discussions on systems and equipment should make this book indispensable to engineers involved in various disciplines of heat recovery. This book in the authors view would be an invaluable addition to the library of engineers and consultants involved in operation, maintenance, engineering, specifying or purchasing waste heat boilers. Since no single book can cover all of the aspects of the subject, the author suggests that serious professionals involved with heat recovery systems and waste heat boilers should acquire the other books and the software COGEN written by himi for more information, please contact the author at : V. Ganapathy, P.O. Box 673, Abilene, Texas 79604, USA.

xxi

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Chapter 1

Waste Heat Boilers

Waste heat boilers (WHBs) or Heat Recovery Steam Generators (HRSGs) as they are often called are used to recover energy from waste gas streams such as those encountered in sulfuric acid or hydrogen plants, refineries, solid, liquid and gaseous incineration systems, power plants and in cogeneration systems using gas turbines and reciprocating engines. With rising fuel costs and limited supply of premium fuels, it is prudent to maximize the energy recovered from waste gas streams whenever possible. Basically HRSGs can be classified into two broad categories: 1. Those which are required to cool gas streams to a desired temperature range from process considerations; examples could be found in hydrogen or sulfuric acid plants; in these plants, the energy recovery aspect is of secondary importance; the exit gas temperature from the boiler has to be controlled within a narrow range of temperatures for further downstream process purposes and methods such as gas bypassing would be used to achieve this objective. 2. In the other category of waste heat boilers, the objective is to maximize energy recovery compatible with considerations of high or low temperature corrosion and economics. Examples could be found in gas turbine based combined cycle or cogeneration systems, incineration plants and flue gas heat recovery in general. There is no standard design methodology or procedure for engineering of waste heat boilers, since one comes across a wide range of gas temperatures, gas analysis, pressures and steam parameters Table 1-1 shows some of the gas streams encountered in the industry. If the gas pressure is high, a fire tube boiler is preferred, Figure 1-1.

1

IV

Table 1-1 Composition of Typical Waste Gases

Gas 1 2 3 4 5 6 7 8 9 10 11 12 13

Temp.,oC

Pressure, atm

300-1,000 250-500 250-850 200-1,100 300-1,100 500-1,000 200-500 300-1,200 100-600 175-1000 250-1350 150-1000 300-1450

1 1 3-10 1 30-50 25-50 200-450 40-80 1 1 1 1 1.5

N2

NO

78-82 80-82 65-67 8-10 70-72 12-13 13-15 18-20 0.2-0.5 70-80 70-75 75-80 65-72 50-55

H2O

18-20 16-18 40-41 34-36

~

A

8-10 10-12 5-7 2-3

502

503

8-11 0.5-1.0

6-8

C~

CO

CI-4

9-10 6-8 13-15

7-9

0.3

0.2-1

H2S

H2

0-0.8

30-32 38-40 56-60 45-49

1-5 0.3-0.5 6-10 8-12 6-10 16-25 20-25

13-16 5-8 3-5 1-3 3-5

4-6 46-48 3-4 10-13 6-8 4-6 5-7 2-3

0.2-0.5

NH3

Hcl

18-20

traces 5-7

:s~

C;

2-3

3-4

1. Raw suller gases 2. ~ from converter 3. Nitrous gases 4. Primary reformer flue gases 5. Secondary reformer gases 6. Converted gases 7. Synthesis gas 8. Shell gasifier effluent 9. Gas turbine 10. Modular MSW incinerator 11. Chlorinated plastics incinerator 12. Fume incinerator 13. Sulfur condenser

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Figure 1-1. Elevated drum fire tube boiler [courtesy ABea Industries]

If the gas flow is large and several levels of steam pressure are used,

a water tube boiler as shown in Figure 1-2 may be appropriate. More discussions on fire tube versus water tube type could be found in Chapters 2 and 3. Another important aspect to the type of boiler is the cleanliness of the gas stream; if clean, extended surfaces could be used and the boiler may be made compact as in gas turbine applications; on the other hand if the gas stream is dirty as in municipal solid waste systems, the tube surfaces should be bare, with provisions for cleaning and ash removal. A large water cooled membrane wall radiant section may be required to cool the gases below the fusion points of eutectics before entering the convection section. Ample consideration should be given to high and low temperature corrosion aspects.

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5

Waste Heat Boilers

HRSGs FOR GAS TURBINE EXHAUST Gas turbines are widely used prime movers in both the low (less than 10 MW) and in the high end (150 MW) of the power spectrum. They have several advantages such as high efficiency in simple cycle mode, low installed cost per kilowatt compared to fossil or nuclear power plants, quick startup capabilities, smaller space and low cooling water requirements. Exhaust gases from gas turbines are usually clean as they burn premium fuels with high excess air. The energy in the gas stream may be used for several applications such as heating of water or air or heat transfer fluids such as Therminol or glycol and in most cases, for generating steam for process or power using HRSGs. Combined with HRSGs, gas turbines can operate in combined cycle or cogeneration mode, Figure 1-3 and 1-4, thereby improving the efficiency of the overall system compared to the operation of the gas turbine alone. In combined cycle mode a large amount of power can be generated if a fired HRSG is used. Several non-operating nuclear power plants in the USA are being replaced with gas fired gas turbine based combined cycle systems. In cogeneration mode, generally low pressure steam is generated and used for process heating or cooling applications. Fuel

Figure 1-3. Combined cycle plant.

Waste Heat Boiler Deskbook

6

EUCIRICITY

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Figure 1-4. Cogeneration systems. [Pennwe11 Books].

NATURAL VERSUS FORCED CIRCULAnON BOILERS Due to the large mass flow associated with gas turbines, water tube boilers are generally used for heat recovery. Fire tube boilers have been used occasionally with very small machines. HRSGs for gas turbines could be of natural or forced circulation type, Figure 1-5 and 1-6. In natural circulation units, the boiler tubes are vertical and the thermal head differential between water and steam-water mixture is responsible for the circulation through the system. The circulation ratio is arrived at by balancing the system resistance to flow and the available thermal head. It could fall in the range of 6 to 30 depending on the system used. The heat flux inside the tubes, steam pressure and circulation ratio are all important variables in any circulation system and the conditions for

7

Waste Heat Boilers

Departure from Nucleate Boiling (DNB) are set by them. One has to be careful particularly with evaporator tubes with extended surfaces as the heat flux inside the tube can be significantly larger compared to the heat flux while using bare tubes.

+

Figure 1-5. Natural circulation boiler. [courtesy ABea Industries].

In forced circulation units, circulating pumps circulate the steam water mixture through the tubes of the evaporator to and from the drum. Since a pump is used to achieve circulation, one can size it to obtain any circulation ratio. It is typically in the range of 4 to 10. Cost of pump operation and economics plays a role in pump sizing. In Europe, forced circulation designs are common. Advantages claimed are smaller floor space and quick startup capabilities. However, natural circulation designs do not need circulating pumps to

8

Waste Heat Boiler Deskbook

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Waste Heat Boilers

9

maintain the circulation of steam water mixture through the evaporator tubes, thereby saving operating cost and concerns about pump failure or maintenance. The difference in cold startup time periods is not largely due to the simple fact that in the transient heat up phase, the bulk of the time is spent on heating the metal and water of the evaporator module, which is nearly the same whether it is a natural or a forced circulation HRSG. The tube side heat transfer coefficient does not impact the overall heat transfer coefficient, which is dependent on the gas flow and temperature alone. If there are numerous hot restarts, there is a small reduction in the hot restart period due to the circulation of hot water in forced circulation designs, but again, this is not a significant factor to change the design concept. In favor of natural circulation, one can say that the concerns about high heat flux and DNB are less due to the vertical orientation of evaporator tubes, which provide a natural path for the steam bubbles to move; in a forced circulation design, the tubes are horizontal and the heat flux for incidence of DNB is much lower. Separation of steam bubbles from the mixture can occur due to low tube side velocities, resulting in stratification and burnout. This is more of a concern if the gas side flow non-uniformity is high, which can result in variations in heat flux across the width of the HRSG. Fired HRSGs have to be critically evaluated for DNB. In summary both natural and forced circulation boilers are widely used in the industry; while the natural circulation design has an edge over the forced as discussed above, the final choice is usually made by the end user based on his experience.

AUXILIARY FIRING Since the exhaust gases contain a lot of oxygen, in the 14 to 16% by volume range, (if there is no significant injection of steam in the gas turbine, see Table 1-1) additional steam can be generated in the boiler by increasing the exhaust gas temperature to the HRSG through the addition of fuel alone. HRSGs for gas turbine can be classified into three broad categories, depending upon the firing temperature. 1. Unfired 2. Supplementary fired 3. Furnace fired

Waste Heat Boiler Deskbook

10

The term supplementary firing is sometimes referred to as auxiliary firing. When the heat input is significant and a membrane wall furnace is used, then the HRSG is referred to as a furnace fired unit. Note that one may use a membrane wall furnace design even for a low firing temperature situation by choice. Hence these terms are loosely used in the industry and one should be aware of the implications of the firing temperature on the selection and design of the furnace, firing duct and the HRSG. Table 1-2 below gives an idea of the steam generation capabilities of each type. Note that no additional air is needed for combustion for any of the three types as the oxygen in the gas stream is generally adequate. If there is steam injection in the gas turbine, additional air, sometimes referred to as augmenting air, may be required. Table 1-2 Inlet gas temperature and steam generation for gas turbine HRSGs Type of System Unfired Supple. fired Furnace fired

Gas Inlet Temp. of

Gas/Steam Ratio

1000 -1700

5.5 to 7 2.5 to 5.5

1700 - 3200

1.2 to 2.5

800 - 1000

UNFIRED HRSGs When the plant steam requirements are such that the energy in the exhaust gases is adequate, an unfired HRSG is selected. Typical gas temperature entering the HRSG ranges from 800° to 1050°F, depending on the gas turbine used. Figure 1-7a shows a two-gas pass design with a superheater, evaporator and economizer. This design occupies a small floor space. The evaporator tube bundle is divided into two portions by using a baffle plate. The gases flow across the superheater and the bottom half of the evaporator, make a 180 degree turn and then flow across the top half of the evaporator tubes into an economizer. Various gas inlet and exit configurations are possible; for instance the gas can enter at the top and flow downwards or exit at the side. Insulated downcomersare located at the turning section.

~--

~----

--------

11

Waste Heat Boilers

Jl.' BY PASS STAC.K

SILENCER

lf4TERTUBE BOILED.

TJrQ

PASS UNFIRED

H~

Figure 1-7a. A two-pass unfired HRSG with superheater, evaporator and economizer. [courtesy ABeD Industries].

The casing is internally insulated with 4 to 5 inches of mineral fiber or insulation and is protected from the hot gas stream by a stainless steel or corten or carbon steel liner, which is designed to move or expand in the plane of the casing. Lower grade liner material is used in low gas temperature regions. The economizer can be located at the gas exit as shown. Designs that help minimize steaming concerns are discussed later. Figure 1-7c shows a single gas pass design. The stack could be self standing or mounted over the turning section.

------~-~--

~

Waste Heat Boiler Deskbook

12

Figure 1-7b. A two-gas pass HRSG installation. [courtesy ABeD Industries.]

AUXILIARY FIRED HRSGs Supplementary firing of exhaust gases is done to raise the temperature of the gas stream entering the boiler to a maximum of 1700°F, so that additional steam can be generated without major modifications to the unfired boiler design. The 1700°F limit is set by the design of the casing, which consists of several layers of high temperature insulation inside the casing, varying from 6 to 10 inches thick, which is protected from the hot gases by a suitable high grade alloy liner material. The liner should cover as much of the insulation fiber as possible and yet be free to thermally expand in both the directions. Liner materials whatever their grade begin to warp above 1700°F, resulting in the insulation being blown off by the

r

-

----~-~------

-- -

---------~--..,

J Waste Heat Boilers

13

gas during the course of time; hence the limit on supplementary firing temperature. Appendix B can be referred to for calculations of casing temperature and heat loses through multi-layer insulation. The casing design described above is used for the duct work leading to the HRSG from the burner and also in the HRSG portions having a high gas temperature. Carbon steel liners and lesser insulation thiCkness may be used in low temperature portions of the HRSG and economizer. The firing transition is about 10 to 15 feet long, depending on the boiler size and the transition angle between the burner and the HRSG. The HRSG does not differ much in concept from the unfired design, except for the sizing of drum and other heat transfer surfaces and components to handle the higher steam generation. The HRSG steam output can be easily doubled as seen from 'Table 1-2 by the additional fuel input. Figure 1-7d shows a supplementary fired HRSG with a superheater.

OUTLET TBANSlTJON

ECONOMIZER

EVAPORATOR

SCREEN

SUpERHEATER

Figure 1-7c. Single pass unfired HRSG with superheater, screen, SCR, evaporator and economizer. [courtesy ABeD Industries.]

14

Waste Heat Boiler Deskbook

The burner is located between the superheater and evaporator due to design considerations. Due to the high firing temperature, the evaporator has to be designed with varying fin configuration. The first few rows are bare, followed by tubes with increasing fin density. This is done to minimize the tube wall and fin temperatures and heat flux inside the tubes. Methods of evaluating the performance of unfired and fired HRSGs are discussed in Chapters 3 and 4.

Figure 1-7d. HRSG for Cheng cycle application. [courtesy ABeD Industries.]

,..,.--------

- - - -

Waste Heat goiler Deskbook

16

convection section, which as before consists of a combination of bare and finned tubes. The superheater, if used, could be located within the convection bank

SlLEgER

SECTION

.j A- A"

SIDE Vln

Figures 1-9a & 1-9b. Furnace fired HRSG [drawing and photo courtesy

Industries.]

ABea

Waste Heat Boilers

17

Line burners and duct burners typically have a low gas pressure drop, on the order of 0.3 in wc. If steam is injected into the gas turbine either for NO x control or for increasing the power output as in the Cheng cycle system, the oxygen content in. the gas stream reduces. This fact along with high water vapor content in the gas stream forces one to use an augmenting air supply to ensure flame stability in the burner. In furnace fired HRSGs, register burners are used to fire the exhaust gas stream to the maximum possible extent, namely adiabatic combustion temperature, depending on the oxygen content in the exhaust gas. The burner resembles a conventional burner used in fired boilers, Figure 1-10 with windbox and throat for burner elements. The gas pressure drop in the windbox can be as high as 4 to 5 in wc and hence imposes a penalty on gas turbine power output. It may be noted that each 1 in wc reduces the power output of the turbine by nearly 0.25%. One of the advantages of furnace fired designs is that it can even handle solid fuels. The gas turbine acts as a fan supplying hot combustion air to the boiler. The boiler has a fully water cooled radiant section of membrane wall design followed by a convection bank and economizer. The superheater, if used, can be buried within the boiler bank as shown. When the capacity of the HRSG is large, a shop assembled D-type boiler may not be adequate and a field assembled HRSG resemblirtg a fossil fuel fired utility or industrial boiler may be required. The cut off steam flow is approximately 150,000 pph. Another important aspect that should be considered in the design of the firing system is the maldistribution or non-uniformity in the velocity profile across the duct cross section as the gases exit the gas turbine. Also due to the size of the duct work between the gas turbine and the boiler (nearly 30 to 50 feet high and 10 to 15 feet wide) and the manner by which the gases exit from the turbine and enter the HRSG system, additional maldistribution can occur. One of the ways of dealing with this situation is to perform a model study and incorporate turning or guide vanes in the duct work ahead of the HRSG and the burner; in addition, the transition angle of the ductwork connecting the burner and HRSG must be carefully reviewed. If it is too small, it minimizes the maldistribution but adds to the cost; some suppliers use a distribution grid, which

~I 1 I

I

..... 00 S4 ",

12'-0·

PLAN YIn

37'-0·

~

{

j

g

\.

i

1----.:;

'\

U

~

: ~

-JJ i

I

1--

1= P

I_ F-

(f

12'-0·

END rIE![

10'-0·

Ur

~

ir)

UJ ...'"

ll

I

SIDE

YIn

Figure 1-10. D-type boiler with superheater and register burner. [courtesy ABeD Industries]

,

-F-··----------~ .. -

..-.-

J

I I

19

Waste Heat Boilers

consists of a perforated plate with approximately 50% opening. This increases the gas pressure drop by nearly 0.5 to 1.00 in we and minimizes the maldistribution to some extent depending on the % opening. A combination of both vanes and distribution grid may be required in some cases. Model analysis is important and could help prevent serious maldistribution problems particularly in units with burners and superheaters.

FRESH AIR FIRING In all of these firing systems, atmospheric fresh air can be introduced to support combustion in case the gas turbine trips. In refineries and in critical applications where steam production should be maintained in the event of a gas turbine trip, this addition is suggested. Figure 1-11 shows the scheme. A double louvered damper with seal air fan prevents hot gases from going to the fan when it is not in operation and the gas turbine is running. A guillotine or damper prevents air from leaking into the gas turbine duct work when the fan is running and the turbine is off.

4 1

o

-HRSG

l.diverter 2.sllole go. te 3.flring duct 4,burner 5.fo.n 6.do.r'lper w/fo.n Figure 1-11. Scheme for fresh air firing.

.1

20

Waste Heat Boiler Deskbook

It should be noted that the burner duty increases significantly when it is on fresh air as it has to be raised from ambient temperature to

the firing temperature. The sizing concept for the fresh air fan is discussed later, as it affects the system efficiency. One of the concerns in several plants is the time for switching over from gas turbine mode to fresh air firing mode. This could be on the order of a few minutes. It is possible to reduce the time duration between the gas turbine trip and start of fresh air supply by using an induced draft fan behind the HRSG, which operates all the time and handles part of the HRSG resistance when the gas turbine operates and the complete HRSG resistance when on fresh air. The fan curve and the HRSG performance must be studied in depth before selecting the system.

COMPUTING FUEL REQUIREMENTS Engineers have to frequently compute the energy required to raise the temperature of gas turbine exhaust to a desired level; also the concern sometimes is whether there is sufficient oxygen in the gas stream to achieve the objective without additional combustion air. Presented below is an analysis and a chart that may be used to perform quick estimates of oxygen depletion and fuel consumption. Supplementary firing of gas turbine exhaust gases using natural gas or distillates is frequently done to generate additional steam in waste hear boilers. With chart shown here, it is possible to estimate: 1. Energy that must be added in the burner system to raise the temperature of a given exhaust gas quantity Wg from Tl to T2·

2. The quantity of fuel required to do this in scf/hr and in lb/hr. 3. Maximum fuel input that is possible with a given oxygen content of the exhaust gas (which usually varies from 15 to 19%).

Also, the chart may be used in reverse to determine the gas temperature after the burners if fuel input is known.

-------------~---

~~~-~--~-.----------------------

21

Waste Heat Boilers

supplementary fuel parameters c-~.

o.JO I

500 I

MMBTU nr

"J"

100

200

300

I

I

I

1.000

50

5.000 100

,; 10.000 200

250 15.000 :

1.100

1.000 900 T1. 'f.

800

I

I

I

1 iii

J

10 11 1213 U 15 16 17 18

300

0,. \;

Figure 1-12. Chart for supplementary fuel parameters [Oil and Gas Journal.]

NOMENCLATURE A C 0 LHV and HHV

Qs and Qmax

= Amount of combustion air in turbine exhaust, lb/hr. Constant for combustion, depends on type of = fuel,lb/MMBtu. = Oxygen content in exhaust gas, vol%. = lower and higher heating values of fuels, Btu/lb Supplementary fuel input and maximum fuel = input possible, MMBtu/hr

--

-~

Waste Heat Boiler Deskbook

22

= Fuel quantity, lb /hr for oil fuels and scf/hr for natural gas

= Turbine emaust gas quantity, lb/hr = Exhaust gas temperature entering and leaving burner system, OF. = Enthalpy of gas entering and leaving burner,

Btu/lb. Theory. The energy required to raise Wg lb/hr of turbine exhaust from Tl to T2 is given by: (1-1)

Where hl and h2 are the enthalpies of the gas at Tl and T2, respectively. With the chart, there is no need to look for hl and h2; it is adequate if Tl and T2 are known. Wf, the fuel quantity required, is given by: Wf= Qs/LHV

(1-2)

LHV the lower heating value of the fuel, is 18,000 Btu/lb for distillate oils and 1,000 Btu/ scf for natural gas. Hence Wg is in lb /hr for oils and in scf/hr for gas. Also, it is often desirable to know the maximum amount of fuel that can be fired for a given amount of oxygen in the gases, especially when the turbine exhaust is used as combustion air in conventional boilers. This can be found as follows. Let 0 be the vol% of oxygen in the exhaust gas. The total combustion air available in Wg lb/hr of gas may be shown to be: A

=100 x Wg x 0 x 32/(23 x 100 x 29.5)

(1-3)

Molecular weight of exhaust gases was taken as 29.5, that of oxygen, 32. One million Btu of fuel fired requires a nearly constant amount of air. For oils, this is 745lb and for natural gas it is 730 lb. So the amount of fuel that can be fired with "A" lb of air will be: Maximum

= 106Af(C xHHV)

-~------

-

-

--------------._-

23

Waste Heat Boilers

Where C is the constant referred to earlier. Converting to fuel input in MMBtu/hr on an LHV basis, the maximum energy that can be input as fuel is: Qmax

= 106 xA xLHVj(C x

HHV)

(1-4)

Now LHVj(C x HHV) can be shown to be nearly 0.00124 for both oils and natural gas. Using this: Qmax= 58.4 xWg x 0 x 10-6.

(1-5)

Example 1: A waste heat boiler for turbine exhaust gases handles 530,000 lb/hr of gas at 900°F. To obtain additional steam, it is desired to raise the gas temperature through supplementary firing to 1,400°F. Determine: 1. Fuel energy required to raise the exhaust from 900 to 1,400°F. 2. Fuel quantities required to do this in lb/hr (oils) and in scf/hr (natural gas) 3. Maximum fuel input possible if the oxygen content in the exhaust as it leaves the gas turbine is 15% by volume. Solution. Go up from T1 = 900 to cut T2 = 1,400 and move right. Connect with Wg = 530,000 and extend to cut Qs scale at 76 MMBtu/hr. (Use about 80 for sizing the burner system.) If distillate is used, fuel quantity is read off the same scale as 4,300 lb/hr. If natural gas is used, it is 76,000 scf/hr. Connect oxygen = 15 with Wg =530,000 and extend to cut Qmax scale at 465 MMBtu/hr. By working in reverse, T2 can be obtained if Qs is known. The oxygen depletion is obtained by proportion. For 76 MMBtu/h, the oxygen depletion = 76 over 465 x 15 = 10.86%.

BASIC PROGRAM COMPUTES FUEL INPUT, FIRING TEMPERATURE For accurate computations one may use the program presented below. Supplementary fuel firing of turbine exhaust gases is frequently utilized in combined cycle and cogeneration projects (Figure 1-13) to generate additional steam from heat recovery boilers.

,

Waste Heat Boiler Deskbook

24

w,

Ffullg.

T, hg,

(W,

Aifltum/nll ,1xhllUlt

+W,/, T2

h91

Duct burner system Ftiel gal W" LHV

Turbine exhilult g8$ or fresh lIif

Figure 1-13. Burner arrangement showing the mass and energy balance.

Typical turbine exhaust has 14-16% free oxygen before entering the combustion system. Part or all of it may be utilized to increase the final combustion temperature to desired levels. Duct burners typically raise the exhaust gas temperature to approximately 1700°F maximum. In furnace-fired heat recovery boilers the final combustion temperature could be as high as adiabatic combustion temperature, which is the temperature obtained by consuming all of the free oxygen in the exhaust gas. Natural gas is the most widely used fuel. In order to analyze the performance of the burner system and the boiler-and 'to calculate the cycle efficiency-the following data are required: 1. Fuel required to raise the temperature of the incoming turbine exhaust gas to a particular value. Sometimes fresh air is used instead of turbine exhaust gas. 2. Given a particular fuel input, the combustion temperature (reverse of case 1) that results. 3. Fuel data such as density and heating value in Btu/lb and BtQ./scf.

25

Waste Heat Boilers

4. Flue gas analysis, molecular weight. 5. Maximum fuel input possible to consume all the oxygen. With the BASIC program presented here one can perform all the calculations with ease. The program runs on IBM PCs and compatibles. Figure 1-14a. Program for computing fuel input, combustion temperature. [Power Engineering.] 1 CLS:KEY OFF 2 REM:AUTHOR·V.GANAPATHY 5 DIM C(6ol.N(6o·.0/Sol.w/6ol.GASsI161.AI16 .S(60) 10 FOR T=2 TO 34 STEP 2 15 READ C(T)W(TI.N(TI.O(T).SfT':NEXT T 20 DATA 29.6.62.4.34.6.30.9.21.5 75.2.154.4.6:76.5,54.4,124.3.249.7 .13S.4 .123.6.89.5.17S.5.348.5, 188.8.172.1.126.5.231.4.450.4 .242.3.221.8.1651.288.7 ,555.8,296.8.272.61051 25 DATA 347.9 664.5.352.3.324.3246.2.408.6 776.6.408.8.376.8.288,470.5.89Z1.486.3,429.9.330.3, 533.2.1011.2 524.7 .483.4,372.7 .596.3.1133.4 584.1,5371,415,659.5,12591 30 DATA 644.4 591.2.456.9.722.3.1388.5.705.7 645.1,498.1,784.3,15211.767.S .698.9.538.3,8451, 1657.4.831,75:.3,577 .1.904.6.1797 .895.80:2.614.4,962.1 ,19401,959.8.85i .5.649.8 35 FOR 1=1 TO 14 40 READ GASSd':NEXT I 45 0 ATA "methane·" ,"ethane·" ,"propane-", "butane--","isobutane·" ,"pentanf.'· .'·1soptntane·" .flhexane-H~ "hydrogen." ,"carbon monoxide-".' carbon ciio.ide.","nitrogen.","water vapnr.'· 46 DATA "hydro,en sulfide·" 48 CLS:PRINT" FUEL GAS ANALYSIS ·~VOL ":PRINT"" 50 FOR 1=1 TO 14:PRINT I.GASS(l):NEXT I 55 LOCATE 22.5:INPUT"what is the maximum .umber of gas=";MG 60 CLS:PRINT" FUel ANALYSIS''''Yol . :PRINT" ":FOR 1=1 TO MG:PRIr;T I;GASS(l): LOCATE 1-2.20:INPUT AIII:NEXT I 65 LHVF=(AI1!+-g 13.1+A(2)+1641-A(3'·238~AI4)·3113-t-A(5)·3105+A(6"370rAi7)+3715+AIBI *4412-0-AI9'+275+All01·321.S-AI14'*5961 1DO 70 MWF=I 16.04 +AI 11·30.067*Ai2'-A(3 1*44.0h A(41-t-A (Si)*58.2+(A(S)+A (7 j' ·72.14+A(8)*SS.172.01*A(9;+A, 10'·2S~All1)*44-AI12)·28-A 13)*IS+A(14)*34.08)11 OO:D ENSF=.002645*MWF 72 PRINT" .. 75INPUT"EXHAUST GAS FLOW.LB'H.TEMP.• =";WAIR,TAIR 80 PRINT" .. 85INPUT"EXHAUST GAS ANALYSIS·%VOL OF C02,H20,N2,02=";C02A.WAA.N2A,02A:PRINT" " 90 INPUT"BURIiER DUTY·BTU 'H.FINAL TEMP·Flinputzero for the unknDWn·=";O,FTEMP 951F 0=0 THE" O=WAIR+.3·(FTEMP·TAIRi:RO=.4-0 100 LHVFM=LH VF fDENSF :WF=Q.'LHVFM:R Q=.4·Q 105 MWA=I44+C02A.1S+WAA.,.2S*N2A+32+02A).'100 110 CD2AW=C02A ·.44'MWA:WAW=WAA ".1 e"MWA:N2AW=N2A *.2B/MWA:02AW=02A *.32/MWA 120 C02P=WF+: 16.04+ A111·2.7 44-2.927·30.07" A(2)+2.994*44.09*A(3)+5S.1"'A 14\;·A (5) 1·3.029-72.14·IA 16 ,- A, 7))"3.05+S6.17 "A IS )·3.0&!-A 11 0)-1.571 *2S+A( 11 )-44 1".01 'MWF 130 H20P=WF";Z14S+1S.04+A i 1'-1.79S+30.C7"AI2)+1.634*44.09-A(3)+5S.1"' A14)~A(5)l+1.5&72.14+IA 16-A, 7 '+1.49S~86.17"AIS!+1.4&!-8.937+2*A(9)+A(: ~)·'B-34.0~+AI14)*.529'+ .01/MWF 140 02T=13.99+16.04· All ,.,.30.07" Ai2'"3.72:-44.o9* A(3)·3.629+58.1·(A (4'-A, 5.' ·3.579.,-7Z.14"iA . (S)+A 17) '"3.54S-S6.17" A(S '+3.52S-7 .937+2" A191+A 114) *1.409*34.07 6'".01 !MWF :02 R=02T*WF 150 N2P= 18+" 2+wF IMWF :S02 FG=WF* AI14' "I.SS· .3407/MWF :WFMAX=WA IR·O 2AW/0 2T: QM AX= WFMAX+L HVFM ISO C02F G=CD2AW*W AIR+C02P:H20 FG=H20P-WAW*WAIR:02FG=02AW*WAIR·02R :N2F G= N2P+N2AW ..... AI R:W FG=( C02 FG-H20 FG-02 FG+N2F G+S02F G) 170 TEMP=TAI R:CO=C02AW:WA=WAW:NI=1i2AW:O X=02AW 180 GOSUS 1050 190 HGAIR=HG 195 HFG=(Q-HGAIR*WAIR)/(WF-WAIR)

Waste Heat Boiler Deskbook

26

Figure 1-14a (cont'd) 200 CO=C02F G'YiFG :WA=H20 FG 'WFG:O X=02F GIWFG :NI=N2FGIWFG :SO=S02F GIWFG 210 TEMP=HFG .3:RA=.5+TEMP 220 GOSUB 1050 230 IF ABSIIHG·HFG)!HFG! 2.999999E·03 THEN 25C 240 RA=.5+RA:TEMP=TEMP·SGN,H G·HFG '·Rt. 245 GOTO 220 250 IF FTEMP=O THEN 265 2551F ABSIIFTEMP·TEMPJ.iEMPI 3.000001 E.jJ3 THEN 265 260 RQ=.5+RQ:Q=Q+SGNIFTEMP·TEMP)+RQ:GOTO 100 265 SUM=(CO 14/,-1'1 AlI8+0 XI32~NI:2B+SO/6~ '+1 00:C02V=CO/SUM/.0044:H20 V=WA/SUM/.0018: o XYV=O X.'SUM/.0032:N IV=N I 'SUM/.002E :S02V=SOISUM/.0064 270 MWFG=IC02V+44+H20 V'18-0 XYV*32~N IV*28+S02V*64)*.0 1 275 CLS:PRINT" RESULTS OF COMBUSTION CALCULATIONS":PRINT" .. 280 PRINT"F UEL 0 ENSITY·L B'SC F=";O ENSF~' LHV ·BTUISC F=";L HVF;"LHV-BTU/L B=";LHVFM 285 PRINT" " 290 PRINT"AIR TEMP IN=";TAIR:"FLOW·LB H=";WAIR;"MOL WT=";MWA;"\;VOL·C02=";C02A; "H20=";W AA ;"0 2=" :02A :"N2=":N2A 295 PRINT" " 300 PRINT"FLUE GAS FLOW·LB.'H=";WFG:"WOL·C02=";C02V;"H20=";H20V;"02=";OXYV; "N2=";NIV;"S02=";S02V;"MOL WT=";MWFG 302 PRINT" " 305 PRINT"FINAL TEMP=";TEMP~'DUTY-MM BTU/H=";Q*10··6;"MAX FUL IN·BTU/H·";DMAX: PRINT" " 315 PRINT"FUEL GAS ANAL YSIS·\1VDL":PRlliT" " 316 FOR 1=1 TO 7:PRINT I;GASS(l);A(I) 320 NEXT I 321 FOR 1=8 TO 14:LOCATE hl,24:PRINT I:GASS(l);A(I) 325 NEXT I 330 LOCATE 22.18:END 1050 IF TEMP.200 THEN HG=(TEMP·6OJ+.24:GOTO 1100 1055 IF TEMP 3400 GOTO 1095 1060 X=.01*TEMP:M=2*INTI.S+X! 1070 I=M:GOSUB 2000:HG1=HG lOBO I=M+2:GOSUB 2000:HG2=HG 1090 HG=.S·(X-M'*(HG2·HG1)-HG1:GDTO l1De 1095 HG=CO' (962.1+ .27*(TEMP·3400) )+WA * .1940.2+.742*(T EMP·3400))+NI+(959.B+.33*(TEMP. 3400) )+0 X*(B57.~ .256+(T EMP·3400) )+SO *(649.B+ .16*(TEMp·3400») 1100 RETURN 2000 HG=O (I )*0 X-N (I)'N I"'WII' *W A"C(lI'CO-S( I)'SO 2010 RETURN

Theory: The enthalpy of flue gas after combustion is obtained from! (1-6)

where: WI hgI

WI Q LHV hg2

=flow rate of turbine exhaust gas or fresh air to burner, lb/h =enthalpy of incoming air or exhaust gas, Btu/lb = fuel input, lb /h = Q/LHV

=fuel input to burner system, Btu/h (LHV basis) = lower heating value of fuel in Btu/lb = enthalpy of final products of combustion, Btu/lb

Waste Heat Boilers

27

hgl and hg2 are functions of the air/flue gas analysis and temperature. Based on combustion constants (Ref 1), the oxygen required, nitrogen, carbon dioxide, sulfur dioxide, and water vapor produced are computed and the flue gas analysis for any given fuel input is obtained. A subroutine calculates the enthalpy. Also, the heating value of the fuel is computed based on the analysis. Up to 14 constituents can be handled by the program, as seen in Figure 1-14b. Figure 1-14b. Inputs and results for Example FUEL ANAL YSIS·%vol 1 methane· 2 ethane· 3 propane· 4 butane· 5 isobutane· 6 pentane· 7 isopentane·

? 95 8 hexane· ? 2.5 9 hydrogen. ?0 10 carbon monoxidec ?0 11 carbon dioxide· ?0 12 nitrogen· ?0 13 weter vapor· ?0 14 hydrogen sulfide· EXHAUST GAS FLOW·L3IH,TEMP·F=? 135000,900

?0 ?0 ?0 ?0 ? 1.5 ?0 ?1

EXHAUST GAS AN.ALYSIS·%VOL OF C02,H20,N2,02=? 4,5.3,75.5,15.2 BURNER DUTY·BTU/H,FINAL TEMP·FHnput zero for the unknown}=? 0,1600 RESULTS OF COMBUSTION CALCULATIONS FUEL DENSITY·LB/SCF= 4.430501E·02 LHV·BTU/SCF= 914.43 LHV·BTU/LB= 20639.43 AIR TEMP IN= 900 FLO'IV·LB/H= 135000 MOLWT= 28.718 %VOL·C02= 4 H20= 5.3 02= 151 N2=75.5 FLUE GAS FLOW·LB/H= 136339.1 %VO L·C02= 5.647 H20= 8.615786 02= 11.50723 N2= 74.21283 S02= 1.716053E·02 MOL 'NT= 28.50841 FINAL TEMP= 1599.112 OUTY·MM BTUiH= 28,35 MAX FUL IN·BTU/H= 1133603E;-Qa FUEL GAS ANAL YSIS·WOL 1 methane· 95 2 ethane· 2.5 3 propane· 0 4 butane- 0 5 isobutane· 0 6 pentane· 0 7 isopentane. 0

8 hexane- 0 9 hydrogen. 0 10 carbon monoxide· 0 11 carbon dioxide- 0 12 nitrogen· 1.5 13 water vapor· Q 14 hydrogen sulfide· 1

If the fuel input to the burner in Btu/h is known, the fuel quantity is obtained and the flue gas analysis and quantity are computed. The final enthalpy is obtained and, through an iterative process, the corresponding temperature is found.

Waste Heat Boiler Deskbook

28

If the final desired temperature is given, the fuel input is assumed and iterated to yield the final enthalpy, after computing the flue gas analysis. Quick converging techniques are used. The program also estimates the fuel input to consume all of the oxygen in the incoming air or turbine exhaust gases.

CalCulation Examples Two examples illustrate the versatility of the program. Example 2: 135,000 lb/h of turbine exhaust gas at 900°F, having the analysis C02 = 4%, H20 = 5.3%, N2 = 75.5%, and 02 = 15.2% all by volume, enters duct burner and has to be raised to 1600°F using natural gas having the following analysis: methane = 95%, ethane = 2.5%, nitrogen =1.5% and hydrogen sulfide =1.0%, all by volume. Determine: 1. Fuel heating value in Btu/lb and in Btu/ scl. 2. Fuel quantity required in lb/h and in Btu/h. 3. Flue gas analysis and molecular weight. 4. Fuel input to consume all of the free oxygen. Step-by-step instructions for running the program: Key in the program from steps 1 to 2010 as shown in Figure 1-14a. 2. Key the letters R,U, N (RUN mode) and press the "enter" key. 3. You will see the 14 gases printed on the screen, one below the other, starting from methane and ending at hydrogen sulfide (see the output in Figure 1-14b for the various gases, 14 in all). 4. Then the question is asked at the bottom of the screen, "What is the maximum number of gas?" By this is meant the number of the constituent in the fuel gas that has the maximum number (1 to 14). If for instance, we have methane, propane, and nitrogen in the fuel, nitrogen has the number 12 (see the list of gases in the results, Figure 1-14b) and is the largest. Hence one would enter 12 in this case. 5. Once this number is entered, the list of gases up to the maximum number appears on the screen, one by one. Enter the volume % of each as shown in Figure 1-14b. If a particular gas is not present, enter zero for its volume. 1.

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29

6. Once the gas data are fed in, the next question asked is "Exhaust gas flow and temperature." Input the gas flow and the tem~ perature, separated by a comma, as seen in Figure 1-14b. 7. Next the gas/air analysis is fed in. Volume % of C02, H20, N2, 02, are inputted, each separated by a comma, as seen in Figure 114b. 8. The next set of data to be fed in are: "duty, final temperature."

Now, input zero for the unknown, and the known value, each separated by a comma, as shown in Figure 1-14b. If the final temperature is 1600 and the duty is to be calculated, input 0,1600. If the duty is, say, 2,000,000 Btu/h and final temperature is to be found, input 2000000,0. This ends the data inputting. The rS!sults appear as shown in Figure 1-14b. Gas analysis and other data can be seen. The fuel gas and flue gas analysis, final temperature, maximum possible fuel input, and heating values of fuel are all printed out. The fuel input is 28.35 MMBtu/h and maximum possible fuel input is 123.3 MMBtu/h. Flue gas quantity is 136,339Ib/h and 11.5% oxygen is still available in the exhaust gas. Lower heating value in Btu/lb is 20,640 Btu/lb.

Example 3: 100,000 lb/h of fresh air at 80°F enters a burner and 100 MMBtu/h of fuel is inputted. Determine the final temperature. Fuel has 95% methane, 3% ethane and 2% propane, all by volume.

Solution: A procedure similar to Example 2 is followed. The analysis used for air is 79% nitrogen and 21 % oxygen by volume. It is seen from the printout that the final temperature is 3150°F and maximum possible fuel input is 126.1 MMBtu/h. See Figure 1-14c. Analysis of the flue gas and other data may also be obtained from the printout.

Waste Heat Boiler Deskbook

30

Figure 1-14c. Input and results for Example FUEL ANALYSIS·%vol

1 methane· 2 ethane· 3 propane-

? 95 ? 3' ?2

EXHAUST GAS FLOW·L9iH,TEMP·F=? 100000,80 EXHAUST GAS ANALYSIS·%VOL OF C02.H20,N2,02=? 0,0,79,21 BURNER DUTY·BTU,H FINAL TEMP·Ftinputzero for the unknown)=? 100000000.0 RESULTS OF COMBUSiION CALCULATIONS FUEL DENSITY·LB/SCF= H02269E·02 LHV·BTU/SCF= 964.375 LHV·BTU/L8= 21419.76 AIR TEMP IN= 80 FLOW·LB/H= 100000 MOL WT= 28.84 %YOL·C02=

°H20= 002= 21 N2= 79

FLUE GAS FLOW-LB,H= 104668.4 %YOL·C02= 7.824417 H20= 15.1437 02= 4.019827 N2= 73.01206 S02= 0 MOL WT= 27.39833 FINAL TEMP= 3149.942 !JUTY·MM BTU/H= 100 MAX FUL IN·BTU!H= 1.261224E-'l8 FUEL GAS ANAL YSIS·WOL 1 methane· 95 2 ethane- 3 3 propane· 2 4 butane- 0 5 isobutane- 0 6 pentane· 0 7 isopentane· 0

8 hexane- 0 9 hydrogen· 0 10 carbon monoxide- 0 11 carbon dioxide· 0 12 nitrogen' 0 13 water vapor· 0 14 hydrogen sulfide- 0

SUPPLEMENTARY FIRING AND SYSTEM EFFICIENCY The efficiency of the HRSG system improves with firing. This can be seen from the example given below. The reason is that with the same oxygen content entering the burner, more fuel is being fired thus reducing the excess air leaving the stack; also, with an increase in inlet gas temperature the exit gas temperature from a HRSG with an economizer usually decreases. This is due to the significantly larger ratio of water to gas flow in the fired mode compared to the ratio in the unfired mode. The gas flow remains nearly constant, while the steam production and the water flowing through the economizer increases, depending on the extent of firing. This fact is partly responsible for the improvement in efficiency. More information on HRSG heat balances and temperature profiles and the use of the software COGEN can be seen in Chapter 4.

....

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Example 4: A supplementary fired boiler generating 200 psig steam operates as shown in Table 1-3. (the data have been obtained from CQGEN software described in Chapter 4.) Determine the system efficiency using ASME PTC 4.4.

1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11.

Table 1-3: Data for Supplementary Fired Boiler case 2 3 1 gas flow, pph 150,000 150,000 150,000 inlet gas temp, OF 900 900 900 firing temperature 900 1290 1715 burner duty, LHV MMBtu/h 0 37.60 17.30 steam flow, pph 22,780 40,000 60,000 steam pressure, psig 200 200 200 feed water temp, OF 240 240 240 exit gas temp, OF 327 315 310 steam duty, MMBtu/h 22.67 39.90 59.90 system efficiency % 68:7 84.90 79.2

Note: gas analysis in :% vol C02 = 3, H20 = 7, N2 = 75, 02 = 15 blow down = 3% Solution: Using COGEN program and the methodology discussed in Chapter 4, the design and off-design performance at various loads were obtained; the results are shown in Table 1-3. Gas enthalpy at 900 F =220 Btu/lb (from Appendix E). Efficiency = output/input =39900000/(17300000 + 150000 x 220) = 79.20 % for case 2. ' As seen from the above table, the efficiency increases with additional fuel input. The additional fuel input generates nearly 90 to 95% of its energy equivalent as steam.

FRESH AIR FIRING AND EFFICIENCY Fresh air firing is often used in gas turbine plants in case the gas turbine fails or is shut down and steam should still be generated. How does one select the fresh air fan capacity? The efficiency of the HRSG system increases with a decrease in stack gas temperature and mass flow, as will be shown in Chapter 4.

,

.,

Waste He.at Boiler Deskbook

32

Hence an obvious method of improving the system efficiency is to select as Iowa mass flow as possible for the fan compatible with firing temperature considerations. That is if the capacity of the HRSG can be maintained at a lower fresh air fan flow without exceeding the firing temperature that may call for a different type of HRSG (membrane wall versus insulated casing) then that is the best choice. In order to illustrate this, let us see how the HRSG in Example 4 performs with fresh air fan while generating 40,000 pph of steam. Example 5: There are two fan selections-one with 130,000 pph capacity and another with 150,000 pph. Determine the system efficiency, when 40,000 pph at 200 psig is generated, feed water temperature is same as before. Solution: Using COGEN the results are obtained for both the cases, see Chapter 4.

1. 2. 3. 4. 5. 6. 7.

Table 1-4: Fresh Air Firing Performance airflow, pph 130,000 150,000 inlet temp, OF 60 60 firing temp, OF 1424 1294 exit gas temp, OF 308 314 steam flow, pph 40,000 40,000 burner duty, MMBtu/h 48.72 50.32 efficiency, % 81.8 79.25

The following points may be seen: 1. As the capacity of the fan reduces, the efficiency improves as the heat loss from the system is lower. 2. The stack gas temperature is also reduced due to the higher firing temperature with the lower air flow, adding to the improvement in efficiency. 3. The cost of the fan will be lower as the capacity and hence the gas pressure drop are lower. 4. The firing temperature is higher but if the design can handle it without a major change, lower fresh air flow is preferred. 5. The burner duty is naturally more as the air is raised from the ambient temperature to the firing temperature, compared to gas turbine operation.

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HRSG DESIGN FEATURES Finned Surfaces and Design

Finned surfaces (solid and serrated) are extensively used in HRSGs for gas turbine applications. Their use makes the HRSG compact and weigh less; the gas pressure drop, which is an operating penalty, is also reduced with finned surface design. Appendix A discusses the method of calculation with finned tubes and also compares a HRSG design with and without finned tubes. When natural gas is fired in the gas turbine or burner, the clean gases permit one to use as high a fin density as 5 or even 6 fins/in for evaporators or economizers. Tubes may be arranged in line or staggered fashion. With distillate fuels, the fin density should be lower say 4.5 or 5 fins/in. Fin height ranges from .5 to .12 in., depending on fin tip temperature calculations. Higher the fin density or ratio of external to tube internal area, higher will be the heat flux inside the tubes, tube wall and fin tip temperatures and the gas pressure drop. Superheaters on the other hand have a low tube side heat transfer coefficient and hence it is prudent not to use a high fin density as discussed in Appendix A; if so done, the tube wall and fin tip temperature will be higher; the surface area will appear to be high with a high fin density, but due to the low overall heat transfer coefficient. the duty transferred could be the same or even less. Hence one has to be careful while evaluating the performance of finned surfaces and not go by surface area alone, as it is misleading. The product of surface area and overall heat transfer coefficient determines the duty and not the surface area alone. More information on finned tubes and their optimization can be found in Appendix A. When the gas turbine burns a dirty fuel such as heavy oil, the fin density should be lower, say not more than 2 or 3, depending on the presence of slagging constituents in the fuel ash; more clearance should be allowed between tubes; inline arrangement is preferred as soot blowers could be more effective. If the fuel oil is extremely dirty, a bare tube deSign, though expensive, may be the only choice. With fired HRSGs the gas temperature along the gas flow path will change. Hence the evaporator should have combinations of bare and finned tubes of varying fin densities so as to minimize the heat flux and tube wall and fin tip temperatures. Usually the tubes

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Waste Heat Boiler Deskbook

at the front end would have a low fin density followed by high fin density section. The thermal design procedure should account for the different fin combinations. If a superheater is used, it is preferable to bury it within the convection section when the firing temperature is very high. This helps lower the metal temperatures and hence increase its life. Methods of minimizing superheater metal temperatures are discussed in Chapter 3. Usually HRSGs for small gas turbines come in packaged designs, while for large machines (gas flow above 250,000 pph) each surface such as superheater, evaporator or economizer is built in a separate module for shipping purposes.

STEAMING ECONOMIZERS As will be seen in Chapter 4, one of the problems associated with low load or low ambient operation of a gas turbine is steaming in the economizer. Formation of steam in the steam water mixture obstructs the flow of the mixture when it flows in the downward direction and may result in vibration or water hammer concerns. It is preferable to have the tubes which are steaming to be oriented in such a way as to aid the flow of bubbles, namely upwards. Figure 115 shows a horizontal tube economizer design that has multi-streams in which the water in the last section flows from the bottom to the top header. Steam bubbles if any should be easily removed with this configuration. The economizer could also be designed as a two gas pass unit as shown in Chapter 3 with the water in the exit section having an upward flow. The other options are: 1. Reverse the flow direction of water using valves, which is cumbersome; in non-steaming mode the economizer operates in counter flow configuration, while in steaming mode, it operates in parallel flow configuration. 2. The exhaust gas may be by-passed around the economizer to decrease its duty and thus prevent its steaming. This is a loss of energy. 3. Some boilers are designed so that the gas flow to the boiler itself is bypassed during steaming conditions; this is not recommended

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35

as it results in a significant loss of energy by virtue of the evaporator not handling the entire gas stream, , 4, Bypass a portion of the economizer surface on the water side so that the surface area participating in heat transfer is reduced and hence the duty or enthalpy rise decreases thus avoiding steaming, The author prefers the design shown in Figure 1-15 as it does not result in loss of energy in any mode of operation,

rh

I I

I I

I I

IIr I

I

I

iii

I

BY OTBRI§'i i i ' I

I

I

I I

I

t

I I

I

I

I

I

I

I

J

J!'~,...---.-.-.-.-.~.-I-l-·-·-·-·-·-·

+-I I

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.

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....

g

.".

4'-0·

6'-6"

to'-O·

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-

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nu

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~

lVdXERTllBE BQ1LrB

S'-O·

6'-0·

to'-O·

2'-6·

Figure 1-15. Arrangement of HRSG with horizontal gas flow economizer. [courtesy ABeD Industries,]

Emissions of NOx and CO Firing of fuel in the turbine and auxiliary burner introduces pollutants such as NO x and CO into the exhaust gas stream, Stringent pollution regulations dictate the amount of NO x and CO that can leave the stack of present day boilers, Pollutants are often

36

Waste Heat Boiler Deskbook

specified in ppmvd (parts per million volume dry) with reference to 15% oxygen in the gas for emission monitoring and control purposes, while burner suppliers often suggest their contributions in mass flow units. The procedure outlined below may be used by engineers for relating the mass flow rate of NO x and CO to ppmvd and vice versa. Emission-Data Conversions Monitoring combustion emission products-NO x and CO in particular-is increasingly important, due to tightening regulations. Gas turbine cogeneration or combined-cycle plants, for example, produce both of these emissions, and regulations pertaining to them are very strict. Because of this situation, it is desirable to have a quick and easy way of converting emission data from mass flowrates (lb/h) to ppm(v), dry basis, which is the usual way that they are reported to regulatory bodies. An oxygen concentration of 15% is the usual reference in industry. With the enclosed chart it is possible to accomplish this purpose. The chart provides a conversion factor, P, which allows both the lb/h-to-ppm(v) conversion and the reverse. Nomenclature V = volume or volumetric ratio; Vn, volumetric ratio of NO x (ppm); Vc, volumetric ratio of CO (ppm) w = flowrate of a constituent gas in the exhaust stream, lb/h W = total exhaust stream flowrate, lb/h = conversion factor = molecular weight of exhaust stream, lbs = percent volume of water in the wet exhaust stream = percent volume of oxygen in the exhaust stream Underlying Theory If w, in units of lb/h, is the flowrate of NO x (usually reported as N02) in a turbine exhaust stream of W lb/h flowrate, the

following equation gives the volumetric ratio on a dry basis. (The molecular weight of N02 is 46.)

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Waste Heat Boilers

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37

v =[(w/46)/(W/MW)] X 100/(100 -%H20)

(1-7)

In this equation, %H20 is the volume of water vapor in the exhaust stream, and MW is the molecular weight of the gases. V n must further be corrected for the 15% oxygen, dry basis, as follows: " """'()] Vn [mppUl v -_

6

10 V (21-15) _ 21- [100/(100- %H20] x %02

(1-8)

In this case, %02 is the percent volumec>f oxygen in the wet exhaust gas. Similarly, the CO emissiot\ rate in ppm(v); 15% 02, dry basis, is: (1-9) Vc [in ppm(v)] =1.642Vn Figure 1-16 solves for the concentrations of both NO x and CO; the conversion factor, P, is obtained from it as a function of %H20 and %02. A molecular weight of 28.2 was used for the construction of the chart. Then, Vn [in ppm(v)] =(w n/w) x 106 xF Vc [in ppm(v)] =1.642 x (wc/W) x 106 x F

(1-10)

Sample Problem Determine the NO x and CO concentrations, in ppm(v), 15%, dry basis, if 251b/h of NO x and 151b/h of CO are present in 550,000 lb/h of turbine exhaust gas. The percent volume of H20 is 10, and the percent volume of 02 is 11 in the wet exhaust gas. Solution: From the figure, Fat 10% H20 and 11 % 02 is 0.46. Then,V n equals (25/550,000) x 106 x 0.46 = 21. Also, V c equals 1.642 x (15/550,000) x 106 x 0.46 = 20.6. Methods of Reducing Pollutants Gas turbine combustor modifications such as staged combustion, premix burning and proper distribution of air and fuel often referred to as dry NO x control methods are being done to reduce NO x and CO. Siemens, a European supplier of gas turbines is introducing a

38

Waste Heat Boiler Deskbook Figure 1-16. Chart for converting NOx , CO emissions. [Chemical Engineering.] EMISSION-DATA CONVERSION CHART

0.35

0.40

0.45

j.: 0.50 c 0.55 o

'j!! CII

> c

0.60

8 0.6& 0.70 0.75 0.80 0.85 0.90 0.95 1.00

1.25

32

26

24

20

16

12 10 8

6 4

The Intersections of the vertical and diagonal lines yield conversion factors for calculating the volumetric ratios (In ppm) of CO and NOx

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machine guaranteeing 9 ppmvd NO x with no steam or water injection with CO at less than 8 ppmvd. Some other suppliers have come out with machines with emissions less than 25 ppmvd NO x . However the popular method is steam or water injection in the gas turbine. Adding steam or water reduces the flame temperature and suppresses the formation of NOx by up to 70%. However water injection can increase the CO and UHC (unburned hydrocarbons) especially at low loads, increase the heat rate nearly 5% to make up for the heat of vaporization lost to stack and possibly shorten the life of the turbine components. Steam' injection does not exact as high a heat rate penalty as water injection but the quantity of steam injected is nearly 50% more. However the electrical power output increases as the mass flow and the specific heat of the exhaust gas increases. These methods typically reduce NO x to 40 to 50 ppmvd range. Note that the addition of water or steam affects the exhaust gas analysis and hence the duty of the HRSG. Gas turbine suppliers can furnish more information on the characteristics of their machines with steam or water injection. The water used for injection should be of high quality, preferably demineralized. If steam is used, it should have a good purity and as low total dissolved solids as possible. In the Cheng cycle system discussed later, a combination of drum internals plus an external steam separator was used to achieve the desired steam purity of less than 50 parts per billion of solids. If the steam or water is not of good purity, the solids can deposit on the turbine blades resulting in poor performance and even turbine failure. With NO x requirements in the range of 6 to 10 ppmvd in several locations, SCRs (Selective Catalytic Reduction Systems) are seen as the only proven solution to meeting these levels of pollutants, even though they are expensive. Figure 1-17 shows an arrangement of a HRSG with SCR for NO x reduction. The base metal catalyst (vanadium) is sandwiched between modules in an operating temperature regime of 600 to 750°F (for some catalysts). This may require splitting up of the evaporator module so as to obtain the window of temperature at different loads and gas flow conditions. An example to compute the gas temperature at the SCR at different gas inlet conditions is given in Chapter 3. Ammonia is injected upstream of the catalyst and is mixed with the gas stream

Waste Heat Boiler Deskbook

40

-,

9

5

4

1

9

5 I.HRSG cOIl

2.0.11"

6.duci: burner

3.0.1')1'Ion10. to.nk

7.o.I"IMOnlo. grid

4.controller

5.90.S turbine

8. SCR or< CO co.to.lyst

9.sto.ck

1. HRSG coil 2. air 3. ammonia tank. 4. controller 5. gas turbine 6. duct burner 7. ammonia grid 8. SCR or CO catalyst 9. stack Figure 17-a & b. Arrangement of catalysts for NOx and CO reduction.

before it reaches the catalyst, where the NO x is converted to N2 and H20. The excess of ammonia called ammonia slip should be minimized by proper controls. Base metal catalysts are suitable for higher temperature operation in exhaust gas streams containing higher levels of 50 x' In oil fired HRSGs the excess ammonia can react with 503 to form ammonium sulfate which can deposit on surfaces at low temperatures and cause fouling and corrosion. One has to check the HRSG performance at different loads and ambient

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41

conditions to ensure that the temperature window for proper operation of the catalyst is maintained. Typically catalysts have 80 to 90% removal efficiency. Catalyst for CO removal could be located in a higher temperature zone, even ahead of the superheater. If there is a burner, it is desirable to shield the catalyst from the radiation by locating it downstream of a screen section or superheater. One supplier offers a precious metal catalyst located in a low temperature zone namely 430 to 530°F for removal of both NO x and CO. The advantage is that the catalyst can be located beyond the evaporator and ahead of the economizer and splitting up of the evaporator is avoided. The gas pressure drop can also be lower due to the use of a single catalyst versus two. It is also claimed that the excess ammonia or slip reacts directly with the precious metal catalyst to form nitrogen, water and nitrous oxide, which is not a pollutant; however it is a greenhouse gas. An additional gas pressure drop df 2 to 3 in we due to the SCR has to be considered in the overall evaluation of the system performance as it affects the gas turbine power output. As mentioned elsewhere, every 4 in we reduces the gas turbine output by nearly 1%. With machines reaching the 150 MW range, the loss in power cannot be ignored.

BYPASS DAMPERS One of the major auxiliaries used in gas turbine plants is the bypass damper, Figure 1-18. The objective of using this is to be able to isolate the boiler from the gas turbine and also modulate the gas flow during startup of the boiler so that the steam pressure in the boiler can be raised slowly. However there are several large gas turbine HRSGs which are connected directly to the turbine. If seal air fans are not used, there is likely to be leakage from the gas side to the atmosphere, resulting in loss of energy; this could be 0.5 to 1.0% of the energy in the gas stream, which is not insignificant. Sealing efficiency of a damper could be specified in two ways and it is important to note the difference. One way is specify on the basis of the area. If the leakage efficiency is 99.5% on this basis, it

Waste Heat Boiler Deskbook

42

means that .5% of the area is not sealed and gases can leak through this area.

Damper arrangement Bypass stack

Gas turbine

Waste heat boiler

Figure 1-18. Bypass damper scheme for gas turbine HRSGs.

The amount of gas that can leak is a function of the differential across the damper, the area of damper and the gas density. If the leakage flow is say 1000 pph when the total flow is 100,000 pph, the. sealing efficiency on flow basis is: 100 - (1000/100000) x 100 =99%. A simplified approach to estimating the leakage across dampers is given below. Considering the flow situation similar to that of an orifice we can obtain an expression for the gas velocity for the leakage flow as follows: Let us derive an expression for the leakage of gas across a damper, stating the assumptions made. Most of the dampers used for isolation of gas or air in ducts are not 100% leakproof. They have a certain percentage of leakage area, which causes there to be a flow of gas across the area. Considering the conditions to be similar to those of flow across an orifice, we have

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(1-11) where Vg Hw

Pg'Pw g Cd

= =

=

=

=

gas velocity through the leakage area, fps differential pressure across the damper, in. WC density of gas and water; lbI cu ft acceleration due to gravity, ftl sec 2 coefficient of discharge, 0.61

The gas flow Win lblh may be obtained from V W=3600pg A(100-E) - g

100

(1-12)

Where E is the sealing efficiency on an area basis (%). Most dampers have an E value of 95 to 99%. This figure is to be provided by damper manufacturers. A is the duct cross section, ft2. Substituting for Cd = 0.61 for Pg =40/(460 + t) and simplifying, we have W = 2484A(100 - E)

~

Hw

460 + t

(1-13)

where t is the gas or air temperature, °P. Example 6: A damper used in a HRSG system is claimed to have a sealing efficiency of 99%. If the gas flowing in is 120,000 pph at 10000P and the differential between the gas and atmosphere at the damper is 7 in we, determine the leakage flow and the loss in energy. Assume that the damper area is 10 sq ft. Solution: Using Equation (1-13) W = 2484 x 10 x (100 - 99) x '17/1460 = 1720 pph The sealing efficiency on flow basis = 100 - (1720/120000) x 100 =98.56%.

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Waste Heat Boiler Deskbook

The loss in energy with reference to 60°F = 1720 x 260 = .45 MMBtu/h.

RECENT TRENDS AND THEIR IMPACT ON HRSG DESIGNS With improvements in firing temperature in the gas turbine to 2400-2500°F range and higher pressure ratio, the exhaust gas temperature is likely to be higher, on the order of 1050 to 1150°F. This may necessitate a larger HRSG with additional heat recovery surfaces and possible reheaters, as in conventional utility boilers, which can improve the combined cycle efficiency by 2 to 3%. Gas turbines are also being used with solid fuels such as coal and wood. The low Btu gas from the gasifier is fired in the gas turbine as shown in Figure 1-19 and 1-20 after being cleaned in a hot gas cleanup system. Pressurised fluidised bed combustion/ gasification system may also be coupled to gas turbine combined cycle systems, Figure 1-21. In another concept, the hot coal combustion products transfer energy to air using a ceramic heat exchanger. The hot air in turn drives the gas turbine. The ceramic exchanger unlike conventional metal exchangers, can withstand 2000 degree C. The hot air can be raised to 1200 degree C, typical inlet temperature of a high efficiency gas turbine. The exhaust gases transfer energy to steam via a HRSG. The system is expected to be 50% more efficient than today's conventional steam turbine plants. Landfill gas from refuse sites is being widely used in cogeneration projects; reciprocating engines as well as gas turbines use this fuel, which typically has 40-50% methane, 35-45% carbon dioxide, some nitrogen and water vapor. Due to the lower heating value, their combustion temperature is lower and hence generate lower NO x levels. Sulfur compounds and chlorinated/fluorinated hydrocarbons present in the gas can cause low temperature corrosion in parts exposed to the gas. Figure 1-22 shows a typical system for collecting and burning LPG gas in an engine.

. ..

--~ ~

-- - - - - - - - - - 1

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Waste Heat Boilers

45

Coal

-..

INTEGRATED GASIFICATION/COMBINED CYCLE Stack

Second-stage combustor 4111

..

First-stage combustor

t

Air~l~ d~Compressor

Coal

DIRECT COAL-FUELED TURBINE

Coal Sorbent

ADVANCED PRESSURIZED FLUIDIZED-BED COMBUSTION

Figure 1-19, 1-20, & 1-21. Coal based combined cycle systems [Power]

Typicallandlill-gas-to-energy project consists of well-fu.eled gas collection and treatment system, and engine/generator

Figure 1-22. Scheme for collecting landfill gas for engines.

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--1

I

47

Waste Heat Boilers

HRSGS FOR STIG AND CHENG CYCLE SYSTEMS Steam injected gas turbines (STIG) are widely used to improve the power production capabilities of gas turbine systems. This requires modifications to the gas turbine combustion chamber. Unlike steam injection for NO x reduction, STIG systems inject a significant amount of steam into the gas turbine. This alters the mass flow and gas analysis significantly. One of the popular systems which is used to balance the thermal load and electrical power is the Cheng cycle, Figure 1-23. In this cogeneration system, the saturated steam that is generated is consumed for process, while low pressure superheated steam is injected into the gas turbine to increase the power output; variations in process steam requirements are handled by varying the fuel input to the duct burner located between the superheater and evaporator. This system offers more advantages compared to a combined cycle system. rll*l---+ PROCESS STEAM

INJECTION STEAM

/

GAS TURBINE

SUPERHEATER

DUCT BURNER

EVAPORATOR

Figure 1-23a. Cheng cycle heat recovery system.

1. Equal or slightly more electrical power output compared to a combined cycle system with the same gas turbine. 2. Absence of major auxiliaries such as steam turbine and condenser, which results in less complexity and cost. 3. Very little cooling water requirements compared to a combined cycle system.

Waste Heat Boiler Deskbook

48

0'-3' DD

I t

SUPERHEATER

Figure 1-23b. HRSG for Cheng cycle system. [Courtesy ABCO Industries]

4. The flexibility of varying process steam and power output without wasting energy. In a conventional gas turbine cogeneration plant, a diverter would have to be used to bypass gas if the steam demand reduces, resulting a waste of energy. When less steam is required for process in Cheng cycle system, the excess steam can be

49

Waste Heat Boilers

superheated and injected into the gas turbine, thus increasing the electrical power. 5. As a incidental advantage, the NO x emission is significantly lower in the injection mode. The only drawback is the large amount of good quality feed water that is lost to the atmosphere along with the exhaust gases in the injection mode. The superheater is designed so that it can run dry if required; . this mode would occur when maximum process steam is required. Four modes of operation are possible with this system, as seen in Figure 124. 6~Q

______________________-{b

Q.

I::l CJ

...J ct

....D'U t;

______________

3.5 DERATING WITH STEAM INJECTION

~

SUPPLEMENTARY FIRING WITH STEAM INJECTION

.: ::l

c

DERATING WITH SUPPLEMENTARY FIRING

~ w

10,750

20,870

STEAM OUTPUT, KG/H Figure 1-24. Operating modes for Cheng cycle.

1. unfired, full injection; no process steam; this mode is attractive for plants which have no requirements for steam or have little cooling water but which would like to maximize the power output using a HRSG and gas turbine alone. Due to the increased mass flow and the higher gas specific heat, the gas turbine power output is increased by more than 60%. 2. fired, full injection; maximum power and process steam are generated.

Waste Heat Boiler Deskbook

50

3. fired, no injection; maximum amount of process stearn is generated. The superheater runs dry. 4. unfired, no injection; all.the stearn goes for process. The HRSG can operate not only in the above four extreme points but also in between with any combination of power and process stearn by varying the injection stearn, and the fuel input to the HRSG. The electrical power output based on an Allison 501 KB5 machine varies from 3.5 to 5.5 MW, while the process stearn can vary from nil to 50,000 pph as shown in Figure 1-24. The HRSG is designed to handle the significant variations in stearn load, firing temperatures, gas analysis and exhaust mass flow; a feed water exchanger is located between the deaerator and economizer to preheat the make up water and hence improve the system efficiency. As mentioned earlier, the superheater is designed to run dry. In order to minimize concerns with exfoliation of metal oxides into the turbine due to the varying wet and dry operation, the superheater tubes are alonized on the inside. Table 1-5 shows the gas data with and without injection. This system is in operation in several plants which have varying stearn and electrical load, Figure 1-7c.

Table 1-5, HRSG Perfonnance in Various Modes A

B

C

D

1. Gas inietTamp., F (C)

948(509)

948(509)

1031(555)

1031(555)

2. Gas Temp. 10 Burner, F (C)

833(445)

833(445)

1031(445)

1031(555)

3.. Gas Temp. 10 Evap., F (C)

833(445)

1553(845)

1550(843)

1031(555)

4. Gas Temp. 10 Eco, F (C)

422(217)

443(228)

430(221)

421(216)

5. Slack Gas Temp. F (C)

269(132)

217(103)

198(92)

226(108)

6. Tot. Sleam, pph (kg/h)

21000(9528)

61000(27670)

46000(20870)

23700(10753)

7. Sleam Press., Psig (kpag)

........................... 265(1826).........................

8. Superhld. Sleam, pph (kg/h)

19900(9029)

9. Sleam Temp., F (C)

19800(8924)

855(457)

855(457)

10. Waler Temp. to Eco, F (C)

119(48)

146(63)

139(59)

124(51)

11. Gas Pro Drop, in WC (mmwc)

6.9(175.3)

8.7(221)

6.3(160)

5.4(137.1)

CASE A: Unfired, injeclion; B: Fired, injeclion; C: Fired, dry; D: Unfired, dry

~~~~--~------------~----~-~~-----------~-

Waste Heat Boilers

51

OTHER APPLICATIONS OF GAS TURBINE EXHAUST There are several used to which the exhaust from gas turbines are put such as direct drying of wood products or cement. The exhaust may be used as combustion air in fired heaters. Compressed air or heat transfer fluids such as therminol, glycol may be heated by the exhaust gases. Recuperators or air-heaters resemble HRSG modules except for the fact that compressed air flows inside the tubes, not steam/water. Due to the low tube side heat transfer coefficient, a low fin density is recommended as explained in Appendix A.

ENHANCED OIL RECOVERY (EOR) One of the methods of enhancing the output of oil is through injection of steam into the wells. This. is called secondary oil recovery. Wet steam of 80% quality at high pressures, on the order of 2500 psi, depending on the depth of the well, is generated using gas turbine exhaust in EOR (enhanced oil recovery) applications. The water available for these applications could be of poor quality having on the order of 1000 to 5000 ppm total dissolved solids. The HRSG resembles an economizer module or a forced circulation evaporator and is generally of the once through type. Drum type units have not been used due to their stringent water quality requirements and also the poor response to load changes. Water enters the coils at one end and come out as wet steam at the other end. Wet steam ensures that solids in the water, which may be of poor quality, are not deposited inside the tubes but dissolved and carried along with the wet steam. One has to be concerned with heat flux inside the tubes particularly with fired units. Hence selection of fin configuration along the gas flow path has to be made with great care. The tubes at the hot gas end are bare or have a lower ratio of external to internal area and the fin density can be increased at the lower gas temperature regions. With small capacities (on the order of 50 to 100,000 pph), the design is of single stream; multiple streams or parallel passes are avoided so that concerns with two phase flow

~,

52

Waste Heat Boiler Deskbook

instabilities are minimal. In large units, methods to improve two phase flow stability, such as increasing the inlet restriction by using an orifice in each stream, are employed. The units could be unfired or fired. Low temperature systems such as Kalina cycle, in which the medium inside the HRSG tubes is a varying mixture of ammonia and water are being developed as also organic cycles.

RECIPROCATING ENGINE HEAT RECOVERY Generally the exhaust from reciprocating engines has a lower temperature and hence is used for heating hot water or for generating low pressure saturated steam. At full load a reciprocating engine converts about 33% of the available energy into useful work and the remaining 67% is lost to jacket water, to the exhaust gas and loss to atmosphere. The effect of ambient temperature on power output is not as significant as with a gas turbine, whose output could drop by nearly 3% to 5% with a lOoP increase in ambient temperature. With larger gas turbines evaporative coolers are used to alleviate this concern, but this is an added cost. Reciprocating engines also have a higher part load efficiency and require fuel gas at a much lower pressure. Though they are more expensive, they find applications in systems requiring higher power to heat ratios, hot water or low pressure steam. Figure 1-25 shows a system for generating low pressure steam on the order of 10 to 15 psig, from jacket water and exhaust heat. In an ebullient cooled engine, water is allowed to boil under controlled conditions inside the engine jacket.The resulting steam water mixture goes to an external steam separator where the steam is separated and water circulates back. The exhaust gas heat recovery system operates in parallel with engine jacket as shown.

WASTE HEAT BOILERS IN HYDROGEN PLANTS In the process of manufacture of hydrogen through steam reforming of natural gas or naphtha several waste gas streams are

J

__

II[IIIiII"'-------------~~-

--------

- -

53

Waste Heat Boilers

Steam header

c::=====:;=::: I

From steam

Figure 1-25. Heat recovery scheme with reciprocating engine.

generated. The steam reforming process converts a mixture of hydrogen, methane, carbon dioxide, carbon monoxide and water vapor in the presence of a catalyst. See Table 1-1 for the gas analysis and temperature ranges. Figure 1-26 shows the scheme for manufacture of hydrogen or ammonia. Reforming process requires that the reaction products at a high pressure be maintained at a high temperature in a reaction furnace heated by flue gases. The flue gas exiting the furnace at temperatures of 1600 to 1800°F are used to generate high pressure steam in a waste heat boiler. Maximum heat recovery compatible with considerations of high and low temperature corrosion is the objective of the flue gas waste heat boiler. The reaction products in the furnace called the reformed gas, on the other hand, are cooled in a reformed gas boiler from about 1650°F to 650°F. The exit gas temperature should be controlled within a narrow range at all loads so that the gas stream can be used for further processing; this boiler is called the reformed gas boiler; steam generation is of secondary importance. Figure 1-27 shows a boiler for this application with internal gas bypass control At low gas flow conditions, the exit gas temperature reduces (see Chapter 2 for methods of evaluation and method of computation of bypass flow). Hence a portion of the gas has to be bypassed and mixed with the cooler gases to obtain the desired exit gas temperature. Since the gas pressure is high, a fire tube boiler is appropriate for this service. The materials of construction have to be chosen considering the hydrogen damage pOSSibility.

--1

.....

Waste Heat Boiler Deskbook

54

NH,

Figure 1-26. Scheme of a hydrogen plant.

Figure 1-28 shows the Nelsons chart which is used as a guide to material selection. The presence of hydrogen increases the tube side heat transfer coefficient nearly 4 to 5 times that of a flue gas stream or airi refer to Chapter 2 for methods of sizing boilers. Due to the high heat flux associated with the presence of hydrogen, on the order of 80 to 120,000 Btu/ sq ft h, an elevated drum fire tube boiler in which the tube sheet can be kept cool with water-steam mixture is preferred over a single shell boiler which has an integral steam space, Figure 1-29.

Waste Heat Boilers

55

Figure 1-27. Fire tube boiler with internal gas bypass for a hydrogen plant. [courtesy ABea Industries.]

Ferrules are also used, Figure 1-30, to transfer the heat flux away from the tube sheet region into the tubes, which are cooled by the circulating steam-water mixture. The front tube sheet is also protected by refractory. This reduces the temperature differential across the tube sheet. Design of refractory is critical at the front end. This boiler operates by itself in a small plant or could be connected to a common steam drum catering to several other boilers as in a large hydrogen or ammonia plant. Flue gas boilers are generally of water tube design in large plants generating high pressure high temperature steam for steam turbines and process use; they are of fire tube design in smaller plants. In some plants the steam drum is made common to the flue gas and reformed gas boilers through a system of downcomers and risers. Figure 1-31 shows a boiler in a small hydrogen plant in which the flue gas stream is cooled in a water tube boiler, while the reformed gas and converted gas streams are cooled in fire tube boilers. The steam drum is common to all the boilers. A large hold up time of 5 to 8 minutes from normal level to empty is used in these plants, so

- - SURFACEDECAR aUIUZATION

'400

---

noo

.... w

..

...a.

c

800

w

iDO

1II

~

C-i.O. . Cr.

0.5./."', STEEL

-----------

---,-7---- ------c-:3.~'

1000

.00

ill

~~~~~~~~~~

noo

ill

;:)

:::--::--_

lJOO

JIiiI-,.Cr

--.

G.5j....,nm

21.

100

c;,

1U1."'e

0.51."' e STEEL ----

STEEL

u·J. ... STEEL -CAIIIBOH STEEL

500

- 0 . 2 5 . "Q STEEL

WELDE D OR HOT FORMED

400 JDD

0

1000

1500 HYDROaEH

2500 PARTIAL P"US URIE

Figure 1-28. Nelsons chart

--~---------------~-------~--------~--~~-------

Waste Heat Boilers

57

that in the event of feed pump failure there is adequate time to shut down the system or take corrective measures.

Figure 1-29. Single shell fire tube boiler for incineration heat recovery. [courtesy ABeO Industries.]

BOILERS FOR SULFURIC ACID PLANTS In the manufacture of sulfuric acid through contact process sulfur gases resulting from the combustion of raw sulfur are cooled from about 2100°F to 750°F in a waste heat boiler before being sent to a converter. The exit gas temperature from the boiler has to be maintained within a narrow range of temperature generally by using a gas bypass system, which could be external or internal to the boiler. The bypass duct is refractory lined. Figure 1-32 shows a water tube boiler for a large 2000 tpd sulfuric acid plant. Due to the size, an external steam drum with downcomers and risers was used. The casing may be of membrane wall design which maintains the casing at the saturation temperature and hence alleviates the need for a hot casing, which is generally used for this corrosive

Waste Heat Boiler Deskbook

58

0,3'(131'11'1) MIN, SA-316-70 "'(102M) VSL SO

16-1/2')(3'x3 3/16' (12,7)(761<131,8",,) SA-316-70 GUSSETS

0,~'(13~",)

MIN,

SA-SI6-70 0.73'(191'1"') NOM, SA-387-11 C\ 2

"'(loa",,,,) GREENCAst 94

INCOLIlY 800 FERRULES

TUBES ARE: ATtACIf£Il BY LIGHT RIlL.L.ING, MUL. tI-PASS STRENGTH VEL.DING AND REROL.L.ING,

Figure 1-30, Arrangement of ferrules, inlet vestibule for high gas temperature application, [courtesy ABCa Industries,]

~9

Figure 1-31. Combination fire tube, water tube boiler for a small hydrogen plant, [courtesy ABCa Industries,]

S'-d'

8'-

SECTION "A -A"

SIDE VIEW

Figure 1-32. Large water tube boiler with external drum for a 2000 tons/day sulfuric acid plant. [courtesy

ABea Industries]

60

Waste Heat Boiler Deskbook

service More information on hot casing design is given in Appendix B. Extended surfaces are used to make the design compact. Due to the high gas inlet temperature (lnd resulting heat flux, the design consists of a few rows of bare tubes followed by tubes with low fin density and then tubes with high fin density. This varying fin configuration also minimizes the tube wall and fin temperature. The gas pressure in some sulfuric acid plants are high, on the order of 3 to 5 psig, which calls for a heavily reinforced casing design; alternatively one could locate the heat transfer surfaces in a pressure vessel. If economizers are used, the feed water temperature has to be very high, on the order of 3500 P to minimize dew point concerns, particularly if sulfur trioxide is present. Carbon steel tubes with extended surfaces have been used successfully in several plants in the USA. Some plants, particularly those in Europe and Asia use cast iron gilled surfaces. This design probably offers more protection from corrosion due to its heavy thickness.

SULFUR CONDENSERS Sulfur condensers are waste heat boilers generating low pressure steam used for recovering energy from the combustion of hydrogen sulfide gas in a reaction furnace. In these boilers, high temperature gases with a typical analysis as in Table 1-1 enter the boiler, usually a fire tube type at very high temperature, in the range of 2300 to 2600 o P. A two pass design is used. The turn around section is designed for a gas temperature of about 900 to 12000 P and the exit gas temperature is about 600o P. Some sulfur is removed as liquid at the end of the first pass through drains, which are kept hot. The boiler is slightly inclined to facilitate draining. The boiler duty has sensible as well as latent heat components due to the condensation of sulfur. Due to the presence of sulfur, sulfur dioxide and water vapor care is taken to design the refractory at the inlet and exit end of the boiler so as to keep the casing at a high temperature to avoid dew point corrosion concerns. External insulation to form a hot casing (see Appendix B) may also be needed. Pigure 1-33 shows a two pass elevated drum fire tube boiler for this service. Several different gas streams may also enter and leave a sulfur reclaimer.

,-SIf'. JIANlfAYS ,/ (TIrO PROVIDED)

...I:i k .(' -GRZENCAS'l' 9 t"-VSL-50

I

I END VIEW'

4'-({'

19'-1f'

S'-If'

81M VIEW'

Figure 1-33. Two gas pass fire tube boiler with elevated drum for sulfur recovery application. [courtesy ABeD Industries]

-

62

- -

--------------------~-.....,

Waste Heat Boiler Deskbook

FLUIDISED BED CAT CRACKER HEAT RECOVERY Heat recovery boilers are used to recover energy from waste gas streams from fluidised bed catalytic crackers used in refineries. Since the gas stream has particulates, bare tubes are preferred for the heat transfer surfaces unless the particulate concentration is very low. Fig. 1-34 shows a boiler for this application.The gas inlet temperature is on the order of 1400°F. Superheated steam at 600 psig, 750°F is generated. As the gas inlet comes from a far away location with several turns, a flow modelling study was done to ensure that a non-uniformity in gas velocity profile across the cross section was minimal.

Figure 1-34. Large boiler with superheater, evaporator and economizer for fluidised bed cat cracker application. [courtesy ABeD Industries.]

INCINERATION AND HEAT RECOVERY Solid Waste Incineration Various types of wastes such as MSW (municipal solid waste), pathological wastes, liquid wastes and sewage sludge are generated by communities and organizations. It is estimated that each

--------------------------~-------

Waste Heat Boilers

63

individual will be generating a ton of waste per year, which has to be disposed of. In addition about 250 metric tons of hazardous waste are generated each year by chemical, electronic and manufacturing industries. Several methods have been used in the past to handle the waste disposal problem: a. landfills

b. composting

c. recycling

d. incineration

The age old method of using lan<;ifills for burying the waste, though in vogue, cannot handle all of the volume; due to various state and local regulations and the NIMBY syndrome (not in my back yard), the landfills have to be located very far away from communities with the result that it is expensive to transport the waste; hauling and tipping fees are on the increase; landfill sites are becoming scarce; these aspects have forced several cities to look for other options to get rid of various wastes. One of the methods is composting or size reduction. This applies to products such as leaves, grass, tree trimmings, MSW and sewage sludge. Compo sting consists of three steps: material separation and sizing, aerobic decomposition and curing. The recyclables are removed and the remainder is shredded, screened and allowed to decompose in air-controlled environment at 150°F, sustained for three days to destroy pathogens-followed by several weeks of curing. Recycling/materials recovery is economically feasible for most communities and dictated by market for the recovered products. Five materials are targeted for removal from the waste stream: paper, glass, ferrous metals, aluminium and plastic. But recycling can do only so much and involves public cooperation and support. About 35% of the waste probably can be recycled. Recycling programs are prevalent in several US cities. Incineration is another option widely used for disposing of several types of wastes. An advantage of incineration is that there is a 70 to 90% reduction in volume of waste which reduces the cost of tipping and landfill space requirements .. In addition, energy is available from the products of combustion for generating hot water or steam for power or process. Steam may be used for generating electrical power via a steam turbine generator. Process steam may be used for heating or cooling applications. Hot water has a few

Waste Heat Boiler Deskbook

64

advantages over steam. It can be transported over greater distances and smaller pipes can be used compared to steam. Hospitals, commercial institutions and residential users consume hot water. There are however a few concerns with incineration. The ash is considered by some to be toxic and concerns on how it is being disposed of playa role in the location of the incineration facilities. There are environmental concerns regarding the emissions but technology is available today to reduce the pollutants to acceptable levels. The decision whether to locate the incinerator at site or far away is taken after discussions with community groups. There are legal and social issues which have to be handled, making incineration siting a costly and long drawn affair. Thus an integrated approach consisting of all of the options discussed above including incineration is probably the best approach to the waste disposal problem. Table 1-6 shows typical composition of MSW. It may be shown that the heating value is significant, 4500 Btu/lb on average on as received basis, and hence a large amount of energy can be recovered from combustion and flue gas heat recovery. Table 1-6. Average Composition of MSW Nationwide, 1981-1990. a Composition (%) Type of Material Glass Metal (ferrous) (aluminum) (other nonferrous) Paper and Paper Products Plastic Rubber, Leather, and Textiles Wood Organics (Food and yard waste) Misc.llnorganics

1981

1985

1990

9.7 8.8 (7.5) (1.0) (0.3) 31.9 4.1 4.5 3.9 35.5 .1..Q 100.0

9.5 8.7 (7.1 ) (1.2) (0.4) 31.7 4.4 5.1 3.7 35.1

9.3 8.5 (7.0) (1.2) (0.3) 31.0 4.9 5.3 3.7 35.0

il

~

100.0

100.0

aSource: Franklin, W. et aI., "Municipal Waste Generation and Composition to 1990'Solid Waste Management and the Paper Industry, Solid Waste Council of the Paper Industry, February 1979.

~------

---- --

--

---

--------;--

Waste Heat Boilers

65

Table 1-7 shows the heating values of a few wastes. Depending upon the heat content of the waste stream between 4 and 6 million Btu/ ton can be recovered as steam. Health care industries generate waste that are high in heat content consisting of plastics, paper, rubber. Table 1-7. Heating Values of Industrial Wastes Vary Widely. Solid.

Average heat

content, Btu/lb Bark 4500-520.0. Bitumen waste 16,570. Brown paper 7250. Cardboard 6810. Corrugated paper (loose) 70.40. General wood wastes 450.0.-6500 Latex 10.,000. Nylon 13,620. Paraffin 18,621 Plastic-coated paper 7340. Polyethylene film 19,780. Polypropylene 19,860. Polystyrene 17,700 Polyurethane (foam) 17,580. Resin-bonded fiberglass 19,500 Sawdust and shavings 4500-7500 Tarpaper 11,50.0. Wax paper 11,500 Wood 9000.

Gaaea

Blast-furnace CO Coke-oven Refinery

1139 575 19,700 21,800

LIquId.

Black liquor Dirty solvents Industrial sludges Oily waste, residue Paints and resins Spent lubr~cants Sulfite liquor

370.0.-420.0. 10.,000-16,000. 3700-4200 18,0.0.0. 6000-10.,0.00 10.,000-14,000. 4200

.....

-----

66

-~

---------------------......,

Waste Heat Boiler Deskbook

Not all materials are acceptable for use as waste fuels unless the facility is specifically designed to accept them. Not acceptable for example as MSW feed are: 1) materials that may cause a waste energy facility to violate an air or water quality effluent standard; 2) pathological and infectious wastes, radio active wastes, poisons, acids and bases, human remains, paints, refrigerators, stoves, airconditioners, bath tubs, sinks, oil sludges to name a few. There is technology available today to incinerate solid, liquid and gaseous wastes including municipal solid waste, pathological waste from hospitals, sewage sludge, hazardous wastes in chemical plants, petroleum coke in refineries, wood wastes and coal mining wastes. Technology is also available to handle the emission problems arising out of these facilities. The various types of incinerators used are: fluidised bed combustors, Figure 1-35, rotary kilns, Figure 1-36, fixed and multiple hearth furnaces and moving grates, Figure 1-37. Infrared incineration technology uses electricity as the energy source and the heat generated by high resistance infrared heating elements is used to burn hazardous wastes; however this technology is still under development. There are two basic technologies for disposing of solid waste, namely mass burning and RDF (refuse derived fuels). Basically two types of designs are used for the combustion chambers of mass burning systems; they are water wall and refractory lined incinerators; these are generally used for large plants in the range of 400 to 2000 tpd capacity. Modular incinerators are used for small units, up to 200 tpd. The ad vantage of these units is the short lead time for their installation. In mass burning systems, the waste is not processed except for removal of bulky items such as mattresses, iron and steel products. Waterwall incinerators, Figure 1-37, use a fully watercooled membrane wall design as in a conventional fossil fuel fired plant. The refuse is transferred from the storage pit by an over head crane into the hopper/chutes; the waste slowly is fed by rams into the stoker assembly. The stoker provides the mechanism for tumbling and mixing of refuse. Excess air of 100 to 120% is used. About 60 to 70% of the ash is collected in the bottom and the rest goes with the flue gas and is called fly ash.

Stack

Feed building

Figure 1-36. Rotary kiln incineration.

69

Waste Heat Boilers

Byproduct

Steam

•

Flue

Gas

Ash

Figure 1-37. Large water tube boiler with moving grate for MSW application.

The products of combustion are cooled in the radiant waterwall section and then enter the convection section for further heat transfer to saturated or superheated steam. Due to the presence of corrosive substances in the flue gas such as chlorides and halides the steam parameters have to be chosen with care. High temperature corrosion is a major problem in MSW plants. Superheated steam temperature unlike utility boilers should be much lower. 700 to 7500 P is the suggested upper limit. The superheater also has to be located in a cooler zone so as to limit the tube wall temperatures. Higher steam temperatures if required are better handled in an externally fired steam superheater.

-------------------------~~---

1

Waste Heat Boiler Deskbook

70

In refractory lined incinerators, the waterwall is replaced by a refractory lined chamber in which combustion takes place. Higher excess air, on the order of 200%, is required so as to reduce the gas temperature entering the convection section or the waste heat boiler. Due to the nature of construction there will be leakage of gases around openings and the heat loss will be higher than a waterwall unit. Waste to energy plants generally prefer waterwall units. Controlled air modular incinerators are widely used for small capacity MSW plants from 20 to 200 tpd. The modular unit consists of a primary and a secondary chamber connected to a waste heat boiler, Figure 1-38. The chambers are refractory lined. Minimal field installation is required on these units. Several modules are used in larger plants each with its own waste heat boiler but with a common steam turbine generator system. natural draft dump stack

stack

t

steam

damper

t

leakage~

{)

() feedwater

() C) secondary combustion chamber

C) SCCair

{)

PeC overfire leakage

under fire air ash

Figure 1-38a. Arrangement of modular incinerator for MSW application.

Waste Heat Boilers

71

Figure 1-38b. Arrangement of multiple boilers behind modular incinerators. [courtesy ABeD Industries.]

Hydraulically operated ram feeders transfer the refuse into the combustion chamber. A small reciprocating grate or stoker moves the waste through the chamber. Combustion air usually much below stochiometric levels is provided by a forced draft fan; the air is directed below the grate and comes into contact with the refuse. The solid waste is pyrolytically decomposed producing a reactive and combustable gas. The charge cooks for several hours. Finally the char is introduced tangentially or other wise into the secondary chamber, where additional excess air and auxiliary fuel such as natural gas or oil are introduced. The combustible mixture is held in the chamber for 1 to 2 seconds to ensure combustion. The flue gases are then drawn into the waste heat boiler using an induced draft fan. Typical gas exit temperature from the incinerator is 1800 to 2000°F.

72

Waste Heat Boiler Deskbook

REFUSE DERIVED FUELS (RDF) RDF and mass burning are distinguished by the mode of refuse preparation. In mass burning, the refuse is used as it is received. Large objects non-combustibles and hazardous materials are removed either manually from the tipping floor or remotely by crane prior to burning. In RDF burning the fuel is prepared from the refuse. It is processed by different means to yield a high quality shredded fuel and other resalable or recyclable products. Hazardous and large bulky materials and non combustables are removed prior to the processing system. The RDF is fired in a boiler furnace using a traveling grate, fluidised bed, rotary combustor or other systems. Table 1-8 shows the analysis of typical MSW, RDF and coal along with the heat losses and efficiency. While mass fired furnaces typically operate at 80-100% excess air, RDF units can, due to the higher quality of fuel, operate at 30 to 40% excess air, resulting in a more efficient unit. Figure 1-39 shows a recovery scheme and Figure 1-40 an arrangement of boilers for generating steam and power from refuse. , Due to the elaborate fuel preparation system, it is necessary for RDF systems to be very large in size to take advantage of economies of scale. RDF systems may be used when rapid response to load and higher boiler efficiency is desired. RDF may be co-fired with other solid, liquid or gaseous fuels, their contribution being up to 20 to 30% of the total energy input. The air requirements for transporting the RDF along with the combustion air requirements have to be factored into the sizing of the forced and induced draft fans. Greater ash quantities may be generated and the fly ash and bot:om ash handling system may have to be sized accordingly. Capital cost for RDF dedicated boilers is higher than boilers used for co-firing RDF. However co-firing has led to several problems in boilers in the past such as slagging in upper furnace wall, decreased efficiency of electrostatic precipitator, boiler tube corrosion and bottom ash accumulation. Through the processing of MSW into RDF removes much of the unwanted non-combustible materials, serious maintenance problems have been caused particularly with the stoker. Aluminium which is

Waste Heat Boilers

73

Table 1-8. Analyses of MSW and RDF Compared to Bituminous Coal Analyses, % (by weight) Constituent

C H2

02

N2 S CI2 H20 Ash HHV (wet), Btu/lb

MSW

27.9 3.7 20.7 0.2 0.1 0.1 31.3 16.0 5100

RDF

36.1 5.1 31.6 0.8 0.1 0.1 20.2 6.0 6200

Bituminous coal

72.8 4.8 6.2 1.5 2.2 0 3.5 9.0 13,000

Losses per fuel, % Loss items Dry gas loss Moisture in fuel loss Moisture in air loss Unburned combustibles Radiation Unaccounted Total losses Efficiency

MSW

RDF

10.1 14.5 0.2 3.3 0.5 1.5 30.1 69.9

6.3 11.0 0.2 2.2 0.5 1.5 21.7 78.3

Bituminous coal

6.2 4.2 0.2 2.5 0.3 1.5 14.9 85.1

light enough to be carried into the refuse stream can melt on the grate bars and hinder their alignment. Glass and silt which may become impregnated in the refuse during preprocessing will also cause accelerated grate wear. The first few generation plants experienced problems such as explosions, dust in processing plant and storage problems. RDF tends to compress itself such that retrieval of the refuse at the bottom of a pile tends to be difficult. Storing RDF for long periods of time caused the fuel to ferment and spontaneously combust.

Waste Heat Boiler Deskbook

74

Community

Glass & cans (Food & beverage) Plastics PET & HOPE (Soda &milk jugs) Household batteries

To curbside Placed in set-out container

Placed in bag or tied

Newspapers & magazines

1

Trash

Intermediate processing cenler Electrical power

I

Tin & steel baled

New aluminum

New tin & steel

!

Refuse to energy plant

I

Aluminum baled

+

Glass 3 colors crushed

+

New glass

I

Newspaper baled

+

New paper

l

Plastics PET HOPE baled

+

New plastic products

Figure 1-39. Scheme of refuse recovery system.

FLUID BED COMBUSTORS FOR MSW In conventional mass-burn systems the fuel or waste is burned on the surface of the grate or hearth near the bottom of the furnace, while in RDF combustors the fuel is blown in over the grate and burned above it. In fluid bed combustors (FBC) the MSW is burned in suspension held well above the furnace floor by a strong upward flow of air. The fluid bed is a highly turbulent mixture of intensely heated inert bed materials such as sand, ash particles and limestone mixed with a relatively small amount of MSW. There are two types of FBCs, the bubbling bed (BFB) and circulating bed (CFB). Figure 135.

Waste Heat Boilers

75

Slab Receiving & Storage

~h~ Aluminum Magnet 9

l

Residue Aluminum

To Stack

9 Magnets 9 9oensifiers9 Heavy Light Ferrous Ferrous Material Material

o ° Boiler

~~~~~~~~

Figure 1-40. Flow diagram for an RDF based plant.

Both may be used to burn RDF alone or in combination with other fuels such as wood chips, sewage sludge, coal or tires. CFB provides more flexibility in burning fuels singly or in combination and has superior load-following capabilities. FBCs have several advantages: 1. High combustion efficiency on the order of 99% compared to conventional solid waste combustors, which claim 97 to 98%. 2. Stable combustion due to the tremendous amount of heat absorbed and retained in the inventory of hot inert particles-cold spots are minimized. .... / _ 3. Lower combustion temperature of 1500 td 1700°F and lower excess air of 30-90% result in the formation of lower NO x on the order of 100-200 ppm versus 150-300 ppm with mass burned systems. Boiler

------~~~----

76

----

------~

-

Waste Heat Boiler Deskbook

efficiency is also improved due to the lower excess air and better combustion. 4. Sulfur removal is accomplished easily by addition of limestone in the bed. The CFBs have additional advantages such as better turn down ratios and increased residence time in the combustor, resulting in higher efficiency of combustion. The limitations or concerns include the following: 1. Requires prepared fuels such as RDF, though in Japan the MSW has been burned without preparation except for removal of non-combustibles. 2. Problems with melting glass in the bed. 3. Erosion of bed coils or boiler parts due to carryover of sand, ash particles.

HAZARDOUS WASTE INCINERATION Chemical plants generate hazardous wastes such as gases, organic liquids, solids and sludges and refuse which have to be oxidized to carbon dioxide and water vapor at a destruction and removal efficiency of 99.99% at least for selected principal organic hazardous components (POHCs). The combustion chamber for hazardous wastes must not only burn the combustible material but also ensure specific residence times at high enough temperatures (2000 to 2400°F) to ensure thermal decomposition of the dangerous components and to meet applicable environmental regulations. The four criteria relevant to successful operation are time, temperature, turbulence and excess air. For example most chlorinated hydrocarbons require temperatures of about 1000 to 1200°C for at least 1 second at 100% excess air to ensure complete combustion. Fluidised beds have been widely used for incineration of hazardous liquids, sludges or solid wastes. The multiple hearth furnace is suitable for handling sludges, solids containing liquids and granular solids up to 4 in. in size. The rotary kiln, Figure 1-36, is the most versatile incinerator presently available. The kiln is a refractory lined cylindrical chamber about 4 to 15 feet in diameter and 40 to 100 feet long. In

Waste Heat Boilers

77

addition to granular solids and sludges the kiln can handle irregular solids, hospital wastes, tires, plastics. The rotary motion of the kiln creates a tumbling action thereby mixing the waste stream with the co-currently flowing hot combustion gases. By fitting the kiln with liquid and gas burners, several types of hazardous wastes may be treated. The combustion gases are subjected to a minimum temperature of 1100 to 1400°C for a residence time of 1 to 2 seconds. Due to the large volume required, usually kilns are fitted with external after-combustion chambers connected to the kiln outlet. Additional burners are fitted in these chambers to ensure temperature level and destruction efficiency.

HIGH TEMPERATURE CORROSION Gas side corrosion associated with high gas and metal temperatures are common in MSW and other incineration applications. They could be broadly classified into three types: 1. High temperature liquid phase corrosion, caused by molten alkali metal salts such as metal chlorides and their eutectic mixtures that have low melting points. Table 1-9 shows the melting points of a few eutectics found in MSW applications. By reducing the gas temperature to convection sections below the melting temperatures of these salts one can minimize slagging concerns, which results in build up of corrosive deposits. Proper cleaning of boiler tubes either through rapping mechanisms, Figure 1-41, which dislodges the deposits as soon as they are formed or through soot blowing helps minimize this corrosion problem. A large radiant section helps cool the gases to temperatures below slagging temperatures of some alkali metals but over a period of time the radiant section also gets fouled and the gas temperature to the convection keeps increasing thus resulting in slagging at the front end. The remedy would be to shut down the boiler and clean it thoroughly. Flue gas recirculation also minimizes the slagging problems by reducing the gas temperature entering the convection section. 2. Corrosion due to non-uniform furnace atmosphere; caused by partial oxidation, which results in a reducing environment in which

----- - - - - - - - - -

- - - - - - -

-"~"-"~--------

Waste Heat Boiler Deskbook

78

CO and H2S are produced. These gases react with the protective layer of iron oxide formed on the tubes exposing them to a corrosive attack. Sulfidation attack on steels can result at high temperatures. High nickel steels which have good corrosion resistance against chlorine attack may get corroded in reducing environment. High chromium steels may have to be used for the boiler tube supports. The corrosion concerns can be alleviated to some extent by proper combustion and distribution of air to various parts of the grate/bed so that a fluctuating oxidizing/reducing atmosphere is not created.

Table 1-9 Melting Points of Some Salts and Oxides Present in Incinerators Component, MQI!2 FraQtiQn P20 3 0.50NaCI - 0.26Na2S04 - 0.24Na2C03 O.65Na2S04 - 0.35NaCI 0.62Na2C03 - 0.38NaCI NaCI Na2S04 Ca~

Fe~3 Fe~3

0Q 569

1)

OF 1056

612 623 633 801 884 1236 1462 1560

2) 2) 2) 3) 3) 3) 3) 4)

1134 1153 1172 1474 1623 2257 2664 2840

R!2marks

tertiary eutectic binary eutectic binary eutectic

decomposition decomposition

1) Fabian, H.W., P. Reher and M. Schoen. "How Bayer Incinerates Wastes, Hydrocarbon Processing, 185, April 1979. 2) Bergnian, A.G., and A.K. Sementann. "The Tertiary Systems Na, CI, S04, CO and KlCI, S04, C04, Zhur Neorg Khim, 3, (2),388,1958. 3) Dean J.A., editor, "Langes Handbook of Chemistry: Ed 12 McGraw-Hili Book Co., New York, NY, pp 4-48-113,1979. 4) Kirk and Othime, "Encyclopedia of Chemical Technology," Ed 2 Vol 12, P 19. John Wiley and Sons, New York, NY 11964. Source: Reference 26

79

Waste Heat Boilers

Rapper drive shaft

. II

II II

feeder header

Figure 1-41. Rapping Mechanism for Oeaning Boiler Tubes.

3. Corrosion due to Rcl and e12 gases. Rcl is formed due to the combustion of plastics in the waste. It is very corrosive above 800°F, as seen in Figure 1-42. Rcl is also responsible for low temperature corrosion when it attacks steel surfaces below the acid dew point as hydrochloric acid; see Appendix B on low temperature corrosion for estimation of acid dew points of various gases; superheaters if used should be designed to minimize the tube wall temperature and located in a region such that the gas and metal temperatures are not high. Fouling is another serious problem particularly with solid waste fuels. Depending on the type of incinerator and the velocity of combustion air in the combustion zone, ash particulates can get carried away with the flue gas; this could be on the order of a 0.1 to several grains/scf depending on the process. When proper cleaning methods are not employed, the dry ash particulates can deposit on

Waste Heat Boiler Deskbook

80

the tube surfaces and hinder heat transfer. The boiler exit gas temperature will keep increasing with time and when it becomes unacceptable, a shut down cleaning may be warranted. On-line cleaning can help alleviate the situation a little.

A

RANGE I" ELECTROLYTIC CORROSlqN

r

:..:: u

B

- -........1 - - -

t t t

~

I

z

0

Vl

..... /

0

« «

I

-l-- -

0

u

()

32

100

212

lI

200

392

300

572

400

752

500

932·

/

600

BI2

"'-CORROS ION IN THE GAS PHASE 700

1292

800 147~

°c Of

t TEMPERATURE CONDENSATION POINT A • RANGE OF CORROS ION DUE TO FERR IC CHLOR IDE OR ALKALI FERR IC SULFATE FORMATION

B = RANGE OF CORROS ION DUE TO FERR IC CHLORIDE OR ALKALI FERRIC SUlfATE DECOMPOSITION Figure 1-42. High and Low Temperature Corrosion Due to Hydrochloric Acid.

HEAT RECOVERY BOILERS Waste heat boilers for modular incineration heat recovery systems have to be designed with care considering the potential problems of high and low temperature corrosion associated with the products of incineration of solid, liquid or gaseous wastes. Figure 1-43 shows a boiler for a modular 100 tpd incinerator for municipal solid waste. The unit generating 650 psig 650°F superheated steam has been in successful operation for several years. This is a two drum boiler with refractory lined casing. The gases on leaving the boiler enter an economizer. All surfaces are of

Waste Heat Boilers

81

bare tube design with inline configuration. Several aspects were looked into before deciding upon the design.

END VIEW'

SIDE VIEW

Figure 1-43a. Boiler for MSW Application with Buried Superheater [courtesy ABCD Industries]

Figure 1-43b. Boiler for MSW Application with Buried Superheater [courtesy ABCD Industries]

82

Waste Heat Boiler Deskbook

1. The steam temperature was selected as 650 oP, though a

higher temperature would have been better for the steam turbine. The reason is the high temperature corrosion problem associated with Hcl in the flue gases, which was discussed earlier. The superheater is located beyond several rows of convection and screen tubes so as to limit the gas temperature at the superheater to about 1000oP. In addition, a mixed flow arrangement was used for the superheater with the steam exit at the cooler gas end so as to minimize the tube wall temperature. All these design features result in low superheater tube metal temperatures. This helps minimize high temperature corrosion with Hcl. Large plants in which high pressure high steam temperature is necessary from an economic consideration should use externally fired superheaters to raise the steam temperature to the desired level. Another advantage of this design is that the effect of fouling does not have a great impact on the steam temperature. If the superheater had been located at the front end, the variations in steam temperature with fouling would have been higher than with a buried superheater design. High particulate loadings and slagging result in heavy fouling at the front end, which is carried down to the end of convection section if cleaning is not proper. 2. The front end of the boiler has several rows of wide spaced slag screen tubes, which helps minimize bridging of tubes by molten slag; this is a combination of several salts of sodium and potassium which have low melting points. Also, the wide spaced staggered arrangement at the slag screen section helps ash particles to drop out into the ash hopper below. the rest of the convection section has an inline arrangement. 3. Low gas velocities were used to minimize erosion associated with abrasive materials in the waste; typical gas velocities are 30 to 40 ft/s. 4. In order to minimize slagging at the front end with gas temperatures about 18000 P and more, flue gas recirculation was used to temper the hot gases with cooler gases at 4000 P from the boiler rear. The gas temperature at the boiler inlet was reduced to about 16000 P from 1850oP, which helps reduce slagging.

Waste Heat Boilers

83

5. Wide spaced cleaning and access lanes were provided with soot blowers to clean the tubes surfaces. Tube shields were used to prevent erosion near soot blower regions. 6. Large hoppers were provided to collect ash. The design of large boilers for RDF or MSW also has to take into consideration high steam and metal temperature at the superheater to minimize corrosion. Interstage desuperheaters using demineralized water are widely used to keep metal temperatures low. Gas velocities are kept low to minimize erosion. Furnace regions around the grate or fluidised beds are protected with silicon carbide coating or liners to minimize wear and corrosion from low melting eutectics such as Zinc chloride and chlorine attack. Tube thickness around this region is also increased to lengthen life. Figure 1-44 shows an A type boiler for hazardous waste incineration application. It is of membrane wall design, which avoids problems with refractory.

Figure 1-44. A-Type Boiler for Hazardous Waste Incineration Plant. [courtesy ABea Industries]

Waste Heat Boiler Deskbook

84

Single shell or elevated drum fire tube boilers have also been used in MSW or pathological waste incineration heat recovery. Since on-line cleaning is difficult with these boilers, they have to be shut down and cleaned more often than water tube boilers if the incoming gas has slagging tendencies. Sometimes auxiliary firing is required to continue generating steam when waste gas stream is absent. A burner as shown in Figure 1-45 may be added to accomplish this. In order to eliminate problems with controls, it is suggested that the waste gas stream and the fired combustion products do not mix but have separate paths through the fire tube boiler. Steam

Waste gas in lIJ .-----

1

~

Gas exit

Tubes~

( I---"--~~~--------+ ~vv~

First pass

__~~~~~~~__~

t""I

C"

II J.J.

IV

Burner

J

Single-shell firatube HRSG with burner. Figure 1-45. Fire Tube Boiler with Burner for an Incineration Plant

INCINERATION OF WOOD WASTES Figure 1-46 sho}Vs a boiler for recovering energy from products of combustion of wood waste and cardboard. It is an A type boiler, with a radiant section and hoppers. The gas enters at the bottom, moves through the radiant section and makes a turn to enter the convection bank. The pitch of the tubes in the convection section varies from 8 in. at the gas inlet to 4 in. at the exit to minimize plugging and fouling concerns considering the uncertain product mix and the low ash deformation and slagging temperatures. Cleaning aisles and soot blowers are provided.

------

- - - - - - - - - - - - - - - - - - - - - -

Waste Heat Boiler Deskbook

86

Fire tube boilers may also be used for recovering energy from products of combustion of wood waste or solid fuels. The HRT boiler as it is called is located in a refractory lined chamber. Combustion takes place below as shown in Figure 1-47 and the products of combustion flow through the tubes, which are large in size, on the order of 2.5 to 3 in. to minimize plugging. Low pressure saturated steam is generated. Low temperature superheated steam may also be generated, in which case the superheater is located behind the HRT. An economizer may also be added to improve the efficiency of the system.

Figure 1-47. Fire Tube Boiler for Wood Waste Incineration Plant. [courtesy ABeD I nd us tries]

ENERGY FROM TIRES Automobile tires weigh about 20 pounds each with a fuel value of 12000 to 16000 Btu/lb, higher than coal. Reciprocating stoker fire boilers have been used to burn them while generating steam for power or process. Combustion temperature is in excess of 1800°F. The

Waste Heat Boilers

87

boiler consists of a radiant section followed by convection sections including superheaters as required. Flue gas desulfurization system and fabric filters may be used to remove acid gases and 99% of the particulates. The boiler arrangement is similar to mass burn MSW systems.

INCINERATION OF LIQUIDS, FUMES AND VOCs Many processes produce gaseous discharges that cannot be directed into the atmosphere if they contain pollutants or acid gas. On site incineration may be used for such streams. The composition of the stream must be known to determine if it needs auxiliary fuel support or excessive air. For lean fuel mixtures thermal oxidation or direct flame incineration is used. Such waste gases generally are a mixture of hydrocarbons and air and known as volatile organic compounds (VOCs). They are usually found in concentrations below the Lower Explosive Limit and hence need auxiliary fuel to burn. Combustible liquids are generated by various industries. In order to burn them, they must be pumpable and capable of being atomized. The presence of chlorine, sulfur and inorganics influence the design of the incinerator. There will be damage to refractory at high temperatures if the liquid contains alkali salts. The design of the heat recovery boiler will also be affected by the presence of chlorine and alkali salts, as discussed earlier. Incineration of organic liquids, fumes and VOCs (volatile organic compounds) generally result in clean flue gases containing C02, H20, N2, and 02j hence water tube boilers for this application could have extended surfaces and the steam parameters can be based on system requirements and not limited by high temperature corrosion considerations. Figure 1-48 shows a large water tube boiler for a fume incineration system with superheater. Due to the high gas inlet temperature, a screen section was used to shield the superheater from radiant heat. The screen section is connected to the evaporator section through a system of downcomers and risers.

-

~~-------------------

88

------~~--

Waste Heat Boiler Deskbook

Figure 1-48. Large Water Tube Boiler with Screen, Superheater, Evaporator and economizer for a VOC/Fume Incineration Plant. [courtesy ABeD Industries]

Fire tube boilers of single shell or elevated drum type have also been used for heat recovery, Figure 1-29 and Figure 1-49. Gas temperatures of up to 2400°F may be handled in these boilers with elevated drum design. The tube sheet has to be adequately protected with refractory and proper circulation has to be ensured particularly at the front end. Gas recirculation is resorted to reduce the gas inlet temperature if the gas has particulates or salts with low melting points. Combustion of chlorinated plastics produces gas with Hcl and Cl2, see Table 1-1 for gas analysis; Figure 1-50 shows the process. The gas inlet temperature to the boiler is very high on the order of 2200 to 2400°F. Figure 1-49 shows a boiler for this service.

.,

~

'"(i) ::r:: /I)

M

~ 0:1

0 t;:;

Ql

'"

I I ~~;:;:;:-k::L-~-----

'-----( D ) - -............

E !.=:J.-----------1 F I-----.--------I.--{

Figure 1-49. Elevated drum fire tube boiler for high temperature incineration plant. [courtesy ABeD Industries]

IFftlrmmllrillllhvt\oll,

Figure 1-50. Incineration heat recovery scrubber system.

Oft.Gas

n"".v .. "t Fume NaOH Sol. Waste Fume Steam

Feed Water

(88·C) ChI.H.C.

lIIIolno.·cn,l!lm

Fume

2OO0"F • 26()OOF (min.)

(1093°C· 1427"C)

Fume Incinerator

.

- --,., ...

Waste Heat Boiler

Primary Scrubber

Secondary Scrubber

Waste Heat Boilers

91

Chlorine is more corrosive than Hcl on carbon steel at lower temperatures of 400 to 450°Fi low temperature corrosion is also a concern if the gases contain sulfur oxides. Hence boilers for this application do not have superheaters or economizers and the steam pressure is chosen so that the tube wall temperature is in the range of 350 to 420°F. Hence low pressure steam generation only is feasible. If chlorine is absent then the steam pressure could be higher. A superheater or economizer could be added behind fire tube boilers if low temperature corrosion is not a concern.

EMISSIONS FROM INCINERATORS There are basically four types of pollutants from incinerators; particulates, acid gases, particularly Hcl; metals such as cadmium, chromium, arsenic, mercury; toxic organics such as dioxins and furons. Electrostatic precipitators have been used for removal of particulates; newer plants use fabric filters and can handle the emission limit of .01 grains/scf. The fabric filters can be fairly effective in removing a broad range of particulate sizes with relatively low gas pressure drop. The drawback of filters is that the gas temperature must be kept below 500°F to prevent damage to the fabric. Thus if there is a severe fouling problem in the boiler which causes the exit gas temperature to increase, provisions such as spray quench must be made to reduce the gas temperature. Venturi scrubbers can remove both particulates and acid gases but require a higher gas pressure drop and are less effective on fine particulates. Good furnace operation and distribution of air minimizes formation of dioxins and furons.

AIR HEATING APPLICATIONS Tubular and regenerative type air heaters have been used in fossil plants for preheating combustion air before they enter the burners; recuperators preheat air to as high as 1500°F; some of the problems encountered by regenerators are leakage of gas or air, corrosion and erosion. Leakage is not a serious concern in tubular

92

Waste Heat Boiler Deskbook

heaters, but they are large in size and hence heavy due to the low overall heat transfer coefficient. Heat pipes have come to fill a need for a unit that is not large and can operate well in environments Where corrosion and fouling are present. Figure 1-51 shows the concept of heat pipes. This consists of several bundles of heat pipes-each filled with a liquid such as toulune, water or naphthalene and sealed. Heat from the flue gas evaporates the working fluid collected in the lower end of the slightly inclined pipe (6 to 10 degrees) and the vapor flows to the condensing section where it gives up heat to the incoming air. Condensed fluid returns to the evaporative section by gravity, assisted by an internal wick, which is a porous surface or circumferentially spiralled groove of proprietary design. The process continues as long as a temperature differential exists between the gas stream and air. Pipes may be finned on air and gas sides to increase the heat transfer. Several advantages are claimed for heat pipes: 1. Due to the isothermal nature of the fluid the temperature along the length of the pipe, the gas steams will see a nearly uniform surface temperature which can be higher than the acid dew point of the flue gas. Therefore condensation of acid is avoided and hence the corrosion problems. 2. Since the heat transfer is between the liquid (having a high heat transfer coefficient due to phase change) and a gas stream, extended surfaces may be used. Extended surfaces are not justified when the heat transfer coefficient of the fluid inside and outside the tubes are nearly equal as in tubular air-heaters; Appendix A discusses this in depth; since extended surfaces are used for both gas and air streams, the size and hence weight of the unit is reduced. Also, the gas and air pressure drops will be lower. 3. Since gas and air sides are independent, the configuration or area of flow for gas and air sides could be different. If the gas side is dirty, a lower fin density could be used for gas side alone, while the air side could take advantage of higher fin density. There are several other applications of heat recovery systems not discussed here due to lack of space, but an attempt has been made to discuss the important applications and types of waste heat boilers.

93

Waste Heat Boilers

Heat pipe shown here Is !HIed 6to 10 deg from hofizoolal 10 Improve Bellon of wor1
Figure 1-51. Arrangement of Heat Pipes,

REFERENCES AND SUGGESTED READING V. Ganapathy, "Applied heat transfer," Pennwell books, Tulsa, 1982 2. V. Ganapathy, Ben Heil and Jack Rentz, "Heat recovery steam generator for Cheng cycle application," ASME Industrial power conference, 1988. 3. V. Ganapathy, "HRSGs for gas turbine applications," Hydrocarbon processing, Aug 1987 1.

94

Waste Heat Boiler Deskbook

4. V. Ganapathy, "Program computes fuel input, combustion temperature," Power Engineering, July 1986 5. V. Ganapathy, "Evaluating waste heat recovery projects," Hydrocarbon processing, Aug 1982 6. V. Ganapathy, "Quick conversion of NO x and CO emission rates," Chemical Engineering, Dec. 1989 7. Robert Smock, "Gas turbines dominate capacity ordering," Power Engineering, August 1989 8. Jason Makansi, "Gas Turbines," a special report, Power, March 1988. 9. Jason Makansi, "Combined cycle power plants," Power, June 1990 10. Steve Collins, "Factor 10 years of experience into new landfill-gas power plants," Power, July 1990 11. Akber Pasha, "Gas turbine Heat Recovery Steam generators for combined cycles: natural or forced circulation considerations." ASME paper 88-GT-142, 1988 12. ASME, PTC 4.4, "Gas turbine Heat recovery steam generators." 1981 13. Power Engineering, "New gas turbines show high efficiency, low NOx emission," Aug 1990 14. Power( "ID fan maintains HRSG output should gas turbine fail," Jan 1990, p 71 15. "High temperature turbine favors IGCC," Modern Power Systems, Aug 1990, p 19 16. Power, "Energy from wastes," March 1987, special report 17. Power, "Energy from wastes," March 1988, special report 18. Power, "Energy from wastes," March 1989, special report 19. Power, "Energy from wastes," March 1990, special report 20. Douglas Smith, "Integrated waste management systems are the only solution, Power Engineering, July 1990 21. Luis Fuica, "Boost power output from MSW plants with fired superheater," Feb 1990 22. David Hitchcock, "Solid-waste disposal: incineration," Chemical Engineering, May 21, 1979 23. Gershman, Brickner and Bratton, "Small scale Municipal solid waste Energy recovery systems," Von Nostrand Reinhold Company, 1986

Waste Heat Boilers

95

24. Douglas Smith, "Can recycling save waste-to-energy?," power Engineering, Sep 1988 25. H.H. Krause, "Chlorine corrosion in waste incineration," paper 401, Corrosion, March 1987 26. JoAnn E. Ward et aI, "Waste incineration and heat recovery," environmental progress, voI1"Feb 1982 27. S.F. Chou et aI, "High temperature corrosion of tube support and attachment materials for refuse fired boilers," ASME paper 85JPGC-pwr-41 28. P.L .Daniel et aI, "Fire side corrosion in refuse fired boilers," paper 400, Corrosion, March 1987 29. Thomas Ellitt, "Standard handbook of Power plant Engineering," McGraw Hill, 1989 30. Power, "Attraction grows for heat pipe air heaters in flue gas streams," Feb 1989 31. "Heat recovery application manual," Maxim bulletin 1431281 32. Roy Wood et aI, "Rotary kiln incinerators-the right regime," Mechanical Engineering, Sep 1989, p 78 33. Calvin Brunner, "Incineration," Chemical Engineering, Oct 12,1987 34. Carol Zera, "Waste to Energy-what it takes to succeed," Power Engineering, Nov 1988, p 49 35. Kimberly Roy, "Medical waste incineration," Hazmat world, June 1989, p 17 36. Zahid Khan, "Co-incineration of waste water sludge," Resource recovery, vol 3, 1987, p 51 37. Max Monheit, "Heat pipe heat exchangers in utility applications," ASME paper 88-JPGC/34, 1988 38. V. Ganapathy, "Chart estimates supplementary fuel parameters," Oil and Gas journal, June 25,1984.

Chapter 2

Fire Tube Boilers Fire tube boilers are widely used in chemical plants, refineries and in incineration systems, Figure 2-1. In this type of boilers, the hot gas stream which could be at very high pressure flows inside the tubes, while low pressure saturated steam is generated outside the tubes. They could be of single or of multi-gas pass design. In single gas pass design the hot gas stream enters at one end and leaves at the other. If the boiler length is a concern due to lack of space, the tube length could be reduced by going in for a two gas pass design, Figure 22.

Figure 2-1. Elevated drum fire tube boiler. [courtesy ABeD Industries] 97

98

Waste Heat Boiler Deskbook

Figure 2-2. Two gas pass fire tube boiler design. [courtesy ABeD Industries]

Fire tube boilers are generally less expensive for low capacity, lo~ fresSuresream systems compared~to}VateflllBeb()iIer?Tfie-"g1is pressure drop with-fire tube boilers is usually higher c()~pared to water tube type for the same duty, which can be made compact with extended surfaces. Fire tube boilers may be of single shell design or of elevated drum type. Single shell boilers are less expensive compared to elevated drum boilers; if the heat flux at the front end is high due to

-----------------~-----~~~~.

99

Fire Tube Boilers

either a high gas inlet temperature or high heat transfer coefficient (as in a reformed gas boiler in hydrogen plants) or both, or !f~):ligh :e.~:'~!r,,~!:'~~. _i~. ~esirc~~' an elevated. drum boiler with extern~l downcoWer~(l!l~~~is~rs !§jised, Figure 2-1. A separate steam drum permits one to use drum internals ~ required to achieve the desired steam purity. Single shell boilers ~~ the other hand have a small steam space and elaborate purifying equipment cannot be installed within the space available. Hence \ wet steam with a steam purity of 2 to 3 ppm may only be generated. I Guidelines for fire tube boilers: .-.-~ 1. Usually limited to low steam pressures, say a maximum of 1400 psig. ~or the san:~yE~~~!~U.h~_J..hicklle~~!!ir~~LfQr~l!b.es subject to external pressure iSl!}!!c;;!:thigller cOlllE(lte.ci~!Q _thCl.ttQrJ1!12~_., subject to internal pressure~ -The thick~ess~of the t~be sheet also increa:SeSWlth pressure. Thus the tube bundle weight and hence the cost increases steeply at higher pressures compared to water tube designs. 2. Suitable for high gas pressures. In hydrogen plants for example one comes across gas steams at very high pressure on the order of 300 to 600 psig; it is easy to handle these streams in fire tube boilers compared to water tube designs. The boiler surfaces have to be located inside large pressure vessels if a water tube design is used, making it very expensive. 3. If the gas stream is dirty and contains dust it is easier to handle the gas stream in a fire tube boiler as only the tubes have to be cleaned, whereas in a water tube boiler the casing as well as the external surface of the tubes get dirty and are difficult to clean. However if the gas stream has slagging constituents as in municipal solid waste incineration plants, cleaning the boiler is easier with water tube boilers with soot blowers or rapping mechanisms. If the front end of the tube sheet and the tube inlet get plugged in a fire tube boiler, online cleaning is difficult. One way of handling this situation is to use a multi-pass boiler. The first pass can be designed with very large tubes and the subsequent passes can have smaller tubes. 4. When a large duty has to be transferred with a low pinch point as in gas turbine exhaust boilers, the surface area and hence the length required and the gas pressure drop become very large and

I I

~.

____

~_

•• _ . . . -

••

-.,_._.,

c __ ' _ ' "

100

Waste Heat Boiler Deskbook

hence uneconomical for such applications. With water tube boilers ne can use finned tubes for clean situations and hence make the boiler compact with consequent lower gas pressure drop and cost. Par comparable velocities, the heat transfer coefficient inside the tubes is lower compared to that outside bare tubes; hence bare tube water (~be boilers are sometimes preferred to fire tube boilers. 5. Fire tube boilers can handle high gas temperature on the lorder of 24000 P if the tube sheet is properly designed. The tube sheet 'lis lined with refractory and ferrules are used to transfer the heat .I rflux at the tube sheet inlet into the tubes where the water steam mixture keeps them cool. Pigure 2-2 shows the arrangement. The use ,\of refractory minimizes the temperature differential across the tube sheet and hence is recommended when gas temperature exceeds say 1800oP. 6. High steam purity can be obtained by using an external elevated steam drum with internals. 7. :I:lcconomizer ,alld superheater can be added as required. In a water tube boiler it is easy to split up the evaporator and locate the superheater in a cooler temperature zone beyond a screen section, while with a fire tube the choice of location is either at the inlet or exit of the boiler .. A.:ite~nativelyCt~o'fireCt~be cboii~;:s '~o~ld bebuIitwith a superheater in between but this is an expensive proposition. 8. Since the water volume and weight in a fire tube boiler is more compared to a water tube boiler, the response to load changes is better with water tube boilers, while fire tube boilers are sluggish.

n I

".,c<

jIC

~.,~.

,-

~,~-~,.:,

••-

-

-

"",._"

.-,.---,--~

DESIGN PROCEDURE ) The basic information obtained through design calculations are ( the surface aNa, ~g:~, t!:1~£QIlfigurationand gas pressure drop, J which is an operating cost. These variables can thenEe'op'tlmized to ! achieve an overall low life cycle cost. The tube size is usually as~med to begin with; smaller tubes usually result in a shorter boiler; larger tubes are used if the gas stream is dirty. Car;bon steel tubes are usually used as the tube metal temperature is usually within 20 to 40 0 P of the saturated steam temperature; when the gas stream

-----------------------------------~-------~-----------------------------

Fire Tube Boilers

101

contains hydrogen, chrome moly steels are preferred depending on the partial pressure of hydrogen, as discussed in Chapter 1. Tube sizes vary from 1.5 in. to 3.0 in. outer diameter.Larger size is preferred if the gas stream has particulates. The tube length depends on the duty. 28 to 35 feet long boilers are not uncommon.

DESIGN CALCULATIONS The surface area S is given by the equation S = Q/U L1 T

(2-1a) (2-1b)

If the overall heat transfer coefficient U is based on tube inner diameter, the surface area is also based on tube inner surface. U could

also be converted to tube outer diameter basis, in which case the outer surface area should be used; engineers evaluating different options or offerings should remember the basis used or the fact that SiUi = SaUa· The overall heat transfer coefficient Vo is obtained from:

The boiling heat transfer coefficient ha outside the tubes will be very high on the order of 2000 Btu/sq ft hOP and hence the smaller tube side coefficient hi influences Ua.

DETERMINATION OF hi The tube side coefficient hi consists of two components namely the convective heat transfer coefficient he and the non-luminous, hn; hence hi = he + hn : usually the non-luminous coefficient hn is very small for flow inside tubes but may be computed by the method shown in Chapter 3 using tube inner diameter as the beam length; hi

~

102

Waste Heat Boiler Deskbook

is nearly equal to he, the convective heat transfer coefficient if the partial pressure of triatomic gases such as C02, H20 is small. The tube side coefficient is obtained from the basic equation of Dittus-Boelter:

Nu

=.023 Re 0.8 Pr O.4

(2-3)

Nusselt Number Nu = hi di/12k

(2-4)

Reynolds Number Re = 15·2 wl(Jl11i)

(2-5)

Prandtls Number Pr

=Jl Cplk

(2-6)

The fluid properties Jl , Cp,k are evaluated at the average bulk temperature of the gas. Substituting the above in Equation (2-3) and simplifying we have: he = 2.44 w .8 Fl/di 1.8

(2-7)

where

(2-8)

Factor Fl for air, flue gases are shown in Table 2-1. Table 2-1: Gas Data and Fl Factors for Air and Flue Gas Air Flue Gas temp,F Cp k C k Fl Fl Jl Jl p 200 400 800 1200 1600 2000

.2439 .2484 .2587 .2696 .2800 .2887

.05369 .01878 .0632 .02211 .0809 .0287 .0968 .0350 .1109 .0412 .1232 .0473

.1687 .1756 .1865 .2015 .2138 .2235

.2570 .2647 .2800 .2947 .3080 .3190

.0492 .0587 .0763 .0922' .1063 .1188

.0174 .0211 .0286 .0358 .0429 .0499

.1702 .1805 .1991 .2159 .2314 .2456

(flue gas analysis used above: % vol C02 = 12, H20 = 12, N2 = 70,02 = 6) A few examples will illustrate the use of the above table.

Fire Tube Boilers

103

Example 1: 200 pph of air at atmospheric pressure flows inside a tube of inner diameter 1.7 in at an average gas temperature of 600 o P. Determine he. Tube outer diameter is 2 in. Solution: Fl from above table at 600 0 P is 0.18.10. Substituting in (2-7),

he = 2.44 x 200. 8 x .1810/1.7 1.8 = 11.78 Btu/ sq ft h P gas velocity V = .05 WV/di 2 (2-9) density p = l/v = 492 x MW x Pg/[(359 x (4660 + Tg) x 14.7]. The gas pressure decreases by nearly 3.2% for every 1000 feet increase in altitude. MW is the gas molecular weight.

p = 28.90 x 492/(460 + 600)/359 = .03728Ib/cu ft (effect of elevation or gas pressure was neglected in the above computation of density.) Hence v =26.82 cu ft/lb and V = .05 x 200 x 26.2/1.7/1.7 = 92.8 fps The effect of gas pressure on he is not significant at low pressures say up to 500 psig. See Appendix E for more information on computing he at high pressures.

A DESIGN EXAMPLE Example 2: Determine the size of fire tube boiler required to cool 100,000 pph of flue gas from 13000 P to 474°P; gas analysis is % volume C02 = 12, H20 = 12, N2 = 70,02 = 6 and gas pressure is atmospheric. Steam pressure is 150 psig and feed water is at 220 oP. Tube size used is 2 x 1.77 in. (design pressure =200 psig). Assume that the fouling factors on gas side = .002 and steam side = .001 sq ft h P/Btu and the metal conductivity = 25 Btu/ft h P. Steam side boiling coefficient is assumed to be 2000 Btu/sq ft h P. Heat loss including margins may be assumed to be 2%.

Waste Heat Boiler Deskbook

104

Solution: Prom Table 2-1, the gas specific heat at the average gas temperature of.5 x (1300 + 474)= 887°P is 0.283. Note that computation of the tube thickness of tubes subject to external pressure is an involved procedure. Appendix P deals with the subject. Let us assume that the size 2 x 1.77 in. is adequate for the design pressure of 200 psig. The duty = 100000 x .283 x .98 x (1300 - 474) = 22.91 MMBtu/h; steam flow using 5% blow down is: W= 22.91 x 106/[(1195.5-188) + .05 x (338 - 188)] = 22570 pph. Enthalpy of saturated steam, feed water were obtained from steam tables, Appendix E. Let us compute he; from Table 2-1, factor Fl= .2027. A starting point in the selection of tube count N is the mass flow per tube, which ranges from 120 to 200 pounds/hour for a 2 in. tube. Let us use 600 tubes and proceed with the design.

w = 100000/600 = 167. Using Equation(2-7), he = 2.44 x 167. 8 x .2027/1.771.8 = 10.62 Btu/ sq ft h P

The method described in Chapter 3 may be used to compute hn, the non-luminous coefficient. It is small for fire tube boilers as the beam length is the tube inner diameter. Partial pressures of C02 and H20 are 0.12 each; the beam length = 1.77 in. = .1077 ft; for the average gas temperature of 887°P or 1347 R, from Equation(3-17), of Chapter 3, E g = .058. The metal temperature may be assumed to be 20 0 P higher than the steam temperature; this will be verified later but the effect of this is insignificant. hn = .58 x 0.173 x .9 x (13.474 - 8.44)/(1347 - 840) = .49 Btu/sq ft hOP

Hence hi = 10.62 + .49 = 11 .11 Btu sq ft h 0p l/Uo = 2/(1.77 x 11.11) + 002 x 2/1.77 + .001 + In (2/.1.77) x 2/(24 x 25) + .0005 = .1017 +.00226 + .001 + .0004 + .0005 = .105865

-- "1

Fire Tube Boilers

105

The various resistances (in sq ft h of IBtu) are: due to gas side heat transfer = .1017 gas side fouling =.00226 metal resistance = .0004 outside fouling =.001 outside heat transfer = .0005 Hence Uo =9.45 Btu/sq ft hOF In order to compute Ui based on tube inner surface area, use the relation Ui x di = Uo x do hence Ui = 9.45 x 2/1.77 = 10.68 Btu/Sq ft h of The log-mean temperature difference L1 T = (1300-474)lln {(1300-366) (474-366)} = 383°F (the saturation temperature at 150 psig is 366 OF) The required Surface area Si =22.91 x 10 6I (383 x 10.68) =5560 sq ft 3.14 x 1.77 x 600 x L/12; hence the length L = 20 ft

=

In order to evaluate the gas pressure drop, the following expression may be used. L1 Pi = 93 x 10"-6 x fx (L + 5 di) x w2v/df

(2-10a)

The factor 5di is used to account for the inlet and exit losses; a more accurate method would be to compute the losses based on inlet and exit gas velocities. If ferrules are used at the gas inlet the gas velocity at the ferrule will be very high and the inlet loss is taken as .5 times the velocity head; the exit loss from the tube is taken as 1 x velocity head, which is computed as follows. Velocity head VH = P V2 x 12/(62.4 x 2 x 32)

(2-10b)

The friction factor f in tubulent flow depends on tube inner diameter, see Table 2-2 below.

106

Waste Heat Boiler Deskbook

Table 2 - 2: Friction Factors Versus Tube Inner Diameter di, in

.5.75

1

1.5

2

.028 .0245

.023

.021

.0195

2.5

3

4

.018 .0175 .0165

5 .016

Density at the average gas temperature of 887°P and at atmospheric pressure is: p = 28.96 x 492/[359 x (887 + 460)]

= .0294Ib/cu ft; v =34.4 cu ft/lb

f = .02, w = 167, L = 20, di =1.77; substituting in (2-10): L1 Pi = 93 x 10--6 x (167)2 x .02

x (20 + 5 x 1.77) x 34.4/1.775 = 2.96 in we

Computing the losses using inlet and exit velocities, we have the following: Loss through the tube = (20/28.55) x 2.96 = 2.05 in wc Gas density at inlet = 28.96 x 492/(359 x 1760) = .02255Ib/ cu ft V = .05 x w/d2 = .05 x 167/(.02255 x 1.77 x 1.77) = 104.5 ft/ s .5 VH = .5 x 104.5 x 104 .5 x . 02255 x 12/(62.4 x 64) = .37 in we The gas density at exit = .0425 Lb/cu ft The gas velocity = 62.7 ft/ s VH =.5 in we Total loss through the tubes =2.92 in we To this one must add the turning loss due to inlet or exit duct work connection. Once the tube count and length are determined, the shell diameter may be found for elevated drum boilers. Tubes may be in triangular or square pitch. Por triangular pitch it can be shown that Ds = 1.l6Pt'v'1-'[

(2-11)

and for the square pitch, Ds = 1.25Pt'vN

(2-12)

107

Fire Tube Boilers

The above formulae do not take into account the space required for refractory or steam separation. Depending on the thickness of refractory used, the shell diameter could be arrived at. If a single shell boiler were used the shell diameter is determined by the steam space used. If a multi-pass design were used, a tube sheet layout has to be done to arrive at the shell diameter. The gas pressure drop of multi-pass units would be higher, considering the additional turns and the bends the gas would have to take. In order to compute the tube wall temperature, the heat flux must be known. Again this may be based on tube inner or outer surface. As the steam water mixture is on the outside and critical heat flux or heat flux to avoid DNB (departure from nucleate boiling) refers to the steam side, let us compute the heat flux on outer surface basis, qo. (2-13)

One could compute qo along the gas path or at the gas inlet; local Uo and Tg have to be used in these calculations. Average qo = 9.45 x (887 - 366) = 4923 Btu/sq ft h The temperature drops across the various films are computed based on their resistances; see the computation of U o: across the gas film = 4923 x .1017 = 500°F across the inside fouling

=4923 x .00226 =11°F

across the metal wall = 4923 x .0004 = 2°F hence the tube outer wall temperature =887 - 500 -11-1

=374°F

There is no need to compute hn again as the effect of this wall temperature is insignificant. The heat transfer coefficient U 0 at the gas inlet would be higher due to the higher temperature, velocity and gas properties. Hence the heat flux would be higher. However considering the fact

Waste Heat Boiler Deskbook

108

the heat flux to avoid DNB is in excess of 100,000 Btu/ sq ft h, it is not a concern in this design. With high tube side coefficients as in hydrogen plants, it is prudent to compute Uo and qo at the tube sheet inlet. One of the important variables in any design is the operating cost associated with moving the gas through the tubes. Assuming that an induced draft fan is used at the boiler exit, the fan power consumption could be found as follows: P = 1.96 x 10-6 X Wi L1 PJ(1] p)

(2-14)

where Wi = 100,000, L1 Pi = 3.2 . The fan drive efficiency 1] is assumed to be 0.60. The gas density at the exit is:

p = 28.96 x 492/[359 x (474 + 460)] = .0425Ib/cu ft substituting in (2-14) ,we have: P = 1.96 x .1 x 3.2/(.0425 x .6) = 24.6 kW

Using the operating cost and the installed cost, which is impacted by the tube count and length, one could evaluate the life cycle cost; several alternates could be studied by varying the tube size or velocity and one could arrive at various configurations. Life cycle costing could be then used to arrive at the optimum design.

EFFECT OF TUBE SIZE ON DESIGN Example 3: Determine the effect of varying the flow per tube or velocity and the tube size for the same duty. Solution: Using a computer program, several alternatives were studied and are presented in Table 2-3 below. The gas inlet velocity was varied from 98 to 162 fps. The following conclusions may be drawn: 1. As the gas velocity increases the surface area reduces and gas pressure drop increases significantly. 2. Smaller the tube, shorter the boiler and smaller the surface area for the same gas pressure drop. This conclusion is important

109

Fire Tube Boilers

Table 2-3: Effect of Tube Size and Gas Velocity 1. size 1.75 X 1.521 2. vel,fps 98 123 163 3.tubes 1000 800 600 4. len, 1t 15.7516.7518 5. Si, Sq ft 6269 5333 4298 19.4711.0813.70 6.Ui

2.0 X 1.77 98123162 750 600 450 18.752021.5 6513 5558 4480 9.0710.6813.19

2.5 X 2.238 98123162 470 375 280 24.75 26.0 28.50 681257104673 8.73 10.29 12.72

1.973.206.00 15.0 24.4 45.6

1.953.166.00 14.924.045.6

7. APi,

in we 8. P, kW

2.05 3.34 6.23 15.6 25.4 47.4

when one wants to fit a boiler into a small space. By going in for smaller tubes (if the cleanliness permits), one could probably fit it in. Note that the length increases to 26 ft from 16.75 ft for an approximate 3.3 in. wc pressure drop when we increase the tube size from 1.75 in to 2.5 in. Surface area for the smaller tube is 5333 sq it versus 5710 for the larger tube. 3. Alternatively, higher the gas velocity results in a longer boiler for the same size. The length increases from 18.75 to 21.5 for a 2 in. tube when the inlet gas velocity changes from 98 to 162 fps. Note that a smaller shell, lesser tube count could mean a less expensive boiler though the length and the gas pressure drop could be more. The operating cost has to be considered in making the decision as it increases significantly from 15 to 45 kW. 4. In general it could be said that the % variation in length has more effect on the performance compared to the same % change in tUbe count.

SIMPLIFIED APPROACH TO DESIGN A simplified approach has been developed by the author to quickly arrive at the design of fire tube boilers. A few equations have been developed based on the assumption that the tube side coefficient is equal to the convective heat transfer coefficient and that the non-luminous heat transfer is insignificant; the analysis is adequate for engineering purposes.

Waste Heat Boiler Deskbook

110

1. In a fire-tube boiler, gas flows inside of the tubes; steamwater mixture flows on the outside. The gas heat-transfer coefficient is small, about 10-20 Btul (ft 2 ) (h) (OF), compared to the outside coefficient of 2,000-3,000 Btul (ft 2 ) (h) (OF).The metal resistance is also small and, hence, the gas-side coefficient governs the overall coefficient and the size of the equipment. Since the inside coefficient governs U, we can rewrite Equation (2~2) as follows (neglecting lower-order resistances, such as ho, metal resistance, and fouling factors, which contribute to about 5% of U):

U = 0.95 hi di I do

(2-15)

Combining Equations (2-1) (2-2) and (2-7) , we have, after substituting S =3.14diLN /12, and for flow per tube w = Wi/N; I

QI (11 TFl Wi 0.8) =0.606 LN 0.21 di 0.8

(2-16)

This simple equation relates several important variables. Given Q, 11 T, Wi and Fl, ' one can try combinations of L,di and N to arrive at a suitable configuration. Also, for given thermal data, (LN 0.2/di 0.8) is constant in Equation (2-16). is shown in Table 2-1 for fluegas and air. For other gases, Fl may be computed from Equation (2-8) . Fl

When a phase change occurs, as in a boiler, .t1T is written as: .t1T

=[(Tl -

ts)- (T2 - ts)]1

In [(T1 - ts ) - (T2 - t s)]

(2-17)

Combining Equations (2-1), (2-16) and (2-17), and simplifying we arrive at the following expression: In[(Tl - ts)j(T2 - ts)J

Factor Fl

=0.606 (Fl ICp) NO.2 L/Wi 0.2 di 0.8

ICp is given in Table 2-4.

(2-

111

Fire Tube Boilers

Equation (2-18) relates the major geometric parameters to thermal performance. Using this method, one need not evaluate heat,..transfer coefficients. Now, consider pressure drop. The equation that relates the geometry to tubeside pres~ure drop in in. H20 is APi = 9.3 x 10-5 j(WiIN) 2 (L + 5di) v/di 5 =9.3 x 1O-5(Wi/N) 2 K 2V

(2-19) (2-20)

where Combining Equations (2-18) - (2-20) and eliminating N : In[Tl - t s)/(T2 -

where

t s)J =0.24 (Fl/e p) K 11.) 0.1 I APi 0.1

K 1 =(L + 5 di)O.1 Lj.1 Idi 1.3

(2-21) (2-22)

K 1 and K 2 appear in Tables 2-1 and 2-6, respectively, as a

function of tube 1.D. and length. In the turbulent range, the friction factor for cold-drawn tubes is a function of 1.D. Using Equation (2-22), one can quickly figure the tube diameter and length that limit tube pressure-drop to a desired value. Any two of the three variables N , Land di determine thermal performance, as well as gas pressure drop. Let us discuss the conventional design procedure: L Assume w, calculate N. 2. Calculate U, using Equations (2-15) and (2-8) . 3. Calculate L after obtaining S from Equation (2-1). 4. Calculate APi from Equation (2-10). If the geometry or pressure drop obtained is unsuitable, repeat Steps 1-4. This procedure is lengthy.

Some examples will illustrate the simplified approach. The preceding equations are valid for single-pass design; however, with minor changes, one can derive the relationships for multipass units (e.g., use length = L/2 for two-pass units).

Waste Heat Boiler Deskbook

112

Table 2-4. Factor F1 ICp, F2/Cp for Air and Fluegas

Air

Temperature, of 100 200 300 400 600 1,000 1,200

FI/Cp

F2/Cp

0.6660 0.6870 0.7068 0.7225 0.7446 0.7680 0.7760

0.3730 0.3945 0.4140 0.4308 0.4591 0.4890 0.5030

Fluegas Temperature, of 200 300 400 600 800 1,000 1,200

Fl/Cp

F2/Cp

0.6590 0.6780 0.6920 0.7140 0.7300 0.7390 0.7480

0.3698 0.3890 0.4041 0.4300 0.4498 0.4636 0.4773

(Fluegas is assumed to have 12% water vapor by volume) Table 2-5. Values of Kl as a Function of Tube Diameter and Length di, in.

L, ft

8 10 12 14 16 18 20 22 24 26 28

1.00

1.25

1.50

7.09 8.99 10.92 12.89 14.88 16.89 18.92 20.98 23.05 25.13 27.24

5.33 6.75 8.20 9.66 11.14 12.65 14.16 15.70 17.24 18.80 20.37

4.22 5.34 6.48 7.63 8.80 9.98 11.17 12.38 13.59 14.81 16.05

2.00

22.5

2.50

2.75

3.00

3.46 2.92 4.38 3.70 5.31 4.48 6.25 5.27 7.21 6.07 8.17 6.88 9.14 7.70 10.12 8.52 11.11 9.35 12.11 10.19 13.11 11.00

2.52 3.17 3.85 4.53 5.21 5.91 6.60 7.31 8.02 8.74 9.46

2.20 2.78 3.36 3.95 4.55 5.15 5.76 6.37 6.99 7.61 8.74

1.95 2.46 2.98 3.50 4.02 4.56 5.10 5.64 6.19 6.74 7.30

1.75 2.21 2.67 3.14 3.61 4.10 4.56 5.05 5.54 6.03 6.52

1.75

113

Fire Tube Boilers

Table 2-6. Values of K2 as a Function of Tube Diameter and Length din

L. fl

1.00

1.25

1.50

1.75

2.00

2.25

2.50

2.75

3.00

8

0.2990

0.1027

0.Q428

0.0424

0.0109

0.0062

0.0037

0.0024

0.0016

10

0.3450

0.1171

0.0484

0.0229

0.0121

0.0069

0.0041

0.0027

0.0018

12

0.3910

0.1315

0.0539

0.0252

0.0134

0.0075

0.0045

0.0029

0.0019

14

0.4370

0.1460

0.0595

0.0277

0.0146

0.0082

0.0049

0.0031

0.0021

16

0.4830

0.1603

0.0650

0.0302

0.0158

0.0088

0.0053

0.0033

0.0022

20

0.5750

0.1892

0.0760

0.0350

0.0183

0.0101

0.0060

0.0038

0.0025

22

0.6210

0.2036

0.0816

0.0375

0.0195

0.0108

0.0064

0.0040

0.0027

24

0.6670

0.2180

0.0870

0.Q400

0.0207

0.0114

0.0067

0.0042

0.0028

26

0.7130

0.2320

0.0926

0.0423

0.0219

0.0121

o.oon

0.0045

0.0030

28

0.7590

0.2469

0.0982

0.0447

0.0231

0.0217

0.0075

0.0047

0.0031

Example 4: A firetube waste-heat boiler will cool 66,000 lb/h of fluegas, from 1,160 to 440o P. Saturation temperature is 350o P. Molecular weight is 2B.5, and gas pressure is atmospheric. If L is to be limited to 20 ft due to layout, determine Nand t:.Pi for two tube sizes: 1. 2 x 1.77 in. (2 in. O.D. 1.77 in. I.D.) and (2.) 1.75 x 1.521 in. Solution: Use Equation (2-18)to find N. Use 2 in. tubes. (FI/e p) Prom Table 2-4 is 0.73 for fluegas at the average gas temperature of 0.5 (l,i60 + 440) = BOOoP. In [1,160 - 350)/(440 - 350)] = 2.197 2.197 = 0.606 x 0.73 x NO.2 x 20/(66,000 0.2 x 1.77 0.8) = 0.60B9 NO.2, N = 611.

Compute L!Pi using Equation (2-19) . Prom Table 2-5, K2 is 0.035. Compute the gas specific volume. Density (p) =2B.5 x 492/[359 x (460 + BOO)] = 0.031lb/ft3 or v =32.25 ft 3/lb. Substituting in Equation (2-19 we have: L!Pi = 9.3 x 10-5 x (66,000/611)2 x 0.035 x 32.25 = 1.23 in. H20

Waste Heat Boiler Deskbook

114

Repeat the exercise with 1.75-in. tubes; length remains at 20 ft. From Equation(2-18), we note that for the same thermal performance and gas flow, NO.2 L/di 0.8 = a constant. The above concept comes in handy when one wants to quickly figure the effect of geometry on performance. Hence: 611°. 2 x 20/1.nD· 8 = N 0..2 x 20/1,521°.8 or N = 333. With smaller tubes, one needs fewer for the same duty. This is due to a higher heat-transfer coefficient; however, the gas pressuredrop would be higher. From Table 2-5, K2 = 0.076 for 1.521-in. tubes. From Equation (2-19): .1Pi = 9.3 x 10-5 x (66,000/333)2 x 0.076 x 32.25 = 8.95 in. H20.

Example 5: Size heat exchanger for 2.0 in. tubes, with a pressure drop of 3.0 in. H20. For the same thermal performance, determine the geometry. Solution: Using the conventional approach would take several trials to arrive at the right combination. However, with Equation (2-21), one can determine the geometry rather easily: In[(1,160 -350)/(440-350)]

= 2.197 =0.24(FlICp)K1VO.1 / t1Pp·1

From Table 2-4, (FlICp) = 0.73; t1Pi = 3, v = 32.25. Then: In[(1.160-350)/440-350)] =2.197= 0.24K132.250.1 x 0.73/3°.1 =0.222Kl, Kl

=9.89.

From Table 2-5, we can obtain several combinations of tube diameter and length, that have the same Kl value, and that would yield the same thermal performance and pressure drop. For the 1.77 in. I.D. tube, L is 21.75 ft. Use Equation (2-18) to calculate the number of tubes: 2.197 = 0.606 x 0.73 x 21.75 x ~.2/(66,OOOO.2 x 1.nD·8), or N = 402.

115

Fire Tube Boilers

Thus, several alternative tube geometries can be arrived at for the same performance, using the preceding approach. One saves a lot of time by not calculating heat-transfer coefficients and gas properties.

LlFE-CYCLE COSTING Such techniques determine the optimum design, given several alternatives. Here,the major operating cost is from moving the gas through the system, and the installed cost is that of the equipment and auxiliaries, such as the fan. The life-cycle cost is the sum of the capitalized cost of operation and the installed cost. Lee = Ceo +Ie

(2-23)

The capitalized cost of operation: T

Ceo = CaY(l -Y )/(1 - Y)

where: Y

= (1

+ e)/(J + i)

(2-24) (2-25)

The annual cost of operating the fan is estimated as: Ca

=0.001 PHCe

(2-26)

where the fan power consumption in kW is: (2-27)

The above procedure is used to evaluate Lee. The alternative with the lowest Lee is usually chosen if the geometry is acceptable. Example 6:

Evaluate the three options in Examples 4 and 5, and determine the best one. All three provide the same thermal performance (gas flows, alld inlet and exit temperatures are the same). However, the

Waste Heat Boiler Deskbook

116

pressure drops and tube geometries are different. Use the following data: e = 0.06, i = 0.10, T = 10 yr., Ce = 25 mills/kWh, fan efficiency is 0.60, and the system operates for 6,000 h/yr. The installed cost of each is shown in Table 2-7. Calculate Lee for the first case: di =1.77 in., N =611, L =20ft, L'lPi = 1.23 in. H20, v = 32.25 ft3 /lb. P = 1.9 x 10-6 x 66,000 x 1.23 x 32.25/0.6 = 8.29 kW

Ca = 0.001 x 8.29 x 6,000 x 25 = $1,245/yr Y =1.06/1.1

=0.9636

Ceo=1,245 x 0.9636 x (1 - 0.9636 1°)/0.0364

=$10,212.

Hence Lee = 10,212 + 75,000 =$85,212. Table 2-7 summarizes the results. Alternative 3 has the lowest Lee and hence should be selected, if the geometry is acceptable.

Table 2-7. Life-Cycle Costing Finds The Most Economical Alternative Alternative

1 2x1.77 611 20

2 1.75 x 1.521 333 20

3 2 x 1.77 402 21.75

Power, kW Ca $/yr

1.23 8.3 1,245

Cea ,$· le,$ Lec ,$

10,212 75,000 85,212

8.95 60.3 9,045 84,410 50,000 134,410

3.0 20 3,000 24,609 58,000 82,609

Tube size, in. Number of tubes Length, ft. Gas pressure drop,

in.H(lO

Fire Tube Boilers

117

PREDICTING THE PERFORMANCE OF A GIVEN BOILER Once a boiler has been designed, its geometry is fixed. Often a plant operates under different load conditions which results in different gas flow, inlet temperature and gas analysis or even stearn pressure. How does the given boiler perform under these conditions? What is the new exit gas temperature or duty or stearn production? There is another important need to predict the boiler performance under different gas or stearn conditions. That is to check if there is fouling or accumulation of deposits on the outside or inside boiler tubes which can reduce the duty transferred. If the exit gas temperature is higher than predicted, the reason could be fouling on the gas or stearn side. The following example illustrates the procedure to predict the performance of a given boiler. Example 7: The boiler designed in Example 1 has to operate under the following conditions: Gas flow = 75000 pph; inlet temperature = 1400oP; Stearn pressure = 175 psig; feed water temperature = 220oP. Determine the exit gas temperature, duty and stearn production. Let us use the 2 x 1.77 tube size, 600 tubes 20 ft long design. Assume that the fouling factors and gas analysis are same as before. Strictly speaking, a trial and error procedure is required to solve for the performance; it consists briefly of the following steps: 1. Assume the exit gas temperature; calculate the assumed duty 2. Determine U at the average gas temperature. 3. Determine the log-mean temperature !:iT. 4. Compute the transferred duty Qt = US!:iT. 5. If Qa and Qt are equal the iteration stops else steps 1 to 5 are repeated using a revised exit gas temperature based on Qt.

Waste Heat Boiler Deskbook

118

However an iteration can be avoided as seen below: (2-28) (2-29) combining (2-28) and (2-29) and simplifying, we have:

= Uo So lin [ ( Ti - ts ) I (T2 -ts)] or in[( Ti -ts) I ( T2 -ts)] = Uo So I WiXCp X hi

WiX CpX hi

(2-30)

In the above equation T2 is the unknown; U o is computed assuming that the average gas temperature is same as before and then corrected; since variations in average gas temperature of up to 30 to 40 0 P do not affect Uo significantly, an iteration can be avoided. Solution: Saturation temperature at 175 psig = 378°P. Assume that the exit gas temperature is 4S0 o P. Por the average gas temperature of 925°P, Fl =.024, Cp =.2845; w =75000/600 =125lb/h. he= 2.44

X

125,8 x .204 / 1.771.8 = 8.48 Btu/sq ft h F

Let us assume that hn does not change, which is reasonable. hn = .49. Hence, 1/Uo - (2/(1.77 x 8.97) + .002226 + .0004 + .001 + .0005 =.13 Uo = 7.68 Btu/sq ft hF (as other resistances do not change) The external surface area So = 3.14 x 2 x 600 x 20/12 = 6280 sq ft Log-mean temperature difference = L1 T

=[(1400 - 378) - (450 - 378)]/ln[(1400 - 378)/(450 - 378) =358°P

Assumed duty Qa = 75000.98 x .2845 x (1400-450) = 19.86 MMBtu/h

J

Fire Tube Boilers

119

Transferred duty =Qt = 7.68 x 6280 x 358 = 17.27 MMBtu/h It appears that the boiler cannot transfer the assumed duty but something less than that. Another iteration may be performed with a lower Qa. Now let us use Equation (2-30) and see how iteration is avoided. In [(1400 - 378)/(T2 - 378)] = 7.68 x 6280/(75000 x .2845 x .98) =2.3065 or (1400 - 378)/(T2 - 378) =exp (2.3065) =10.04

In order to check this, compute Qa and Qt. Qa =75,000 x 2854 x .98 x (1400 - 480) =19.25 MMBtu/h tJ.T = [(1400-378) - (480-378)]/ln (1022/102) = 399 Qt =399 x 6280 x 7.68 = 19.23 MMBtu/h and hence close.

The enthalpy addition to steam is: (1197.6-188) + .05 x (351-188) =1017.75 Btu/lb. Hence the steam production =19.23 x 106/1017.75

=18895 pph.

One can revise the calculations based on the new average gas temperature but the variation would be marginal.

SIMPLIFIED APPROACH TO PREDICTING PERFORMANCE One can simplify Equation(2-30) further and arrive at the following equation in order to predict the performance of fire tube boilers. A few assumptions have to be made: 1. The fouling factors and the non-luminous heat transfer coefficient are not high and that the overall heat transfer coefficient is proportional to the convective coefficient he ..

Waste Heat Boiler Deskbook

120

2. The gas temperature profile does not vary significantly in the boiler so that the effect of variations in gas properties can be neglected. The above assumptions are fairly reasonable. Let us apply them to arrive at the following expression. From (2-30) assuming that Uo is proportional to Wi· 8 and treating all other terms on the right hand side as a constant for a given boiler, we have: (2-31)

Example 8: Use Equation (2-28) to arrive at K factor from Example 1 and predict the performance of the boiler for conditions in Example 5. Let us compute K factor. Using the design data of Example I:

In [(1300 - 366) /(474-366)] = In (8.648) = 2.1573 = K x 100000 -.02 or K = 21.573.

Let us use the new conditions:

Using (31), In [(1400 - 378)/(T2 - 378)] =,21.573 x 75000 -.2 = 2.285 or T2 = 482°F, which is close to what we obtained earlier. Thus Equation (2-31) gives reasonably good results for purposes of quick estimates. The pressure drop may be computed as before. Checks for Fouling: A boiler may deteriorate in performance due to accumulation of deposits inside or outside of tubes over a period of time. This may be due to the nature of the gas, which contains dusty or fouling components, which cannot be removed on line or due to the build up of salts outside the tubes due to poor water chemistry.

121

Fire Tube Boilers

One has a tendency to use poor water chemistry in fire tube boilers as they are low pressure boilers. If one analyzes Equation (21), one can see that fouling factors do influence the overall heat transfer coefficient and hence the performance. Poor water chemistry leads to the formation of scale and sludge. Scale is a relatively hard and adherent deposit, while sludge is softer and less adherent. The buildup of boiler scale is most severe in those areas in which the maximum heat transfer occurs. Scale buildup is associated with compounds whose solubilities decrease with increasing temperatures. Conversely sludges are precipitated directly from the boiler water when their solubilities are exceeded. Scale and sludge increase the resistance to heat transfer and decrease the overall heat transfer coefficient and most importantly increase the tube wall temperature. The thermal conductivities of some scale materials are given in Table 2-8 below. Table 2-8 Thermal Conductivities of Scale Materials Material

Thermal Conductivity, Btu/sq ft h f/in.

Analcite Calcium Phosphate Calcium Sulfate Magnesium Phosphate Magnetic Iron Oxide Silicate scale (Porous) Boiler Steel Firebrick Insulating brick

8.8 25 16 15

20

.6 310

7

.7

EFFECT OF SCALE ON BOILER PERFORMANCE Example 9: The boiler in Example 2 is assumed to have a .03 in thick silicate scale. Determine the performance when 100,000 pph of gas at 13000 P enters the boiler. Stearn pressure as before is 150 psig and feed water temperature is 220oP.

Waste Heat Boiler Deskbook

122

Solution: The outside fouling factor ffo is obtained first. ffo = thickness/conductivity = .03/.6 = .05 sq ft h F/Btu. Let us assume that hc,h n remain unchanged, though they would be slightly more due to the higher average gas temperature in the boiler. Compute U o using the other resistances as before. l/Uo =2/(11.11 x 1.77) + .05 + .00226 + .0005 + .0004 = .1532; Uo = 6.53. Using Equation (2-30) :

=6.53 x 6280/(100000 x .2845 x .98) =1.47 or (1300-366)/ 4.349 =(T2 - 366) or T2 = 580°F.Hence Qt = 100000 x .2845 x (1300-580) = 20 MMBtu/h. Average heat flux qo = 6.53 x (940-366) = 3748 Btu/sq ft h. Drop across outside heat transfer film = .0005 x 3748 = 2°F. Drop across scale = .05 x 3748 = 188°F. Hence outer tube wall temperature =366 + 188 + 2 =556°F. In[(1300-366)/(T2-366)]

This is significantly higher than before. Fouling inside the tubes on the other hand will also result in lower U o, lower duty and possibly increase in gas pressure drop, which can be monitored. For example in the above example when the gas flow was 75000 pph and inlet temperature was 1400°F, if the exit gas temperature had been measured to be 500°F versus the 480°F computed, surely one has to analyze the performance closely and see if the fouling is due to the gas side or water side. A shut down cleaning may be necessary. Fouling is an operational cost and affects the duty and steam production and even the cost of moving the gas through the boiler and must be attended to. One can also compute the fouling factor by using the following procedure. The expression for U may be written as: (2-32)

Compute A,K and ff based on three sets of performance data over a period of time. Keep recording data and continue to estimate ff over a period of time using K,U and A. If it shows an increasing trend, there is fouling and a shut down cleaning is warranted.

-----------~---~·~----------------~-~-----~-~--~""'1

123

Fire Tube Boilers

HYDROGEN PLANT BOILERS As discussed in Chapter I, fire tube waste heat boilers are used in hydrogen plants to cool reformed gas and maintain a desired exit gas temperature at all loads, generally through internal gas bypass, Figure 2-4. The high tube side heat transfer coefficient results in a high overall heat transfer coefficient and hence heat flux at the tube sheet inlet. The following example illustrates the fact. Example 10: In a hydrogen plant, 50,000 pph of reformed gas at 1650°F is cooled in a boiler to 650°F. Gas analysis is: % vol C02 = 5, H20 = 30, CO = 10, H2 = 52, CH4 = 3, N2 =.1. Gas and steam pressures are 285 psig and 635 psig respectively. Determine the heat flux and boiler performance when 190 chrome moly steel tubes of size 1.5 x 1.14 in. and 20 feet long are used. Feed water is in at 400°F. Fouling factors on gas side = .003 and steam side = .001. Solution: The gas properties must be evaluated. Using the methodology discussed in Appendix E, the following data were arrived at:

Table 2-9 Data For Process Gas Temp, of 1650 1150 650

Cp

Ji

k

.7517 .7080 .6640

.089 .0723 .0535

.1315 .1031 .0751

Let us estimate he. W = 50000/190 = 263. Hence he = 2.44 x 263. 8 x(.708/ .0723)·4 x .1031. 6/1.141.8 = 106 Btu/ sq ft hF. hn may be shown to .88 using methods described earlier. l/Uo = (1.5/1.14/106.88) + .001 + .003 x 1.5/1.14 + .0005 + .00069 = .0123 + .001 + .003947 + .0005 + .00069 = .01845. (steam side coefficient = 2000 Btu/sq ft h F and metal resistance = .00069). Hence Uo = 54.2 Btu/sq ft h F.

124

Waste Heat Boiler Deskbook

Substituting in (2-30), In [(1650-495)/(T2_495)] = 54.2 x 1492/(50000 x .99 x .708) = 2.307 (heat loss = 1% and So = 3.14 x 1.5 x 190 x 20/12 = 1492 sq ft. Hence T2 = 610 oP; since the desired exit temperature was 650 oP, it appears that the boiler is oversized and hence gas has to be bypassed. The average heat flux is: U o x (tg-t s) = 54.2 x (1150-495) = 35500. As the inlet it can be shown that the clean Uo = 78 Btu/sq ft h P, neglecting the fouling factors. Hence the maximum heat flux at inlet = 78 x (1650-495) = 90000 Btu/ sq ft h. Heat fluxes of up to 100,000 to 150,000 Btu/sq ft h can be permitted depending upon the steam pressure and cleanliness of water. It should be noted that the tube wall temperature will be higher if scale forms outside the tubes. The average tube wall temperature is estimated as follows: Drop across the steam-water mixture = 35500 x .0005 = 18°P. Drop across steam side fouling layer = 35500 x .001 = 36°P. Hence outer tube wall temperature = 495 + 18 + 36 = 549°P. If there is scale formation, the tube wall temperature will be much higher as shown in an earlier example.

GAS BYPASS CALCULATIONS In process boilers such as those used in hydrogen or sulfuric acid plants, the exit gas temperature from the boiler has to be maintained within a range at all loads. This is usually achieved by bypassing a portion of the gas around or through the boiler. When gas is internally bypassed, it transfers little energy to the steam water mixture as the bypass pipe is usually protected inside by a liner material. Hence the bypass gas remains at a high temperature close to the gas inlet temperature; the rest of the gas is cooled through the boiler tubes and is at a low temperature; the hot bypass gas then mixes with the cooler boiler gas and achieves the desired exit gas temperature. The bypass flow is decided by a temperature controller at boiler exit. Pigure 2-4 shows a typical arrangement. The exchanger diameter increases due to the inclusion of the bypass pipe within the tube bundle.

Fire Tube Boilers

125

When an external bypass is used, it works in the same way with the difference that the bypass pipe is external to the boiler; the exchanger diameter is smaller; however the bypass pipe operates at a much higher gas temperature and has to be protected with refractory; the damper is also expensive. Example 11: Determine the amount of gas to be bypassed in Example 10 to achieve 650 0 P exit gas temperature.

Solution: Let the gas flow through boiler = 45,000 pph and 5000 pph through bypass. Estimate T2 using (2-30): he = (45/50)·8 x 106 = 97.4; h n = .88; U o = 51 Btu/sq ft h F. In [(1650-495)/(T2_495)] = 51 x 1492/(45000 x .99 x .708) = 2.408. Hence T2 = 599°P. The approximate mixed temperature = (45000 x 599 + 5000 x 1650)/50000 =704°P; this is higher than 650 0 P required. (A more accurate method of estimating the mixed temperature is by using the gas enthalpies; however the above is for illustration purposes only). Since the exit gas temperature is higher, more gas must be cooled in the boiler. Try 48000 pph through boiler and 2000 pph through bypass. Uo = 52.73; T2 = 606°P; mixed temperature = 648°F. This is acceptable. Hence the bypass flow is 2000 pph and has to be sized accordingly. One must also check the boiler performance at part loads and determine the bypass flow for each load and see if the control mechanism can handle the flow based on the resistance of ~ain and bypass paths.

DETERMINING HEAT LOSSES FROM BOILER We have been assuming a heat loss on the order of 1 to 2% in the boiler. However the loss could be computed more accurately as follows. Example 12: A fire tube boiler is 6 ft in diameter and 20 ft long and operates at 150 psig. Estimate the heat loss from the shell. 3 in. of mineral

Waste Heat Boiler Deskbook

126

fiber insulation is used. Ambient temperature is 70°F and wind velocity is 100 fpm. Casing emissivity is .9. Boiler duty is 23 MMBtu/h. Solution: Using the program described in Appendix B, for a saturation temperature of 366°F it maybe shown that: heat loss from casing = 29.6 Btu/sq ft h and the casing temperature is 86°F. See printout below, Figure 2-5. The total loss from the casing = 3.14 x 6 x 20 x 29.6 = 11155 Btu/h. The heat loss = (11155/23 x 106) x 100 = .05%. If the inlet and exit vestibules are considered along with an external drum, downcomers, risers, the heat loss will be more and can be computed by summing up the heat loss in each section. The total heat loss used in the calculations may include a margin depending upon the performance guarantees.

REFERENCES

1. V Ganapathy, "Applied heat transfer", Pennwell Books, Tulsa, 1982. 2. V Ganapathy, "Simplified approach to designing heat transfer equipment", Chemical Engineering, April 13, 1990. 3. V. Ganapathy, "Evaluate the performance of waste heat boilers", Chemical Engineering, November 16, 1981, Pg 291. 4. V. Ganapathy, "Steam plant calculations manual", Marcel Dekker, NY, 1984.

NOMENCLATURE

Ca Ce Cp Ceo

do,di

e

- Annual cost of operation, $/yr - Cost of electricity, mils/kWh - Specific heat of fluid, Btu/lb F - Capitalized cost of operation, $

- Tube outer and inner diameter, in. - Escalation rate

Fire Tube Boilers

127

- Friction factor - Fouling factors, inside and outside, sq ft h F /Btu F1 - Factor defined in Equation (2-8) R - Period of operation, h/yr L1Rs - Enthalpy absorbed by steam, Btu/lb he - Convective heat transfer coe#icient, Btu/ sq ft h F hi - Heat transfer coefficient inside tubes, Btu/ sq ft h F hn - Non-luminous heat transfer coefficient, Btu/ sq ft h F hz - Heat loss factor; 1% loss equals a hz of .99 ho - Heat transfer coefficient outside tubes, Btu/ sq ft h F Ie - Installed cost, $ i - Interest rate K1' K2 - Factors defined in Equation (2-22) and(2-20) k - Thermal conductivity of fluid, Btu/ft h F Km - Thermal conductivity of metal, Btu/ ft h F L - Length of tube, ft Lee - Life cyde cost, $ MW - Molecular weight of fluid N - Total number of tubes Nu - Nusselt number P - Electrical fan power consumption, kW LiP - Pressure drop, in we Pr - Prandtl number q - Heat flux; subscripts i and 0 startd for inside and outside tubes, Btu/sq ft h Q - Duty, Btu/h; subscripts a and t stand for assumed and transferred Re - Reynolds number S - Surface area, sq ft; subscripts i and 0 stand for inside and outside tubes T - Life of equipment, yr T1, T2 - Gas temperature in and out, F L1 T - Log-mean temperature difference, F t - Average gas temperature, F ts - Saturation temperature, F U - Overall heat transfer coefficient, Btu/sq ft h F; subscripts i and 0 stand for inside and outside v - Specific volume, cu ft/Lb

f

tfi, !fo

128

Waste Heat Boiler Deskbook

w - Flow per tube, Lb/h W p Jl

11

-

Total mass flow, Lb/h Gas density, Lb / Cll ft Gas viscosity, Lb/ft h Efficiency of drive, fraction

I econoMizer

5uperheo.ter r-

----

I

I

I-

.-

I

fire tube boiler

L-

Figure 2-3. Fire tube boiler with superheater and economizer.

:'~

';;

Fire Tube Boilers

129

GAS IN

GAS

OUT

BOILER TUBES DAMPER

---------.-----::------t

BYPASS PIPE

GAS IN loolLer

)

~

+----1-

bYP(ksS pipe

do.Mper

Figure 2-4. Arrangement of internal and external gas by pass scheme.

~

130

Waste Heat Boiler Deskbook

eeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeee RESULTS-INSULATION PERFORMANCE___________________ Project: NAME

Casing minfb

THICK-IN 0.00 3.00

flat surface

example

TEMP-F 85.63 366.00

TEMP 1

COND1

TEMP2

0.00 200.00

0.00 0.30

0.00 400.00

COND2 0.00 0.42

HEAT LOSS -BTU/ft2h= 29.53123 Number of layers of insulation= 1 AHB TEMP= 70 WIND VEL-fpm= 100 EMISS= .9 MAX LOSS-BTU/FT2H= 1170.664

eeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeee Figure 2-5. Printout of casing heat loss.

Chapter 3

Water Tube Boilers Water tube boilers are more difficult to design compared to fire tube boilers due to the complex arrangement or disposition of heating surfaces such as superheaters evaporators and economizers. Figure 31 shows the various configurations that are feasible for a single pressure system. 'The superheater may be located ahead of the evaporator or as discussed inChapterl~when-thegasTnIe1:-temperature"isvery' high, it may be located between a screen section and an evaporator. Whep. the degree of s~h~t i~_yeIy§mall, it_rna)" be located even behind ~ . . '---~"-"~~---~'~---------------'-----'--"----------the evaporator. Also, t~S~!'h~l!tg:L~Q.lJld })~ in several stages, each with a different fin configuration or mech;;ti~a.i-arra~gementto ensure appropriate tube wall and fin tip temperatures and degree of superheat. The evaporator bundle could have varying fin configurations or tube pitch along the gas flow direction to handle the gas temperature or slagging concerns. It could also be built in two parts with a SCR (Selective Catalyst Reduction system) in between them ifS in the case of gas turbine applications. If we add to this a multipressure HRSG situation or the presence of a radiant section with different operating parameters and fuel firing, the design procedure would be indeed difficult without a computer. This chapter discusses the methodology of design for various types of heating surfaces such as superheater, evaporator and economizer, which are the building blocks of a complex water tube HRSG; simplified procedures are also suggested, followed by procedures for evaluating their performance under different gas inlet conditions. .

-

131

Waste Heat Boiler Deskbook

132

(1)1 [VAPI

[}ill I [VAPII ECON

I

(2)/EVAPI [}illIEvAPI (3) [}B] I EVAP II..-E-C-Ot\----.lI (4)

IEVAP I [}ill

IECON I

(5)

[}illIEVAPI

(6)1 EVAP/

[}B]

(7)

Ir--E-V-Ap----', IECON I

(8)

IEVAPI

(9)

1 E-C-0N---', r-

(10) (11 )

IECONI Key EVAP

~

ECON =

SH

~

Evnporntor Econo~zer

Superhenter

Figure 3-1. Possible Configurations for a Watertube HRSG.

133

Water Tube Boilers

GUIDELINES FOR WATER TUBE BOILERS 1. They are suitable for high steam pressure and temperature applications and large capacity units, even exceeding a million pounds per hour of steam or gas flow. 2. Exten~~~o~,~f~~~~S~!1:,~~'=!~,~(!.to make the design compact it ~"E~~~~~am is c!ea!}~ Compared to a fire tube boiler, a water tube with extended surfaces is much smaller and will weigh less, particularly if the gas flow is large, say exceeding 100,000 pph. 3. Various types of fuels can be fired with ease including solid and dirty fuels. The water cooled membrane wall enclosure, Figure 32, makes an excellent furnace and can be sized to match any firing or

~sh~9i~~ps:>~~f"~gitIpiil~~9 4. If the gas stream is dirty, provisions can be made for cleaning th~jub~es by using rapping mechanisms or soot blower~: Access lanes as required may be easily provided. Wide spacing may be provided at the front end of the boiler, where there is more chance for slagging or fouling, followed by sections with smaller spacing at the cooler end. In a fire tube boiler, it is very expensive or not feasible to build the boiler in two parts or use different tube sizes to accommodate slagging or temperature concerns. The best that can be done is to use a multi-pass design with different tube sizes in each pass. On-load cleaning is difficult in fire tube boilers. 5. Th~su£er~~ater! i(useci!~,!ll,aY_1>~ locate
tube

134

also to be wary of

Waste Heat Boiler Deskbook waterl!:!Q~J:>oil~!:~~~~t!u~xtended~sur£~_~~ which

~rH.t~_j!La~mJl<:!tlligh~J:"heat "~fltl)( y~nd

tube wall temperature compared to bare tube designs or fire tube boilers. 9. Due to the h!gher heat transfer coefficients associated with ~~tlow__~"-~Jh~~~-;~te-;-t~be-boIIe;;-~quire--les~s~~facearea and hence the gas pressure drop can be lower than in a fire tube boiler. 10. A water tube boiler will be more expensive in the smaller gas flow rangesay §o,qoq pph or less compared to fire tube typ~-but less-expensive-for larger ~ass flows. For some situations such as gas turbine exhaust where the ratio of gas to steam flow is high and the pinch point is low, water tube boilers with extended surface may be the only choice as the fire tube equivalent will have a very high gas pressure drop and will be extremely large.

Figure 3-2. Water-Cooled Membrane Wall D Type Boiler Burning Waste Fuels. [Courtesy ABCO Industries].

135

Water Tube Boilers

Heat Transfer Calculations The basic equation for energy transfer for any surface is given by: (3-1)

The overall heat transfer coefficient for finned tubes may be estimated by the methods discussed in Appendix A. The overall heat transfer coefficient Uo for bare tubes is given by: l/Uo = l/ho + (l/hi)d o/di + ffi (do/di) + ffo + (d/24K m) In(d o/di)

(3-2)

i

The tube side coefficient hi may be estimated using the correlation given in Chapter 2 for single phase fluids such as steam or water. For two-phase fluids such as mixture of water and steam, correlations cited in Reference 1 may be used. Since boiling heat transfer coefficient inside tubes is very high on the order of 2000 to 3000 Btu/ sq ft h F, even a 20% error in this assumption will not affect U much as the gas side coefficient governs U. The gas side coefficient ho

=he + hn

(3-3)

The non-luminous coefficient hn can be obtained by the methods discussed later. There are several correlations for determining the convective heat transfer coefficient he, see Reference 1. This chapter will address two of the well known correlations. Convective Heat Transfer Coefficient The Grimson's Equation is widely used for determining he . Nu=BReN

(3-4)

Coefficients Band N for bare tubes in inline or staggered arrangements is given in Table 3-1. The gas properties are evaluated at the gas film temperature, which may be taken as the average of the gas and fluid temperatures. The maximum gas velocity is to be

136

Waste Heat Boiler Deskbook

used while determining the Reynold's number. Figure 3-3 shows inline and staggered configurations. It may be noted that in the case of staggered tube bundles the maximum velocity may occur in the diagonal or transverse plane depending on the transverse and longitudinal pitch. For the maximum velocity to occur in the diagonal plane, the diagonal pitch x need be only half of the transverse pitch. (St - d) L = 2 (x-d) L or x= (St + d)/2 from the right angled triangle ABC,

If St = 4 in, d = 2 in, then 51 should be less than 2.3 in, for this to occur. Sz is usually on the order of 3 in, or more and hence the maximum

velocity occurs in the transverse section. Reynolds Number Re = Gd/12Jl

(3-5a)

Nusselt Number Nu = he do /12k

(3-5b)

where G = Wo x12l[ Nw(St-d)LJ

(3-6)

Table 3-1. Grimson's Values of B and N 1.25

ST/d St/d

(2)

1.5

3

B

N

B

N

B

N

0.518 0.451 0.404 0.310

0.556 0.568 0.572 0.592

0.505 0.460 0.416 0.356

0.554 0.562 0.568 0.580

0.519 0.452 0.482 0.44

0.556 0.568 0.556 0.562

0.522 0.488 0.449 0.421

0.562 0.568 0.570 0.574

0.348 0.367 0.418 0.290

0.592 0.586 0.570 0.601

0.275 0.250 0.299 0.357

0.608 0.620 0.602 0.584

0.100 ~ 0:1(1) 0.229 0.374

0.704 0.702.1 0.632 0.581

0.0633 0.0678 0.198 0.286

0.752 0.744 0.64£ 0.60P.

B

N

Staggered 1.25 1.50 2.0 3.0

In-line 1.25 1.50 2.0 3.0

d

~I_ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _

137

Water Tube Boilers

Table 3-2. F Factor from Fishenden and Saunders. In-Line Banks

Staggered Banks

Sdd

1.25

1.5

2.0

3.0

1.25

1.5

2.0

3.G

Re 2,000 8,000 20,000

1.06 1.04 1.0

1.06 1.05 1.0

1.07 1.03 1.0

1.0 0.98 0.95

1.21 1.11 1.06

1.16 0.99 1.05

1.06 0.92 1.02

0.fl5 0.!J3

1.5

2,000 8,000 20,000

0.95 0.96 0.95

0.95 0.96 0.95

1.03 1.01 1.0

1.03 1.01 0.98

1.17 1.10 1.04

1.15 1.06 1.02

1.08 1.00 0.98

1.02 0.96 0.!l4

2.0

2,000 8.000 20.000

0.73 0.83 0.90

0.73 0.83 1.0

0.98 1.0 1.0

1.08 1.02 1.0

1.22 1.12 1.09

1.18 1.10 1.07

1.12 1.04 1.01

1.08 1.02

2,000 8,000 20,000

0.66 0.81 0.91

0.66 0.81 0.91

0.95 1.02 1.01

1.0 1.02 1.0

1.26 1.16 1.14

1.26 1.15 1.13

1.16 1.11 1.10

1.13 1.06 1.02

ST/d 1.25

3.0

0.~i6

o.~n

Another widely used Equation is that of Fishenden and Saunders. (3-6)

where prandtl number Pr = I1Cp/k

(3-7)

Factor F correcting for inline and staggered arrangements is given in Table 3-2. It will be seen later that with bare tubes, inline arrangement is preferred as the increment in heat transfer coefficient with staggered arrangement is not that significant compared to the increase in gas pressure drop. Example 1:

150,000 pph of flue gases having an analysis of % C02= 12, H20= 12, N2 = 70 and 02 = 6 by volume flow over a bank of boiler tubes of size 2 in., outer diameter with a transverse ~nd longitudinal pitch of 4 in., and arranged in inline fashion. Tubes/row = 18, length = 10 ft; assume that number of tubes deep exceed 10.

- .----~

.---~~------------------~

138

Waste Heat Boiler Deskbook

w

s,

w ~~~-,--+----~~---+--.----

s,

T

Figure 3-3. In-Une and Staggered Tube Bundles.

d

Water Tube Boilers

139

If the saturation temperature inside the tubes is 353°F and the average gas temperature is 700°F determine the convective heat transfer coefficient using both the correlations.

Solution: From Appendix E on gas properties, at the film temperature of .5(353 + 700) =526°F, Cp =.2695, Jl =.0642 and k = .02344, G =150000 x 12/[18 xlO x (4-2)] =5000lb sq ft h, Re = 5000 x 2/12/ .0642 =12980. From Table 3-1, B = .229 and N =.632, Nu = .229 x 12980.632 = 91 or he X 2/(12 X .02344)/~ 91. Hence he =12.80 Btu/Sq ft h F. Let us now determine he using Equation (3-6). From Table 3-2, F = 1. Pr = .0642 x .2695/.02344 =.738, Nu =.35 x 12980.6 x .738.3 = 93.85 or he = 13.2 Btu/sq ft h F. We will see the effect of tube pitch and configuration on heat transfer later. Determination of Tube Side Coefficient hi The tube side coefficient for the flow of steam, water or boiling steam-water mixture plays a role in the estimation of overall heat transfer coefficient U, as seen from Equation (3-2). The basic equation for obtaining hi for single phase fluids is that of Dittus-Boelter, see reference 1. NU = .023 RdPr,4 (3-8) The above equation may be simplified as shown in chapter 2 and hi may be written as: (3-9)

Factor C is given for steam in Table 3-3. For water flowing through tubes at a temperature T in DegreeOF, C =10 [-1·318 + .2141n en]

(3-10)

Example 2: Determine hi when 5000 pph of superheated steam flows inside a tube of inner diameter 1.78 in. in a superheater. The steam pressure

---~-

----~---~------~~-------~---

140

Waste Heat Boiler Deskbook

is 1000 psia and the average temperature is 800°F. What is hi when the steam is in saturated condition? Solution:

From Table 3-3, C = .345. Hence hi = 2.44 x 5000.8 x .345/1.781.8 = 271 Btu/ sq ft h F. If the steam is saturated, then C = .49 from Table 33. Hence hi = 385 Btu/sq ft h F. The mass velocity inside the tubes Gi = 5000 x 144 x 4/3.14 x 1.782 = 289480 Lb/Sq ft h. Mass velocity of steam inside superheater tubes is dependent on pressure, temperature and steam pressure drop to ensure uniform flow through parallel circuits. It could vary from 250 to 600 Lb / sq ft h. In the case of two-phase flow or boiling inside tubes a value of 2000 to 3000 Btu/ sq ft h F may be used for hi. Its effect on U as explained above is not significant. Example 3:

Determine hi when 10,000 pph of 500 psia water at an average temperature of 400°F flows in an economizer tube of inner diameter 1.738 in. Solution:

From (10), C = 10 [-1·318 + ·2141n(400)] = .921i hi = 2.44 x 10000. 8 x .921/1.7381.8= 1316 Btu/sq ft h F. NoJgthClLtyp!c:c'!t",ater velgcity-in-, ~_~~zer tubes ranges-from 2Jg__ §J!!~From steam tables,-Appendix E, the specific volume of compressed water at 400°F = .01185 cu ft/Lbi i water velocity = .05 x Wv/d 2 = .05 x 10000 x .01855/1.738 2 = 3.07

£tis. NON-LUMINOUS HEAT TRANSFER COEFFICIENT Due to the partial pressures of triatomic gases present in the flue gases such as carbon dioxide, sulfur dioxide and water vapor, significant amounts of energy can be transferred between the gas stream and the boiler surface due to non-luminous radiation heat transfer, particularly if there is a large amount of steam injected into the exhaust gas stream.

Water Tube Boilers

141

The variables that influence the non-luminous heat transfer coefficient are: 1. Partial pressure of C02, H20, 502,503. 2.Beam length L, which is dependent on the pitch and arrangement of the tube bundle. 3. Temperature of the gas stream and the surface temperature of the bundle. A number of factors enter into the calculation of non-luminous radiation heat-transfer coefficients between hot gases and tube bundles. Net interchange of radiation between gases and tube bundles is given by the basic equation. 4

4

QIA = (j[ E gTg -ag To]

(3-11)

For quick practical estimates, a good assumption that simplifies the above equation is E g= ago Then 4

4

QIA = Eg (j[Tg - To]

(3-12)

From this, the non:..luminous heat-transfer coefficient hN may be obtained as follows: (3-13)

Where

Eg

= Ec + 1J E W-.1E

(3-14)

The emissivity of hot flue gases is dependent upon the emissivity of carbon dioxide, E c, emissivity due to water vapor, E w and correction terms 1\ and AE . These values may be obtained from the well known Hottels charts, Figures 5a, 5b, 5c, 5d, once the partial pressures of carbon dioxide and water vapor, the beam length of the heat receiving surface and the gas temperature are known. The beam length, L of a tube bundle is given by: L

= j·08(STSL- ·785d2)Id

(3-15)

Waste Heat Boiler Deskbook

142

Figure 3-4b gives L if ST,SL and d are known. Hence it may be noted that an estimate of hN requires information on partial pressures of C02 and H20 in flue gas, which in turn implies availability of results of combustion calculations of fuels. This is a lengthy procedure. However, a simplified procedure has been developed based on detailed calculations with various fuels. Equation (3-15) is rewritten as: hN =KEg

(3-16)

Egis calculated for various fuels at normally used excess air factors and presented in Figure 3-4a along with K. The accuracy of this figure has been checked through detailed calculations and a maximum error of 5 percent was observed. Hence for quick engineering estimates of hN, Figures 3-4a and 3-4b may be used. The advantages are: •

Detailed combustion calculations for fuels need not be carried out and thus a lot of time is saved (sometimes even the fuel analysis may not be readily available).

• Reference to Hottels' charts is not necessary, again saving time .

•

Quick comparisons of E g as a function of fuel used may be made.This is an important factor as furnaces of boilers, fired heaters etc., have to fire different fuels during their operation, and performance with alternative fuels must be known.

An example is worked out to illustrate the versatility of the charts. Example 4: A boiler is fired with No. 2 oil at 10 percent excess air. Determine the non-luminous radiation heat transfer coefficient over a tube bank when the average flue gas temperature is 1650°F and tube wall temperature is 400°F.

ST = 5.0 in., SL = 3.5 and d =2.0 in.

j

Water Tube Boilers

143

Solution: In Figure 3-4b, connect 5L = 3.5 with 5T =5 and extend to' cut line 1. Connect with d =2 and extend to cut L scale at 7.8 in. In Figure 3-4a go up from tg = 1650°F to cut to = 400°F and move right to cut K scale at 26. Extend the vertical from tg = 1650 to cut L = 7.8 and move left to cut the reference lin~. Connect with point 3 (No.2 oil) and extend to cut E g scale at 0.105. Hence hN = 0.105 (26) = 2.74 Btu/ft2hoF. Exc:euar%

Fuel 8lIumlnou, coal,

3

20 10 10 10 10' 30

Naturalga.

~

No.2 011 No. 6011

7

Blpl furnace gu 81asl furnace gal No. 2 011

70

60

40

20

0.1

"<\,,.

30

",

,

'0.2

50

20 4,7

"3~'X~ 2

'.,

5,\

40

i

~

12

'"

10

It

30

""'\'

2 '/ 20

0.3

10

1)

400 •

eoo

1000

1200

1400

To, 'F

Figure 3-4a, Nomograph for Determining hN

1600

1800

2000

---------~------~-------------------------------------------~~~~----

144

Waste Heat Boiler Deskbook

2

3

4

10-' 12

14

'" Figuare 3-4b. Chart helps evaluate beam length.

The above charts, Figure 3-4a and 3-4b are suitable when the gas analysis are not known and the flue gases are products of combustion of fossil fuels. In case the gas analysis is known, it is preferable to use the Hottels' charts, Figures 3-5a, 3-5b, 3-5c and 3-5d to estimate the gas emissivity, see reference 1. A good estimate of the gas emissivity may also be obtained from the following equations.

145

Water Tube Boilers

ABSOLUTE TEMPERATURE('R'

Figure 3-5a. Emissivity of Water Vapor. [Penwell Books]

.., ~l

~

~-

~~

.....

- - - -'. ~s

'·10

..

tot

'

OOS

UC

...

\IJ

",

:-.....

f ... ••••:-:~ tit ~ .... ~F:§ §§. ~"'" ..!"', ....,j

.

......

841 OllIS

8-010

--=

...... " .... t-- .....-......... ...........

~,

--

0-0 .. "01 0-007 "01

I

f

~

.

~

847

84. 'S:::

I

-- .... ~

-.....; II.~

.........

........ ~ .......... r............... ... '

"II~

,.

III~ .,;:-: "

~

.!:..I, ... .... ....r-.......... o.~" r-.............. -........: t--... . . . ;-;., ... ~.... ,:I, ...."" "'," ~~ ~

r-......

ro.... ......

1-0..""' ....." ~

..... ro....

...... :"'0..

......

~,~

/I,~

;~....-:

." ' I" ' :I. . ~

t': ~ ~ .c!

-."'"

o-oos

~

'/1'111.

....

''', ... ~ '''''/1'' ..... " ',0. ""., "I ..... '"

UOC

~

....

0-001

SOO

1000

ISOO

1000

1500

'000

1500

CIIOO

4100 5000

ABSOLUTE TEMPERATURE (,oR)

Figure 3-5b. Emissivity of Carbon Dioxide. [Pennwell Books]

146

Waste Heat Boiler Deskbook

1.8

U 1.4 1.2

"

0.8 0.6

0.4 0.2 0.0 0.0

0.2

Figure 3-5c. Corrections Factor for Emissivity of Water Vapor. [f'enwell B ooksl

AE

0-01

o ~

Pc· Pw

02

0"'

0,' Pw

0&

-;;c:p;-

I'w

;;c;-;;w

Figure 3-5d. Correction Term Due to Presence of Water Vapor and Carbon Dioxide. [Penwell B ooksl

j

Water Tube Boilers

Eg

147

=0.9(1 - e-KL)

(3-17)

K = (0.8 + 1.6pw) (1 - 0.38Tg/1 000) '{(Pc + pw)L

X

(pc + pw)

(3-18)

Tg is in K. L is the beam length in meters and Pc and Pw are the partial pressures of carbon dioxide and water vapor in atm. L, the beam length, may be estimated fora tube bundle by Equation (3-15). The beam length for a bare tube bundle is given by the above equation. However for a finned bundle or a cavity, it may be estimated as: L = 3.4 volume/surfa<;e. Due to the much larger surface, the beam length is smaller with finned bundles and hence hn will be lower. Example 5: In a boiler superheater with bare tubes, the average gas temperature is 1600°F and the tube metal temperature is 700°F. Tube size is 2.0 in., and transverse pitch 5t = longitudinal pitch = 511= 4.0 in. Partial pressure of water vapor pw = 0.12, of carbon dioxide = 0.16 =pc Determine the non-luminous heat transfer coefficient. Solution: From Equation (3-15) the beam length L is calculated. L = 1.08 )(.4 x4 - 0.785 x2 x2 = 6.9 in. = 0.176 m 2

Using Equation (3-18) with Tg = (1600 - 32)/1.8 + 273 = 1114 K. we obtain Factor K = (0.8 + 1.6 xO.12) x ( 1 - 0.38 x 1.114) x0.28 = 0.721 '10.28 xO.176 From Equation (3-17):

Eg

=0.9 x[l- exp(-O.721 xO.176)] =0.107

Then, from Equation (3-13): 4

hN = 0.173 xO.107 x10-8

>2060 -1160

4

= 3.33 Btuljt 2 hrF

1600 - 700

j

--~--~

-~~~

Waste Heat Boiler Deskbook

148

Table 3-3. Factor C for Steam. Pressure (psia) : Sat. :

100

500

1000

2000

0.244

0.417

0.490

0.900

Temp.

(F) 400 500 600 700 800 900 1000

0.271 0.273 o• 281' r"./ 0.291 " 0.305 0.317 0.325

0.360 -------~ ,'0.322--'\ \0.316/ . / 0:'320 0.327 0.340

, ___ ------:'::L""

1)~0.413 )

D.)I

0.358/ \0.345 0.347 0.353

0.520 0.420 0.394 0.386

GAS PRESSURE DROP IN TUBE BANKS

Tube bundle I ST I

Nw ", number (If

NH

=

tube~

wide

number of tubes deep

L '"

length of tube

W '"

Width of tube b!lndle

o s~' .:c. ,

Tube diameter

'

0

,0

0

0 0

0, 0

o

0

0

0

0

o

0

"

0

0

L

0

In"line arrangement

Figure 3-6. Typical Tube Bundle

d

Water Tube Boilers

d, In. 1.25

1.50

149

G, Ib/ft2h 16,000 14,000

2.0 2.25 2.5 2.75 3.0 3.5 4.0

0.03

12,000 10,000 9,000

1.75

f.lg

0.025

a.oOO 7.00

Re 30,000 25,000

0.04

t, of. 0.06

6,DOO 5,000 4,500 4,000 3,500 3,000

0.05

0.07 10,000 9.000 8,000 7,0006,000 5,000

0.08 0.09 0.10 0.12

2,500

4,000

2,000 1,800 1,600

3,000

0.14 0,16 0.18 0.20

2,000-

0.25

FigUre 3-7. To Find Reynolds Number.[Oil and Gas Journal]

Tube bundles used in boil~r superheaters, economizers, fired heaters, or process heaters offer resista;nce to the flow of flue gases or air flowing over them. ~epending on the tube geometry, pitch arrangement, temperature, and gas velocity used, the gas pressure drop will vary. Gas pressure drop can be costly in the long run as fan power consumption converted over the life of the equipment can be substantial. For example, if the additional fan power is 10 kw because of using high gas velocities, the capitalized cost could be about $20,000, depending on cost of electricity, escalation, and interest rates. Several correlations are available in the literature to predict the gas pressure drop. Each organization develops its own data based on site operation and experience, which should be used to give accurate results. In the absence of such data, the gas pressure drop over a bundle of plain tubes may be predicted as follows for estimation purposes.

Waste Heat Boiler Deskbook

150

Plain Tubes (3-19)

Where the friction factor, fg, is given for 2000
fg =Re -D.lS 0.08 (STId) ] 0.044 + [ (STld _ 1) (0.43 + 1.13 dSIJ

(3-20a)

For a staggered arrangement: fg

=Re-O·l~0.25 +

l

0.1175 1)8] (STId-l)

1

Re, the gas Reynolds Number is given by:

(3-20b)

Re

= Gd

12J.l

(3-5a)

Where G is the gas mass velocity, The maximum value of G is to be used in the above equations. This maximum occurs in the plane transverse to the gas flow. G=

12Wg L[w - Nw(ST - d)]

(3-6)

Figure 3-6 is a typical arrangement of tubes. Through the equations are not particularly difficult, more time is spent on obtaining the viscosity and density of flue gases in order to estimate the Reynolds Number and pressure drop. Figure (3-7) has been developed to give Re if gas mass velocity G, tube diameter d, and flue gas or air temperature are known. One need not look up gas property tables for viscosity. Also,if viscosity is known, Re may be found. The chart is reasonably accurate for air as well as flue gases.

Water Tube Boilers

151

Example 6: Flue gases flow over a tube bundle with a gas mass velocity of 7,000 Ib/ft2h. Tube diameter is 2.0 in., and gas temperature is 800°F. Estimate the Reynolds Number. Solution: Connect d = 2.0 with G = 7,000 and extend to cut line 1. Connect with t = 800 to cut Re scale at 16,000. Figures 3-8 and 3-9 give the friction factor, fg, for staggered and in-line arrangements.

Example 7: If the ratio O'r = (ST/d) = 2.2 and (jl = Sz)d = 2.0, estimate the friction factor in staggered and in-line tube bundles of a boiler evaporator. Solution: For staggered tube bundles, in Figure 3-8, go up from O'r = 2.2 to cut Re = 16,000 (an error of 10% in Re hardly matters), and move left to cut fg scale at 0.074. This is the friction factor for staggered arrangement; (jL is not used. In Figure 3-9, for in-line bundles, go up from O'T = 2.2 to cut O'L = 2.0. Move left to cut reference line. Connect with Re = 16,000 and extend to cut f g scale at 0.043. Figure 3-10 gives the gas pressure drop per row of tubes if G, t, and f g are known based on Equation (3-19). Also the velOcity, Vg, is obtained based on: Vg

=

G 3600 pg

(3-21)

There is no need to calculate the gas density Pg. Example 8: In a boiler economizer, flue gases at 800°F average temperature flow over 16 rows of a tube bundle with a gas mass velocity of 7,000 Ib/ft2h. Friction factor found using Figure 3-8 for the tube geometry is 0.085. Estimate the gas velOcity and the gas pressure drop.

152

Waste Heat Boiler Deskbook

0.16 0.14 0.13 0.12 0.11

19 0.10

Re

0.09

2,000

0.08

$,000

0.07

10,000

0.06

20,000 3.5

0,06

4.0 1fT == Srld

Figure 3-8. Friction Factor for Staggered Arrangement.

0.01 0.02 0.03 ~e~

t9 0.04

COO

~, ~t;}"

/

').\J~~\J\J

O.OS

,.../:/

0.06

.'<~-

0.07 0.08 0.09

1.6

2.0

2.5 O'T ;:

3.0 ST/d

3.5

Figure 3-9. Friction Factor for In-Line Arrangement.

4.0

Water Tube Boilers

153

Solution: In Figure 3-10, connect t = 800 with G = 7,000 to cut line 1 at A. AhlO, read off V g = 63 ft/sec as gas velocity. Connect point A with fg =0.085 to cut (..1 pglNH) scale at 0.12 in., wc. Hence, total gas pressure drop = 16 x 0.12 =1.92 in., wc.

1'9. 0.10

0.09 I,"F. 60· 0.,01 100

0.06

0.06

0.01

Figure 3-10. Gas Velocity, Pressure Drop.

IN-LINE VERSUS STAGGERED ARRANGEMENT Inline arrangement is generally preferred with bare tubes as the g-ain in heat transfer with staggered configuration is not significant compared to the increase in gas pressure drop. The choice of tube spacings is based on considerations of slagging or bridging, distribution of external radiation from flame or cavities, gas pressure

--------------

-----~---

~-------

154

Waste Heat Boiler Deskbook

drop, heat transfer and finally the mechanical ligament efficiency in case of convection sections using boiler drums. As the longitudinal pitch increases, so will the gas pressure drop and the length of the section. The longitudinal pitch should be low enough to result in a low gas pressure drop but large enough to yield a good ligament efficiency. The following example compares inline versus staggered arrangement and analyses the effect of longitudinal pitch. Example 9: 150,000 pph of flue gases at 900°F are cooled to 500°F in a boiler generating 125 psig stearn. Neglecting the effect of non-luminous coefficiEmt , study the effect of using longitudinal pitches of 3, 4 and 6 in. Tube diameter is 2 in., and transverse pitch is 4 in. Boiler duty is 16MMBtu/h.

Solution: Using the procedure given in example 3, he was computed for the various configuration. Non-luminous radiation was not considered. The number of rows deep was estimated assuming that the other resistances were negligible. The gas pressure drop was computed for each case using the equations given above. The results are shown in Table 3-4.

Table 3-4. Study of In-line Versus Staggered Arrangements Slid

1.5 In-line Stagg

10.96 11.90 Friction Factor .0396 11.50 he, Average 49 No. of Rows L1Po, in we 1.32

, hc, Crimson hc, Fishenden

13.79 14.35 .0807 14.07 40 2.12

2.0 In-line Stagg

3.0 In-line

Stagg

12.81 13.20 .04928 13 43 1.44

12.91 13.30 .0686 13.10 43 2.00

12.69 13.20 .0807 12.95 43 2.33

13.13 13.56 .0807 13.35 4.2 2.29

Water Tube Boilers

155

The following points may be noted: 1. The staggered arrangement does not have a significant increase in he over the inline when the longitudinal pitch to diameter ratio exceeds 2. It appears attractive below 1.5. However this ratio is not widely used in industry due to the fact that the ligament efficiency is affected if a drum is used. Also one has to look at the bridging of tubes by ash particles or slag and cleaning considerations if tubes are so close. 2. The gas pressure drop is much higher for the staggered arrangement. There may be a trade off if 511 d is less than 1.5 but above this ratio, the heat transfer coefficients are nearly the same for both inline and staggered but the gas pressure drop with staggered arrangement is much higher, resulting in an operating penalty. There is some increase in beam length with a larger 511 d ratio but the overall benefits are minimal. Hence an inline configuration with a 5t l d and 511 d ratio of 2 is generally used unless there are other reasons such as external radiation or bridging of tubes with molten particles from ash, which may require a large 51 and 5t .

DESIGN OF EVAPORATORS The starting point in the design of an evaporator bundle, Figure 3-11, is the estimation of overall heat transfer coefficient. The cross sectional data such as the number of tubes wide, spacing and length of tubes are assumed. Based on the mass velocity, U is estimated. The surface area is then determined followed by the number of rows deep. The gas pressure drop is then evaluated. This is only a feasible design. 5everal options are possible depending on the size and operating cost, which is influenced by the gas pressure drop. Optimization is then resorted to, preferably using a computer program. The following example illustrates the design approach.

Example 10:

. 150,000 pph of flue gas is required to be cooled from 900 to 520°F in an evaporator generating saturated steam at 125 psig with 230°F

156

Waste Heat Boiler Deskbook

feed water. Blow down is 5%. Gas analysis is as follows: % volume COZ = 12, HZO = 12, NZ = 70 and Oz = 6. Heat loss from the casing = 1.0%. Fouling factors on steam lind gas side = .001 sq ft h F/Btu. Solution: Let us assume the following: Tube size = 2 x .105 in. (inner diameter = 1.77 in.); number wide = 18; tube length = 10 feet; transverse and longitudinal pitch are 4 in., each. Average gas temperature = .5 x (900 + 520) = 710°F = 650 K. Fluid temperature inside tubes = 353°F. Hence film temperature = .5 x (353 + 710) = 531°F. The gas properties from Appendix E are: Cp= .270, J1 = .0645 and k= .02345. The gas specific heat at the average gas temperature = .277. Duty = Q= 150,000 x.~9 x.277x(900520) = 15.60 MMBtu/h. The steam enthalpy change = (1193-i98) + .05 x (325-198) = 1001.4 Btu/Lb Hence steam generation = 15.60 x 106/1001.4 =15,580 Lb/h. G = 150000 x12/[18 x 10 x (4-2)] = 5000 Lb/sq ft h. Re = 5000 x2/ (12 x .0645) = 12920. Using Crimsons Equation, for a spacing of 4 in. inline, Nu = .229 x 12920·63Z = 90.8 = he X 2/ (12 x .02345). Hence he = 12.78 Btu/ sq ft h F. Let us compute the non-luminous coefficient; partial pressures of COZ and HZO = .12, Beam length L = 1.08x (4 x 4 - .785 x 4)/2 = 6.95 in. = .176 m, K = (.8 + 1.6 x .12) (1- .38 x .650) x .24/ (.24 x .176)·5 = .872, Eg = .9 x (l-e -.872 x .176 ) = .128. Assume that the surface temperature is 400°F. A reasonable assumption is 30 to 40°F above the average fluid temperature. hn =.173 x.9 x .128 x [11.74 - 8.64]/ (1170860) = .85. (An additional factor of .9 was used to account for the emissivity of the surface.) Using a hi value of 2000 Btu/sq ft h F and fouling factors of .001 for both inside and outside the tubes and a tube metal conductivity of 25 Btu/ft h F, l/U = 1/(.85 + 12.78) + .001 + .001 x 2/1.77 + (1/2000) x 2/1.77 + 1 x In (2/1.77)/ (24 x 25) = .07326 + .001 + .0011 + .000565 + .0004 = .07633. Hence U = 13.1 Btu/ sq ft h F. Log-mean temperature difference = [(900-353) - (520-353)]/ln (547/167) =320°F. Required surface area = S =15600000/(320 x 13.1) = 3721 sq ft =3.14 x (2/12) x 18 x 10 x Nd ; or Nd =39.5 use 40. Provided surface area = 3768 sq ft.

Water Tube Boilers

157

Let us compute the gas pressure drop. Gas density = (MW /359) x 492/(460 + Tg } = 29 x 492/(359 x 1170) = .034 Lb/cu ft;/= 12920-. 15 [.044+ .08 x 2] = .0493. LiPo = 9.3 x lO-10 x 50002 x 40 x .0493/.034 = 1.35 in we. The average heat flux q (based on inner diameter) = Ux (Tg-ts) = 13.1 x (710-353) x 2/1.77 =5285 Btu/sq ft h. The temperature drop across the various resistances are computed as in Chapter 2. Drop across inside film = 5285/2000 = 2.7 F. Drop across inside fouling = 5285 x .001 = 5.3 F. Drop across tube wall = (.0004 x 1.77/2) x 5285 = 1.9 F. Hence the outer wall temperature = 353 + 2.7 + 5.3 + 1.9 = 363 F. Note that this is only an average outer wall temperature. The heat flux and the tube wall temperature should be evaluated at the gas inlet, considering the non-uniformity in gas flow and temperature profile across the boiler cross-section. Note also that the heat flux with bare tubes is low compared to that with finned tubes as will be shown later. Also, an evaporator section may have several combinations of tube spacings and fin configurations. This design would be used in clean gas applications when the inlet gas temperature is very high where due to heat flux and tube wall or fin tip temperature concerns, a few bare tubes would be used followed by tubes with increasing fin density. Also in dirty gas applications, the tube spacing would be wide at the gas inlet due to slagging or bridging concerns. In these cases, an analysis of each sort of arrangement for heat transfer and pressure drop may be warranted.

SIMPLIFIED APPROACH TO DESIGN Whenever gas flows outside a tube bundle-as in water-tube boilers, economizers and heat exchangers with high heat-transfer coefficients on the tubes ide-the overall coefficient is governed by the gas-side resistance. Assuming that the other resistances contribute about 5% to the total, and neglecting the effect of nonluminous transfer coefficients, one may write the expression for U as:

--------------------------------~

158

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Waste Heat Boiler Deskbook 9'-d'

jOj !I

!

!I

!I

II I

I

I

!I II

! I

!I

I I!

I

II I

I

j

I -r---------------

l ______________ _

END VIEW Figure 3-11. Milti-gas Pass Evaporator Bundle [ABeO Industries]

U =0.95 ho

where the outside coefficient, Nu = 0.35 R~·6p~.3

(3-22) hOI

is obtained from

-----------~--

159

Water Tube Boilers

Equation (3-6) is valid for both inline (square or rectangular pitch) and staggered (triangular pitch) arrangements. For bare tubes, the difference in ho between inline and staggered arrangements at Reynolds numbers and pitches found in practice is 3-5%. Substituting Equations (3-5) , (3-7) in Equation (3-6) and simplifying: (3-23) (3-24)

Table 3. Factors F2 and F3 For Air and Flue gas Air Temperature, of 100 200 300 400 600 800 1,000 1,200

Temperature, OF 200 300 400 600 800 1,000 1,200

F2

F2/C p

0.0897 0.0952 0.1006 0.1056 0.1150 0.1220 0.1318 0.1353

0.3730 0.3945 0.4140 0.4308 0.4591 0.4750 0.4890 0.5030

F2 0.0954 0.1015 0.1071 0.1170 0.1264 0.1340 0.1413

0.5920 0.6146 0.6350 0.6528 0.6810 0.6930 0.7030 0.7150

Flue gas F2/C p 0.3698 0.3890 0.4041 0.4300 0.4498 0.4636 0.4773

0.5851 0.6059 0.6208 0.6457 0.6632 0.6735 0.6849

(Flue gas is assumed to have 12% water vapor by volume)

.

Waste Heat Boiler Deskbook

160

(3-25) F2 is given in Table 3-5, Gas-transport properties are computed at the film temperature. (3-26) Combining the above with Equation (3-1) and simplifying: QIAT = US = O.9Go. 6F2doN"N,Jj(12ct·4)

=0.235F2GO.6NvNaLd~·6

(3-27)

Substituting for G from Equation (3-6) (3-28) The above Equation relates thermal performance to geometry. When there is phase change, as in boilers, further simplification leads to: (3-29) If the tube diameter and pitch are known, one can estimate N d or G

for a desired thermal performance. Let us now account for gas pressure drop. The Equation that relates the gas pressure drop to Gis: (3-30)

For inline arrangements, the friction factor is obtained from: (3-31) where: X = [0.044 + (O.08Si/do)I(Stido _1/°·43 Another form of Equation (3-30) is:

S

+ 1.J3d,J I)j

(3-32)

Water Tube Boilers

161

Substituting for f in Equation (3-30) and combining with Equation 3-29) we can relate LlPo to performance in a single Equation: &'0 = 4.78 x 10

-10 2.25

G

.

0.75

(S t - do)ln[(T1- ts)/(T2 - ts)J X/[do F3PJ

(3-34) (3-35)

F3 is given in Table 3-5. With the above equation, one can easily calculate the geometry for a given tube bank so as to limit the pressure drop to a desired value. Two examples will illustrate the versatility of the technique.

Example 11: In a water-tube boiler, 66,000 lb/h of flue gas are cooled from 1,160 to 440o P. Saturation temperature is 350°F. Tube D.D. is 2 in., and an inline arrangement is used with 5Z = 5il= 4 in. Determine a suitable configuration to limit the gas pressure to 3 in., H20. Let us use Equation (3-34) . Pilm temperature is 0.5(800 + 350) = 575°P. Interpolating from Table 3-5 at 575°P, F3 = 0.643. Gas density at 8000 P is 0.031Ib/ft3. L1Po = 4,78 x 10-10 G 2.25 x (4 -2)ln [(1,160 - 350)/(440 - 350)][0.044 +

0.08 x 2]/[20 .75 x 0.643 x 0.031] = 128 x 10-10 G 2.25 = 3. Hence, G = 5,200 Ib/(ft2)(h). Prom Equation (3-6) one can choose different combinations of N w and L: NwL = 66,000 x 12/(2 x 5,200) =76. If N w = , 8, then L = 9.5ft. Calculate Nd from Equation (3-29). In[(1160 -350)/(440 - 350)J = 2.197 = 2.82 (F2/Cp )NdI[Go.4(Srldo -1)d~·4J,

or 2.197 =2.82 x 0.426 NdI(5,200°.4 xl ><20.4), or Nd = 74. Thus, the entire geometry has been arrived at.

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162

Example 12: If in the previous case, Nd .is to be a maximum of 60 due to space limitations, determine the geometry and pressure drop. Thermal performance is unchanged.

'.

Solution: Using Equation (3-29), let us calculate G: 2.197 = 2.82 x 0.426 x 60/(G 0.4 x 1 x 20.4), or G = 3,080. Prom Equation (3-6): 3.080 = 66,000 x 12/[Nw L(4 - 2)], or NwL = 128.5. If we choose N w = 12, then L = 10.7 ft. Compute the gas pressure drop, using Equation (3-34). F3 =0.643; X =0.204; P= 0.031Ib/ft3. APo =4.78 x 10-10 x 3,0802 .25 x 2 x 2.197 x 0.204/[0.643 x 20.75 x 0.031] = 0.91 in. H20. Once G or APo is fixed, the entire configuration is fixed. Of course, adjustments can be made to N w and L to obtain a suitable geometry.

PERFORMANCE CALCULATIONS Let us see how the above boiler behaves when. the gas inlet temperature and flow change, while the steam parameters remain unchanged. Example 13: Predict the performance of the boiler in example 10 when the gas flow = 130,000 pph and inlet gas temperature = 1500oP. Gas analysis is unchanged. Solution: Performance calculations are more involved than design calculations, as we have a given surface area and the heat balance has to be arrived at through a trial and error procedure. The basic steps are: 1. Assume the exit gas temperature. 2. Calculate U.

Water Tube Boilers

163

3. Compute Qa, the assumed duty = Wo xCp x(Tl- T2) x hZ .. 4. Calculate ,1T, the log-mean temperature difference. 5. Compute Qt, the transferred duty = U xS x,1T. If Qa and Qt do not agree, go back to step 1 and try another exit gas temperature. Por evaporators this is not a tedious calculation, but for economizers or superheaters, particularly when there is more than one stage, the calculations are time consuming. Let the exit gas temperature = 700oP. The average film temperature = [.5 x (1500 + 700) + 353] =725°P. Gas properties at 725°P from Appendix E are: Cp = .277, Jl = .07256 and k = .027. G = 130000 x 12/[18 x 10 x (4-2) = 43331b/sq ft h. Re =4333 x 2/ (12 x .07256) =9954 Nu = .229 x 9954. 632 = 77 = he X 2/(12 x .027); Hence he = 12.47 average gas temperature = .5 x (1500 + 700) = llOOoP =866 K.. K = (.8 + 1.6 x .12) (1- .38 x 866) /x (.24 x .176).5 = .782. Eg =.9 x (1- e-·782 x .176) =.1286. hn = .173 x .9 x .1286 x [15.44 - 8.64]/(1540 -860) = 1.49. 1/ U =1/(12.47 + 1.49) + 1/ (2000) x 2/1.77 + .001 x 2/1.77 + 2 In(2/1.77)/24/25 = .0747; U = 13.38. The assumed duty = 130000 x .99 x (1500 - 700) = 29.8 MMBtu/h (C p at average gas temperature = .289). Log-mean temperature difference =[(1500 - 353) - ( 700 - 353)]/ In [(1147/347) ] =669°F. Transferred duty Qt = 3768 x ,13.38 x 669 = 33.72 MMBtu/h. Since "~"- \;1 ---the discrepancy is large, and the transferred duty is higher than the assumed duty, another trial is required. Try 650 0 P as exit gas temperature. There is no need to compute U again as the difference ~ill be marginal. Assumed duty Qa = 130000 x .99 x .289 x (1500 - 650) = 31.61 MMBtu/h. ,1T =629; U =13.38; S =3768. Transferred duty Qt = 629 x 13.38 x 3768 = 31.71 MMBtu/h. This is close enough. Hence the new duty is 31.71 MMBtu/h and the corresponding steam flow =(31.71/15.60) x 15,580 =31670 Lb/h. Gas pressure drop is computed as before. Average gas temperature =1075°P =1535 R; P= 29 x 492/359/1535 =.0259 Lb/cu ft; f =9954-. 15 x (.044 + .08 x 2) = .0512. ,1Po =9.3 x 10-10 x 43332 x 40 x .0512/ .0259 = 1.39 in we. Tube wall temperature and heat flux may be computed as before.

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164

A SIMPLIFIED APPROACH TO PREDICTING EVAPORATOR PERFORMANCE: Using the approach discussed in Chapter 2, Equations (3-25) to (3-27), we have:

In [(Tl -ts)/(T2 -ts)] = U S/(Wo x Cp x hI)

(3-36)

Substituting the values, we have:

In [(1500 - 353)/(T2 - 353)] 1.355. Hence T2 =649°P.

= 13.38 x 3768/(130000 x .289 x .99) =

If the effects of non-luminous heat transfer and fouling are not significant and the gas analysis or temperature do not change much to warrant changes in gas specific heat and other properties, one can simplify the above equation further. U is proportional to W o·6 and S is a constant for a given boiler.

Hence In [(Tl- ts)/(T2 - ts)] = Y/Wo·4

(3-37)

The above equation may be used to predict evaporator performance keeping in mind the assumptions made. Constant Y is first estimated based on design conditions. Then for any other case of gas flow or inlet temperature, the exit gas temperature and duty can be predicted. This procedure is useful for quick estimates and when one does not have the time to compute U. It may also be used to predict the performance of a given boiler at another condition, based on field data for a particular condition. In the case of Example 13, there could be some errors in using the above equation, as the gas temperature is significantly different from the design case. However the method may be used for estimation purposes as illustrated below. Example 14:

A boiler operates with a SCR. As discussed in Chapter I, the gas temperature at the SCR should be within a bandwidth of

165

Water Tube Boilers

temperatures to obtain the desired performance of the catalyst. The SCR has been designed for a gas temperature of 650°F when the gas flow is 550,000 pph and gas inlet temperature to the boiler is 1120°F. Steam pressure is 300 psig. Determine the gas temperature at the SCR when the gas flow is 518,000 pph and the gas inlet temper<}ture to the evaporator is 1240°F. Assume that the gas analysis and extent of fouling have not changed much. Solution: The saturation temperature is 422°F. Using Equation (3-37), In [(1120 - 422)/ (650 - 422)] = Y/550000. 4. Hence Y = 221.27; For the new conditions, In [(1240 - 422)/(T2 - 422)] = 221.27/518000.4 or T2 =682°F. This is a good estimate for engineering purposes. If this is within the recommended bandwidth of temperatures, the original design is acceptable.

SELECfING DESIGNS WITH LOW PINCH AND APPROACH POINTS: The evaporator generates more steam if the pinch and a~aclipQmIs1!r~~f~~~ras discussed in Chapter 4. BQVY~V~! low~r these values, more will be the surface area required and higher will --b~the ga~-:Er~s~siire drop. The optimum choice is based on an analysis of fixed and variable-costs. Shown below is a study for an evaporator with finned tubes for gas turbine HRSG. For calculations with finned tubes and methods of determining overall heat transfer coefficient and pressure drop, see Appendix A. By selecting designs which have a low pinch and approach point, one can maximize the energy recovery and generate more steam. However a few aspects should be considered. a. The surface area of the evaporator gets larger as the pinch point decreases and reaches a point where it becomes counter productive in terms of cost and size. Typical unfired pinch point ranges from 15 to 30°F for evaporator with extended surfaces. Once the unfired pinch point is selected, the performance at other ambient conditions as well as in fired modes can be evaluated, as seen in ---'~~'--'-------

--

-'"-

"-

--~--,--

~\-,'«'

i~-

__________________________________________________________________ u

I

166

Waste Heat Boiler Deskbook

Chapter 4. As the gas inlet temperature increases, the pinch point increases and vice versa. b. The gas pressure dr()p also increases as the pinch point decreases resulting in an operating penalty. Approximately an additional 4 in. WC gas pressure drop results in a 1% drop in electrical power output of the gas turbine. c. The approach point is initially selected based on unfired cold ambient conditions. Steaming is a concern at low ambient conditions when the gas inlet temperature is low as in gas turbine applications. Problems associated with steaming and methods of avoiding the same in the economizer are discussed in Chapter 1. Table 3-6 shows two designs for two different values of pinch points. The gas inlet conditions are the same. The HRSGs are designed based on unfired pinch points of 41 and 22°P and a 15°P approach point and then their performance is evaluated in the fired condition when they are required to generate 30,000 pph of steam. Note that in gas turbine exhaust boilers, the oxygen content in the exhaust gas ranges from 14 to16% by volume typically if steam is not injected and hence additional fuel can be fired to generate more steam The performance calculation procedures are discussed later in this chapter and also in Chapter 4. HRSG A is designed with a pinch point of 41°P, while B with 22°P. The steam production and gas pressure drop are higher with design B. In the fired mode, while making 30,000 pph steam, A requires 8.1 MMBtu/h additional fuel, while B only 6.9 MMBtu/h (on lower heating value LHV basis). Assuming that the unit operated 50% of the time in unfired and fired modes, the following analysis can be made to evaluate each alternative. Let the fuel cost = 2.7 $/MMBtu (LHV), cost of steam = 3$/1000 pounds and electricity =5 c/kWh. The improvement of design B over A in dollar terms is as follows: • due to extra steam in unfired mode =878 x 3 x 4000/1000 =$10536 • due to added fuel in fired mode =1.2 x 4000 x 2.7 =$ 12960 • due to loss in electrical power output = (using the approximation that 4 in. pressure drop is equivalent to 1% less gas turbine power output) =- 1.2 x 4500 x .05 x 8000/400 = $5400. Hence design B offers $ 18096 (10536 + 12960 - 5400) in value over design A per year.

167

Water Tube Boilers

Based on the differential cost, one can compute the payback. Generally the payback ranges from 8 to 20 months. Selecting design A based on first cost alone would have been a poor decision. Similar analysis may be made to arrive at the optimum pinch and approach points for any evaporator. Table 3-6. Performance of Alternate Designs

Design A Unfired Fired

Design B Unfired Fired

1. Gas flow, pph 2. Gas temp to evap, F 3. Temp to eco, F 4. Stack temp, F 5. Gas dp, in we 6. Steam (150 psig sat) 7. Feed water temp, F 8. Temp to evap, F 9. Burner duty, MMBtu/h 10. Gas turbine output, kW

900 407 332 4.20 22107 < 351 0 <

(Nominal) 11. Surface area-evap, Sq ft -Eco, sq ft 12. Pinch point, F 13. Approach point, F

<-- 13227 --> <-- 16?34--> <-- 5948 --> <:--8922--> 41 22 27 53 15 29 14 26

150,000 1086 419 329 4.60 30000 230 337 8.10 4500

900 388 309 5.40 22985

1062 393 302 5.80 30000

352 0

340 6.90

>

>

COMPARISON OF BARE VERSUS FINNED EVAPORATORS Appendix A compares the design of an evaporator bundle for the same gas parameters with and without finned tubes. The conclusions are that !l fillE~ci1:>ull<.i1~ is ll1ore~()Il1l''!~~~_i~we~_!9-~~, of tubes _re[)~1ting_!I\}owergaspressuz:e drop and operating~cost and weighs less. Howevertheh~tflux ~_~t,g;h.hlgher wit!!linned tubes. and care should be taken to ensure that DNB conditions are not c

168

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reached. This can be handled by designing a few rows at the high gas temperature portions with a lower fin density and increasing the fin density as the gas gets cooler. Note also that as discussed in Appendix A, the effect of fouling on tube side is magnified several fold with a finned bundle compared to a bare tube bundle.

RADIANT HEAT TRANSFER SURFACES Radiant sections using partially or fully water cooled membrane wall designs are used to cool gas streams at high gas temperatures, Figure 3-12. They generate saturated stearn and may operate in parallel with convective evaporators if any. The design procedure is simple and may involve an iteration or two. The higher the partial pressures of triatomic gases, the higher will be the nonluminous radiation and hence the duty. If a burner is used as in the radiant section of a furnace fired HRSG, the emissivity of the flame has also to be considered. As explained in Chapter 1, radiant sections are necessary to cool the gases to below the softening points of eutectics if present so as to avoid bridging or slagging at the convection section. They are also required to cool gases to a reasonable temperature at the superheater if it is used. Example 15: 150,000 pph of flue gases at 1700°F has to be cooled to 1500°F in a radiant section of cross section 8 feet x 10 feet. Saturated stearn at 150 psig is generated by the gases. Determine the length. Gas analysis is: % volume of C02 = 12, H20 = 12, N2 = 70 and 02 = 6. Solution: Let the length = 25 ft. Beam length (see above) = 3.4 x volume/surface = 3.4 x 8 x 10 x 25/(8 x 25 x 2 + 10 x 25 x 2+2 x 10 x 8) = 6.4 ft = 2 m. Average gas temperature = 1600 F = 1144 K. Using (18), K = (.8 + 1.6 x .12) 0-.38 x 1.144) x .24/(.24 x 2)·5 = .194; Eg = .9 x (1-e-.194 x2) = .29.

Water Tube Boilers

169

Let average surface temperature of radiant section = 400o P. Surface area for heat transfer = (8 + 10) x 2 x 25 = 900 sq ft. Transferred energy = Qt = .173 x .9 x .29 x (20.64 - 8.64) x 900 = 7.1 MMBtu/h. Required duty = 150,000 x .99 x .308 x 200 = 9.15 MMBtu/h. Since there is a large discrepancy and the required duty is more, a larger surface area is required. As the furnace length does not significantly change the beam length and hence the gas emissivity, we can assume that the transferred duty is proportional to the surface area. Hence the required furnace length = (9.15/7.1), x 25 = 32 ft.

drUM

GAS

IN .,

/'

rc.dlo.nt sectlotl

GAS

OUT

Figure 3-12. Radiant Section

170

Waste Heat Boiler Deskbook

Iffi.SG Configurations and Circulation Figures 3-13a-i shows various types of evaporators and the configurations they can assume depending on availability of space, layout considerations, gas and steam parameters and cost. Figure 3-13a shows a two gas-pass design. In this type of design, the gas makes a 180 degree turn after passing through the first pass. The bottom and top half of the evaporator tubes are separated by a baffle plate. This design offers a lot of flexibility in terms of gas inlet and exit locations. The gas inlet can be from the bottom or side or even from the top and the outlet can be at the top or side and bottom. This design occupies a small floor space. When the gas temperature is high as in fired gas turbine exhaust applications, the first half can consist of bare or low-finned tubes, while the top half can have tubes with a higher fin density. The economizer is located at the gas exit section. The superheater can be located at the gas inlet. The unheated (insulated) downcomers are located at the gas turning section. These pipes carry the water to the mud drum and are responsible for the circulation. Figure 3-13b is a single gas pass cross flow design. The height has no limitations but the width depends on shipping clearances. The downcomers are located on either side of the casing external to the gas path. The boiler tubes act as risers. In larger units the steam drum can be external to the headers and external downcomers and risers should be used to promote circulation. Figure 3-13c is a widely used design, particularly in D or A or 0 type boilers or evaporators for dirty gas applications, where a large number of bare tubes are required. The last few rows of the evaporator act as downcomers and the rest as risers. No external circulation system is required unless this module is in parallel with another section. In Figure 3-13d several evaporator modules are connected in parallel through a system of downcomers and risers to the steam drum. This design is used when a screen section is needed as discussed in Chapter 1. A superheater may be sandwiched between the two evaporator modules. Also, if a large number of evaporator rows are needed and each module can be built with only say 22 to 24 rows deep, multiple modules may be required.

,

clowncoMers

oJ

b

0.

c

-aownCOMers f Figure 3-13 a-f. Various Configurations for Evaporators

172

Waste Heat Boiler Deskbook

Figure 3-13g. Single Gas Pass Evaporator [ABeO Industries]

·s-.--..

IF

------------------------------------------------------------------------------~·~·-----~

Water Tube Boilers

Figure 3-13h. Two Gas Pass Evaporator [ABCO Industries]

173

174

Waste Heat Boiler Deskbook

Figure 3-13e shows a forced circulation design with external downcomers and risers and circulation pumps to ensure flow of steam water mixture through the evaporator tubes. Figure 3-13f shows a radiant section with external downcomers and risers.

Figure 3-13i. Horizontal Gas Flow Evaporator [ABCO Industries]

CIRCULATION CALCULATIONS T~~_OJ2j~-f!!"\,:e of circulation calculatiol1s is to ensu!'e tl1CiJ !h~ he'!tect:_!!~e!!ub.~.~ are· aa:equatelicooled and whether the heat flux -inside the tubes >is·~fow enough to' p-revent condition known as Departure from Nucleate Boiling (DNB). Reference 3-1 shows the calculation procedure to evaluate circulation and DNB conditions. In order to perform circulation calculations, which is an iterative process for natural circulation boilers, the ~herma.L,

Water Tube Boilers

175

head available to move the ~e~unwate~m!xtureJhrough the system is balanced a;ainstthe~vari~us losses such as:__ '-"'--_ _ _~_:::o_._=,~~

1. Friction losses in down comer pipes. 2. Friction losses in riser pipes and boiler system. 3. Gravity losses in boiler and riser system. 4. Acceleration losses due to phase change. 5. Losses in drum internals. To start with, ~..s.ir,!:Yl~tiQll=t.
...,

176

Waste Heat Boiler Deskbook

DESIGN OF SUPERHEATERS Superheaters are used in waste heat boilers to superheat the steam that is generated in the waste heat boiler; sometimes additional steam from other boilers may also be superheated. Saturated steam may often be taken off for process from the waste heat boiler evaporator drum and the balance could be superheated. These situations arise in chemical plants and refineries. Hence one has to understand the process requirements before designing the equipment. Unlike utility boilers, the steam temperature and pressure are generally low, varying from 600 to 1500 psig and 600°F to 900°F. If steam is used for injection into a gas turbine or for use in steam turbine, high purity steam is required, on the order of 50 to 200 ppb solids in steam. Hence good steam purifying equipment may be required. Depending on the type of waste heat boiler used and the space or layout considerations, several types of superheater designs are feasible. Figure 3-14a-f shows a few types. Figure 3-14a shows a superheater with vertical headers and horizontal tubes, with the gas flow direction being horizontal. This design fits well in a short but very wide boiler, such as the twogas pass design shown in Fig. 3-13a. The headers are drainable. Baffles could be used in the headers to match the number of parallel streams. Figure 3-14b shows a unit which has horizontal tubes with horizontal headers and is designed for horizontal gas flow. The headers are baffled to account for the stream requirement; the unit is drainable and finds application in short boilers which have layout limitations. Figure 3-14c shows a drainable inverted loop superheater; this is used for small degree of superheat in units where space is a concern. The unit could be buried within a convection bundle or space. Figure 3-14d shows a design that may be used within a convection bundle. The advantage is that it could be located in a convenient gas temperature zone so as to obtain a bandwidth of steam temperature without the need for control; an example would be a fired D-type or a-type boiler. One can also avoid incidence of

-

177

Water Tube Boilers

radiant energy if necessary. The unit is drainablej the gas flows along the header direction. Figure 3-14e could be used in any situationj this design has horizontal tubes but vertical gas flow direction. Figure 3-14f shows a superheater for tall, less wide units, as in gas turbine applications. The tubes are vertical and drainable using headers at the bottom of each tube. By splitting the header, multiple streams could be arranged. The gas flow is horizontal. In addition, these designs could be arranged in parallel, counter or mixed flow configuration. The unit could be split up into two or more modules for steam temperature control purposes and could have single or multiple stream depending on steam side pressure drop and tube wall temperature. Also, the units could be of bare tube design or made compact by using finned tubes, depending on the cleanliness of the gas streamj soot blowers as required may be used within the superheater bundle or ahead of or beyond it.

~o o~ 0000 0 0000 0000

looog 0000 0000 0000

1001001

..

.. 0.

[] d

c

10

e

--~-

Figure 3-14a-f. Various Configurations for Superheaters

f

....

---~

178

Waste Heat Boiler Deskbook

Figure 3-14g. A Drainable Superheater [ABCD Industries]

Figure 3-14h. A Large Superheater Under Erectwn [ABCD Industries]

d

179

Water Tube Boilers

The superheaters could be located within a refractory or insulated duct or inside a membrane wall enclosure. Depending on the~r location, ~~:r:he~J~J;:s--couldl;>g~J(lssified as.l'adia!lt,~~Ir\i~ ~ia~~~Q!,~SQl!YE!~tive. In general, if the stearn temperature is not high, say below 900oP, superheaters are designed as convective units.

DESIGN PROCEDURE A sup~rh~
External radiation. Qr may be Q!ie to presence ota flame in the furna~e-or:d~~ to a cavity. Later we~halrseeh~w t~ estimate the contribution from this source and the effect of tube pitch on energy absorbed by each row. Qc and Qn are computed as shown earlier by estimating the convective and non-luminous heat transfer coefficients. The surface area required is then obtained from: (3-39)

U is computed as shown earlier after computing he, h n and hi. The stearn side pressure drop is obtained from: (3-40)

An example illustrates the design procedure. Example 16: A superheater has to be designed to superheat 30,000 pph of saturated stearn at 635 psia to 782°P using 150,000 pph of flue gases at 1000oP. Gas analysis is: % volume C02 =12, H20 =12, N2 =70,02 = 6. Tube size is 2 x .120 and tubes are of carbon and low alloy steel, SA 192 A and SA 213 Ttl materials. Tube pitch is 4 in., square. Assume fouling factors of .001 sq ft h P/Btu on each side and a 1% heat loss.

180

Waste Heat Boiler Deskbook

Assume also that the superheater is of convective type and the configuration is counter flow. Solution: The enthalpy of steam at exit from steam tables = 1397.5 Btu/Lb and that at inlet is 1202.6. Hence the duty = 30000 x (1397.5 1202.6) =5847000 Btu/h. The gas temperature drop =5847000/(150000 x .99 x .2851) = 138°P. Hence the exit gas temperature = 862°P. The average gas temperature = 931°P; the average steam temperature is 637°P; hence the gas film temperature = 784°P. Note that the steam inlet temperature is 492°P. Prom Appendix E, the gas properties at 784°P are: Cp = .2795, J.l = .07566, k = ,02826. Assume that there are 18 tubes/row, and the effective length is 9ft. Use 9 streams for steam flow. Hence w = 33331b/h, G = 150000 x 12/[18 x 10 x (4 - 2) =5000; Re =5000 x 2/12/ .07566 = 11015. Using Grimsons correlation, Nu = .229 x 11015.632 = 82.1 = he X 2/(12 x .02826); Hence, he = 13.90. h n using the equations (3-17) and (3-18) can be shown to be 1.5. hi using a C factor of ·335 = 2.44 x 3333.8 x ·335/1.7381.8 = 198. l/U = 1/(13.9 + 1.5) + .001 ~2/1.738;i+ 2 In (2/1.738)/24/20 + .001 + 2/(1.738 x 198) = .074; hence U = 13.5 Btu/sq / L fthF. hI, Log-mean temperature difference for counter flow arrangement is: (1000 - 783) - (862 - 492) /In (370/217) =287°P. Hence surface area S required =5847000 /(13.5 x 287) = 1509 sq ft = 3.14 x 2 x 18 x 10 x Ndj12; hence Nd = 16 rows. In order to compute the pressure drop inside the tubes, the effective length should be known. Since there are 9 streams and 18 sections and 16 rows, Le = 10 x 2 x 16 x 32 x 2.5 x 1.738 = 459 ft, using 2.5xdi for the effective length of each tube bend. Using Equation(340), L1Pi =3.36 x.02 x459 x 3.332 x .935/1.7385 = 20.3 psi where .935 is the specific volume of steam at the average temperature. Gas pressure drop may be obtained from the charts or by using Equations (3-19) to (3-21). Gas density at the average temperature of 931°P. P= .0285 Lb/cu ft;f= 11015-.15 x (.044 + .08 x 2) = .05. L1Po = 9.3 x 10-10 x 50002 x 16 x .05/ .0285 = .66 in. we.

Water Tuba Boilers

181

PERFORMANCE CALCULATIONS The procedure for evaluating the performance involves trial and error as discussed in Chapter 4 and is illustrated by the following example. Example 17: Determine the performance of the superheater designed above when the gas flow is 125,000 pph and inlet temperature is 950 o P. 25,000 pph of saturated steam at 635°F enters the superheater. Gas analysis is the same as before. Solution: The procedure involves trial and error. The steps are: 1. Assume an exit temperature for steam; compute assumed duty Qa and exit gas temperature. 2. Compute U and transferred duty Qt = U xS xLiT. 3. If Qa and Qt are close, the iteration stops or another trial is warranted. Let the exit steam temperature = 750°F; the enthalpy of steam is 1378.5. Assumed duty = Qa = 25000 x (1378.5 - 1202.6) = 4397500 Btu/h. The gas temperature drop = 4397500/125000/ ·99 / ·283 = 125°P. Hence the exit gas temperature = 825°P. Log-mean temperature difference LiT = 261°P. he = (125000/150000)·632 x 13.92 = 12.4; hn = 1.3; hence U = 12.12. Transferred duty Qt = 1507 x 261 x 12.12 = 4770000 Btu/h. Since the transferred duty is higher than that assumed, the steam exit temperature will be higher. It can be shown that the duty balances at 4570000 Btu/h; the exit steam temperature = 761°P and exit gas leaves at 820oP. The gas and steam pressure drops may be obtained as before.

NTU METHOD OF PERFORMANCE CALCULATIONS Another method of predicting the performance of superheaters or economizers is by using the NTU (Number of Transfer Units)

182

Waste Heat Boiler Deskbook

method. This procedure avoids iteration in the sense that once U is determined, the duty can be arrived at without assuming the exit temperature of either fluid. It should be noted however that in order to estimate U, one should have a feel for the gas/steam temperature profiles. The duty Qt =Cmin (Ti - tsi )

E

(3-41)

=

Where

1 - exp [-NTU(l - C)J 1 - C exp[ -NTU ( 1 - C)J

(3-42)

for counter flow arrangement is given by: NTU

= US/Cmin

(3-43)

and C = (WCp) min/(WCP)max

(3-44)

Example 18:

Solve example 6 using the NTU method. Solution: (WCP)min

= 25000 x .68 = 17000. (The specific heat of steam in the temperature band is estimated as .68 Btu/Lb F). Let us assume that U has been computed; U = 12.12 Btu/sq ft h F. NTU = 12.12 x 1507/17000 =1.064. C =17000/35071 =.485. E

= [1- e- 1.064 X(1-.485)J/[l

-

Ce- 1.064 >«1-.485)J = .586

hence Qt = .586 x 17000 x (950 - 491) = 4570000 Btu/h. The gas and steam exit temperatures may' be computed as before.

METAL TEMPERATURE CALCULATIONS M~tal temperature calculations are important for stlP~J::heate:r~ as th~ f~~~ the basis for· selection of materials and thickness of~ J~~~~,:c The allowable stress is computed at the mid-waIf temperatu~

183

Water Tube Boilers

of the tube and is used in ASME code calculations to determine the tube thickness as discussed in Appendix F. The outer wall temperature determines the material to be used from oxidation considerations. Table 3-7 shows the maximum outer wall temperatures acceptable for some commonly used superheater materials. Table 3-8 (pg 184) gives the tpermal conductivity data to be used in computing the wall temperatures. Table 3-7. Ferrous Tube Material Specification and Maximum Allowable Metal Temperatures Material. ASME Specification SA-192 SA-210 SA-210 SA-178 SA-178 SA-209 SA-213 SA-213 SA-213 SA-213 SA-213 SA-213

gr gr gr gr gr gr gr gr gr gr

Al C A C Tla T2 T11 T22 T9 TP 304 H gf TP 321 H

Maximum Allowable Temperature, of 950 950 850 900 950 975 1,025 1,050 1,125 1,200 1,400 1,400

There are several aspects to be looked into before performing the metal temperature calculations. !yhil~ _!he over~11hea!transfer coefi!cient determined eculieEis used for obtaining the surface area, ~. mate~aI~~J~~t~OE· is base~Qll. Aocal tube :",all temI':~~~u~esC1nd !teat tra?sf~r._~.gefficie~t! which vary along the length of the tube; also due to the arrangement of headers, Figure 3-15/mal~distribution can occur in the flow through each stream of a multi-pass superheater. It can be seen that a central inlet and outlet arrangement minimizes the non-uniformity in flow, while an end inlet and exit help maximize flow non-uniformity. Multiple entry and exit help minimize flow non-uniformity. A low pressure drop through the superheater is also not preferred as at low loads, the maldistribution can be significant. Also, depending on the effective length

......

~

Table 3-8. Thermal Conductivity of Metals BTU!/'f x ft'WFlft] 200

300

400

500

600

700

800

900

Carbon Steel

30

29

28

27

26

25

24

23

Carbon Moly (WYo] Steel

29

28

27

21;

25

25

24

23

1% Cr. '.11% Mo

27

27

26

25

24

24

23

21

21

2V.% Cr. 1% Mo

25

24

23

23

22

22

21

21

20

20

5% Cr. '1,% Mo

21

21

21

20

20

20

20

19

19

19

12% Cr

14

15

15

15

16

16

16

16

17

17

Material

Temp of

1000

1100

1200

1300

17

18

1400

1500

Chrome Moly Steels

:s '" CIl

it

Austenitic Stainless Steels 18% Cr. 8% Ni

9.3

9.8

25% Cr. 20% Ni

7.8

8.4

Admiralty Metal

70

75

10 8.9 79

11

95 84

11

12

12

13

13

14

14

14

15

15

10

11

11

12

12

13

14

14

15

15

89

::c In ~

tJ;:j

0

~ tj In CIl

~

00 0

~

I

Win

W'n

III 11111111

1-ctE:--

I

IIIII1IIII1

lUe.• I ___

I

o

!

o

(0)

'....a-

11111111111

IIIIII 11111

o

wex

i~

0

(b)

(C)

(d)

Figure 3-15. Static pressure distribution along the length of a header depending on the method of supply and discharge of a single-phase flow.

J.

186

Waste Heat Boiler Deskbook

of each stream, the flow can vary, like the current flowing through several different resistances arranged in parallel. In addition, one has to be careful about the non-uniformity in gas flow across the superheater. If the bundle is located in a turning section or a bend there is likely to be a large variation in the velocity profile across the duct cross-section, resulting in high local heat fluxes in some portions. In order to compute the tube wall temperature, the local maximum heat flux at the gas temperature zone desired should be obtained. qmax = 12QrFI(N...dI) + (Tg-ts)l[JI(hc+ hn) + ffi(doldi (3-45)

The second term is obvious; it is the product of overall heat transfer coefficient and the difference between the gas and steam temperature. In this calculation, the local maximum convective and non-luminous coefficients should be used. While estimating hi, the actual flow through the tube in question should be used after accounting for mal-distribution due to header arrangement and different tube lengths of a parallel muti-stream assembly. The first term involves external direct radiation Qr, which can be due to presence of flame in the furnace or radiation from a cavity upstream of the superheater. The factor F accounts for the effect of tube pitch and the fraction absorbed by each row of tubes in a bundle.

EXTERNAL RADIATION Qr If the superheater is located at the exit of a furnace or a cavity and is capable of receiving radiation from the flames or hot gases, Qr may be estimated as follows.

(3-46)

Egis determined as discussed earlier; the beamlength used would be that of the cavity or space ahead of the superheater bundle. A is the opening to the tube bundle in sq ft.

187

Water Tube Boilers

The fraction F absorbed by each row depends on the shielding effect of the tube rows and is affected by the pitch and diameter of the tubes. The first row facing the radiation would naturally receive .the maximum energYi the factor F decreases as we cross tube rows.

DISTRIBUTION OF RADIATION TO TUBES External radiation may be estimated by the following formula: 2

a = dl2st -d1St[sin-1(dlst) + {(sJd) _IrS - stld] (the term sin-1 (dlst) is in radians.)

(3-47)

The second row would absorb a(1-a) of the energy and the third 1-[a+(1-a)aJ and so on.

It can be seen that significant amount of external radiation gets absorbed in the first few rows particularly if the stld ratio is small. Also it can be seen that the radiation to the first row can be minimized by using a large stJd ratio. This is one of the reasons for using a wide pitch in radiant sections. The energy gets distributed in a more even manner if stJd is large.

Example 18: A superheater is located beyond a cavity of dimensions 3 ft wide, 10 ft high and 5 ft deep, with the opening of 50 sq ft to the superheater. Gas temperature is 1600°F and % volume C02 and H20 = 12 each. If the average surface temperature of the superheater tubes is 700°F, determine the external radiation to the bundle. If the tube size is 2 in. OD and the pitch is 8 in. wide, inline, determine the energy to each row if there are 6 rows deep. What happens if the transverse pitch is 12 in? Solution: We have to obtain E g and Qr first. The beam length of the cavity = 3.4 x volume/surface = 1.7/(1/10 + 1/3 + 1/5) = 2.68 ft =.817 m. Partial pressures of C02 and H20 = .12. Average gas temperature = 1600°F = 1144 K. K = (.8 + 1.6 x .12) (1-.38 x 1.144) x .24/(.24 x .817) .50 = .303. Eg = .9 x (1-e -.303 x .817) = .198. Qr = .1173 x .9 x .198

x (20.6 4 -11.64) x 50 = 250,000 Btu/h.

188

Waste Heat Boiler Deskbook

Using a pitch of 8 in., st/ d = 4; substituting in (3-46) a = 3.14/2/4 - .25 x [sin-1(.25) + 15.5 -4] = .361. row 2 would receiv~: (1-.361) x .361 =.231 row 3 = 1 - [1-(.361 + .231) x .361] = .147 row 4 = 1- [1-(.361 + .231 + .147) x .361] = .094 Hence the first row receives 250000 x .361 = 90250 Btu/h and so on. If st/d =6, a = .248, row 2 = .186, row 3 = .140, row 4 = .106 and so on. Note that the first row receives less energy now and hence the heat flux will be lower. It can also be seen that a substantial amount of energy gets absorbed by a fewer rows if st/d is small. As a result, the heat flux will be higher when the stld ratio is smaller. Heat flux due to external radiation in row 1 with st/d = 4 will be, assuming a height of 10 ft and 7 tubes/row, q =90250 x 12/(2 x 7 x 10) = 7735 Btu/ sq ft h. To this must be added the heat flux due to convection and nonluminous coefficients in order to arrive at the total heat flux, which will be used in the estimation of tube wall temperatures. qmax is computed by using the local maximum convective and non-luminous heat transfer coefficients. Then, the temperature drop across each resistance is estimated in order to obtain the tube wall temperature. For example in order to compute the maximum outer wall temperature in example 6, the gas inlet conditions should be used with the gas temperature of 1000°F and steam temperature of 783°F. Example 20: Determine the maximum outer wall temperature for the superheater in design case. Neglect external radiation. Solution: At the gas inlet zone the gas temperature = 1000°F and steam temperature = 783°F. he = 14.11 and hn = 1.935. Using a 20% nonuniformity in he , the local U may be computed. he = 1.2 x 14.11 = 16.95. llU = 1/(16.95 + 1.935) + .001 x 2/1.738 + .001 + .000585 + 2/1.738/198. Hence U = 16.24; q = 16.24 x (1000-783) =3524 Btu/sq ft h. Temperature drop across gas film = 3524/(16.95 + 1.935) = 187°F; drop across the fouling layer = 3524 x .001 = 3.5°F. Hence the outer wall = 1000-187-3.5 = 810°F; since the drop across the tube wall is

h

Water Tube Boilers

189

hardly 2°F, (3524 x .000585), the above value may be used for material selection. Note that if external radiation had been present, the tube wall temperature will increase. The above calculation has to be performed along the length of the superheater tube and based on the tube wall temperatures, different thickness and materials may be chosen.

FLOW IN PARALLEL STREAMS In order to determine the tube wall temperatures, the actual steam flow through each tube and not the average should be used. If the effective length of each stream or parallel pass is the same, then the flow through each tube will be the same; however due to the type of construction and configuration used, the parallel passes may have different effective lengths, which can result in different flow through the parallel passes. The tubes with the lowest flow should be looked at carefully while determining the wall temperatures. Example 21: A superheater has two parallel passes each with 30 sections. The effective length of one is 50 ft and the other, 60 ft. If the total flow in the superheater is 300,000 pph, determine the flow in each pass. All the tubes are of same size. Solution: There are a total of 60 parallel tubes handling 300,000 pph of steam; for the estimation of hi, a value of 300,000/60 = 5000 pph would have been used. However for metal temperature estimation, the flow in each pass is required. Let us assume that the nonuniformity is due to the difference in the effective length of the passes. Let us assume that 30 tubes have Le = 50 and the other 30 have Le = 60 ft. Since the pressure drop between the headers is the same, L1P = Jw2Le, where J is a constant. Let Wa and Wb be the flow in the pass with an effective length of 50 and 60 ft respectively. Then, wi x 50 = Wb 2 x 60 or Wa = 1.095 Wb. Also, 30 x (wa + Wb) = 300000 or (w a+ Wb)

190

Waste Heat Boiler Deskbook

= 10,000. Hence Wb = 4722 pph and Wa = 5278 pph.

Thus the tubes with an effective length of 60 ft have a flow of 4722 pph through them, while 5278 pph flows through the shorter tubes. The effect of variations in specific volume was neglected in the above analysis; another iteration could be made by computing the pressure drops using the actual conditions in each pass.

MINIMIZING SUPERHEATER WALL TEMPERATURES Superheater tube wall temperatures could be reduced by locating them in a cooler gas zone; however this would reduce the log-mean temperature and thus increase the surface area requirements. Parallel flow arrangement could be used versus counter flow. The high steam temperature section will then be at the cooler gas region, resulting in lower wall temperatures. The tube side coefficient could be increased by using a higher flow per tube; this increases the steam pressure drop and also improves the flow distribution at low loads, although this may incur an operating penalty in the form of additional pump power consumption; smaller diameter tubes minimize the non-uniformity around the tube periphery and hence reduce the wall temperatures. Wherever radiant superheaters are used, it may help to use a few rows on convection screen ahead of them to absorb the direct radiant energy from cavity or furnace. Header arrangements should be such that the mal-distribution is minimized. A central inlet and exit is preferred to end connections. Multiple inlets and exits also minimize the mal-distribution. Whenever the steam temperature has to be controlled, it is desirable to split up the superheater into two segments and use a desuperheater spray or some other control mechanism between the stages. This ensures that the tubes are not overheated. One has to be more careful with finned superheaters as the tube wall and fin tip temperatures can increase significantly if a high fin density is used. The procedure for computation of wall and fin tip temperatures for finned tubes is outlined in Appendix A. Steam purity is particularly important. If the drum internals are not properly designed, wet steam can be generated and solids can

-r-----=,--- - - - - - - - - - - - - - - -

___

---~

c~c

_ _ _ _ _ _ _ _ _ _ ~-_~~~~~-~--~- . . . .

Ir Water Tube Boilers

191

get deposited inside the tubes; as discussed in chapter 2, the tube wall temperature can increase as a result of any deposit or scale formation. Hence good drum internals are required if steam temperature is high or if the superheater is located in a high gas' temperature zone. Good steam purity also minimizes deposit formation in steam or gas turbine blades.,

STEAM TEMPERATURE CONTROL As in utility or industrial boilers, the steam temperature may have to be controlled in waste heat boilers. Several options are available for doing this such as the cooling of a controlled amount of steam in the drum, bypassing.)l9me steam flow aroMo,g, the sU2t::rh~~~E~,~_£ that its duty'ls reduced and tne use of .~~prBy<};Y,.e!~r for interstage desuperheating. Spray attemperation or desuperheating is widely used in waste heat boilers. One has to ensure however that the water that is being injected into the steam has a good purity. Demineralized water or condensate is preferred. A rule of thumb is that the total solids in the feed water used for injection should be nearly the same as the total dissolved solids expected in the steam. Computation of spray water is important from heat and mass balance considerations. The flow through the superheater ahead of the attemperator will be lower by the amount of spray quantity as also the flow in the economizer and hence affects the gas and steam temperature profiles. Presented below is the method of computation of spray quantity and a BASIC program that permits rapid evaluation of the same. ~.=->"O"'~="'

','

-""-_~ .••~

","_"'r--q"'\.~-'C~-);="'='~_.-?O"~:r4_-<-<'··-

,

-,~,-,7T""~':--'

DETERMINE SPRAY WATER TO DE SUPERHEAT STEAM In steam plants, steam is often desuperheated in spray type attemperators to control the final steam conditions Figure 3-16. With the help of this program written BASIC, one can determine several important variables such as:

.....

I:S

\;/s

Tl PI

g

\\If

\.If

\;/s + \.If

Tf

T2 P2

~

Tf VIs

TI

Figure 3-16 a & h. Arrangement of Desupemeaters for Steam Temperamre Control a-interstage; b-at exit of supemeater

VIs + \;If

TZ

--------------~--

Water Tube Boilers

193

•

Spray water required to reduce the steam temperature to a desired value. • Steam temperature obtained by spraying a given quantity of water into steam (reverse of first case). • Maximum amount of water that may be sprayed to saturate the steam. e Initial and final enthalpies of steam and water and saturation temperatures. The equation for energy balance around a desuperheater is: (3-48)

From the above equation;, one can solve for H2, the final steam enthalpy, and hence the steam temperature. Given Wf or by obtaining Wf and given the final steam temperature, one can determine the enthalpy. Example 22: 100,000 lb per hr of superheated steam at 1500 psia and 900°F is to be reduced to 775°F in a spray type attemperator. Determine the spray water required at a temperature of 300°F. Solution Use the program listing that appears in Figure 3-17. In the RUN mode, the screen asks for steam pressure (1500), temperature (900), steam flow (100,000), and spray water temperature (300), which are inputted as shown in Figure 3-17. . The computer prints out the saturation temperature (597°F at 1500 psia pressure) and the maximum quantity of spray water to saturate the steam (28,899 lb per hr). This information is helpful since steam temperature cannot get lower than the saturation temperature in normal spray operations. Do not inject more than the maximum spray water quantity, which is 28,889 lb per hr. The screen asks for the spray water flow or the final steam temperature, which is inputted as shown. The final results are printed out. For given conditions, spray water required is 7743 lb per hr. Enthalpies of steam and water are also shown in Figure 3-17.

(

Waste Heat Boiler Deskbook

194

Figure 3-17. Program Listing with Results [Heating, Piping, Air-Conditioning] 2000 T=273.1+(T-32)/1.8:P=P/14.696 2010 K4=80870 I /T/T:L1=10 K4* (-2641:62/T) :M1=1. 89+L1:N1=M1*P*P/T/T 2020 01=2+(3724201/T/T):Q1=Ol*L1:R1=1.89+Q1:U1=(.21828*T-1269701/T):V4=2*U1*R1-( M1/T)*1264601 2030 W1=82.54-1624601/T:Y1=2*W1*R1-(M1/T)*1624601 2040 Z=775.6+.63296*T+1.62467E-04*T*T+20.5697*LOG(T) 2050 Z=Z+.043557*(Rl*P+.5*Nl*(Yl+Ml*(Wl+V4*Nl») 2060 P=14.696*P:T=(T-273.15)*1.8+32 2070 RETURN A

DATA STEAM PRESS-PSIA,TEMP,FLOW,SPRAY WATER TEMP=? 1500,900,100000,300 SAT TEMP= 596.9012 MAX WATER FLOW= 28889.84 SPRAY WATER FLOW,FINAL STEAM TEMP(INPUT ZERO FOR UNKNOWN)=? 0,775 RESULTS STM FLOW IN= 100000 STM TEMP IN= 900 STH TEHP OUT= 775 SAT TEHP= 596.9012 SPRAY WATER FLOW= 7743.298 WATER TEMP= 300 ENTHALPIES:STM IN= 1429.198 STH OUT= 1346.058 WATER IN= 272.36 SAT STH= 1169.9 STM PRESSURE-PSIA= 1500 MAX WATER FLOW TO SATURATE STEAM= 28889.84

Ok

DATA STEAM PRESS-PSIA,TEMP,FLOW,SPRAY WATER TEMP=? 800,750,50000,250 SAT TEMP= 518.5016 MAX WATER FLOW= 8669.253 SPRAY WATER FLOW,FINAL STEAM TEHP(INPUT ZERO FOR UNKNOWN)=? 2000,0 RESULTS STM FLOW IN= 50000 STH TEMP IN= 750 5TH TEMP OUT= 678.3801 SAT TEMP= 518.5016 SPRAY WATER FLOW= 2000 WATER TEMP= 250 ENTHALPIES:STM IN= 1369.13 STM OUT= 1324.816 WATER IN= 220.48 SAT STM= 1199.4 STM PRESSURE-PSIA= 800 MAX WATER FLOW TO SATURATE STEAM= 8669.253

~--------------

Water Tube Boilers

195

1 CLS:KEY OFF 2 REM:PROGRAM RELATES DESUPERHEATING PARAMETERS 5 REM:AUTHOR-V.GANAPTHY 10 DIM P(30),HGS(30),HF(8,28) 20 FOR 1=1 TO 15 30 READ P(I),HGS(I) 40 NEXT I 50 DATA 200,1198.3,400,1204.6,600,1203.7,800,1199.4,1000,1192.9,1200,1184.4,1400 ,1175.3,1600,1164.5,1800,1152.3,2000,1138.3,220q,1122.2,2400,1103.7,2600,1082,28 00,1055.8,3000,1020.3 60 FOR TF=1 TO 5 70 FOR P=O TO 25 STEP 5 80 READ HF(TF,P) 90 NEXT P 100 NEXT TF 110 DATA 68,69.36,70.68,72,73.73,74.61,168,169.2,170.32,171.46,172.6,173.75,270, 270.53,271.46,272.36,273.33,274.28,374.8,375.4,376,376.6,377.2,377.8,488,487,487 .5,487.4 120 DATA 487.3,487.3 125 PRINT" DATA ":PRINT" " 130 INPUT"STEAM PRESS-PSIA,TEMP,FLOW,SPRAY WATER TEMP=";P,T1,WS1,TF 135 PRINT" " 150 GOSUB 1000 160 TSAT=116*PA.224:IF P<200 THEN HGS=1100+18.95*LOG(P):GOTO 190 170 X=P/200:Y=INT(X)+1 180 HGS=HGS(Y)-(Y-X)*(HGS(Y)-HGS(Y-1» 190 T=T1:GOSUB 2000 • 200 H1=Z:WMAX=(H1-HGS)*WS1/(HGS-HF) 205 PRINT" SAT TEMP=";TSAT;"MAX WATER FLOW=";WMAX:PRINT" " 210 INPUT " SPRAY WATER FLOW,FINAL STEAM TEMP(INPUT ZERO FOR UNKNOWN)=";W,T2 215 IF T2=0 GOTO 240 220 T=T2:GOSUB 2000 230 H2=Z:W=(H1-H2)*WS1/(H2-HF):GOTO 320 240 H2=(WS1*H1+W*HF)/(W+WS1) 250 IF WMAX<W THEN A$="FINAL STEAM IS WET:CHECK SPRAY WATER QUANTITY" : GOTO 310 260 T2=.5*(T1+TSAT):RA=T2-TSAT 270 T=T2:GOSUB 2000 280 H2=Z:WC=(H1-H2)*WS1/(H2-HF) 290 IF ABS«W-WC)/WC)<.005 THEN 320 295 IF RA<2 THEN RA=.3*(T1-TSAT) 300 RA=.5*RA:T2=T2+SGN(WC-W)*RA:IF T2
196

Waste Heat Boiler Deskbook

Example 23: 50,000 lb per hr of steam at 800 psia and 750°F is to be sprayed with 2000 lb per hr of feedwater at 250°F. Determine the final steam conditions. Solution Key in the data listed in Figure 3-17. Final steam temperature is 678°F. The maximum spray water to saturate the steam is 8669 lb per hr. Input data and results are shown in Figure 3-17.

DESIGN OF ECONOMIZERS The design procedure for economizers is similar to that for superheaters and hence will not be elaborated. The arrangement can be as in Figure 3-14. Steaming at part loads is a concern in gas turbine applications. Methods of overcoming this concern are outlined in Chapter 1. The tubes can be bare or finned. Low temperature corrosion is a concern and Appendix B discusses this topic in detail. One may use a heat exchanger to preheat the feed water in case the wall temperature is low enough to cause condensation of acid vapors on the tubes. It should be borne in mind th
-- 1

197

Water Tube Boilers

In Figure 3-18c the economizer is in the horizontal gas pass, with the last few rows of the tubes aiding upward flow of water, thus minimizing steaming concerns. This configuration results in a compact arrangement. The stack could be self-standing.

o 0 0 o 0 0

a

(

D

(

I:

0 0 00 0 0 00

10

0.

~

-

1

2 l.evo.poro. to~

0000 0000

J

a.econoMlzer

2

c

Figure 3-1Sa-c. Arrangement of Economizers (a-c)

PERFORMANCE OF COMPLETE HRSG The performance of the complete HRSG unit with superheater, evaporator and economizer requires several iterations as discussed in Chapter 4. T~~(lk!llru:ion.st\~~_l1lc!
-

---,,--~~---,-,.............,--~-

-

.-

~-

-

~

198

Waste Heat Boiler Deskbook

Figure 3-18d. A Large Economizer Coil [ABCO Industries]

j

Water Tube Boilers

Figures 3-18 e & f. Finned Tube Economizer and Boiler With Vertical Gas Flow Economizer [ABCO Industries]

199

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200

Waste Heat Boiler Deskbook

Figure 3-19. Printout of Results From HRSG Program WHB PERFORMANCE •• V.GANAPATHY ---PROJECT

•• SAMPLE

GAS FLOW-PPH = 400000 GAS TEMP IN -F =1500 GAS MW =28.24 GAS PRES-PSIA= 14.5 DRUM PRES-PSIA = 655 SAT TEMP-F = 496 BLOW ON = .040 FW TEMP-F =250 EXT DUTY-MMB/H = 9.00 HEAT LOSS-%= 2.00 FOUL FTR IN=.0010 PROS STM-PPH= 0 CO2 6.00

H2O 12.00

N2 70.00

HCL 0.00

H2S 0.00

DO 2.000 NW 40

DI 1.738 NO 5

EVAPORATOR CPG DUTY TG1 TG2 16.95 1500 1356 0.3001 H B L N 9.00 0.00 0.000 0.000

U 19.34 WS 0.000

ARRGT=IN

TG1 1356 1264

NO

L 12.50 12.50

6 8

10 1. 738 1.738

CONF OD CF 2.000 CF 2.000 FOUL FTR = .001

DI 1.738 1.738 1.738

NW

NO

40 40 40

2 2 12

N 0.00 0.00

H2 0.00

U 16.71 16.24

CO 0.00

SURFP 942 ST 3.750

16.945 GAS DP =

SUPERHEATER DUTY CPG 0.2963 10.64 13.02 0.2929

TG2 1264 1150

ARRGT=IN

S02 0.00

FOUL FTR =0.001 DUTY=

NW 24 24

DO 2.000 2.000 2.000

02 12.00

SURFP 942 1256

FOUL FTR =0.001 DUTY=

DELT DELPG MAXVL 930 0.75 110.8 SL TFIN TWAL 4.000 546 546

DELT· 676 638

DELPG 0.46 0.58

ST 4.500 4.500

SL 4.000 4.000

DUTY STM IN STM OUT STH HTC 566 701 280 10.64 13.02 496 641 291

U 14.62 10.76 7.76 WS 0.000 0.000 0.000

64.508 GAS DP =

MAXVL

86.2 81.9 PROP STRMS 11.41 24 12.58 24

TWAL TFIN 774 774 706 706

ARGT=IN STM FLO = 123227

EVAPORATOR TG1 TG2 CPG DUTY 1150 1104 0.2903 5.20 1104 928 0.2865 19.81 928 563 0.2771 39.50 L N H B 13.45 0.00 0.000 0.000 11.50 3.00 0.750 0.102 11.00 4.00 0.750 0.050

S03 0.00

0.749

H B WS 0.000 0.000 0.0000 0.000 0.000 0.0000

STH PRES -PSIA= 625

CH4 0.00

SURFP 563 3573 25987 ST 3.750 3.750 3.750

SPRAY= 5665

DELT DELPG 631 0.12 515 0.71 196 2.51 TFIN SL 4.000 521 4.000 841 4.000 759

MAXVL

60.9 93.9 77.5 TWAL

521 619 576

3.339

ECONOMIZER DO 2.000 NW 40

01 1.738 NO 20

TG1 563 L 12.00

WAT TEMP IN= 250 ECO FOUL FTR= .001

TG2 CPG DUTY 312 0.2666 26.30 B N H 4.00 0.750 0.050

OUT =

454

ARRGT=IN

U SURFP DELT DELPG MAXVL 7.05 44791 83 1.99 44.7 WS ST SL WATDP STRMS 0.157 4.000 4.000 16.77 10

WAT HTC= 1451 % STH = 0

BOILER DUTY= 131.413 TOT GAS PROP=

WAT FLOW= 122264

STM SURF= 0

SPRY TEHP = 250

7.11 STEAM GEN= 123228 TOT BLR DUTY= 140.413

##############################################################################

Water Tube Boilers

201

automatically arrives at the firing temperature required to generate the desired final steam quantity if the fuel analysis is specified. The mechanical configuration is inputted along with gas inlet conditions entering the burner system (if used). The steam and gas side pressure drops, tube wall temperatures, spray water, firing temperature, gas analysis and properties are all automatically computed in seconds; provision exists to superheat steam from other boilers or to withdraw a certain quantity of saturated steam from the waste heat boiler for process purposes and superheat the balance quantity of steam. Fouling factors or tube configurations such as pitch, size could be varied for each section to simulate operating conditions.

REFERENCES 1. V. Ganapathy, "Applied Heat Transfer," Pennwell Books,

Tulsa, 1982 2. V. Ganapathy, "Charts can help give quick engineering estimates of gas pressure drop in tube banks," Oil and Gas Journal, March 1, 1982 3. V. Ganapathy, "Charts simplify estimation of non-luminous heat transfer coefficients," Hydrocarbon processing 4. V. Ganapathy, "Basic programs for steam plant engineers," Marcel Dekker, New York, 1984 5. V Ganapathy, "Simplified approach to designing heat transfer equipment," Chemical Engineering, April 13, 1987 6. V Ganapathy, "Determine spray water to desuperheat steam," Heating, Piping, Air-conditioning, December 1987 7. V. Ganapathy, "Steam Plant calculations manual," Marcel Dekker, New York, 1984

NOMENCLATURE C - factor used in Equation (3-9) and (3-44) Cp - gas specific heat, Btu/lb F d - tube diameter, in-subscript i and 0 refer to inside and outside f - friction factor for pressure drop; subscript g stands for gas.

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202

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Waste Heat Boiler Deskbook

if - fouling factor, sq ft h F /Btu -

subscripts i and 0 refer to inside and outside tubes. F - fraction of energy absorbed by direct radiation F1,F2" F3 - factors used in equations (3-25), (3-35) G - gas mass velocity, Lb/sq ft h H - enthalpy, Btu/lb L1 denotes difference in enthalpy. Subscript s stands for stearn h - heat transfer coefficient, Btu/ sq ft h F; subscripts c-convective, iinside tubes, n- non-luminous, o--outside hI heat loss factor; 1 % heat loss means hI =.99 k - gas thermal conductivity, Btu/ft h F K -factor used in equations (3-16) , (3-18) Km - tube metal thermal conductivity, Btu/ft h F L - length, ft or m; subscript e stands for effective length, ft. Nd, NH - number of tube rows deep N w _ number of rows wide NTU - Number of Transfer Units Nu - Nusselt Number Pc, Pw - partial pressure of carbon dioxide and water vapor Pr - Prandtl Number ~p - pressure drop; subscripts 0 or g stand for gas, in wc and subscript i for tube side, psi q - heat flux, Btu/sq ft h; subscripts i and 0 stand for inside and outside Q - duty, Btu/h; subscripts a and t stand for assumed and transferred: c, nand r for convective, non-luminous and radiant Re - Reynolds Number 5 - surface area, sq ftsl, stlongitudinal and transverse pitch, in L1 T -log-mean temperature difference, F T1, T2 - temperature entering and leaving, for gas Tg - average gas temperature, F or K ts - saturation stearn temperature, F U - overall heat transfer coefficient, Btu/ sq ft h F; subscripts i and 0 - stand for inside and outside tubes. v, V - specific volume, cu ft/Lb and velocity, ft/ s Wo, Wg - gas flow, pph Wi, Ws, Wf - Flow inside tubes, pph, inside, stearn and feed water E - effectiveness of exchanger, Equation (3-42)

J

--------

Water Tube Boilers

gas emissivity p-- gas density, lb/cu ft cr - Steffan Boltzman constant Y - factor defined in Equation (3-37) Eg-

203

~

Chapter 4

HRSG Design and Performance Simulation IMPORTANCE OF HRSG SIMULATION Water tube HRSGs (Heat Recovery Steam Generators) or WHBs (Waste Heat Boilers) as they are sometimes called are widely used in various applications as discussed earlier in Chapters 1 and 3. Methods of performing their design and performance calculations were elaborated in Chapter 3. The mechanical configuration of the WHB had to be known in these calculations and the heat transfer coefficient U was evaluated in each case in order to predict the duty and heat balance. However when consultants and system engineers are in the early stages of developing a project they need some basic information about HRSG performance and capabilities at different plant loads and operating points. For example in a gas turbine cogeneration plant, if the HRSG performance can be simulated at different· ambient conditions and loads, they can study the complete system performance knowing the steam generation, auxiliary fuel consumption, system efficiency etc., and use the information to select plant auxiliaries such as steam turbine, condenser and deaerator or evaluate the overall plant economics. They may obtain this information from HRSG suppliers; however, during the early stages of a project, a lot of time can be wasted by consultants if they depend on HRSG suppliers for this information for the simple reason that too many alternatives or gas turbines may be involved and HRSG suppliers may take a long time to design a HRSG for each alternative and then provide the 205

206

Waste Heat Boiler Deskbook

performance data for different operating conditions; also, they may not respond as quickly as the consultants would like to, particularly if it is a study and the gas inlet conditions are not firm. Also it would be in the interests of consultants to be able to simulate the HRSG thermal performance and not be tied to any particular HRSG design or supplier. Since they are only interested in the thermal performance, the mechanical details are not important at this stage. Once the study is completed and a particular gas turbine or heat source is selected, then the consultants can approach different HRSG suppliers for proposals, knowing what kind of performance and steam production they can expect for their system; this also makes them knowledgeable and the process of evaluation of HRSG bids can be quickened. Another reason for consultants to perform this simulation themselves is that they can optimize the steam system (whether to maximize HP or IP steam, where to take off the steam for deaeration, how much LP steam should be generated etc.,) as they are more familiar with cost of steam, fuel and the utilities and the . steam needs of the customer. Is there a method of evaluating HRSG performance without knowing the mechanical design details such as tube size, length, fin density etc? Can one obtain information on steam generation capabilities and temperature profiles without doing a mechanical design? The answer is YES and this chapter will elaborate the method, which can be effectively used by plant engineers, consultants and even HRSG designers to simulate and optimize a HRSG, be it a single pressure system or a complex multi-pressure unit and integrate it into the steam plant.

DESIGN AND PERFORMANCE CALCULATIONS There are two basic types of calculations performed while selecting a HRSG; one is the "design" calculation; in this mode, the basic configuration or the disposition of various surfaces is arrived at including the gas and steam temperature profiles and the steam production. Once a unit is designed, it means that the surface areas of

j

r 2rJ7

HRSG Design and Performance Simulation

the various components such as the superheater, evaporator and economizer and the configuration are frozen. Two important variables namely the PINCH point and APPROACH point determine the complete design, the design temperature profiles and the steam production, as explained below. The "performance calculation" tells the user how the HRSG that has been designed performs at different other gas flows, inlet temperatures, or steam parameters. This is a complex iterative procedure and will be discussed later. There is only one design point but there are several off-design performance points.

DESIGN CALCULATIONS AND DESIGN TEMPERATURE PROFILE Figure 4-1 shows the temperature profile for a simple HRSG for a single pressure case with a superheater, evaporator and economizer. The desired superheated stea:n:Lpressu~e aEd temperature ani kno'Y.-n as al~ thelee
gas

Q12 = duty of superheater plus evaporator bdx(hf-hwl)] = Wg xC p x(Tgl-T g3) x hlf

= Wsd x[(hs2-h!!94) + (4-1)

From the above, Q12 can be computed as Tgl, Tg3 and Wg are known. C p is obtained from the gas analysis. hlf is the heat loss

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Waste Heat Boiler Deskbook

208

factor-typically ranging from .98 to .99 depending on the size of the HRSG and insulation or casing design. Example 8 of Chapter 2 shows how one can compute this loss .. Wsd, the design steam flow rate may be solved for; saturated steam enthalpy hsI' superheated steam enthalpy hs2 and enthalpy of feed water hw2 are obtained from steam tables. The superheater duty::;: QI ::;: Wsd x (hs2 - hsI)::;: Wg xC p x (TgI-Tg2) xhZf

(4-2)

Hence T g2 may be obtained. Since Q 12 ::;: QI + Q21 Q2 the evaporator duty may be obtained; also, Q2 ::;: Wg xC p x (Tg2 - Tg3) x hI! (4-3) The economizer duty is then::;: Q3::;: Wsd x (1 + bd) x (h w2 - hwI)::;: Wg x Cp x (Tg3 - Tg4) x hlf

(4-4)

Hence one can solve for Tg4 from the above equation. Thus the complete gas and steam temperature profiles and the steam flow are arrived at. (hf is the enthalpy of saturated water.)

Tg1

ts2 ts tw2

Evaporator

Economizer

Superheater

Figure 4-1. HRSG Temperature Profiles

,.--

tr

------------

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HRSG Design and Performance Simulation

209

GUIDELINES FOR SELECTING PINCH AND APPROACH POINTS A lot of confusion exists among engineers regarding the above two important variables. There are several important facts that should be kept in mind before selecting them. 1. Pinch and Approach points cannot be arbitrarily selected. If done so, a temperature cross situation can arise as discussed below (4-5) (4-6) In the above equations, blow down and heat loss were neglected. Neglecting variations in Cp, we have, dividing (4-5) by (4-6) : (Tg l- T g3)j(Tg l-Tg4) = (hs2 -hw2)j(hs2 -hwl) = X

(4-7)

where X is a constant for the steam parameters in consideration. Two conditions must be met for steam generation to occur: Tg3 > ts and Tg4> twl

(4-8)

If the pinch point or approach is selected arbitrarily, it is likely that T g4 can be lower than tw l, causing a temperature cross situation. Substituting ts for T g3 and twl for T g4 , we can obtain a critical inlet temperature T glc, above which the feed water temperature governs the temperature profile and below which, the pinch point governs the temperature profile.

From (4-7) , (Tglc-ts)j(Tglctwl)= X or Tglc = (ts-Xtw l)j(l-X)

(4-9) (4-10)

Example Let steam pressure = 585 psig, steam temperature = 700°F, Feed water temperature = 250°F. Let approach point = 20°F. ts = 488°F; hence tw2 = 468°F. From steam tables, hs2 = 1351.8, hwl = 219.5, hw2 =

210

Waste Heat Boiler Deskbook

450.7. then X = .796. Prom (4-10) Tglc = (488-.796 x 250)/(1-.796) = 1416°P. Above 1416°P, the feed water temperature governs the profile. To illustrate this, let us assume that we selected a pinch point of 30 0 P when Tgl = 1600oP. The Tg3 =488 + 30 =518°P. Then from (4-7), (1600-518)/(1600-Tg4) = .796 or Tg4 = 240 oP, which is below 250oP, resulting in a temperature cross situation, which is not practical. Let us say we selected 2800 P for Tg4, then: (1600-T g3)/(1600280) = .796 or T g3 = 550 oP, a pinch point of 62°F. This is a feasible value. With a higher approach point, Tlgc increases and the pinch point would not be governing the temperature profile. Let us select tw2 = 400 0 P in the above case. Then X = .862. Hence Tglc = 1974°P. Since Tgl = 1600oP, the pinch point can now be lower. Try 30oP. Then (1600-518)/(1600-Tg4) = .862 or Tg4 = 345°P, which is higher than T wl and hence thermodynamically feasible. However whether it is feasible from a mechanical size consideration has to be seen. Thus temperature cross situations can easily be handled by increasing the pinch or approach points or both. However the point that has been brought out is that they cannot be arbitrarily selected, and the inleegas temperature plays a role. Hence one has to be careful in selecting pinch and approach points in the case of fired units or in HRSGs where the gas inlet temperature is high above 1000o P. It is suggested that pinch and approach points not be selected in the fired mode (in the case of gas turbine HRSGs) for the above reason. Note that the gas flow does not matter. 2. Pinch point has to be selected such that the HRSG evaporator is of reasonable size or one that can be built and shipped. If a very low value is selected, then the surface area required may be too much and the HRSG will be cost prohibitive. If a very high value is used, then one may not be able to recover adequate energy from the gas stream. See Chapter 3 for an example on optimum pinch point selection. Choice of Pinch and approach points are also affected by the type of surface used, whether bare or finned. With bare tubes, it is impractical to obtain a pinch point of less than 100 to 150°F. The number of tube rows would be too many and hence the gas pressure

211

HRSG Design and Performance Simulation

drop would be very high. Appendix A compares two designs with bare and finned tubes for the same pinch and approach points . . The following table may be used as a guide for selecting pinch and approach points. The author suggests that for gas turbine HRSGs, these values be selected in the unfired mode even though the unit may be operating in the fired mode .. Table 4-1. Suggested Pinch and Approach Points Pinch Point, F a. Evap type

bare

Approach Point,F

finned

b. Inlet gas temp, F

1200-1800 750-1200

130-150, 30-60 80-130 10-30

40-70 10-40

Por a gas turbine application, pinch and approach points of 15 to 20 0 P are reasonable in the unfired mode. Por a dirty gas application such as MSW, with a gas inlet temperature of 1500 to 1800o P, a pinch point of 130 to 1500 P and approach of 40 to 700 P may be reasonable. 3. Pinch and Approach points should be selected in the unfired mode for gas turbine applications, even though the unit may be operating in the fired mode all the time. The following are the reasons: a. A temperature cross situation can result as discussed above. Consultants who are not familiar with HRSG performance sometimes call out the design pinch point or approach point without realizing the implications of the gas inlet temperature or the practicality of the value chosen. b. Por example if a pinch point of 20 0 P is selected with a firing temperature of 1600oP, the boiler size would be huge and the cost and gas pressure drop would be unreasonable. It is difficult to visualize a HRSG size in the fired mode; having designed several units,the author recommends that the pinch and approach points be selected in the unfired mode and the performance evaluated in the fired mode. Through experience, it can be said that pinch and approach points of 10 to 20 0 P are feasible for unfired gas turbine units.

212

Waste Heat Boiler Deskbook

c. Steaming in the economizer is a concern with gas turbine units. This occurs during cold ambient unfired conditions when the gas flow is higher and the inlet gas temperature lower than normal conditions. The reason for this will be given later with an example. Hence if one selects the pinch and approach to avoid steaming at the unfired cold ambient conditions, steaming can be avoided at any other operating condition. Note that as the gas inlet temperature increases, the pinch and approach points will increase and vice versa. d. If the selection of steam temperature is done in the fired mode, it will not be achieved in the unfired mode due to the lower gas inlet temperature. In units where the steam temperature has to be controlled over a wide load range, the steam temperature has to be achieved in the unfired as well as in the fired modes. The excess surface area of the superheater or the amount of spray water required for steam temperature control cannot be visualized if the temperature profiles are selected in the fired mode. Also, several performance checks have to be made to ensure that the desired steam temperature range is being achieved at different unfired and fired conditions. On the other hand if the steam temperature is selected in the unfired mode, it can be certainly achieved in the fired mode and through some steam temperature control methods, we can obtain the desired value at any other higher gas inlet temperature. 4. Once the pinch and approach points are selected and the design temperature profile is arrived at, the design is nearly complete. If the HRSG operates under a different mass flow, inlet temperature, gas analysis or steam pressure, the HRSG performance or the temperature profile and the steam flow would change. This information is obtained through "performance" calculations, which is discussed later. The pinch and approach points would change with different case of inlet conditions. They fall in place and have to be evaluated using complex iterative procedure. The important thing to keep in mind is that pinch and approach points are not constants but vary with variations in gas flow, inlet temperature, gas analysis, steam pressure and feed water temperature. Hence consultants and engineers should not use these terms loosely but associate them with a particular gas inlet condition and

- - - - - - - - - - - - - - - - ---

HRSG Design and Performance Simulation

213

steam parameters. We will now illustrate the design and performance calculation procedure with examples.

DESIGN TEMPERATURE PROFILE AND CALCULATIONS The superheater and economizer are assumed to be in counterflow arrangement, which is the widely used configuration. Example 1.

A gas turbine HRSG is to be designed for the parameters shown in Table 4-2. Determine the gas/steam profiles and the steam flow. Let the gas pressure drop = 6.0 in. We. Let superheater pressure drop = 7 psi. The drum pressure = 450 + 7 = 457 psig. The saturation temperature is 460oP. Gas temperature leaving the evaporator = 460 + 20 = 480 oP. Compute the gas properties for the given analysis. The data are shown in Table 4-3. Using an instantaneous specific heat of 0.267 for the range 900 to 480 0 P and a heat loss factor of 0.99, the duty in the superheater and evaporator is: Q1 + Q2 = 1qO,OOO (0.267) (0.99) (900-480) = 16.65 x 106 Btu/h = Wsd[(l,330.8-431.2) + 0.02(442.3431.2)] = 899.8 Wsd. Where 1,330.8 = enthalpy of superheated steam at 450 psig, 650 oP, 442.3 = enthalpy of saturated water at drum pressure, 431.2 = enthalpy of water entering the evaporator at 450 oP. 0.02 is the blow down factor. Prom the above, Wsd = 18,510 pph. Superheater duty, Q1 = 18,510(1,330.8-1,204.4) = 2.34 x 106 Btu/h, where 1,204.4 is the enthalpy of saturated steam. Gas temperature drop in the superheater = 2.34 x 106/(150,000) (0.273) (0.99) = 58°P Hence, gas temperature to evaporator = 900-58 = 842°P. Q2 = Evaporator duty = 16.65-2.34 = 14.31 106 Btu/h. Economizer duty = 18,510 (1.02) (431.2209.6) = 4.19 x 106 Btu/h, where 209.6 is the enthalpy of feed water at 240o P.

~

Waste Heat Boiler Deskbook

214

Table 4-2. Data for "Design" and "Performance" Calculations

1. 2. 3. 4.

Case no. Gas flow, pph Exhaust temp., of % vol CO2 H2 O

N2

02 5. 6. 7. 8. 9. 10. 11. 12. 13. 14. 15.

Steam press., psig Steam temp., of Feed water temp., OF Slowdown, %

Design

Perf

1 150,000 900 3 7 75 15

2 165,000 840 3 7 75 15

450 650 240 2

450

? 240 2

Process steam, pph Heat loss + margin, % SH press. drop, psi Pinch point, of

7 20 10

Approach point, of Steam flow, pph ? Ambient temp., of 80 Natural gas used: % vol Cl = 96, C2

? ? ? ?

Perf

Perf

3 4 165,000 , 165,000 840 840 3 3 7 7 75 75 15 15 450 300 ? 650 240 240 2 2 2,500 1

? ? ?

? ? ?

26,000 26,000 50 50 50 = 2, C3 =2. Note that steam is required at a controlled temperature of 650 0 P in case 4. In cases 2 and 3 it is uncontrolled. Also, in case 4, 2,500 pph of saturated steam is taken off the drum and the balance of 26,000 pph is to be superheated to 650 oP. The steam exit pressure is 300 psig in case 4. It will be seen later that cases 3 and 4 are fired and cases 1 and 2 are unfired. I

Table 4-3. Part a-Gas Properties-Unfired Gas (% vol C02 = 3, H20 = 7, N2 =75, 02 = 15) Temp., of Cp J1 k 900 650 400

0.2736 0.2658 0.2584

0.083 0.0724 0.0612

0.0304 0.0261 0.0218

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215

HRSG Design and Performance Simulation

Part b-Gas Properties-Fired Case (% vol C02

=3.45, H20 =7.87, N2 = 74.65, 02 =14.01)

Temp.,

Cp 0.2800 0.2689 0.2583

of

1,050 700 350

Jl

k

0.0887 0.0743 0.0586

0.0330 0.0267 0.0208

Units: Cp-Btu/lboP, Il-lb/ft h, k-Btu/ft hOP (Interpolate for gas properties at intermediate temperatures) Gas temperature drop in the economizer = 4.19 x 106/(150,000) (0.26) = 109°P. The gas specific heat at the average gas temperature in the economizer, obtained from Table 4-3 by interpolation, is 0.26. Hence, the exit gas temperature = 480-109 = 371°P. The temperature profile is shown in Figure 4-2. Using a similar approach, the temperature profiles'for any other pinch or approach points can be obtained.

EVAP

SH

ECO

Resuhs-deslgn case,-unlired Amb. lemp. - OF = 50, ReI. hum. - % = 0, Heal loss - % = I. Gas lemp. 10 HRSG - of = 000 Gas IIow·pph = 150,000, % vol. CO2 = 3, H20 = 7, N, = 75, 0, = 15, SO, = 0 Gas lemp. in-oul - OF SH EVAP 000 ~ ECO 481

I 8~

481 372

Wat./stm. in-out - OF 461 451 240

I

650 461 451

Duty Press. Flow pph MMb/h psig 2.35 14.38 4.22

450 457 467

Pslm. %

18,610 100.0 18,610 100.0 18,982 0.0

Pinch OF

Apprch. OF

20

10

Figure 4-2. Design Case 1 Results [Hydrocarbon Processing]

-4

216

Waste Heat Boiler Deskbook

To proceed with the performance calculations for case 2 shown in Table 4-2, a few parameters should first be computed, as discussed in insert B. These parameters help relate the heat transfer coefficients in the "design" mode to those in "performance." . Por the superheater: KI = QIj(,1TI) (WgO.65) Fg) where ,1TI = log-mean temperature difference = [(842--460)-(900-650)]fln[(842460)/(900-650)] = 311°P, Fg = CpO.33 kO.67/j.l0.32 = 0.135, using a Cp = aO.273, k = 0.029,and j.l = 0.0826. Hence KI = 2,340,000/ 150,000°.65 /311/0.135 = 24.10. Similarly for the evaporator K2 = 387.6 and K3 = 218.4 for the economizer. KI, K2 and K3 will be used to compute (US)p, the product of U and 5 in the performance modes as discussed in insert B.

PERFORMANCE CALCULATIONS Let us see how the unit performs when the conditions are as shown in case 2, Table 4-2. The gas flow is 165,000 pph at 840oP. The gas analysis, feed water temperature and steam pressure remain the same as earlier. The performance of the HRSG is arrived at through an iterative process described in Inserts A and B. Trial 1. As a first approximation, assume that the steam flow is proportional to the gas flow and temperature drop. Ws = 18,510 (165,000/150,000) (840-371) /(900-371) = 18,050 pph. Superheater Performance. Let tS2, the steam exit temperature = 640 oP. Then, from steam tables, the enthalpy = 1,325 Btu/lb. The assumed duty = 18,050 . (1,325-1,204.4) = 2.177 x 106 Btu/h. Gas temperature drop = 2,177,000/(165,000) (0.99) (0.271) = 49°P. Hence,gas temperature leaving the superheater = 840--49 = 791°P. Compute the transferred duty, Q1t, using Equation (4-12) in insert B. Fg = 0.135, Wg = 165,000, K1 = 24.1, WSd = 18,510, Ws = 18,050. Hence (US)p = 165,0000.65(0.135) (24.1) (18,050/18,510)0.15 = 7,974. ,1T = log-mean temperature difference = [(840-640)-(791460)]fln[(840-640)/(791--460)] = 260o P. Hence, Q1t = 7,974 (260) = 2,074,000 Btu/h. This is close to the assumed value. If it were not, we

r

------------

HRSG Design and Performance Simulation

217

would have to assume another steam temperature and repeat the steps. Let us continue. Evaporator Performance. Compute Fg at the average gas temperature in the evaporator. Fg = 0.129, K2 = 387.6. Then, (US)p = 165,0000.65 (0.129) (387.6) = 123,123. Using Equation (4-18), [(791-460)/(Tg3-460)] = e(123,123/165,000/0.99/0.266) = 17.00. Hence Tg3 = 480 o P; Q2 = 165,000(0.99) (0.266) (791-480) = 13.522 x 106 Btu/h. Note that the gas properties have to be interpolated for the values at the average gas temperature in the section. Economizer Performance. Let the water temperature leaving the economizer be 450 o P. hW2 = 431.2 from steam tables. Assumed duty Q3a = 1.02 (18,050) (431.2-209.6) = 4.08 x 106 Btu/h. The gas temperature drop = 4,080,000/165,000/0.99/0.26 = 96°P, exit gas temperature = 480-96 = 384°P. Fg = .120, K3 = 218.4, Hence (US)p = 218.4 (165,0000.65 ) (0.120) = 64,535. Transferred duty = Q3t = 64,535(72.7) = 4.69 x 106 Btu/h, where 72.7 is the log-mean temperature difference. Since the transferred duty is more than the assumed, let us repeat the calculations with say tW2 = 457°P. Q3a = 18,050 (1.02) (439-209.6) = 4,230,000 Btu/h. The exit gas temperature = 381. LiT = 65. Then Q3a, = 64,535 (65) = 4,190,000 Btu/h. Since this is closer to Q3t , let us continue. The total transferred duty = Q1t + Q2t + Q3t = 2.07 + 13.52 +4.19 = 19.78 MMBtu/h. The corrected steam flow, Wsc = 19.78 x 106/[1,325-209.6 + 0.02(442-209.6)] = 17,660 pph, per Equation (4-24). Since this is not close to the assumed value of 18,050 pph, another trial is warranted. Try Ws = 17,770 pph. Trial 2. Let the revised steam flow = 17,700 pph. Pollow a similar procedure as before. Superheater Performance. Let t52 = 640 o P. Q1a = 17,700 (1,325-1,204.0) = 2.134 MMBtu/h. Gas temperature drop = 2,134,000/(165,000) (0.99) (0.271) = 48°P. T g2 = 840-48= 792°P. LiT = 260o P. Fg = 0.135. Kl = 24.1. Then, (US)p =

----

.

218

Waste Heat Boiler Deskbook

165,0000.65(0.135) (24.1) (17,700/18,510)°·15 = 7,957. QIt = 7957(260) = 2.07 MMBtu/h. SinceQ1t is less than Q1a, try a lower steam temperature, say 635°P. Then Q1a = 17,700(1,332-1,204.4) = 2.081 MMBtu/h. Gas Temperature drop = 47°P. Tg2 = 840-47 = 793°P. LiT = 264 °P. Hence, QIt = 7,957(264) = 2.1 MMBtu/h. This is close enough. Continue.

Evaporator Performance. Solve for Tg3 as before. [(793--460)/(Tg3-460)] = 17.00; hence Tg3 =480 o P. Q2 = 165,000 (0.99) (0.266) (793-480) = 13.6 MMBtu/h. (The factor 17 computed from Trial 4-1 is unchanged.) Economizer Performance. Let tW2 = 455; hW2 = 436.8; Q3a =17,700(1.02) (436.8-209.6) = 4.1 MMBtu/h. Gas temperature drop = 96°F. Tg4 = 480-96 = 384°P. Li T = 68°P. Using the same (US)p as before, Q3t = 64,535(68) = 4.36 MMBtu/h. Since the variation between Q3a and Q3t is large, try tW2 = 458°P. Then, Q3a = 4.14 MMBtu/h. Tg4 = 383°P. f).T = 64.6°P. Hence, Q3t = 64.6(64,535) = 4.16 MMBtu/h. This is quite close. The total transferred duty = 2.1 + 13.6 + 4.16 = 19.86 MMBtu/h. The corrected steam flow, Wsc = 19.86/[(1,322-209.6) + 0.02(442-209.6)] = 17,770 pph.Since this is close to the assumed value of 17,700, let us stop here. The final temperature profile is shown in Pigure 4-3. The gas pressure drop, using Equation (4-25) = 6(165,000/150,000)2[0.5(840 + 383) + 460]/[0.5(900 + 371) + 460)] = 7.1 in. We. Performance Check-Fired Case. Let us check the performance for case 3 shown in Table 4-2, where it is desired to make 26,000 pph of steam. The steam temperature is uncontrolled. It is obvious that with same inlet gas conditions as in the earlier case, we need additional fuel input to the HRSG to generate 26,000 pph. The procedure is similar to the earlier one. However, additional steps are necessary to iterate for the firing temperature, as discussed in Insert A. The method of computing the fuel input, firing temperature and gas analysis is discussed elsewhere. Let us only check the final results which are shown in Pigure 4-4.

219

HRSG Design and Performance Simulation

Results-performance case-unfired Amb. temp. - of • 50, Rei. hum. - % - 50, Heat loss - % = 1, Gas temp. to HRSG - of = 840 Gas flow-pe~_~J.\!9~ti % vol. CO, = 3, H,O = 7, N, = 75, 0,.15, SO,. 0 Gas temp.

in-out - of SH EVAP ECO

Wat.lstm. In-out - OF

840

634

793

461

460

456

MM blh pslg

Pstm. %

450 456 486

100.0 100.0 0.0

Duty

2.08 13.61 4.19

Press. Flow pph

Pinch

Apprch.

of

of

19

3

Figure 4-3. Performance Case 2 Results

240 Results-performance case-fired Amb. temp. - of 50, ReI. hum. - % = 50, Heat loss - % 1, Gas temp. to HRSG - of = 840 ~~OOO~ . % vol. CO, = 3, H,O = 7, N, = 75, 0,. 15, SO, . .

=

Gas temp. In-out - of BURN 840 1,035 SH EVAP 954 ECO 490

1,035 954 490 382

=

Wat.istm. In-out - of 0 462 436 240

0 677 462

436

Duty Press. Flow MMb/h psig pph 9.29 3.69 20.52 5.45

Pstm. %

0 434 0 450 g§.963 100.0 464 25,963 100.0 474 28,462 0.0

=0

Pinch Apprch.

of

of

27

26

Gas flow after HRSG = 165,434, % vol. CO, = 3.45, H,O = 7.87, N, = 74.65, 0, = 14.01, SO, = 0.00 Fuel - gas: analysis - % volume, 1 methane = 96, 2 ethane = 2, 3 propane. 2, Ihv-Btu/cu ft = 957, Ihv-Btu/lb = 21,398, aug. alr-pph = 0

Figure 4-4. Performance Case 3 Results

220

Waste Heat Boiler Deskbook

Superheater Performance. Table 4-3b shows the gas properties for the gas analysis after combustion. From the printout, Figure 4-4, it is seen that the HRSG gas inlet temperature is 1,034°F and the burner fuel input is 9.29 MMBtu/h (LHV basis). Wg = 165,430; Ws = 26,000; ts2 = 677°F. Fg at the average gas temperature is 0.142. The saturation temperature is 462°F, at the corrected drum pressure of 463 psig. Q3a = 26,000(1,346.0-1,204.3) = 3.69 MMBtu/h. Gas temperature drop = 3,690,000/(165,430) (0.99) (0.278) = 81°F. Exit gas temperature, T g2 = 1,034-81 = 953°F; ,1T = 420°F. Kl = 24.1. (US)p = 165,4300.65(0.142) (24.1) (26,000/18,510)0.15 = 8,840. Then, Q1t = 420(8,840) = 3.71 MMBtu/h. Evaporator Performance. Fg = 0.135; K2 = 387.6; hence (US)p =165,4300.65(0.135) (387.6) = 129,437. Using Equation (4-18), [(953-462)/(T g3-462)] = e129 ,437 /165,430/0.99 /0.27) = 18.67. Hence, T g3 = 489°F. Q2 = 165,430(0.99) (0.27) (955-489) = 20.52 MMBtu/h. Economizer Performance. tW2 = 435; hW2 =414.45; Q3a = 26,000(1.02) (414.45-209.6) = 5.43 MMBtu/h. Gas temperature drop = 128°F; T g4 = 489-128 = 361°F. ,1T = 83°F. K3 = 218.4; Fg = 0.120; {US)p = 165,430°·65 (218.4) (0.120) = 65,000. Hence Q3t = 83(65,000) = 5.4 MMBtu/h. Total energy transferred = 3.71 + 20.52 +5.4 = 29.63 MMBtu/h. Wsc = 29.63 x 10 6 /[(1,346.7-209.6) + 0.02 (442.6-209.6)] = 25,970 pph. The gas pressure drop could be corrected as before. This gives an idea of the complexity of performance calculations if fuel firing is involved. Several iterations of performance calculations would be required before the correct firing temperature is arrived at. Also, if the steam temperature has to be controlled, the superheater has to be split up into two stages with a spray desuperheater in between. The method of computing the spray water for steam temperature control is discussed in Chapter 3. In such an HRSG, more iterations are involved before the spray water flow and the final temperature profiles are arrived at. Without a computer it would be extremely tedious and time

221

HRSG Design and Performance Simulation

consuming. Figure 4-5 shows the results of case 4 where steam temperature control and fuel firing are involved.

Fuel

840

~

__ __-+____________-4______ ~

~240

Resu~s-performance case-fired Amb. temp. - OF = 50, Rei. hum. - % - 50, Heat loss - % = 1, Gas temp. to HRSG - of = 840 Gas flow-pph = 165,000, % vol. CO, = 3, H,O = 7, N, = 75, 0, _ 15, SO, = 0

--

Gas temp.

in.-out - of

Wat./stm. In-out - of

BURN 0 840 1,067 SH 1,067 979 427 DESH 1,032 1,032 587 EVAP 979 458 405 ECO 343 240 458

0 650 542 427 405

Duty Press. Flow MM blh pslg pph

lo.a2 4.05 0.00 23.03 4.90

0 300 309 318 328

506

25,968 608 g~d§!t

28,610

Pstm. %

0 100.0 0 100.0 0.0

Pinch

of

Apprch. OF

31

22

Gas flow after HRSG = 165,5Q6, % Vol. CO,. 3.52, H,o = 8.01, N, - 74.60, 0, = 13.84, SO, = 0.00 Fuel - gas: analysis - % volume, 1 methane = 96, 2 ethane = 2, 3 propane = 2, Ihv-Btu/cu ft = 957, Ihv-Btunb • 21,398, aug. air-pph =0

Figure 4-5. Performance Case 4 Results

Note that gas inlet temperature is l,067°F. The spray quantity has been arrived at based on a split i,n the ratio of 6:4 in design U times 5 values between the first and second stages of the superheater. This ratio is built into the program. Slight changes in the temperature profile and spray quantity can result due to a different split in the surfaces between the two stages of the superheater while actually building the HRSG. Also, note the higher steam pressure drop in the superheater due to the lower steam pressure. The economizer flow includes the 2,500 pph saturated steam taken off the drum.

222

Waste Heat Boiler Deskbook

A note of caution on V, S and V times S values. Note that US values could be computed for each surface from its Q and ,1T data. For instance in the "design" case, for the superheater, US =234,000/311 = 7,524. These would naturally change depending upon the gas flow, analysis and temperature profile. Hence, these values should be interpreted with caution. After arriving at the US values, some engineers try to split up the U an S values and compare alternate designs based on S values alone. This can lead to very misleading conclusions and the author strongly recommends against it. particularly if extended surfaces are used. With finned tubes, the gas side heat transfer coefficient and fin efficiency are affected by variables such as fin density, height, thickness and fin or tube material. By using tubes with high fin density, say six, one could show more surface in the HRSG, but due to the lower U associated with it, it does not mean that the energy transferred is more compared to a design which has a lower fin density, say two to four, and hence, lower S. Lower fin density should be used whenever possible to increase U and minimize gas pressure drop and fin and tube wall temperatures. This is more important in surfaces with low tube side heat transfer coefficients such as superheaters. One could show that S can be 100 to 200% more by using six fins/in.,compared to two, but due to the higher U, the duty can be the same or even more. The author has performed studie:; on optimization of finned tubes and advises engineers against comparing and selecting HRSGs simply because the surface area, S, is more compared to another design which used lower fin density. Vnless the engineer is familiar with all aspects of heat transfer with extended surfaces and the impact of each variable on U, comparisons of S alone can be misleading and should be avoided. Limitations and Software. The approach discussed has a limitation. It cannot be used in HRSGs which have a radiant section. However, the author is of the view that 80 to 90% of HRSGs fall under the category discussed in Figure 4-6, and hence the methodology discussed can be applied to a wide variety of HRSGs used in the industry.

223

HRSG Design and Performance Simulation

Process steam Feed water

f

Gas

ECO

a. SH + EVAP + ECO [module 1)

b. SH + EVAP [module 2)

Steam

c. EVAP [module 3)

d. EVAP + ECO [module 4)

~*~I' e. SH [module 6)

f. ECO [module 5)

Figure 4-6. Various modules can be combined to represent multiple pressure and complex HRSG configurations. [Hydrocarbon Processing)

While the method of predicting performance using U values based on actual tube geometry, fin configuration, etc., gives accurate results, this methodology has been checked against several designs and operating results. For the purposes of engineering analysis, trend projections, evaluation of alternate designs and for studying the effect of different gas/steam parameters on performance, this approach is very effective and hence a powerful tool.

;;

;;:;;;;

1

224

Waste Heat Boiler Deskbook

Considering the complexity of the calculations and iterative nature of the procedure, particularly if multipressure HRSGs are involved, program COGEN has been developed by the author for HRSG design and performance evaluation. For more information on the software and its availability, contact the author at P.O. Box 673, Abilene, Texas 79604, USA.

INSERT A Performance calculation procedure

The procedure is discussed for a single pressure HRSG. Figure 46 shows the various configurations of HRSGs considered. The first case is quite involved. The methodology for this case will be discussed. The gas flow, gas inlet temperature and analysis, steam pressure and feed water temperature are assumed to be known. The design calculations, which are the basis of establishing an initial design, are assumed to be done and the results available, along with KIf K2 f K3 factors. 1. Assume the steam flow. A good estimate is obtained by using a ratio of the "performance" to "design" gas flows and temperature drop. 2. Solve the superheater performance. This is an iterative process. See insert B, see Equations (4-11 )to (4-15). If the transferred and assumed duty are not equal, repeat with another steam temperature or else continue. 3. Solve the evaporator performance. Obtain the duty and exit gas temperature using Equations (4-16 )to (4-19) . 4. Solve economizer performance using Equations (4 - 20) to (4-23). This is again an iterative procedure. Calculate the total transferred duty. 5. The steam flow is then corrected based on the total transferred duty and enthalpy rise, Equation (4-24). If this is close to the assumed steam flow in step 1, continue or else repeat steps 1 to 5. 6. If the final steam temperature is greater than that desired, the steam flow is corrected for the desired steam temperature.

HRSG Design and Performance Simulation

225

7. If the desired steam flow is zero (unfired mode) or less than the corrected flow, proceed to step II. 8 .. If the desired steam flow is larger than the corrected flow, calculate the fuel input required to raise the gas temperature to the required level to achieve the desired steam flow. This again involves several iterations, and for each firing temperature, all the steps from 1 to 8 have to be repeated until they match. 9. If the final steam temperature is higher than desired, calculate the interstage spray quantity based on a split superheater 10. Another round of fine tuning is done to check the temperature profiles and steam flow. 11. It can be easily seen that a lot of iterative calculations are involved. For each round, the gas and steam properties have to be computed based on the gas analysis and temperature. If there is steaming in the economizer, the economizer is split up into two stages, a small evaporator and an economizer and calculations are done to evaluate the extent of steaming. It is obvious that without a computer, the calculations can be overwhelming, particularly if there are several alternate performance conditions, and steam is generated at several pressure levels.

INSERTB Equations used in performance calculations Superheater performance. Assuming that the steam flow = Ws, from energy balance we have: Qla = Ws(hs2-h sl) = Wg(C p) (hI! )(TglTg2)

(4-11)

where ts2 = exit steam temperature and hS2, the enthalpy. Compute the exit gas temperature, T g2, from the above. The transferred duty is then: Q1t = (US)p t1T (4-12a) ~ T = log-mean temperature difference (4-12b) ~T = [Tgl-ts2)-(Tg2-tsl)]/ln[(Tgl-ts2)/(Tg2- tsI)] assuming counter flow configuration, which is widely used. (US)p is the product of 5 and U in performance mode and is obtained from the

226

Waste Heat Boiler Deskbook

(US) value in the design case by adjusting as follows for the gas properties and flow.

(4-13) K1

is obtained from Q1, t1T, Wg and Fg values in design case:

K1 = Q1j(t1T(Wi· 65 ) (Fg) )

(4-14)

Fg = (Cp0.33f
(4-15)

If for the assumed steam temperature Q1a and Q1t do not come close (say within 0.5%), another iteration is warranted. All of the above steps are repeated until Q1a and Q1t match.

Evaporator Performance. From energy balance, Q2

= Wg(C p) (hIf) (Tg2-Tg3) = (US)p t1T

(4-16)

where t1T = [(Tg2 -ts)-(Tg3 -ts)}/In [(Tg2-ts)/(Tg3-tS)) = (Tg2 - T g3)fIn [(Tg2_ts)j(Tg3-tS)]

(4-17)

From Equations (4-16) and (4-17) after simplification, we have: [(Tg2-tS)/(Tg3-tS)}

= e [(US)pf(Wg Cp hziJ

where: (US)p = Wi· 65 Fg K2 K2

(4-18) (4-19)

is computed as in Equation (4-14) from the design conditions.

F g is computed for the performance conditions. Tg3 is solved from

Equation (4-18) without iteration. Q2, the duty, can be obtained from Equation (4-16) . Economizer Performance. Assume tW2, the water exit temperature. Then, (4-20)

227

HRSG Design and Performance Simulation

Obtain Tg4 and then the LlT, assuming counter flow conditions (4-21)

Transferred duty: Q3t

=(US)p t1T

(4-22)

Where (4-23)

K3 is obtained as in Equation (4-14) from design conditions. If Q3a and Q3t are close, continue or else the iteration continues from Equations (4-20) and (4-23) with a different tW2 . The steam flow is then corrected as follows: (4-24)

If W5C is not close to the assumed flow, W5, the calculations are

repeated starting with the superheater. The gas pressure drop is corrected for performance conditions: ,1p= (t1P)d (Wg/Wgd)2 [(Tavg + 460)/(Tavgd + 460)1

(4-25)

From the above calculations we can draw some conclusions: 1. In a HRSG with an economizer, the pinch and approach points decrease with a decrease in inlet gas temperature and vice versa; for a gas turbine HRSG, this means that as the ambient temperature decreases (case 2), the pinch and approach points decrease. The reason is that the energy transferring ability of the economizer has not reduced (namely U x S), as it is a function of the surface area and overall heat transfer coefficient, which has increased due to the larger mass flow. With a reduced steam flow, due to the reduced duty of the evaporator, the economizer can now bring the water closer to

228

Waste Heat Boiler Deskbook

saturation temperature than before. Hence the caution that steaming is likely in the cold unfired ambient gas turbine exhaust conditions. It can also be seen from cases 3 and 4 that as the firing temperature increases, more steam is generated and the pinch and approach point increase. 2. The other point to keep in mind is that due to the lower water flow, the exit gas temperature from the economizer will be higher in case 2 versus case 1. So the HRSG efficiency is affected. It will be also be seen later that as the inlet gas temperature increases, more steam is generated and the stack temperature reduces. These trends are important and specification writers and consultants should be aware of them. 3. The superheater duty is not obvious in the fired mode as seen from case 4. It is higher compared to the situation where the steam is simply raised from saturation to 650°F due to the spraying of injection water for temperature control. Hence the suggestion that temperature profiles should be selected in the unfired mode and the performance checked in the fired mode.

SOFTWARE FOR HRSG SIMULATION - COGEN Predicting the off-design performance of a HRSG is a complex procedure involving several iterative loops even for a single pressure HRSG as seen from above. If more than one pressure level is involved or if the fuel input has to be arrived at automatically or if a common economizer feeds several pressure levels, the HRSG performance cannot be evaluated manually unless one is prepared to spend several hours on calculations. The author h;as developed a software, "COGEN", which is available from him. This may be used to arrive at a design and then predict its performance in a matter of minutes for complex HRSGs. Several alternates may be studied and the system optimized. These features will be explained through examples. Figure 4-6 shows the various configurations of HRSGs that can be combined to form a complex single or multi-pressure HRSG for simulation.

HRSG Design and Performance Simulation

229

With this software, consultants, plant engineers and cogeneration system analysts can predict the HRSG performance under different conditions of gas and steam parameters. Based on known plant operating data or assumed pinch and approach points, a "DESIGN" is simulated; gas/steam temperature prOfiles and steam flow are computed; then, for any other jnlet gas flow, temperature, steam pressure or feed water temperature, the revised gas/steam temperature profiles and steam flow are arrived at based on a quick converging iterative logic developed by the author; this is the "PERFORMANCE"; if the steam flow desired is more than that in the unfired mode, a gas or oil fired burner kicks in if there is sufficient oxygen in the gas stream and automatically computes the fuel input and firing temperature and revised temperature profiles. Different configurations as seen in Figure 4-6 can be handled by this program; by combining the modules, multi-pressure units can be evaluated in ONE run. In order to use this program one need not know how to size a HRSG or calculate the heat transfer coefficients; the tube size, pitch or fin parameters, etc., need not be known. In short, this is a powerful tool for consultants, plant engineers and cogeneration system designers, who would only like to simulate the HRSG performance under different conditions but do not want to get involved in detailed thermal or mechanical design aspects, which is best done by HRSG suppliers. Special Features of the Program. 1. The software arrives at a HRSG "DESIGN" based on plant

operating data or initially suggested pinch and approach points or and then evaluates its "PERFORMANCE" under different inlet gas conditions and steam parameters in unfired or fired conditions. Single or multi-pressure HRSGs can be evaluated in ONE run. There is no limitation on the number of modules. Up to six pressure levels can be easily analyzed. The HRSG configuration is selected by combining the various modules shown in Figure 4-6. Complex HRSG systems can thus be modelled. Each module can operate at its own pressure, feed water or steam conditions. Supplemental firing can be introduced ahead of the HRSG or between modules.

----

--------------------------------------------------------~-~-----

230

Waste Heat Boiler Deskbook

2. One need not actually design a HRSG in order to evaluate its off-design performance. Tube configuration, layout, surface area, boiler size etc. need not be known. Heat transfer coefficient U is not computed; instead the product of U and surface area S is computed and then corrected for different gas flows, analysis and temperature. 3. Unfired or supplementary fired HRSGs can be analyzed. Gaseous (up to 14 constituents) as well as oil fuels can be used in the burner system. The program automatically arrives at the firing temperature, burner fuel input if the required steam flow in any evaporator is more than that obtained in unfired mode. Source of oxygen can come from the exhaust gas or from atmospheric air (fresh air fired case); augmenting air can also be added if there is insufficient oxygen in the exhaust gas; this is important in HRSGs for Steam Injected Gas Turbines, which have low levels of oxygen and high water vapor content. 4. The effect of gas analysis is considered; (U x S) values and specific heat values are corrected for different gas analysis and temperature making the evaluation realistic. 5. If the final steam temperature is controlled and if for given gas inlet conditions, the steam temperature exceeds the desired value, an interstage desuperheater is simulated and spray water flow required to achieve the final temperature is computed along with the gas and steam temperature profiles. This can be done for any module in multi-pressure units. 6. The extent of "Steaming" in any economizer can be computed, if it does occur under different gas and steam conditions. The "DESIGN" can then be easily revised if desired. 7. The HRSG is modelled such that some saturated steam from the evaporator can be taken off for process heating or other applications if required and the balance superheated. Saturated steam from other boilers can also be introduced into the superheater thus increasing its flow. 8. The common economizer concept can be used to supply feed water to any number of modules; the program arrives at the water flow and exit water temperature through several rounds of complex iterations. 9. One can arrive at a design and fine tune it based on its offdesign performance within minutes.

HRSG Design and Performance Simulation

231

EXAMPLES TO ILLUSTRATE USEOFCOGEN Supplementary Firing and HRSG Efficiency Example 2: A gas turbine HRSG generates satutated steam at 200 psig using feed water at 240°F; blow down is 3%. The gas flow is 150,000 pph and inlet gas temperature is 900°F. Using a pinch and approach of 20°F design the HRSG and study the effect of fuel input on performance while generating 40,000 and 60,000 pph of steam. Fuel used is natural gas with 95% methane and 5% ethane. Gas analysis: % vol C02 =3, H20 =7, N2 =75, and 02 =15. Solution: Using COGEN, the design and performance calculations were done. Results are shown in Figure 4-7. The program automatically arrives at the firing temperature, fuel required and the exhaust gas analysis after firing within seconds. The data and results are shown in Example 1-5, Chapter 1 and the system efficiency is computed in the unfired and fired modes. It was shown that the system efficiency as estimated by ASME PTC 4.4, increases with additional fuel input.

IMPROVING EFFICIENCY OF HRSGs COGEN may be used to improve the efficiency of HRSG systems in two ways: a. Use add-on surfaces such as condensate heater, deaerator coil or heat exchanger to lower the stack gas temperature. Figure 4-8 shows a HRSG with several add-on coils to improve the energy recovery. With a condensate heater, the efficiency is improved through reduced demand of deaeration steam. b. Rearrange location or disposition of heating surfaces to arrive at a lower stack temperature with the steam production remaining unchanged in the fired mode or more steam generation in unfired mode.

~~, %':;~y

Ity\

Waste Heat Boiler Deskbook

232

RESULTS •••• DESIGN CASE

UNFIRED

amb temp-f= 60 rel hum-%= 0 heat loss-%= 1 gas temp to HRSG= 900 gas flow-pph= 150000 %·vol C02= .:j H20= 7 N2 = 75 02 = 15 S02= 0

gas temp in-out-F EVAP ECO

900 408

408 327

wat/stm in-out-F 368 240

388 368

duty press MM b/h psig 19.57 3.10

200 210

flow pph

pstm pinch apprch % F F

22779 100.0 23462 0.0

20

(a)

20

RESULTS ••• PERFORMANCE CASE ••• FIRED amb temp-f= 60

reI hum-%= 60 heat loss-%= 1 gas temp to HRSG= 900 gas flow-pph= 150000 % vol C02= 3 H20= 7 N2 = 75 02 = 15 S02= 0 gas temp in-out-F BURN 900 1289 EVAP 1289 419 419 315 ECO

wat/stm in-out-F 0 336 240

gas flow after HRSG 0.00 fuel

0 388 336

duty press MM b/h psig 17.30 35.80 4.07

0 200 210

flow pph

pstm pinch apprch % F F

808 0 40029 100.0 41230 0.0

(I)

32

52

150808 % vol c02= 3.92 h20 = 8.79 n2=74.30 02=12.98 s02=

GAS: vol %

1 methane= 95 2 ethane= 5 lhv-btu/cu ft= 949 lhv-btu/lb = 21422 aug air-pph = 0

RESULTS ••• PERFORMANCE CASE ••• FIRED amb temp-f= 60 reI hum-%= 60 heat loss-%= 1 gas temp to HRSG= 900 gas flow-pph= 150000 % vol C02= 3 H20= 7 N2 = 75 02 = 15 S02= 0

gas temp in-out-F

wat/stm in-out-F

duty press MM b/h psig

flow pph

pstm pinch apprch % F F (c)

BURN 900 1714 EVAP 1714 430 ECO 430 310

0 315 240

gas flow after HRSG = 0.00 fuel

0 388 315

37.53 55.16 4.74

0 200 210

1752 0 60128 100.0 61932 0.0

42

73

151752 % vol c02= 4.97 h20 =10.84 n2=73.50 02=10.67 s02=

GAS: vol %

1 methane= 95 2 ethane= 5 lhv-btu/cu ft= 949 lhv-btu/lb = 21422 aug air-pph = 0

Figure 4-7. Printout of results of example 2 from COGEN software,

233

HRSG Design and Performance Simulation

m ·ffi I

i i i

i! ~1I:l=-:::f:=~U._

!

-'-4

i i

i

i ........L i ~ ......

iI

i Figure 4-8. HRSG with condensate heater and glycol heater. [ABeD Industries]

Improving HRSG Efficiency Using Secondary Surfaces Figures 4-9, 4-10, 4-11, 4-12 show the concept of adding a condensate heater, a heat exchanger or a deaerator to improve the efficiency. The system consists of a HRSG generating superheated steam for producing electrical power via a steam turbine condenser system. The steam for deaeration is taken from an extraction point in the turbine as it is more economical to do so than take it off the HRSG. The reason for this will be given later. Condensate heater. This is basically an economizer used for preheating the mixture of condensate returns and make up before it enters the deaerator, Figure 4-10. A 20°F approach is used; that is if the deaeration temperature is 240°F, the design condensate heater exit temperature is 220°F. The following aspects of this option should be considered. 1. The log-mean temperature difference between the gas and water is higher than in the deaerator option described below; as a result the surface area requirements will be lower.

Figure 4-9. Basic arrangement: Option 1 has heat recovery only provided by HRSG [Power]

Figure 4-10. Condensate heater is added downstream of HRSG economizer in Option 2 to preheat condensate reducing deaeration steam. [Power]

Figure 4-11. In Option 3, heat exchanger is used to preheat makeup/condensate mixture before entering deaerator. [Power]

Figure 4-12. A low pressure evaporator can be used to generate low pressure deaeration steam, maximizing electrical output. [Power]

236

Waste Heat Boiler Deskbook

2. Due to the low water temperature at the inlet, the tube wall temperature will be lower and hence condensation of water vapor can occur resulting in corrosion .. It is suggested that the tube wall temperature be above the water vapor dew point of the exhaust gas. If there is steam injection in the gas turbine, the water vapor dew point is higher and hence suitable precautions have to be taken such as better choice of tube materials. If the exhaust gases contain sulfuric acid vapor or sulfur dioxide, then this option is not recommended. 3. The gas pressure drop will also be lower due to the lower surface area requirement.

Table 4-4. Results of Study of Deaeration Steam and Electric Output * Option Design reference Exhaust gas, 103 Ib/hr Gas-inlet temperature, F Stack-gas temperature, F Steam to turbine, 1031b/hr ** Steam to deaerator (from turbine), Ih/hr Feedwater temperature, F Mixture temperature (to deaerator),F Electric power, kW

(2) Cond htr 550 975 310 80

(3) Heat exch 550 975 323 80

(4) L-p evap 550 975 297 80

10,250 1730 240 240

3400 151

0 240

220 6830

200 6770

107 6890

(1 ) Base case 550 975 374 80

107 6528

* Assumptions: Natural-gas fired gas turbine; gas analysis, % vol: COz =3.5, HzO = 10, Nz = 74, Oz = 13.5; pinch point = 20°F; approach point = 20°F, blowdown = 2%; deaerator pressure = 10 psig; makeup temperature = 60°F; condenser pressure = 2.5 in. Hg abs 0.22 psia); heat loss in HRSG = 1%. **Conditions: 620 psig, 650°F.

HRSG Design and Performance Simulation

237

Heat Exchanger. Figure 4-11 shows a heat exchanger used to improve the efficiency of a HRSG system. A water to water exchanger preheats the make up using the feed pump discharge before it enters the economizer. The make up water temperature increases thereby reducing the de aerating steam requirements. Since the feed water to the economizer is lower, the stack gas temperature reduces. 1. If the gas stream contains sulfuric acid, then this scheme is unsuitable due to the low tube wall temperature at the economizer inlet. Better tube materials may have to be used if water vapor dew point is high. 2. This scheme is attractive if the make up water temperature is low say 40 to 60°F. Then, the feed water can be cooled further in the exchanger. The size of the exchanger will also be smaller due to the larger log-mean temperature difference. Also, if the make up temperature is very high, on the order of 150 to 180°F, the exchanger may not be feasible due to the resulting low log-mean temperature and possible temperature cross conditions in the exchanger. 3. Due to the lower feed water temperature, the size of the economizer will increase. This cost has to be evaluated along with the cost of the exchanger, piping and associated valves. Deaerator coil. Figure 4-12 shows the deaerator coil used to recover additional energy from the gas stream. The bundle generates low pressure steam for de aeration and is essentially another evaporator. This may be an integral deaerator or simply an evaporator generating low pressure steam, which is taken to a deaerator. 1. The surface area required will be more due to the low logmean temperature difference compared to the condensate heater. 2. This scheme is suitable even if the gas stream contains sulfuric acid as the evaporator pressure can be raised to a value close to or above the acid dew point. The tubes will be at least 5 to 10°F above the saturation temperature. 3. This system is more expensive than the others due to the size and the use of drums and associated trim and controls. 4. The gas pressure drop will also be higher than the other options, due to the size.

238

Waste Heat Boiler Deskbook

USE OF COGEN TO ANAL YZE THE OPTIONS Example 3: A HRSG generates about 80,000 pph of superheated steam at 620 psig and 650°F from 550,000 pph of gas turbine gases at 975°F; the steam is expan9.ed in a steam turbine to generate power, with deaeration steam taken from an extraction point in the steam turbine. Study the options assuming that the gas stream contains no sulfur dioxide. Solution: Figure 4-13 shows the results from COGEN in the design mode for all of the options. A 20°F pinch and approach were used. The steam required for deaeration was estimated for each case through heat balance as described below. The power output from the steam turbine was obtained through computation of turbine steam rates. A Basic program for this is given later. Table 4-4 shows the summary of results. Since the deaerator option generates the maximum steam output to match de aeration requirements, the electrical power is the highest. One may also predict the HRSG performance at any other condition using COGEN and see how the system behaves.

DEAERATION STEAM CALCULATIONS Estimating steam quantity for deaeration is an important aspect of plant mass and energy balance calculations. Steam for de aeration should preferably be taken from an extraction point in the steam turbine if available. This results in a better system efficiency compared to the case where the steam is taken from the HRSG exit. Figure 4-14 shows the two schemes.

COMPUTING DEAERATION STEAM AND ELECTRIC OUTPUT To illustrate the procedure, consider two sources for deaeration steam for the base case (Option 1), Example 3. Assume, first, that the

239

HRSG Design and Performance Simulation

RESULTS •••• DESIGN CASE

UNFIRED

8mb temp-f= 60 reI hum-If;= 0 heat 10ss-%= 1 gas temp to HRSG= 975 gas.f10w-pph= 550000 'I; vol C02= 3.5 H20= 10 NZ = 73 02 = 13.5 S02= 0 gas temp

in-out-F

wat/stm in-out-F

duty press MM b/h psig

flow pph

pstm pinoh approh 'I; F F (a)

SH EVA!> ECO

975 914 515

914 515 374

495 475 240

650 495 475

RESULTS •••• DESIGN CASE

9.31 59.57 20.36

620 635 645

80085 .100.0 80085 100.0 81687 0.0

20

20

UNFIRED

amb temp-f= 60 reI hum-%= 0 heat 10ss-%= 1 gas temp to HRSG= 975 gas flow-pph= 550000 % vol C02= 3.5 H20= 10 N2 = 73 02 = 13.5 S02= 0 gas temp

in-out-F

wat/stm in-out-F

duty press MM b/h psig

flow pph

pstm pinoh approh % F F (b)

SH EVAP ECO

975 914 515

914 515 324

495 475 151

650 495 475

RESULTS •••• DESIGN CASE

9.31 59.57 27.65

620 635 645

80085 100.0 80085 100.0 81687 0.0

20

20

UNFIRED

amb temp-f= 60 reI hum-%= 0 heat 1099-%= 1 gas temp to HRSG= 975 gas flow-pph= 550000 % vol C02= 3.5 H20= 10 N2 = 73 02 = 13.5 S02= 0 gas temp in-out-F

wat/stm in-out-F

duty press MM b/h psig

flow pph

pstm pinch apprch F F

"

SH EVAP ECO

975 914 515

914 515 374

495 475 240

650 495 475

9.31 59.57 20.36

620 635 645

80085 100.0 80085 100.0 81687 0.0

ECO

374

310

107

220

9.02

650

79870

RESULTS •••• DESIGN CASE

(c)

20

20

0.0

UNFIRED

amb temp-f= 60

reI hum-%= 0 heat loss-%= 1 gas temp to HRSG= 975. gas flow-pph= 550000 % vol C02= 3.5 H20= 10 N2 73 02 = 13.5 S02= 0 gas temp in-out-F

=

wat/stm in-out-F

duty press MM b/h psig

flow ppb

pstm pinoh approh F F

'"

(d)

SH EVAP ECO

975 914 515

914 515 374

495 475 240

650 495 475

9.31 59.57 20.36

620 635 645

80085 100.0 80085 100.0 81687 0.0

20

EVAP

374

297

107

240

10.93

10

10050 100.0

57 132

20

Figure 4-13. a, b, c, d: Results from COGEN program for various options.

240

Waste Heat Boiler Deskbook

Makeup. Y 60F (17001b:hr)

C"ndensate: 'OOF (OO,OOO·X /b/hr)

Mixture: '07F

Feedwater: 240F• • 01,700 /b/hr

Condensate: 10SF

Figure 4-14. Two alternatives for deaeration steam take off location; steam from HRSG (top) and from steam turbine (bottom), [Power]

HRSG Design and Performance Simulation

241

steam is taken from the HRSG exit (Figure 4-14 left) rather than from the turbine. Makeup temperature is 60°F and condensate enters the mixing tank at 108°F (2.5-in. Hg condenser pressure). Neglecting flashed and vented steam, a mass balance around the mixing tank gives: 81,700 - X = Y + (80,000 - X), where X is the deaeration-steam flow and Y the makeup flow. Very simply, Y = 1700 lb/hr. (Flashed steam from blowdown and vent steam are neglected in this analysis.) An energy balance is then performed around the deaerator, giving: (81,700) (208) = (1700) (28) +(80,000 - X) (76) + 1319X, where 1319 Btu/lb is the enthalpy of steam at 620 psig and 650°F, 28 Btu/lb the make-up enthalpy, 76 Btu/lb the enthalpy of condensate returns, and 208 Btu/lb that of feedwater after deaeration. Solving this equation, X =87411b/hr steam: using 87851b/hr allows for losses. To obtain the electric output, we must compute the actual steam rate (ASR) for the expansion from 620 psig and 650"F to 1.22 psia (2.5 in. Hg) A program developed for this purpose gives ASR = 11.14 Ib/kWh. Allowing for 4% mechanical losses, P = 0.96 (80,000 - 8785) /11.14 = 6137 kW. Next, consider deaeration steam taken from the turbine extraction point (Figure 4-14 right) at about 30 psia. For an expansion efficiency of 70%, this corresponds to an enthalpy of 1140.6 Btu/lb (calculated from the program referenced). An energy balance in this case gives: (81,700) (208) = 1140.6 X + (80,000 - X) (76) + (1700) (28). Solving for X gives 10,206 lb/hr for deaeration-steam flow. To include losses, use 10,250 lb/hr. The program then gives 19 Ib/kWh for the ASR from 620 psig to 30 psia, and 11.14 Ib/kWh for the entire expansion. For the net electric output, P =0.96 [(10,250/19) + (80,000 10,250) /11.14] = 6528 kW. Thus, greater electric output is obtained by taking steam from the steam turbine, and so this is the preferred procedure for all other options. Results of similar calculations for those options are summarized in the Table 4-4.

STEAM TURBINE CALCULATIONS Engineers involved in cogeneration projects and power plant studies often need to calculate the steam properties during expansion in a steam turbine to evaluate the theoretical and actual steam rates

I Input P" rio

p" E

I Yes

Figure 4-15a (left): Expansion of steam in a turbine. Figure 4-15b (right): Logic diagram to obtain expansion results [Hydrocarbon Processing]

Sat steam properties

HRSG Design and Performance Simulation

243

and hence, the electrical power output. With the help of this program written in BASIC, one can quickly evaluate all the pertinent data. Correlations used for steam property evaluation are also presented. Theory: Figure 4-15a & b shows the typical expansion process in a steam turbine.

TSR

= 3,413/(h1 -h2s)

(4-26) (4-27) (4-28)

Superheated steam enthalpy is computed from the equations given P, T as shown in Appendix E. For wet steam enthalpy, use: (4-29) h = xhv + (1 - x) hf where the saturated vapor and liquid properties are obtained from Appendix E given the pressure. The dryness fraction, x, is computed for each state from either the enthalpy relation in (4) or from the entropy relation:

s = XSv + (1-x)sf

(4-30)

The logic used in the program is shown in Figure 4-15b. Example 4: Two examples illustrate the use of the program. Example 4: Superheated steam at 650 psia and 750°F is expanded in a steam turbine to 150 psia with an expansion efficiency of 80%. Calculate the steam properties at inlet and exit as also the actual and theoretical steam rates.

Solution: Key in the program in Figure 4-16a. In the RUN mode the screen asks for the inlet pressure and temperature. If saturated, key 0 for

244

Waste Heat Boiler Deskbook

10 REM,PROGRAM COMPUTES TURBINE STEAM RATES 15 DIM A(8) ,B(8) ,C(8) ,0(8) ,E(8) ,F(8) ,G(8) ,HW(8,28), VW(8,28) 20 FOR 1=1 TO 8,READ A(I) ,B(I) ,C(I) ,0(1) ,E(I) ,F(I) ,G(I) ,NEXT I 25 DATA -.17724,3.83836,11.48345,31.1311,8.762969E-5,-2.78794E-8,86.594,-5.28012 6E-7, 2. 99461E-5, 1. 521874E-4, 6. 62512E-5, 8. 408856E-I0, 1. 86401E-14, .01596, -. 48799,3 04.717614,9.8299035 30 DATA -16.455274,9.474745E-4,-1.363366E-6,19.53953,2.662E-3,457.5802,-.176959, .826862,-4.601876E-7,6.3181E-l1,-2.3928,-.15115567,3.671404,11.622558,30.832667, 8. 74117E-5 35 DATA -2.62306E-8,54.55,-.14129,2.258225,3.4014802,14.438078,4.222624E-5,-1.56 9916E-8,1100.5,-1.67772E-4,4.2726S8E-3,.0104S04S,.95SO1509,9.101291E-S,-2.7592E-· 11, • l1S01 40 DATA -1.476933E-4,1.2617946E-3,3.44201E-3,-.OS49412S,6.S913SE-S,-2.4941E-11,1 .97364 45 PRINT" STEAM PROPERTIES AFTER EXPANSION •• BY V.GANAPATHY "'PRINT"" 50 INPUT" INLET PRESS-PSIA,TEMP-F(IF SATURATED INPUT 0 FOR TEMP)=";P1,Tl,PRINT" 55 INPUT" EXIT PRESS-PSIA,EXPN EFF-'=";P2,EF 60 IF T1=0 THEN GOTO 100 70 P=Pl,T=Tl,GOSUB 300 80 H1=Z,Sl=SV,GOTO 120 100 P=P1,GOSUB 400 110 Hl=HV,Sl=SV,T1=TSAT 120 P=P2,GOSUB 400 130 IF Sl>SV THEN GOTO 200 140 X=(Sl-SL)/(SV-SL),H2S=HV*X+(1-X)*HLIQ 150 H2=Hl-.0l*EF*(Hl-H2S) 160 IF H2>HV THEN GOTO 200 170 XF=(H2-HLIQ)/(HV-HLIQ),DELH=H1-H2,T2=TSAT,GOTO 290 200 T2=TSAT,RA=1000-TSAT 210 P=P2,T=T2,GOSUB 300 220 S2C=SV,IF ABS«Sl-S2C)/Sl)<.002 THEN 240 230 RA=.5*RA,T2=T2+SGN(sl-S2C)*RA,GOTO 210 240 XF=l,T=T2,GOSUB 300,H2S=Z 250 H2=Hl-.01*EF*(H1-H2S) 260 T2=TSAT,RA=1000-TSAT 270 P=P2,T=T2,GOSUB 300 280 H2C=Z,IF ABS«H2-H2C)/H2)<.002 THEN 290 285 RA=.5*RA,T2=T2+SGN(H2-H2C)*RA,GOTO 270 290 ASR=3413/(HI-H2),TSR=3413/(HI-H2S):PRINT" "'PRINT" INLET PRESS-PSIA=";P1;"TE MP -F=";Tl;" ENTH =";Hl;" ENTRPY =";Sl:PRINT" W 294 PRINT" EXIT PRESS =";P2;" TEMP =";T2;" ENTH =";H2;"SAT TEMP=";TSAT:PRINT" " 295 PRINT" ASR =";ASR;" TSR =";TSR;" EXPN EFF-% =";EF;" EXIT QLTY =";XF:PRINT" "

297 END 300 T=273.1+(T-32)/1.8,P=P/14.696 310 K4=SOS701/T/T,L1=10'K4*(-2641.62/T),Ml=1.89+L1,Nl=Ml*P*P/T/T 320 01=2+(3724201/T/T),Q1=Ol*Ll,R1=1.S9+Ql,U1=(.2182S*T-1269701/T),V4=2*U1*Rl-(M 1/T)*1269701 330 Wl=82.54-1624601/T:Y1=2*W1*Rl-(Ml/T)*1624601 340 Fl=«(U1*Ml*N1+W1)*N1/P+1)*M1+4.5504*T/P)*.01601S5 350 Z=775.6+.63296*T+l.62467E-04*T*T+20.5697*LOG(T) 360 Z=Z+.043557*(Rl*P+.5*N1*(Y1+M1*(W1+V4*N1») 370 SV=«(U1*Ml-2*V4)*.5*Ml*N1-Y1)*.5*N1+(Ml-Rl)*P)/T*(-.0241983)-.355579-11.427 6/T 3S0 SV=SV+1.8052E-04*T-.11022*LOG(P)+.35164*LOG(T) 390 P=14.696*P,T=(T-273.15)*1.8+32 395 RETURN 400 FOR 1=1 TO S:RA(I)=A(1)*P+B(1)/P+C(I)*P'.5+D(I)*LOG(P)+E(I)*P*P+F(I)*P'3+G(I ) ,NEXT I 410 TSAT=AA(1):HLIQ=AA(5):HV=AA(6),SL=AA(7):SV=AA(8) 420 RETURN

Figure 4-16a. Listing of program for expansion of steam [Hydrocarbon Processing]

HRSG Design and Performance Simulation

245

the temperature and the program computes the saturated steam temperature. Then, the exit pressure and expansion efficiency are input as shown in Figure 4-16a. The results are printed out. It can be seen that the final steam is superheated. Actual and theoretical steam rates are shown. To compute the electrical power 01.l.tput with, say, a steam flow of 100,000 pph, simply divide the steam flow by the actual steam rate. Here it is 100,000/26.9 = 3,717 kilowatts. Example 5:

Saturated steam at 1,000 psia is expanded to I psia in a turbine. Determine the data for isentropic expansion. Solution: Input as before using 0 for steam temperature and 100 for efficiency of expansion. Results are shown in Figure 4-16b. Actual and theoretical steam rates are the same.

USE OF COGEN TO OPTIMIZE HRSG TEMPERATURE PROFILES Example 6:

This example shows that by rearranging the heating surfaces one may lower the stack temperature while generating the same steam output. Table 4-5 shows the data for a gas turbine HRSG generating steam at 625 and 205 psig. There are two options, shown in Figure 4-17 and 4-18. In Figure 4-17, the HP stage is followed by the LP stage. In Figure 4-18, the HP and LP stages are followed by; a common economizer, which handles the water for both HP and LP, thus increasing the capacity of the heat sink and thus lowering the stack gas temperature. Hence, we need a lower inlet gas temperature in Figure 4-18 in order to generate the same quantity of HP and LP steam or we can generate more steam (HP or LP) with the same gas inlet conditions.

~~~~~~~~~~~-

~

-~~~~~~-~~-

Waste Heat Boiler Deskbook

246

Figure 4-16b Data and results for examples

Example 4: STEAM PROPERTIES AFTER EXPANSION - BY V. GANAPATHY INLET PRESS-PSIA, TEMP-F (IF SATURATED INPUT 0 FOR TEMP) = ? 650.750 EXIT PRESS-PSIA, EXPN EFF - % = ? 150,80 INLET PRESS - PSIA = 650 TEMP - F = 750 ENTH = 1377.067 ENTRPY = 1.601213 EXIT PRESS = 150 TEMP = 458.7688 ENTH = 1250.208 SAT TEMP = 358.5408 ASR = 26.90375 TSR = 21.523 EXPN EFF -% = 80 EXIT QLTY = 1

Example 5: STEAM PROPERTIES AFTER EXPANSION BY V. GANAPATHY INLET PRESS-PSIA, TEMP-F (IF SATURATED INPUT 0 FOR TEMP) = ? 1000.0 EXIT PRESS-PSIA, EXPN EFF -% = ? 1,100 INLET PRESS-PSIA = 1000 TEMP - F = 547.2928 ENTH = 1193.038 ENTRPH = 1.392013 EXIT PRESS = 1 TEMP = 101.7387 ENTH = 776.87 SAT TEMP = 101.7387 ASR = 8.201008 TSR = 8.201008, EXPN EFF -% = 100 EXIT QLTY = .682389

Solution:

The results are shown in Figure 4-19, output from COGEN. Since the burner duty is small and the firing temperature less than 900°F, only the design modes are analyzed to show how the rearrangement of surfaces is helpful. The gas analysis was assumed to be close to the inlet gas analysis. The saving in fuel input is about 12 MMBtu/h for the improved option; assuming that the HRSG operates in this mode for 7000 hours/year, the annual savings is $210,000 at $2.5/million Btu.

247

HRSG Design and Performance Simulation

"-

~

w

H-P APPROACH 15'

Q:

....=><:

I

Q:

W

a. J;; W ....

36S'F

I

300 200

L -P ECONOMIZER

230'F

L -P EVAPORATOR

100

900 800 700 ~

w 600

Q:

:.J ....
49S'F

500

Q:

W

a.

J;; 1,,'

....

400

H-P SUPERHEATER

37S'F

390'F

300 200 100

l

H-P EVAPORATOR

COMMON ECDNOM[ZER . L-P EVAPORATOR

, 230"F

Figure 4-17 (top): Temperature profile for HRSG with two separate economizers. [Power]

Figure 4-18 (bottom): Temperature profile for HRSG with combined economizer. [Power]

Alternatively we could have generated more HP or LP steam with the configuration in Figure 4-18, if the gas inlet temperature and flow were the same as in Figure 4-17.

Or

248

Waste Heat Boiler Deskbook

RESULTS •••• DESIGN CASE

UNFIRED

amb temp-f= 60 re1 hum-%= 0 heat 10ss-%= 1 gas temp to HRSG= 900 g,!s flo,w-pl'~~-1-1.~~~og. % vol C02= 2.6 H20= 21 N2 ~ 63 02 = 13.4 gas temp in-out-F

wat/stm in-out-F

duty press MM b/h psig

flow pph

pstm pinch apprch 'Ii F F

SH EVAP ECO

900 880 721

880 721 661

495 480 230

600 495 480

6.47 57.12 21.12

625 635 645

z.u.u,..100.0 77372 100.0 79693 0.0

EVAP ECO

661 410

410 363

375 230

390 375

86.32 15.60

205 215

101337 100.0 104377 0.0

RESULTS •••• DESIGN CASE

226

15

20

15

UNFIRED

amb temp-f= 60 re1 hum-%= 0 heat 10ss-%= 1 gas temp to HRSG= 864 gas f10w-pph= 1215000 % vol C02= 2.6 H20= 21 N2 = 63 02 = 13.4 gas temp in-out-F

wat/stm in-out-F

502= 0

duty press MM b/h pdg

flow pph

S02= 0

pstm pinch apprch % F F

SH EVAP

864 846

846 662

495 375

600 495

6.34 64.99

625 635

75798 100.0 75798 100.0

EVAP

662

410

375

390

86.76

205

101986 100.0

ECO

410

328

230

375

27.32

650

183118

167 120 20

14

0.0

Figure 4-19. Results from COGEN for example 6 for the 2 options shown in Fig 4-17 and 4-18.

Table 4-5. Data for HRSG design Exhaust gas flow, pph = 1215,000 Exhaust temperature, F = 770 HP steam flow, pph = 75-77,000 HP steam pressure;, psig = 635 HP steam temperature, F = 600 LP steam flow, pph = 95-100,000 LP steam pressure, psig = 205 LP steam temperature, F = 390 (sat.) Feed water temperature, F = 230 Exhaust gas analysis = % volume CO2 = 2.6, H2 0 = 21, N2 = 63, O2 = 13.6

249

HRSG Design and Performance Simulation

RESULTS • • . . DESIGN CASE amb g~s

60 re1 hum-%- 0 heat loss-%- 2 gas temp to HRSG- 1400 f10w-pph. l~E~_ 1 vol CO2. 7 H2o.. 12 N2. 75 02. 6 S02. 0

t~mp-f.

gas temp in-out-F EVAP

UNFIRED

duty

wat/stm in-out-F

press psig

14M b/h

34.84 7.17

1400 578 408 448 578 398 250 408

400 410

pinch

flow pph

pstm

42398 -43of0

100.0 0.0

1

apprch

F

F

130

40

RESULTS ••. PERFORMANCE CASE ••• UNFIRED amb temp-f. 60 re1 hum-%- 60 heat loss-%- 2 gas temp to HRSG. 1300 gas flow-pph. 165000 1 vol CO2. 7 H2o.. 12 N2 _ 75 02 _ 6 S02- 0 gas temp in-out-F EVAP ECO

1300 544

544 384

duty

wat/stm in-out-F

HH b/h

396 414 240 396

34.76 6.97

press psig

flow pph

pstm pinch 1 F

275 285

41795 43048

100.0 0.0

130

apprch F 18

1400

~

~98

~2S0 • •• •

Design

1300 ~conomlzer

5 « - - - 384 414

~ uu

Evaporation

...............

240

Performance

Economizer

Module 3

Figure 4-20. Results from COG EN and temperataure profiles for Example 7.

This example illustrates how COG EN may be used to simulate the design of an operating HRSG from field data and then predict its performance at a different condition.

Waste Heat Boiler Deskbook

250

USING FIELD DATA TO PREDICT HRSG PERFORMANCE Example 7: Table 4-6. Data For Design and Performance

This is an example where the design is simulated from plant operating data. Table below shows the operating data for a HRSG for incinerator exhaust gases. By playing with pinch and approach points, a design is arrived at matching the operating data. Then its performance is predicted for case 2 conditions. The HRSG consists of an evaporator and economizer, module 4 and generates saturated steam. 1 2 1. Case 150,000 165,000 2. Gas flow, pph 1400 1300 3. Gas inlet temp, F 275 400 4. Steam pres, psig 240 250 5. Feed water temp, F ? 42,400 6. Steam flow, pph 7. Exit gas temp, F ? 400 3 3 8. Blow down %

[Gas vol %-C02 = 7, H20 = 12, N2 = 75, 02 = 6,Heat loss % = 2J Results from COGEN are shown in Figure 4-20.

The results of COGEN may also be used to check if there is fouling of heat transfer surfaces. The results from COGEN are based on the fact that there is no appreciable fouling. For example in the above case 2 if the stack gas temperature were 420°F versus 384°F predicted, then one can assume that there is some fouling in the HRSG. If the gas temperature at the economizer inlet were measured and found to be much higher than 544°F then the fouling could be in the evaporator.

MULTI-PRESSURE HRSG DESIGN AND PERFORMANCE As a final example, the design and performance of a muItipressure HRSG with supplementary firing will be simulated. This example shows how one can simulate the behavior of complex and

HRSG Design and Performance Simulation

251

fired HRSG, which are used in large cogeneration systems. Data is shown in Example 8 and results in Figures 4-21a, 4-21b and 4-21c.

G

I EVAPII ECO I Module 1

G

IEVAPI

Module 2

I EVAPI

Module 3

I EVAPI I ECO I Module 4 I ECO I Module 5

Module 2 Module 2 Module 5 Module 3 HP SH + EVAP IP SH + EVAP COM. ECO LP EVAP

Module 6 Sui table HRSG Configuratuons can Be Ardved AT By Combining Modules. Problem 1 USes Modules 2 •. 2,. 5 and 3 Burner Can Be Placed Ahead Of Or In Between Modules. More than 1 Burner Can be Sim ulated

Figure 4-21a. HRSG configuration for problem. [Sci-Tech publishers]

Example 8: This is a multipressure HRSG. A common economizer provides feed water to both HP and IP evaporators. The configuration is obtained by combining four modules, namely 2,2,5, and 3. Both HP and IP steam are superheated, while LP steam is saturated. The design case assumes that 375°F water is obtained from the common eco. Since its flow is unknown to begin with, inputting of zero for the water flow will enable it to be computed automatically. In the performance case, the desired HP steam is obtained by firing the duct burner, which automatically arrives at the firing temperature and fuel input. Note that augmenting air is also used for the burner. A Desuperheater is also simulated. The program also automatically arrives at the flow through the common economizer and its water exit temperature, which is the inlet feed water temperature to HP and IP steam systems. All these complex iterations are completed within 30 to 40 seconds with a 386 processor. The design case takes less than 5 seconds.

Waste Heat Boiler Deskbook

252

UNFIRED

RESULTS .•• DESIGN CASE

heat loss-~ 1 gas temp to HRSG - 900 rel hum-~ 0 amb temp-f-60 gas flow-pph- 1215000 't. vol C02= 2.6 H20 - 22 N2 - 63 02 - 12.4 S02=0 press psig

wat/stm duty in-out-F MM b/h

gas temp in-out-F

flow pph

pstm 't.

pinch apprch F F

SH 900 877 EVAP 877 642

492 600 375 492

8.38 83.70

600 615

97865 97865

100.0 100.0

150

116

HP

642 625 SH EVAP 625 414

394 500 375 394

5.71 73.08

205 215

85876 85876

100.0 100.0

20

18

IP

ECO

414 336

240 375

26.11

·650

187415

0.0

EVAP 336 255

190 240

27.08

10

27015

100.0

COM.ECO 15

.49

LP

RESULTS ... PERFORMANCE CASE ••• FIRED amb temp-f- 60 rel hum-'t. - 60 gas f1ow-pph- 1215000 't. vol C02 wat/stm duty in-out-F MM b/h

gas temp in-out-F BURN SH DESH EVAP

=

900 1003 1003 971 991 991 971 678

heat 10ss-'t. = 1 gas temp to HRSG • 900 2.6 H20. 22 N2 - 63 02 - 12.4 S02 =0 press psig

pstm 't.

pinch apprch F F

0 493 560 366

0 0 45.77 600 11.51 600 0.00 611 552 493 107.15 623

2135 0 124936 100.0 732 0 124936 100.0

185

127

98342 100.0 98342 100.0

23

28

SH 678 EVAP 658

658 417

395 366

508 395

6.96 205 84.72 218

ECO

417

332

240

366

29.41

650

EVAP 332

255

190

240

26.51

10

gas flow after HRSG = 1242294 't. vol co2 s02=0.00 fuel -GAS:

flow pph

226575

=

0.0

26413 100.0 2.83 h20 - 22.10 n2

97

2.

ethane

2

3.

propane

15 =

63.11

lhv-btu/cu ft = 942 1hv-btu/1b = 21438 aug air-pph = 25000 NOTE:

IP COM.ECO

analysis -'t. volume

1. methane

HP

TEMPERATURE PROFILES ARE NOT GENERATED BY SOFTWARE. fuel input is on LHV basis

Figure 4-21b. Printout of results from COGEN. [Sci-Tech publishers]

50

LP

02=11.94

HRSG Design and Performance Simulation

253

900

394 375

240 190

HP SH I HP EVAP I

IP SH I IP EVAP

J

COM. ECO

J

lP EVAP I

240 Case 2

190

Figure 4-21c. HRSG temperature profiles. [Sci-Tech publishers]

l. Case no. 2. Remarks 3. Heat source 4. Gas flow, pph 5. Gas temp to HRSG, F 6. HP steam pressure, psig 7. IP steam press, psig 8. LP steam press, psig

1 design gas turbine 1215,000 900 600 205 10

2 performance gas turbine 1215,000 900 600 205 10

----------

- - - - - - - - -

Waste Heat Boiler Deskbook

254

9. 10. 11. 12. 13. 14. 15. 16. 17. 18. 19. 20. 21. 22.

Feed water temp, F Feed water to HP, F Feed water to IP, F HP pinch point, F IP pinch point, F LP pinch point, F HP steam flow, pph HP steam temp, F IP steam flow, pph IP steam temp, F HP SH press drop, psi IP SH press drop, psi Feed water temp to LP, F LP steam flow, pph

240 375 375 150 20 15

240

?

125,000 600

600 ?

? ? ? ? ?

? ? ? ?

500 15 10 190

190

?

?

[fuel used in case 2:N. gas:% vol Cl = 97, C2 = 2, C3 = 1, augmenting air flow = 25,000 pph; exhaust gas % vol C02 = 2.6, H20 = 22, N2 = 63, 02 = 12.4; heat loss = 1% and blow down = 2% each pressure level]

REFERENCES 1. V. Ganapathy, "Simplify heat recovery steam generator evaluation", Hydrocarbon processing, March 1990, p 77 2. V. Ganapathy, "Analyze options for deaeration steam", Power, September 1989, p 35 3. V. Ganapathy, "HRSG temperature profiles guide energy recovery" Power, September 1988 4. V. Ganapathy, "Win more energy from hot gases", Chemical Engineering, March 1990, p 102 5. V. Ganapathy, "Program computes turbine steam rates and properties", Hydrocarbon Processing, Nov 1988, p 105 6. V. Ganapathy, "COGEN", software published by Sci Tech, New Jersey, 1990

.~

//.4

i~If'

:fJ~

"~~

----

HRSG Design and Performance Simulation

255

NOMENCLATURE bd -:-Blow down fraction; if blow down =2%, then bd = 0.02. Cp-Gas specific heat, Btu/lboF. Fg -A factor accounting for gas properties, defined in Equation 4-5. hI! -Heat loss factor; if heat loss = 2%, then hI! = 0.98. hS2, hSl -Enthalpy of superheated steam and inlet steam, Btu/lb. hWl, hW2 -Enthalpy of water at eco inlet and exit, Btu/lb. k -Gas thermal conductivity, Btu/ft hOF. KI, K2, K3 -Factors obtained from design conditions, Equation 4-14. Ql, Q2, Q3,-Energy absorbed in superheater, evaporator and economizer, Btu/h; subscript a = assumed and t = transferred. Tgl, Tg2, Tg3, Tg4 -Gas temperature distribution, of. S -Surface area, sq ft. .1T -Log-mean temperature difference, of. Tavg, Ta-ogd -Average gas temperature in HRSG in performance and design modes. tWl, tW2 -Water temperature at inlet and exit of economizer, °P. tSl, tS2 -Saturated and superheated steam temperature, °P. U --Overall heat transfer coefficient, Btu/sq ft hOP. (US)p -Product of U and S in performance mode. W g, Wgd -Gas flow in performance and design modes, pph. W s, Wsd -Steam flow in performance and design modes. (.1P) d,p -Gas pressure drop in design and performance, in. We. /l-Gas viscosity, lb/ft h.

~

Chapter 5

Specifying Waste Heat Boilers Waste Heat Boilers are being specified and bought for every sort of application-hydrogen plants, gas turbine exhaust sulfuric acid plants, effluents from incineration of solid, liquid and gaseous waster, fluidised bed cat cracker exhaust etc., to name a few. Having designed several types of fire tube and water tube boilers for the above applications, the author feels that a large number of boilers are being specified inadequately or even incorrectly in some cases by consultants or the specification writers. This chapter will highlight some of the areas which should be emphasized or elaborated so as to result in a boiler that will be functional and cost effective.

1. APPLICATION OR SYSTEM DESCRIPTION The application or the system of which the HRSG is a part should be described completely. Often I come across specifications that simply say waste heat boiler for flue gas heat recovery without mentioning the process or the source of the gas. Depending upon the process the gas stream may have peculiar characteristics which can be inferred if the application is stated. For instance if the flue gas is from a municipal solid waste incinerator, we know that it can be dirty and hence precautions should be taken to have cleaning lanes, soot blowers and avoid extended surfaces. If the flue gas were from a fume incinerator then it is much cleaner and hence extended surfaces may be used to obtain a compact design. If the effluent is reformed gas from a hydrogen plant, one can ensure that provision is made for exit gas temperature control and also design the boiler to handle the high gas pressure. Extreme 257

258

Waste Heat Boiler Deskbook

operating conditions or upset conditions should also be described so that the design can accommodate these conditions. Mention should also be made as to how and where the steam that is generated is being utilized so that steam purity could be given consideration. If steam is used in a steam turbine drive, depending on the pressure, the purity could be demanding. Proper drum internals, Figure 5-1, should be used to give the desired purity. In the Cheng cycle discussed in Chapter 1, where the superheated steam is injected into the gas turbine, a combination of internal and external steam separation devices were used to achieve a steam purity of less than 50 parts per billion solids. In high pressure boilers, a combination of cyclones, and chevrons may be used to obtain a steam purity of less than 100 ppb solids. If steam is used in a steam turbine or injected into a gas turbine for NO x purposes or for increasing its power output, the specification writer should obtain information on steam purity requirements from the suppliers of these equipment so that the boiler drum internals can be properly designed. If the steam is used for process heating or cooling and is at a low pressure, a demister pad which gives 1 to 3 ppm solids may be adequate. A flow diagram showing the flow path of water and steam in the system is helpful, particularly in HRSGs for multi-pressure units. In some boilers saturated steam may be taken off the drum and the balance superheated. In some systems, external superheated steam from another boiler or source may be mixed with the saturated steam from the waste heat boiler and the total may be superheated in the waste heat boiler. In order to understand and properly design the various heat transfer surfaces such a flow scheme would indeed be helpful and avoid confusion later. The take off point for deaeration steam is also an important aspect of any HRSG system; as discussed in Chapter 4 if steam is taken off the steam turbine at a suitable extraction point, the plant efficiency is better than if it had been taken off from the boiler outlet. This is because the steam that is drawn off the turbine has performed some work in the process of expansion. The scheme or flow diagram should suggest the source of deaeration steam or any other steam for process.

259

Specifying Waste Heat Boilers

6

1

I.chevron 2.cyclones 3.downcoMer 4.risers 5.feed wnter pipe 6.steQM

2

Figure 5-1. Arrangement of drum internals.

2. SPACE AND LAYOUT GUIDELINES Waste heat boilers may be located inside a building, where the space is limited or could be located outside but within a plant as part of the expansion scheme or there could be no restrictions as in a new project; anyhow space limitations and layout restrictions if any should be outlined in the specifications. The boiler configuration could be developed based on the space availability. As seen in Chapter I, for unfired gas turbine exhaust a two-pass could be used instead of a single gas pass design to minimize floor space. If height is a concern, a short but wide boiler could be designed. There are cost implications to these special designs and hence the specification should state any requirements in the early stages of the project. As discussed in Chapters 1 to 3, several different boiler configurations could be developed for the same gas and steam parameters.

260

Waste Heat Boiler Deskbook

3. GAS PARAMETERS All of the data pertaining to the waste gas should be stated clearly. a. Gas flow in mass units NOT in volumetric units. One of the common errors made by specification writers is to state the gas flow in cubic feet per minute. This should be avoided. The gas mass flow to the boiler should always be given in pounds per houri one can confuse cfm with acfm or scfrn and corne up with a mass flow that is different from the actual flow due to an incorrect estimation of gas density which is dependent on the gas molecular weight, temperature and pressure. Thus several variables are involved and a wrong assumption of anyone of them could result in a different density and hence mass flow. The author is familiar with a case where a specification mentioned that the gas flow was 5000 efm and did not say whether it was aefm or sefm. Two different boiler designers carne up with designs .in which the mass flows were apart by a factor of 4i one designer had assumed that the flow was in sefm and the other thought that the flow was in aefm. A lot of time was wasted in the process of explaining the basis of design to the consultant, which could have been avoided. If the gas is exhaust from a gas turbine, it is prudent to state the gas flow data at different gas turbine loads and ambient conditions and also specify the design point so that comparison or proper evaluation of boiler performance could be made. As mentioned in Chapter 1, the mass flow and exhaust gas temperature vary significantly with ambient conditions. If a bypass damper is used the leakage flow through the damper should be accounted for. In some wood based cogeneration projects, a portion of the hot gases are taken off for drying the wood products and the balance goes to the HRSG. Hence the gas system should be described adequately so that the HRSG mass flow can be properly evaluated. Gas temperature to the boiler should be stated along with the gas flow. If the duct work between the waste gas source and the boiler is significant, the heat loss in this should be accounted for.

r-

Specifying Waste Heat Boilers

261

b. gas analysis. This is another important aspect which is neglected by specification writers. Depending upon the composition, the gas properties such as specific heat, viscosity and thermal conductivity can vary; Appendix E shows how to compute the properties. If the gas properties are different naturally the heat transfer coefficients and hence the duty and gas temperature profiles will be different as discussed in Chapter 4. The difference in duty could be as high as 10% if the moisture content is significantly different. Hence if a steam injected gas turbine exhaust boiler is being specified it is advisable to state the gas flow, exhaust temperature and analysis pertaining to the mode of operation of the turbine, whether steam injected or dry. In the Cheng cycle, discussed in Chapter I, the gas properties vary significantly between the dry mode (7% volume of water vapor) and the injected mode (25% volume of water vapor). Appendix E shows how the gas properties differ. As the size of the boiler gets larger or the steam production increases, even a 2 to 3% difference in output can affect the project economics. Also, depending upon the process, the gas stream could have traces or small fraction of Hel, e12, 502, H2S, or 503, It is in the interest of the specification writer to mention this as the design would have to consider methods of minimizing low or high temperature corrosion concerns as discussed in Appendix B, even though the heat transfer aspects or sizing may not be affected. In hydrogen plants for example the data on partial pressure of hydrogen is used in the selection of boiler tube material as discussed in Chapter 1. Hence gas analysis is relevant. Gas velocity and pressure drop through the boiler are also affected by gas density which in turn depends on the gas analysis used. Even if an analysis is not readily available, the specification should describe the process of obtaining the flue gas-ego natural gas combustion at 20% excess air-and the boiler designer can perform combustion calculations to arrive at the flue gas analysis and flow.

c. Gas pressure and pressure drop. A fire tube boiler may have to selected if the gas pressure is high; if it is marginally higher than atmospheric pressure and a

262

Waste Heat Boiler Deskbook

water tube boiler is used, the casing would have to be strengthened; or as discussed in Chapter I, the boiler or superheater or economizer may have to be located within '.a pressure vessel. If the gas pressure is very high, on the order of a few atmospheres, the gas properties would also be affected. See Reference 2 on computation of gas properties at elevated pressures. Appendix E shows how gas pressure affects heat transfer coefficients in flow inside or over tube bundles. Gas pressure also affects the density and hence the velocity and gas pressure drop. Gas pressure drop through the boiler is an operating expense. As mentioned in Chapter 1 in the case of gas turbine exhaust each 4 in we of additional pressure drop results in nearly 1% decrease in the gas turbine power output. The specification may suggest a pressure drop but should give credit if a HRSG vendor has a lower gas pressure drop than suggested or penalize the design if it is higher. The life cycle cost of this expense should be evaluated, though the initial cost of the design with a high gas pressure drop may look appealing. D. Nature of gas. The nature of the gas stream should be stated, whether it has particulates, whether it a dirty gas and has potential fouling or slagging concerns. If an ash analysis is provided one can infer from it whether the ash has slagging characteristics; if salts of sodium and potassium are present, then as discussed in Chapter I, we know that the melting points of their salts could be low and can result in slagging problems at the front end of the boiler. Depending on the type of combustion or incineration device and the process, whether it is a fixed bed, moving bed or fluidized bed, the carryover of particulates could be different and hence has a bearing on the boiler design. The particulate concentration in grains/scf should be stated. Bare tubes have to be used in the case of a dirty gas, while extended surfaces could be used if it is clean, thus making the design compact. The author is familiar with cases where extended surfaces were used on boilers which should have had bare tubes and which had to be replaced completely within a few years of startup.

_,..._------------~~-~--~~~----~~~--~---~~----

Specifying Waste Heat Boilers

263

Special cleaning provisions as discussed in Chapter 1 should also be made if the gas stream is dirty. It is also a well known fact that it is easier to clean bare tubes compared to finned tubes.

4. BOILER DUTY The duty or the energy to be recovered from the boiler should be clearly stated. This includes the steam and feed water parameters and the steam generation. This may not be difficult in cases such as single pressure boilers. However, if this is not clear, different suppliers can come up with different combinations of evaporators and economizers and thus have a stack gas temperature that could be different. The specification should alternatively state the criterion used for optimization; the cost of energy, steam and fuel should be provided so that the boiler supplier can make his own economic evaluation. In the case of multiple pressure boilers as discussed in Chapter 4, one can have different combinations of HP, IP and LP steam; the specification writer should analyze the HRSG temperature profiles using software such as COG EN and arrive at the temperature profiles and duty of each section at different operation conditions. Some specifications require that the pinch point or approach point have certain values. This is a poor way of specifying boiler duty. As discussed in Chapter 4, the pinch and approach points are a function of the gas flow and inlet gas temperature and vary with operating conditions and load. There could be temperature cross situations and one cannot arbitrarily select these values. Some specifications call out the desired exit gas temperature from the boiler; now this mayor may not be achieved with the lowest practical pinch and approach points as discussed in Chapter 4. Hence one should be familiar with HRSG temperature profiles and how they vary with gas inlet conditions before suggesting the duty or temperature profiles. Part load conditions if any should also be addressed in the specifications. In the case of incinerator exhaust, the gas exit temperature would be lower at lower loads and concerns about low temperature corrosion should be addressed.

-----

-.J

...

-----~---~-

264

Waste Heat Boiler Deskbook

Some engineers are under the impression that low temperature corrosion in the economizer can be avoided if the boiler exit gas temperature is increased at lower loads by using methods such as bypassing a portion of the economizer. This is incorrect. As discussed in Appendix B, the tube wall temperature of the economizer is not affected by the gas temperature but by the feed water temperature. Hence by bypassing the economizer, we are not increasing the tube wall temperature of the economizer. If the objective is to minimize corrosion in the ductwork between the economizer and the stack, then probably it could be achieved by this method. The stearn temperature varies with gas flow and inlet temperature conditions. It is desirable to know if the stearn temperature could be allowed to float within a range or a constant temperature is required at various gas flow conditions. Oversizing of the superheater may be necessary to achieve the desired stearn temperature at lower loads. Provision should be made for stearn temperature control at higher loads. In the case of gas turbine HRSGs it should be stated clearly if the same stearn temperature is required in the unfired as well as in the fired modes. As shown in Chapter 4, the superheater should then be designed to generate the stearn in unfired mode and the temperature would have to be controlled in the fired mode. If a lower stearn temperature is acceptable in unfired modes, then this should also be stated as it will help reduce the size of the superheater.

5. AUXILIARY FUEL DATA In case supplementary fuel is needed, as in gas turbine applications, the fuel analysis and augmenting air requirements if any should be stated. The burner supplier can provide information on augmenting air requirements based on the gas turbine exhaust gas analysis, the firing temperature and the fuel used. If the burner is required to bum a different fuel at a later date, this requirement must be stated in the specifications. Modifications that may be required to be done later to the boiler configuration could be reviewed early in the design and handled less expenSively .

----------------------------

-r--~

Specifying Waste Heat Boilers

265

6. EMISSION GUIDELINES In the case of gas turbine exhaust, where a SCR or catalyst is used for CO or NO x control, the incoming levels of these pollutants as well as the desired outlet level should be stated. Even if the SCR is likely to be added after several years, the HRSG should be designed with this requirement in mind. Space should be provided for the SCR. In addition, the boiler cross section should be preferably close to that of the SCR in order to minimize the gas pressure drop.

7. FEED WATER ANALYSIS AND BLOWDOWN Often the feed water analysis is omitted from the specifications. This is an important data and is helpful in determining the blow down requirements. The boiler water quality depends on the steam pressure. Both ABMA (American Boiler Manufactures Association) and ASME have issued guidelines, Tables 5-1 and 5-2. In order to calculate the blow down, the extent of make up water and condensate returns should also be known as illustrated in the example below. Example 1: A boiler generates 25,000 pph of saturated steam at 400 psig. 60% ~ of the feed water is condensate returns and 40% is make up water, which has a TDS of 250 ppm. The boiler drum is maintained at 2000 ppm Determine the blow down, make up flow and condensate returns Solution: Let us define a few variables. See Figure 5-2. Let E = evaporation, 25000 pph R = condensate returns, pph M =make up water flow, pph F = feed water flow, pph B =blow down, pph S =steam to process, pph =E-R

Table 5-1. Suggested Water Quality limits [Adapted From ASME 1979 Consensus] Boiler Type: Industrial watertube, high duty, primary fuel fired, drum type Makeup Water Percentage: Up to 100% of feedwater Conditions: Includes superheater, turbine drives, or process restriction on steam purity Saturated Steam Purity Target (9) Drum Operating

~

MPa (psig)

0-2.07

~

2.08-3.10 (301=-15Q)

3.11-4.14 (~

4.15-5.17 (601-750)

5.18-6.21 (751-900)

6.22-6.89 (901-1000)

6.90-10.34 (1001-1500)

10.35-13.79 (1501-2000)

<0.04 sO.050 sO.025

<0.007 sO.030 $0.020

0.007 SO.025 SO.020

<0.007 SO.020 SO.015

<0.007 sO.020 SO.015

<0.007 s0.o10 sO.010

<0.007 S0.o10 s0.o10

-Not detectable9.0-9.6 9.0-9.6

Feedwa\er ill Dissolved oxygen (mg/l 02) measured before oxygen scavenger addition (8) Total iron (mg/l Fe) Total copper (mg/l Cu) Total hardness (m g/lCaC0 ) 3 pHrange@25°c Chemicals for preboiler system protection

<0.04 SO.100 sO.050 $0.300 7.5-10.0

SO.300 7.5-10.0

sO.200 7.5-10.0

$0.200 7.5-10.0

SO.100 7.5-10.0

SO.050 8.5-9.5

Nonvolatile TOC (mg/l c)(6)

<1

<1

<0.5

-As low as possible, <0.2 - - - -

Oily matter (mg/l)

<1

<1

<0.5

<0.5

<0.5

-As low as possible, <0.2

Use only volatile alkaline materials

~

~

Boiler Water Silica (mg/l S102> Total alkalinity (mgfl CaCo3) Free hydroxide alkalinity (mg/l CaC03)<2)

til

g s150 <350(3)

s90 <300(3)

S40 <250(3)

s30 <200(3)

S20

S8

S2

<150(3)

<100(3)

-NotSpedfied(4)-

::c

(!)

~ tl:i

0

:::; Not specified

Not detectab1e(4)-_

Specific conductance (Jlffiho/cm) @ 25°c without neutralization

<1

~

0(!) til

i';"'

<3500(5)

<3000(5)

2500(5)

<2000(5)

<1500(5)

<1000(5)

0"'

S150

Sl00

0 0

i';"'

Notes for Table 5-1

1. 2. 3.

4.

5.

6.

7.

8. 9.

With local heat fluxes >473.2 kW/m 2 (>150,000 Btu/hrlft2 ), use values for the next higher pressure range. Minimum level of OH- alkalinity in boilers below 6.21 MPa (900 psig) must be individually specified with regard to silica solubility and other components of internal treatment. Maximum total alkalinity consistent with acceptable steam purity. If necessarY,should override conductance as blowdown control parameter. If makeup is demineralized water at 4.14 MPa (600 psig) to 6.89 MPa (1000 psig), boiler water alkalinity and conductance should be that in table for 6.90 to 10.34 MPa (1001 to 1500 psig) range. Not detectable in these cases refers to free sodium or potassium hydroxide alkalinity. Some small variable amount of total alkalinity will be present and measurable with the assumed congruent or coordinated phosphate-pH control or volatile treatment employed at these high pressure ranges. Maximum values often not achievable without exceeding suggested maximum total alkalinity values, especially in boilers below 6.21 MPa (900 psig) with >20% makeup of water whose total alkalinity is >20% of TDS naturally or after pretreatment by line-soda, or sodium cycle ion exchange softening. Actual permissible conductance values to achieve any desired steam purity must be established for each case by careful steam purity measurements. Relationship between conductance and steam purity is affected by too many variables to allow its reduction to a simple list of tabulated values. Nonvolatile TOC is that organic carbon not intentionally added as part of the water treatment regime. Boilers below 6.21 MPa (900 psig) with large furnaces, large steam release space and internal chelant, polymer, and/or antifoam treatment can sometimes tolerate higher levels of feedwater impurities than those in the table and still achieve adequate deposition control and steam purity. Removal of these impurities by external pretreatment is always a more positive solution. Alternatives must be evaluated as to practicality and economics in each individual case. Values in table assume existence of a deaerator. No values given because steam purity achievable depends upon many variables, including boiler water total alkalinity and specific conductance as well as design of boiler, steam drum internals, and operating conditions (note 5). Since boilers in this category require a relatively high degree of steam purity, other operating parameters must be set as low as necessary to achieve this high purity for protection of the superheaters and turbines and/or to avoid process contamination.

268

Waste Heat Boiler

v",'.o.U'JUK.

Table 5-2 Watertube boilers Recommended Boiler Water Limits and Associated Steam Purity At Steady State Full Load Operation Drum Type Boilers

fABMA-19821 Drum Pressure

Rarge Total Dlsolved Solids 1 Boier Water ppm (MAX)

psg

Rarge Total Alkalinity 2 Boller Water ppm

0-300 301-450 451-600 601-750 751-900 90HOOO

700-3500 600-3000 500-2500 200-1000 150·750 125-625

140-700 120-600 100-500 40-200 30-150 25·125

1001-1800 1801-2350 2351-2600 2601-2900

100 50 25 15

NOTE (3)

1400 and Above

0.05

Suspended Solids Boiler Water ppm (MAX)

Rarge Total Dlscived Solids, 2, 4 Steam ppm (MAX expected value)

15 10 8 3 2 1

0.2-1.0 0.2-1.0 0.2-1.0 0.1-0.5 0.1·0.5 0.1-0.5

1

0.1 0.1 0.05 0.05

N/A N/A N/A OOQE II:lBQUGI:l BQILEBS

N/A

N/A

0.05

NOTES: 1. Actual Values within the range reflect the TDS In the feed water. Higher values are for high solids, lower values are for low solids In the feed water. 2. Actual values within the range are directly proportional to the actual value of TDS of boiler water. Higher wavlues are for the high solids, lower values are for low solids in the boiler water. 3. Dictated by boiler water treatment. 4. These values are exclusive of silica.

The total dissolved solids (TDS) in blow down water is 2000 ppm; TDS of make up water is 250 ppm; the steam or condensate returns is assumed to have 0 solids. Mass balance gives the following equations: F=;M+R (5~1)

S=E-R

(5~2)

since the condensate returns and make up are in the ratio of 60 to 40%, RIM =1.5 (5~3)

269

Specifying Waste Heat Boilers

s

R

E

2 tv1

F

1 B

Figure 5-2. Scheme of blow down system.

from a mass balance of dissolved solids,

250 xM + 0 x R = solids in feed water = 2000 x B

(5-4)

orM=8B

(5-5)

also, R + M

=25000 + B

(5-6)

substituting (5-3) and (5-5) in (5-6) ,

2.5 M =25000 + B or 20 B =25000 + B hence B = 1316 pph; M

(5-7)

= 10526 pph; R = 15790 pphi F =26316 pph

In case all of the feed water is make up, then R =O. Then:

F=M

(5-8)

250 F =2000 B

(5-9)

Waste Heat Boiler Deskbook

270

(5-10)

F=25000 +B

hence B = 25000/7

=3572 pph and F =28572 pph.

Note also that the deaeration steam requirements will change with the ratio of R/M . The more the condensate returns, less the de aeration steam. In addition to the feed water or make up water analysis, the specification should give the feed water temperature. It is usually the saturation temperature corresponding to the deaerator pressure if thermal deaeration is used.

8. SURFACE AREA AND FIN CONFIGURATION In applications such as gas turbine exhaust, extended surfaces are widely used. Depending upon the fuel used in the gas turbine and HRSG burner, up to 5 or 6 fins/in. may be used for the evaporator and economizer. Lower fin density should be used for surfaces with a low tube side coefficient such as superheaters or air heaters as discussed in Appendix A. Engineers who are not familiar with heat transfer aspects of extended surfaces sometimes fall into the trap of assuming that more surface area means more duty. This can be wrong as shown in Appendix A. As surface area is determined by a combination of factors, specifying the fin density to be used or purchasing a design based on surface area alone should be avoided. In applications where the gas stream could be slightly dirty, the maximum number of fins to be used from cleaning considerations may be specified. Surface area is also affected by the gas velocity used. One should also note that the surface area is sometimes specified on tube inner diameter basis in case of fire tube boilers. Finally, the basis for evaluation should be the overall operating and installed costs and not surface area.

Specifying Waste Heat Boilers

271

9. COST DATA FOR ELECTRICITY, FUEL AND STEAM Infonnation on the cost of utilities such as electricity, fuel and steam should be provided whenever possible so that the HRSG designer can make his own economic evaluation of the design. In a fired HRSG for example, there can be variations in gas pressure drop, fuel consumption and steam production among different designs. If the cost data and the period of operation in each mode (unfired, fired low load, high load etc.) are known, a study may be performed by the HRSG designer and the design may be optimized to yield the lowest life cycle cost. The consultant should also use a life cycle costing analysis to arrive at the optimum design from different bidders and not go by initial investment alone. While selecting an option based on low initial cost may look attractive in the short term, in the long term it may prove to be uneconomical. Hence the cost of major utilities such as steam, fuel, electricity should be stated in the specifications.

10. DRUM SIZING One of the parameters that is often suggested in the. specification is the drum hold up time or the duration between various levels such as normal level to low level or normal level to empty. The objective is to ensure that steam can be generated for a few minutes even if the feed pump or the heat source is cut off without setting off the level trips. The drum sizing is usually based on steam release rates and the steam purity that can be achieved with the required drum internals. In addition, the holdup time criterion has to be met. Figure 5-3 shows the fonnula used for computing the hold up as a function of the liquid level. The formula may also be used to compute the duration between levels as shown in Example 2 below. Example 2: A boiler generating 20000 pph of steam at 400 psig has a 42 in. drum 10 feet long with 2:1 ellipsoidal ends. If the normal water level (NWL) is 2 in. below the center line and the low level cut off (LLCO) is 4 in. below it, detennine the duration available between NWL and LLCO at the design steaming rate.

Waste Heat Boiler Deskbook

272

Liquid volume In straight section = (LR'/231)[(a/S7.3). alnacosa

V.

Liquid volume In each end

V. = O.261H'(3R- H)/231 Total liquid volume V, = 2V•.+ V.

k-______~L~______~.I Figure 5-3. Partial volume of liquids in pressure vessels.

Solution: The specific volume of saturated water at 400 psig = .01934 cu ft/lb. hence evaporation rate =20000 x .0193/60 =6.43 cu ft/min.

L = 120 in., R = 21 in., Hl = 19 in., and H2 = 15 in. It is desired to find the volume Vl - V2, where Vl corresponds to Hl and V2 to H2 ..cOS a = 2/21. hence a = 84.53 . cosa = .09523; sina = .9954 . Vsl = 120 x 2121 x [(84.53/57.3) - .09523 x .9954] = 73052 cu. in. =42.38 cu. ft. Vel = .261 x 19 x 19 x (3 x 21-19) = 4145.7 cu. in = 2.4 cu. ft. Hence Vl = 42.38 + 2 x 2.4 =47.08 cu. ft. Similarly V 2 = 34.1 cu. ft. Hence V 1 - V 2 = 13 cu. ft. and the duration between the level 19 in., and 15 in., = 13/6.43 = 2.02 min. In some applications the drum has to be sized to handle a fast start up or load change requirements. This usually increases the hold up time. These are some of the important aspects to be looked into while developing a specification for waste heat boilers from the process view point. Mechanical details and code aspects are not covered

Specifying Waste Heat Boilers

273

here. Table 5-3 below shows a format that summarizes the above discussions. Table 5-3 Engineering data or check list for waste heat boiler design (or aspects to be considered while developing specifications) 1.

2. 3.

4.

5. 6. 7.

8. 9.

10.

Application: describe process; give flow diagram for gas/steam; source of deaeration steam; distribution of process and superheated HP,IP and lP steam. Space limitations: describe or provide drawings. Is site visit required? a. gas flow, pph: (at different loads/ambient conditions). b. inlet gas temperature, F: (associated with gas flow) c. analysis, % volume: C02, H20, N2, 02, S02, HCl, S03, H2S, CL2, etc. corresponding to each gas flow condition d. gas pressure, psig: e. suggested gas pressure drop: (at a given gas flow and inlet condition) or the maximum value. f. nature of gas: dirty/clean; particulate concentration in grains/set; ash analysis to indicate slagging tendencies. Duty or suggested steam generation/temperature profile or exit gas temperature: part load conditions; % of time in each operating mode; steam temperature control if any; provide flow diagram showing HP/IP/lP steam, make up water, condensate returns at various loads. Auxiliary fuel data: fuel analysis, augmenting air for burners if any. Emission data: NOx and CO at boiler inlet and outlet. Contribution by burner: (from burner vendor): Emission control equipment suggested: Feed water or make up water analysis: % condensate returns if any. Is demineralized water available for spray temperature control? The TDS in injection water should be very low. Cost of fuel, electricity and steam: In addition, the % of time in each operating point in order to optimize life cycle cost. Steam purity requirements: Drum hold up time criteria if any. Quick start up or load change requirements. Special requirements if any.

Waste Heat Boiler Deskbook

274

REFERENCES 1. V. Ganapathy, "HRSG features and applications", Heating,

piping and air-conditioning, Jan 1989, pg 169. 2. ABMA guidelines on water quality for boilers. 3. ASME guidelines on water quality.

r

--~~-----

~~-

-~~-~~~-----~

Appendix A

Extended Surface Heat Transfer Finned tubes are extensively used in boilers, superheaters, economizers and water heaters for recovering energy from clean gas streams such as gas turbine exhaust or flue gas from combustion of premium fossil fuels; if the particulate concentration in the gas stream is very low, finned tubes with a low fin density may be used. However the choice of fin configuration particularly in clean gas applications is determined by several factors such as tube side heat transfer coefficient, overall size, cost and gas pressure drop, which affects the operating cost. This section deals with calculations pertaining to circumferentially finned tubes of solid or serrated type, Figure A-1 and A-2 and the effect of fin geometry on performance of heat transfer equipment. Though several correlations may be found in literature, the widely used ESCOA correlations will be the basis for these studies and conclusions. Finned surfaces are attractive when the ratio between the heat transfer coefficients on the outside of the tubes to that inside is very small. In boiler evaporators or economizers, the tube side coefficient could be in the range of 1500 to 3000 Btu/sq ft h F, while the gas side coefficient could be in the range of 10 to 20 Btu/ sq ft h F. A large fin density or a large ratio of external to internal surface area is justified in this case. As the ratio between the outside and inside coefficient decreases, the effectiveness of using a large ratio of external to internal surface area reduces. For example, in superheaters or air heaters, where the tube side coefficient could be in the range of 30 to 300 Btu/ sq ft h F, it does not pay to use a large fin surface; in fact it is counter productive as will be shown later. 275

Waste Heat Boiler De!;kbl)ok

276

A moderate fin density such as 2 or 3 fins/in. is adequate, while for economizers or evaporators, a 5 or even 6 fins/in. may be justified if cleanliness permits. The other important fact to be kept in mind is that more surface area does not necessarily mean more energy transfer. It is possible to have, through a poor choice of fin configuration, more surface area and yet transfer less energy. One has to look at the product of surface area and overall heat transfer coefficient and not the surface area alone. Overall heat transfer coefficient reduces significantly as we increase the fin surface or use more fins/in. We will discuss these aspects as we go along.

HEAT TRANSFER, PRESSURE DROP The. basic equation for heat transfer coefficient with finned tubes is given by:

(A-1)

The calculation for tube side coefficient hi is discussed in chapters 2 and 3. ho consists of two parts, a non-luminous coefficient h n which is computed as discussed in chapter 3 and he, the convective heat transfer coefficient. Computation of he and involve an elaborate procedure and solving of several equations, as detailed below. Determination of he (or h g)

=Cl C3 C5{(d + 2h)/d ].5[( tg + 460) /

he

(ta + 460) ]-25 GCpk/pCp)/O.67 G

= Wg/{ (St/12) -

Ao

= (d/12)

Ao] LNw

+ nbh/6

Cl to C3 are obtained from Table A-l.

(A-2) (A-3) (A-4)

,....------~----------~-"---------------

277

Appendix A - Extended Surface Heat Transfer

= Gd/12Jl

(A-5)

S = (l/n) - b

(A-6)

Re

Table A-t. Factors <;1 to C6 Solid Fins

Serrated Fins

C1 = 0.25 Re-0.35 C2 = 0.07 + B.O Re -0.45

C1 = 0.25 Re- 0.35 C2

INLINE

= 0.07 + B.O Re -0.45 INLINE

= 0.20 + 0.65 e (-0.25h/s) 15 C4 = O.OB [0.15Stld oi-1.1 (hlslo. ] C5 = 1.1 -[0.75-1.5 e(-0.70N d)]

C3 = 0.35 + 0.50 e(-0.35h/S) 20 C4 = O.OB [0.15Stldo][-1/1 (h/slo. ]

C3

Cs = 1.1 - [0.75 - 1.5e(-O·70 Ndlj [e(-2.0SI/Stlj C6 = 1.6 - [0.75 - 1.5e(-O·7 Nl j [e(-2.0SVS 2 C6 = 1.6 - [0.75 - 1.5e(-O·7 Ndlj [e(-2.0SVStl j

tll

STAGGERED

STAGGERED

C3 = 0.35 + 0.65 e (0.25h/s) 20 C4 = 0.11 - [0.05Stldoj [-0.7 (h/slo. ]

C3

= 0.55 + 0.45 e(-0.35h/S)

C4 = 0.11 _ [0.05Stldoj [-0.7 (h/slo.

23 ]

= 0.7 + [_.70_0.Be(-O·1SNSlj [e(-1·0SVStlj Cs = 0.7 + [_.70_0.Be(-O·1SNSll[e(-1·0SI/Stlj C6 = 1.1 + [1.B-2.1e(-O·1SN&lj [e(-2.0S VStlj

Cs

-[0. 7_0.Be(-O.1SN&lj [e(-O·6S VStl j C6

=1.1 + [1.B-2.1 e(-O·1SN&lj [e(-2.0SVStl j -[0.7_0.Be(-O.1SNSlj [e(-O·6SI/Stl]

Waste Heat Boiler Deskbook

278

FIN EFFICIENCY AND EFFECTIVENESS For both solid and serrated fins, fin effectiveness 1]

1]

is:

=1- (1-E) At/At

(A-7)

For solid fins:

At= 1tn (4dh + 4h Z + 2bd + 4bh)J24 At

=At

(A-8) (A-9)

+ mis (1-nb)/12

E =1/[1 + .002292 mZ hZ ( (d + 2h) /d).5]

Where m

= (24 ho/Kb)'S

(A-I0) (A-I1)

For serrated fins:

At = min[2h (ws + b) + bws] /12ws

(A-12)

At =At + mi (l-nb) /12

(A-13)

E =[tanh (mh) ] /mh

(A-14)

Where m =[24 ho (b + ws) / (Kbws) ]-5

(A-1S)

Gas pressure drop L1Pg= (f + a) CZNd/ (Pg x 1.083 x 109)

f

= Cz

(A-16)

C4 C6[(d + 2h) /d]-S for staggered arrangement

(A-I7)

= Cz C4 C6[ (d + 2h) /d] for inline arrangement

(A-18)

a =[(1 + BZ) / (4Nd)]

(A-19)

Where B

= (free

(tgZ -tg l) / (460

+ tg )

gas area/total area)Z

(A-20)

C2, C4, C6, are given in Table A-I for solid and serrated fins.

-r---------------"~"-~-

- ------- ---- ----------- ---------

--~---~--~~

279

Appendix A - Extended Surface Heat Transfer

Serrated fins

Solid fins

Figure A-l,2. Solid and serrated fins.

Table A-2. Bessel function, 101 11, Ko, and Kl Values for Various Augments X

lo(X)

11 (X)

Ko(X)

K1 (X)

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.2 1.4 1.6 1.8 2.0 2.2 2.4 2.6 2.8 3.0

1.0 1.002 1.010 1.023 1.040 1.063 1.092 1.126 1.166 1.213 1.266 1.394 1.553 1.75 1.99 2.228 2.629 3.049 3.553 4.157 4.881

0 0.05 0.10 0.152 0.204 0.258 0.314 0.372 0.433 0.497 0.565 0.715 0.886 1.085 1.317 1.591 1.914 2.298 2.755 3.301 3.953

8 2.427 1.753 1.372 1.114 0.924 0.778 0.66 0.565 0.487 0.421 0.,318 0.244 0.188 0.146 0.114 0.0893 0.0702 0.554 0.0438 0.0347

8 9.854 4.776 3.056 2.184 1.656 1.303 1.05 0.862 0.716 0.602 0.434 0.321 0.241 0.183 0.140 0.108 0.0837 0.0653 0.0511 0.0402 (Continued)

Waste Heat Boiler Deskbook

280

Table A-2 (Continued). Bessel function, 10 , Iv KO, and Kl Values for Various Augments X

'o(X)

11 (X)

Ko(X)

K1 (X)

3.2 3.4 3.6 3.8 4.0 4.2 4.4 4.6 4.8 5.0

5.747 6.785 8.028 9.517 11.30 13.44 16.01 19.09 22.79 27.24

4.734 5.670 6.793 8.140 9.759 11.70 14.04 16.86 20.25 24.34

0.0276 0.0220 0.Q175 0.0140 0.0112 0.0089 0.0071 0.0057 0.0046 0.0037

0.0316 0.0250 0.0198 0.0157 0.0125 0.0099 0.0079 0.0063 0.0050 0.0040

TUBE WALL AND FIN TIP TEMPERATURES For solid fins the relation between tube wall and fin tip temperatures is given by: (t g - tf)/ (tg - tb) =[ K1(mre)lo(mre) + 11 (mre)Komre)] J K1 (mre) 10 (mro) + Ko (mro) 11 (mre)

(A-21)

The various Bessel functional data are shown in Table A-2. For serrated fins, treated as longitudinal fins: (A-22)

A good estimate of tf may also be obtained for either type of fins as follows: tf

= tb + (tg- tb) x (1.42 -1.4 x E)

(A-23)

tb , the fin base temperature is estimated as follows: tb

=ti + qx (R3 + R4 + R5)

(A-24)

281

Appendix A - Extended Surface Heat Transfer

Where R3, R4 and R5 are resistances to heat transfer of the inside film, fouling layer and the tube wall, and heat flux q is given by: (A-26)

The following example illustrates the use of the equations. Example 1: A steam superheater is designed for the following conditions: Gas flow = 150,000 pph Gas inlet temperature = 1000°F Gas exit temperature =861°F Gas analysis: % volume C02 =12, H20 =12, N2 =70, 02 =6 Steam flow =30,000 pph Steam temperature in. =491°F (sat) Steam exit temperature = 787°F Steam pressure (exit) =600 psig. Tubes used: 2 x .120 low alloy steel tubes, 18 tubes/row, 6 deep, inline arrangement with 4 in. square pitch and 9 streams. Tube inner diameter =1.738 in. Length =10 ft. Fins used: Solid stainless steel, 2 fins/in., .5 in. high and .075 in. thick. Fin thermal conductivity K = 15 Btulft h F. Determine the heat transfer coefficient and pressure drop. Solution: Ao = (2/12) + (2 x .5 x .07516) = .17917 sq /tIft. G =150,0001 [18 x 10 x «4112) - .17917)J =5420 Lblsq ft h. The gas properties at the average gas temperature (from Appendix F) are: Cp= .2851, Jl =.08146 and k =.03094 Re = 5420 x 2 1(12 x .08146) = 11090 Cl =.25 x 11090-. 35 =.0096 s =112 - .075 =.425 C3 =.2 + .65 e (-.25 x .5/.425) =.6843 C5 =1.1 -[.75 -1.5 e (-.7 x 6)J [e (-2 x 4/4) J =1.0015 Assume that the average fin temperature is 750°F. The average gas temperature = 930°F and steam temperature = 640°F. The fin thermal conductivity K is assumed to be 15 Btu/ft h F.

282

Waste Heat Boiler Deskbook

he =.0096 X .6843 X 1.0015 X (3/2) .5 x[ (930 + 460) / (750 + 460 ) ] 0.25 X 5420 X . 2851 X (.03094/.2851/.08146) .67 = 15.74 hn = 1.12 using methods. discussed in Chapter 3. The beam length for finned tubes is computed as 3.4 X volume/surface area. Hence ho = 15.74 + 1.12 = 16.86, m =[24 X 16.86/ (15 X .075) 1.5 = 19. E = 1/[1+.002292 x 19 x19 x .5 x .5 xyT3] = 0.80 AI = 3.14 X 2 x[4 X 2 x .5 + 4 x .5 x .5 + 2 x .075 x 2 + 4 x .075 x .51 /24 = 1.426 At = 1.426 + 3.14 x 2 x (1-2 x .075)/12 = 1.871 Hence = .8 + (1-.8) x 1.426/1.871 = .848 Let us compute hi for steam. w = 30000/9 = 3333 lb/h per tube; From Table 3-3 in Chapter 3, factor C = .337. hi = 2.44 x .337 x 3333.8/1.738 = 200 Btu/sq ft h F. l/U = 1/(16.85 x .848) + 12 x 1.872/(200 x 3.14 x 1.738) + .001 + .001 x 1.871 x 12/(3.14 x 1.738) + 24 x In(2/1.738 x 1.871/(24 x 20 x 3.14 x 1.738) = .0699 + .0211 + .001 + .0041 + .0024 = .0985; U = 10.16 Btu/sq ft h F. Calculation of tube wall and fin tip temperature Heat flux q = 10.16 x (930-640) = 2945 Btu/sq ft h, tb = 640 + 2945 x (.0024 + .0041 + .0211) =722 F. Using the elaborate Bessel functions: mre = 19 x 1.5/12 = 2.38 ft. mro = 1.58 ft Ko (2.38) = .07, K1 (2.38) = .0837, 10 (2.38) = 3.048, 11 (2.38) = 2.295, Ko (1.58) = .186,10 (1.58) = 1.74, Hence (930-tf)f(930-722) = (.0837 x 3.048 + 2.295 x .07 )/(.0837 x 1.74 + .186 x 2.295) = .723, Hence tf = 780°F. Using the approximation tf = tb + (1.42-1.4 x .8) x (930-722) = 785°F. Note that this is only an average base and fin tip temperature. For material selection purposes one should look at the maximum heat flux, which occurs for instance at the gas inlet in counter flow arrangement and also consider the non-uniformity or maldistribution in gas and stearn flow. A computer program may be developed to compute the tube wall and fin tip temperatures at various points along the tube length and the results used to select appropriate materials.

--- - -

Appendix A - Extended Surface Heat Transfer

283

It may be noted from the above that there are a few ways to reduce the fin tip temperature: 1. Increase fin thickness. This reduces factor m and hence tt. 2. Increase thermal conductivity of fin material. This may be difficult as the thermal conductivity of carbon steels is higher than alloy steels and carbon steels can withstand only up to 850°F, while alloy steels can withstand up to 1300°F, depending on the alloy composition. 3. Reduce ho or the gas side coefficient by using a lower gas mass velocity. 4. Reduce fin height or density. 5. In designs where the gas inlet temperature is very high, use a combination of bare and finned rows. The first few rows could be bare, followed by tubes with a low fin density or height or increased thickness and then followed by tubes with higher fin density, height or lower thickness to obtain the desired boiler performance. A row-by-row analysis of the finned bundle is necessary, which requires the use of a computer program.

Computation of gas pressure drop: C2 = .07 + 8 x 11090 -.45 = .191 C4

= .08[.15 x2l-1.1 x (.5/.425) .15J = .3107

C6=1 f = .191 x .3107 x 1 x (3/2) = .089 B2 =[(.33-.17917)/.331 2 = .2089 a =[(861-1000)/(460 + 930) 1 x[1 + .20891/24 =-.005 LlPg = .084 x 5420 x 5420 x6/(.0288 x 1.083 x 109 ) = .53 in we Computer solution to the above system of equations saves a lot of time. However, the author has developed a chart, Figure A-3 which may be used to obtain he (or h g) and 11 values for serrated fins and inline arrangement for various fin configurations and gas mass velocities for gas turbine exhaust gases at an average gas temperature of 600°F. While a computer program is the best tool, the chart will be used to show trends and the effect of fin configuration on performance of finned surfaces. The use of the chart is explained with an example.

-~

Waste Heat Boiler Deskbook

284

!I: t:

"" :.

?;

t::. PG./I0 ROilS 5

88 86

22

4

84

21

3

20

2

19

I

18

0

23

sa eo

78 76 74 72 70 68 66 ~

tf~

17

64

16 15

0

.r;

14 13

fiG, 2 EFFECT [If' FIN CIlNPlGURA nON ON HEAT TRANSfER. PRESSIlRE IDROP

12 11 10

9

5000

G, LB/Sd 6000

7000

8000

9000

n

H <MI\SS 11£-llDCITY)

10000

nooo

12000

Figure A-3, Effect of fin configuration on heat transfer, pressure drop [Chemical Engineering]

Example 2: Determine the overall heat transfer coefficient and pressure drop for a finned tube boiler for gas turbine exhaust under the following conditions: 1. Gas flow =150,000 pph (% vol C02 =3, H20 = 7, N2 =75 and 02 = 15).1 2. Gas inlet temperature =1000°F. 3. Exit gas temperature = 382°F. 4. Duty = 150000 x .2643 x .99 x (1000-382) = 24.25 MMBtu/h (1 % heat loss assumed); Cp = .2643 was taken from Table A-1 above. (See Appendix E also). 5. Steam pressure =150 psig. 6. Feed water temperature =240°F. 7. Fouling factors (in/out) = .001 ft 2 hF/Btu.

285

Appendix A - Extended Surface Heat Transfer

8. Boiler configuration: 18 tubes/row; square pitch = 4.0 in.; Length = 10 ft; serrated fins; 4 fins/in., .75 in., high, .05 in. thick, all carbon steel (Km = 25) surface area of finned tube = 5.35 sq ft/ft. Let us use equation (A-l) and Figure A-3 to arrive at he and Uo. Note that he and hg are used synonymously. The gas properties have to be computed first. Table A-3 below gives the properties along with the factor F used to compute hg.

Table A-3 Data for Gas Turbine Exhaust Gases (% vol C02 = 3, H20 = 7, N2 =75, 02 = 15) temp, f

sp. heat

viscosity

tho cond

F

200 400 600 800

.2529 .2584 .2643 .2705 .2767

.05172 .0612 .0702 .07885 .0870

.0182 .02176 .02525 .02871 .0321

.1152 .1238 .1316 .1392 .1462

1000

To obtain hg the gas mass velocity G must be computed. Using the equations given earlier, Ao = (2/12) + (5 x .75 x .06/6) = .1979 sq fi/ft. G = 150000/(18 xlO x (.33-.1979) = 6308 Lb/sq ft h. From Figure A-3, hg = 11.5; fin effectiveness = .745; gas pressure drop/10 rows = 1.5 in. wc(note that Figure A-3) has been developed for gas turbine exhaust gases for an average gas temperature of 600°F and for serrated tubes; hence corrections for gas data or fin type should be done as required. The gas pressure drop is for 10 rows and corrections for actual number of rows should be made. Let us assume that hi, the tube side coefficient = 2000. The boiling heat transfer coefficient is very high compared to the gas side coefficient and hence does not impact hg . Ratio At/Ai = 5.35/(3.24 x 1.77/12) = 11.55. Then substituting in (A-l), we have:

286

Waste Heat Boiler Deskbook

l/Uo = (1/11.5 X .745) + (11.55/2000) + .001 X 11.55 + .001 + 2 X 10.85 X In(2/1.77)/24/25 = .13946. Hence Do = 7.17; since the average gas temperature in our case is close to 6000 P and the gas analysis is the same as that used for the chart, no correction is required; else factor F should be computed and hg should be corrected; also due to the low gas temperature, non-luminous heat transfer coefficient was neglected. Log-mean temperature difference, LiT = (1000-366) - (382366)/ln{1000-366)/(382-366)] = 168°F. Surface area required = 24250000/168/7.17 = 20140 sq ft = 5.35 X 18 X 10 X Nd, where Nd = number of rows deep = 21. Hence gas pressure drop = (21/10) x 1.5 = 3.15 in. wc.

COMPARISON OF BARE VERSUS FINNED TUBE EVAPORATOR Example 3:

Let us see the advantages of using finned tubes. Try to design the boiler for the same application using bare tubes and compare the two designs. Chapter 2 discussed the methodology for design of bare tubes. Let us use the same cross section and tubes/row, length, tube size, and pitch. G = 150000 x 12/(18 x 10 x (4-2» = 5000 Lb/sq ft h Gas film temperature = .5 x[(1000 + 382)/2 + 366] =528°F. From Table A-I, Cp= .262,jl = .0675, k = .024. Fl = .262.33 x .024.67/.0675. 27 = .1093. he = .9 x 5000. 6 x .1093/2.4 = 12.36; non-luminous heat transfer coefficient is more significant for bare tubes compared to finned tubes; hn can be shown to be about 0.5 Btu/sq ft h F. Hence ho = he + hn = 12.86. From Chapter 3, Equation (3-2), l/U o = 1/12.86 + (2/1.77) x .001 + .001 + (2/1.77)/2000 + 2 In(2/1.77)/24/25 = .08085; Uo = 12.37 Btu/sq ft h F. Surface area required = 24250000/168/12.37 = 11670 sq ft = 3.14x 2 x 18 x 10 x Nd/12 = 94.2 Nd; Nd = 124. Gas pressure drop can be computed as in Chapter 3 and shown to equal 4.5 in. wc. Table A-4 shows the results. Advantages of using finned tubes are obvious.

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Appendix A - Extended Surface Heat Transfer

287

Table A-4 Comparison of bare vs finned tube boiler bare tube 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13.

Gas flow, pph inlet gas temp, F exit gas temp, F duty, MMBtu/h steam press, psig feed water temp, F steam flow, pph surface area, sq It Uo, Btu/sq ft h F gas pr drop, in wc no of rows deep heat flux, Btu/sq It h tube wall temp, F

finned tube

>

150,000

< < < < < < <

,1000

382 24.25 150 240 24500 11670 12.86 4.5 124 9213 385

> > > > > > 20140 7.17 3.15 21 52295 484

Tubes wide = 18,length =1 Oft, square pitch = 4.0 in.; finned tubes use

:4 fins/in. serrated fins, .75 in. high, .05 in. thick.

The advantages of using extended surfaces are obvious. The finned tube design is more compact as it has fewer rows deep; this also results in lower labor cost. The length of drums or casing would also be smaller as a result of fewer rows deep, resulting in savings in material cost. The gas pressure drop is also lower, resulting in lower operating costs. It can also be shown that the weight of the finned bundle is much lower. However the heat flux inside the tubes is much higher with finned tubes, as also the tube wall temperature. This is due to the larger ratio of external to internal surface area. The heat flux difference is more pronounced when the gas inlet temperature is higher, as in fired HRSGs. Hence care must be taken to use appropriate finning. A few bare rows of bare tubes, followed by tubes with low fin density and then with high fin density tubes is recommended.

COMPARISON OF INLINE VERSUS STAGGERED ARRANGEMENT Both inline and staggered arrangements have been used with extended surfaces. The advantages of staggered arrangement are higher overall heat transfer coefficients and lesser surface area; cost

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could be marginally lower depending on the configurationj gas pressure drop could be higher or lower depending on the gas mass velocity used. If cleaning lanes are required for soot blowing, an inline arrangement is preferred. Solid as well as serrated fins are used in the industry. Generally solid fins are used in applications where the deposition of solids are likely. The following examples compare the effect of arrangement on boiler performance. Example 4: 150,000 pph of turbine exhaust gases at 10000P enter an evaporator of a waste heat boiler generating stearn at 235 psig. Determine the performance using solid and serrated fins and inline versus staggered arrangement. Tube size is 2 x 1.77 in.j tubes/row = 18, length = 10 ft. use 2 x .75 x .05 and 5 x .75 x .05 fins. Solution; Using the Escoa correlations and the methodology discussed in Chapter 3 for evaporator performance, the following results shown in Table A-5 were arrived at. Data in column 3 are for a staggered design with a duty close to the inline arrangement. Table A-5 Results of boiler performance (solid/serrated fins and inline/staggered arrangements)

1.

2. 3. 4. 5. 6. 7.

8. 9. 10.

Table 5a-5 x .75 x .05 Table 5b-5 x .75 x .05 x .157 Solid fins Serrated fins 1 2 3 col.no 1 2 3 in arrgt in st st st st 18 18 18 18 18 18 tubes/row 20 no. deep 20 18 20 17 20 length 10 10 10 10 10 10 6.51 7.71 7.71 7.18 8.67 8.69 Uo 4.87 4.38 LlPg 2.76 3.19 4.62 5.45 ,23.30 23.14 23.46 23.28 23.24 23.55 dutyQ 410 415 exit gas,F 418 416 414 408 surface (sqft) 21677 21677 19509 20524 20524 17446

-

Appendix A - Extended Surface Heat Transfer

1. 2. 3. 4. 5. 6. 7.

8. 9. 10.

289

Table 5c-2 x .75 x .05 Table 5d-2 x .75 x .05 x .157 Solid fins Serrated fins 1 col. no 2 2 1 2 3 in arrgt in st st st st 18 18 18 18 18 18 tubes/row no. deep 20 20 18 20 18 20 10 10 length 10 10 10 10 9.75 10.85 10.86 Uo 10.02 11.42 11.45 1.72 2.51 2.27 2.33 1.79 2.59 L1Pg 21.68 22.21 21.70 duty Q 21.59 22.22 21.72 exit gas 455 443 455 454 442 458 surface 9802 9802 8822 8407 9341 9341

duty-MMBtu/h; L1Pg - in we; surlaee-sq It; temp--F; Uo-Btu/sq It h F. The following observations may be made: 1. Staggered arrangement results in lower surface area for the same duty but higher gas pressure drop for both types of fins and fin density if the gas mass velocity is the same. For the same surface area you can transfer more energy with staggered configuration. 2. Serrated fins have a higher overall heat transfer coefficient for the same mass velocity; the surface area is lower than that of solid fins for the same duty; also, the gas pressure drop is slightly higher than that of solid fins for the same duty. Using 5 and 2 fins/in., the above design was revised to obtain a staggered arrangement with a lower pressure drop, closer to the inline configuration, for the same duty. Results are shown below. Table A-5e compares inline versus staggered designs for nearly the same gas pressure drop and same duty. It may be seen that due to the use of a lower gas mass velocity with staggered arrangement (more tubes per row), the gas pressure drop is reduced. It turns out that less surface area is required with staggered arrangement for the same duty and pressure drop. The staggered design could be marginally less in cost but there are other aspects to look into such as the effect of pitch on ligament efficiency, wider headers or longer tubes and above all, cleaning

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290

requirements. Selection of type of arrangement is mostly based on the past experience of the company. Table A-5e Comparison between staggered and inline designs for nearly same duty and pressure drop 1.

2. 3. 4. 5. 6. 7. 8. 9. 10. 11.

col no arrgt fin type fin config tubes/row no deep length

Uo ,1Pg Q surface

2 1 in st serr serr 5x .75 x .05 x .157 20 18 16 20 10 10 8.36 7.18 3.19 3.62 23.24 23.31 20524 18244

4 3 in st solid solid 2x.75x.05xO 18 20 20 16 10 11 9.75 10.02 1.72 1.42 21.68 21.71 9802 9584

FIN CONFIGURATION AND PERFORMANCE Now that we have a feel for the effect of type of fins and arrangement on the performance, the next important question is how should the fin configuration be selected? Does the tube side coefficient influence the choice? Should one select a design simply because it has more surface area than another one? Can we transfer more energy with less surface area? We will answer the above questions in the following sections. 1. Higher fin denSity or height means lower U. From Figure A3, it can be seen that for a given mass velOcity, higher the fin denSity or height, lower the gas side coefficient or effectiveness, which results in lower Uo. 2. Higher fin density or height means higher ,1Pg. Even after adjusting for the increased surface area per row, it can be shown that higher the fin density or height, higher will be the gas pressure drop for mass velocities which are close.

Appendix A - Extended Surface Heat Transfer

291

IMPORTANCE OF TUBESIDE COEFFICIENT A simple calculation may be done to show the effect of tube side coefficient on U o. It was mentioned earlier that higher the tube side coefficient, higher can be the ratio of external to internal surface area. In other words, it makes no sense to use the same fin configuration, say 5 fins/in. fin density, for a superheater as well as for an evaporator. Rewriting Equation A-I based on tube side area and neglecting other resistances: (A-27)

Using the data from Figure A-3, Ui values have been computed for different fin densities and for different hi values. These are shown in Table A-6. Also shown are the ratio of Ui values between and the 5 and 2 fins/in. designs as well as their surface area. Following conclusions can be drawn. a. As the tube side coefficient reduces, the ratio of Ui values (between 5 and 2 fins/in.) decreases. With hi = 20, Ui ratio is only 1.11. With a hi of 2000, Ui ratio is 1.74. What this means is that as hi decreases the benefit of adding more external surface becomes less attractive. We have 2.325 time the surface area but only 1.11 times the improvement in Ui. With a higher hi of 2000, the increase is better, 1.74. b. A simple estimation of tube wall temperature can tell us that higher the fin density,higher will be tube wall temperature. For the case of hi = 100: With n = 2, Ui = 39.28, gas temperature = 900°F and fluid temperature of 600F, the heat flux qi = (900-600) x 39.28 = 11784 Btu/sq It h. The temperature drop across the tube side film (hi = 100) = 11784/100 =118°F. The wall temperature =600 + 118 =718°F. With n = 5, Ui = 53.55, qi =53.55 x 300 = 16065 Btu/sq It h. Tube wall temperature = 600 + (16065/100) = 761°F. Note that comparison is for the same height. The increase in wall temperature is 43°F.

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c. The ratio of the gas pressure drop between the 5 and 2 fins/in. designs (after adjusting for the effect of Ui values and differences in surface area for the same energy transfer) increases as the tube side coefficient reduces. It is 1.6 for hi = 20 and 1.02 for hi = 2000. That is, when hi is smaller, it is prudent to use a lower fin surface .. Effect of fouling factors: The effect of inside and outside fouling factors fti and !fa are shown in Table A-7. Following observations can be made: a. With a smaller fin density, the effect of Iii is smaller. With .01 fouling and 2 fins/in., U o = 6.89 compared with 10.54 with .001 fouling. The ratio is .65. With 5 fins/in., the corresponding values are 4.01 and 7.56, the ratio being .53. That means with increased tube side fouling, it makes sense to use a lower fin density or lower ratio of external to internal surface area. The same conclusion was drawn with a lower tube side coefficient. b. The effect of ffo is less significant as it is not enhanced by the ratio of external to internal surface area.

PERFORMANCE AND SURFACE AREA Let us now study the performance of an evaporator and superheater with different fin configurations to bring out the fact that it is possible to transfer more energy with a lower surface area and at a lower gas pressure drop.HRSGs invariably use extended surfaces if the gas stream is clean as in gas turbine exhaust, air-tofluid heat transfer, and similar applications. With bare tubes, one could probably assume that the greater the surface, more energy will be transferred (for comparable velocities). With HRSGs using extended surfaces, however, one can fall into a trap by evaluating alternate designs or bids based on surface area alone. The reason is that different combinations of fin height, thickness, and density lead to different heat transfer coefficients, ffn efficiencies, and overall heat transfer coefficients. A large surface area does not necessarily mean more energy transfer. The energy

------

293

Appendix A - Extended Surface Heat Transfer

transfer capability depends on the product of the surface area and the overall heat transfer coefficient, not on surface area alone. We shall illustrate this with two examples, one for an evaporator and one for a superheater. Table A-6 , Effect of hi on Vi

[Calculations based on: 2. 0 x .105 tubes, 29 tubes/row, 6 ft long, 0.05 thick serrated fins; tubes on 4.0 in. square pitch; fin height = 0.75 in.; gas flow = 150,000 pph ; gas inlet temp = 10000 P] <-20--> <-100-> 1. hi 2 2 5 2. n, fins/in. 5 5591 5591 6366 3. G, 1b/sq ft h 6366 .01546 .00867 .01546 .00867 4. At/AI/1J hg 2.73 1.31 7.03 5. Uo 4.12 15.28 17.00 39.28 53.55 6: Ui 7. ratio 5 2.325 < <--1.11--> 8. ratio Ui <-1.363 -> 9. ratio LlFg <--1.6--> <-1.3->

<-2000-> 2 5 5591 6366 .01546 .00867 11.21 8.38 62.66 109 > <--1.74-> <--1.02-->

(surface area of 2 fins/in. tube = 2.59 sq ft/ft and for 5 fins/in. = 6.02)

Table A-7a affect of fli, tube side fouling factor (tube side coefficient = 2000) 1.Fins/in., n 2. Uo clean

3·fti 4. Uo dirty 5. Uo as %

2 11.21 .001 10.54 100

2 11.21 .01 6.89 65

5 8.38 .001 7.56 100

5 8.38 .01 4.01 53

-.!

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TableA-7b affect of ff01 outside fouling factor (tube side coefficient = 2000) 1. fins/in. 2. Uo clean 3. 110 4. U o dirty 5. U o as %

2 11.21 .001 11.08 100

2 11.21 .01 10.08 91

5 8.38 .001 8.31 100

5 8.38 .01 7.73 93

Example A-5: Evaporator Let us consider a finned evaporator for a HRSG for gas turbine exhaust. Table 8-A-a sets forth the design data.

Table A-Sa Data for HRSG evaporator design Gas flow, pph Gas inlet temperature, F Steam Pressure, psig Feedwater temperature, F Gas analysis, % by volume Carbon dioxide Water Nitrogen Oxygen Nominal gas turbine power, kW

150,000 900 235 240 3 7

75 15 4,500

We shall evaluate fin densities ranging from 2 to 6 fins per in. and fin heights of 0.5 and 0.75 in. Assumptions include 6 ft long tubes, 2 by 0.105 in., 29 tubes per row, 20 rows deep, serrated fins, 0.05 i. fin thickness, and an inline arrangement with 4 in. square pitch. The methodology of design is presented elsewhere, and the results are shown in Figure A-4.

295

Appendix A - Extended Surface Heat Transfer

STUDY OF EVAPORATOR Tube Size: 2 x .105 in , 29 tubes/row, 6 ft long, 20 deep, 4 in square pitch Fins: 2 to 6 per inch, 0.5 to 0.75 in high, .05 thick, serrated

Tabkl1-oata for HRSG evaporator design.

1lIbIe 2-E1fect of fin geometry on performance forthe same evaporator duty (see Ag. 2 also).

Duly,Q. Gas pressure drop, 4 ~.in.

Duty,MM8ll1h 18.41 18.J8 Gas pressure ckop, ill. WG U4UZ 12,633 1W2 Surface area. sq It OvenH heat transfer coef· ficient Btu pe~sq It.flr·F 8.60 9.42 Case A: 3 Hns per in. 0.15 in. high by 0.05 in.lhiclI

19 18.5 -§

~ 18

::E

Case 8: 4 fins per in., 0.50 in. high by 0.05 in. thick

~ 1

Table 3 -Effect of fin density on performance for the same flri height (evaporator).

17.5

&rIns periL 18.41 19.24 U4 3.86 12,633 23,444 4.50 85.50

3 fins peril.

17

10

16.5...1.--..:;---r----r-....,..-.....,.---'3'

Fins~rin.

Effeci of fin geometry on performance.

6

Duty, MM81ut.

Gas pressure drop, in. WG Surface area, sq It Increase in duly, percent

Increase in area, percent OvtraH heat ltansfer coef·

ficient, ntu per sq ft.flr·F

8.60

6.60

Figure A-4. Effect of fin configuration on evaporator performance. [Heating, Piping, Air-Conditioning]

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296

Figure A-4 shows the effect of fin density and fin height on energy transferred, Q; gas pressure drop, LiP; and surface area, $. As the fin density increases, so do Q, 5, and LiP. However, the rate of increase in Q decreases while that for LiP decreases. One can see that for a particular Q, the surface area and the gas pressure drop are lower for the configuration with the lower external surface area. Table A-8b shows the results for a duty of approximately 18.4 MMBtu h. The higher heat transfer coefficient associated with the lower surface/ft helps in transferring the same energy with less surface and also at a lower gas pressure drop. Hence, with finned tubes, one has to look at the product of surface area and overall heat transfer coefficient and not just the surface area. Evaluating alternates from spreadsheet giving only surface areas, which unfortunately is being done in a number of engineering organizations, may result in a poor selection that will also be more expensive to operate because of the higher gas pressure drop. (Greater fin density and fin height result in greater gas pressure drop.) Typically, a 4 in. WG additional gas pressure drop in a HRSG results in a 1 percent drop in electrical power output of the gas turbine. Greater fin density or fin height can also result in higher tube wall temperature and fin tip temperature. This will be discussed when we take up the example of the superheater. Table A-8b Effect of fin geometry on performance for the same evaporator duty (see Figure A-4 also)

Duty, MMBtu h Gas pressure drop, in. WG Surface area, sq ft Overall heat transfer coefficient Btu per sq ft-hr-F

Case A

Case B

18.41 2.24 12,633

18.38 1.82 11,432

8.60

9.42

Case A: 3 fins per in., 0.75 in. high by 0.05 in., thick Case B: 4 fins per in., 0.50 in. high by 0.05 in. thick

297

Appendix A - Extended Surface Heat Transfer

Table A-8c shows the effect of fin density on the duty and gas pressure drop for the same fin height. Because of the lower heat transfer coefficient, the additional surface area of 86 percent results in only a 4.5 percent increase in duty. Also, the gas pressure drop is much higher. Based on an electrical cost of $0.06 per kWh and a gas turbine power output of 4500 kW, the additional gas pressure drop is worth: C =(3.86-2.24) x 0.06 x 8000 x 4500/400 C =$8740 per yr The additional energy output at $3 per MMBtu h is worth: C =(19.24-18.41) x 3 x 8000 C =$19,920 per yr Based on cost and overall economics, one could arrive at either Option A or Option B. However, selecting an option because there is more surface is simplistic and can lead to wrong conclusions and improper HRSG selection. Table A-Be Effect of fin density on performance for the same fin height (evaporator)

Duty, MMBtu h Gas pressure drop, in. WG Surface area, sq ft Increase in duty, percent Increase in area, percent Overall heat transfer coefficient, BTu per sq ft-hr-F

3 fins per in.

6 fins per in.

18.41 2.24 12,633

19.24 3.86 23,444

4.50 85.50 8.60

6.60

Example A-6: Superheater Let us now consider a superheater for clean gas application. Table A-9 shows the design data. We shall evaluate fin densities of 2 and 5 fins per in., and fin heights of 0.5 and 0.75 in. We shall also

- ---- - - - - - - - - - - - -

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298

vary the number of rows deep to obtain desired duties of approximately 14 and 17.5 MMBtu h. Assumptions include lO.it long tubes, 2 by 0.120 in., 0.075 in. thick solid fins, 22 tubes per row, square pitch of 4 in., 22 streams and counterflow arrangement, and fouling factors of 0.001 hr ft 2 F /Btu on both the gas and steam sides. Table A-9 Data for HRSG superheater design Gas flow, pph Gas inlet temperature, F Gas analysis, % by volume Carbon dioxide Water Nitrogen Oxygen Steam flow, pph Entering steam temperature, F Leaving steam pressure, psig

200,000 1,200 7 12 75 6 100,000 491 600

Table A-lO shows the results of calculations for four cases, two at each of the two desired duties. The following facts can be inferred from Table A-10. 1) The energy transferred is the same for both the 2 and 5 fin per in. designs (Cases 1 and 2). 2) Because of the higher heat transfer coefficient of 11.79 versus 5.5, however, the surface area for Case 2 is nearly 2.16 times that for Case 1. This clearly shows that through poor fin geometry, one can have excessive surface area and still transfer the same amount of energy as a well designed configuration with a much lower surface area. Also, the higher fin density combination results in higher operating costs. The same conclusion is drawn by comparing Cases 3 and 4 and Cases 2 and 3. In Case 3 there is less surface area but more energy is transferred than in Case 2! 3) The tube wall and fin tip temperatures are significantly higher for the higher fin density. In Case 1, the tube wall

299

Appendix A - Extended Surface Heat Transfer

temperature is 836°F versus 908°F in Case 2 while the fin tip temperature is 949°F versus 1033°F. Hence, one could be forced into selecting better grade materials (at higher cost) for the tubes and fins by using a high fin density design, particularly in superheaters with higher steam temperature requirements. Table A-lO Effect of fin geometry on superheater performance Case 1 Case 2 Case 3 Case 4 Duty, MMBtu h Leaving steam temperature, F Gas pressure drop, in. WG Leaving gas temperature, F Fins per in. Fin height, in. Fin thickness, in. Surface area, sq ft Max tube wall temperature, F Fin tip temperature, F Overall heat transfer coefficient, Btu per sq ft-hr-F Tube side pressure drop, psi Number of rows deep Fin efficiency, percent

14.14 14.18 17.43 17.39 747 747 689 691 0.65 1.20 1.15 1.37 951 950 892 893 2 2.5 4 5 0.50 0.75 0.75 0.75 0.075 0.075 0.075 0.075 2471 5342 5077 6549 836 908 931 905 949 1033 1064 1057 11.79 9.0 6 79

5.50 6.5 4 70.8

8.04 11.0 7 63

6.23 9.0 6 68

As a general rule, the lower the tube side coefficient (as in superheaters or air heaters), the lower the ratio of external to internal surface area should be. By calculating the heat transfer coefficient based on inside tube diameter, one can easily show that a high ratio of external to internal surface area does not improve performance. Superheaters using more than 3 fins per in. do not contribute to improved duty and have to be looked at carefully. Also, they can be counter-productive, leading to higher tube wall and fin tip temperatures or gas pressure drops.

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300

The author's advice to potential users of finned surface is: Don't evaluate any finned surface based on surface area along; try to understand the effect of fin configuration on heat transfer coefficients, gas pressure drops, and metal temperatures.

SURFACE AREA OF FINNED TUBES Finned tubes are widely used in boilers, economizers, superheaters and other heat-transfer devices. Estimating their surface area involves solving the following complex equation. (Solid Fins): A =1to[4dh + 4hz + 2bd + 4bh] + l1td) (I-ob) 24 \12

(A-28)

This nomograph can be used to estimate tube surface area in ft2 / ft. It covers the most common configurations in boiler plants: Tubes of 1.25 in. O.D. to 4.0 in. O.D., fin heights of 0.5 in. to 1.0 in., and fin densities of 1 to 6 fins/in. For quick estimates, the effect of fin thickness on surface area can be neglected. For example, the surface area of a 2.0 in. tube with 3 fins/in., where each fin is 0.75 in. high and 0.102 in. thick, is 3.88 ft 2 /ft, whereas for the same configuration but with fins 0.036 in. thick, the area is 3.81 ft 2/ft. Example A-7: Estimate the surface area of a finned tube with an O.D. of 2.0 in. and a fin density of 3 fins/in., where the fins are 0.75 in. high and 0.05 in. thick. Solution: The line on the nomograph illustrates the steps. Start at n = 3 on the right horizontal axis. Draw a vertical line up to the h = O.75in. line, then a horizontal line left to the d = 2.0 in. line. From there, draw a vertical line down to the left horizontal axis to find that A = 3.85 ft 2/ ft.

301

Appendix A - Extended Surface Heat Transfer

2.0 1.75

6

1.25

3 A.1t21ft

o

6

Nomenclature A Surfat,e area flI. ~ tube, /t'/ft; (including area of liDs ad tube """""") b Fill tlUtbess, ilL Ii Tube ..... diaJIIeter, in.

., "

F!nlleicbt.isL Fill cleasity, finsIia.

~+-b

b Figure A-S. Chart for estimating surface area. [Chemical Engineering]

I

I

T

I- ' -

-

l-

-

-

l-

i-

i

~

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302

FIN WEIGHT ESTIMATION To evaluate structural and handling problems, fin weight must be known. While tables are available for figuring weight of bare tubes, tables are not readily available for finned tubes because of he several variables involved in the estimation of fin weight, namely the fin density, tube outer diameter, fin thickness, etc. With the enclosed chart, however, one can rapidly determine the fin weight for commonly used fin configurations in industrial heat transfer equipment. Of course, to determine the total weight of the tube or pipe, one must add the fin weight to the bare tube weight. The chart is based on the formula for solid fins of carbon steel given below: W = 10.68 x (D + H) (H + 0.03) NB where B = fin thickness, in. D =outside diameter of tube or pipe, in. H = fin height, in. (see drawing on chart) N = fin density, fins per in.

(A-29)

Example A-8: A 3 in. schedule 40 pipe is used in a fired heater for recovering energy from flue gases. If fin density is 4 fins per in., fin height is 0.75 in., and fin thickness is 0.06 in., determine the fin weight if fins are of carbon steel. Solution: Go up from D = 3.5 (outer diameter of a 3 in. pipe) to cut H = 0.75. Move left to cut N = 4 fins per in. and move down to cut fin thickness at 0.06 in. Move right to cut fin weight scale at 8.5 lb per ft. Multiplication factors for other fin materials are given in the table within the chart. If fins were made from 316 stainless steel, the fin weight would have been 1.024 x 8.5 =8.71b per ft. The weight of bare 3 in. schedule 40 pipe is 7.58 lb per ft, and hence the total weight of the finned pipe is 7.58 + 8.5 = 16.08 lb per

ft.

~_a.at

6

7 10

11

,,~

-II-B

='

8 9

12 13 14 15

~

i~~

h

;~

01

Muttiplh::ation factors for.

11,

17 ,,~

18 19 20

Typ. 304. 310. 316. 32! Typ. 409.410. 430. 26·1 Nickel20D

Incone1600.625

1.024 0.978 1.133 1.073

IncoloySDO

Incolo), 825 C.""ntor2o.Cb3 Hls:telloyB

1.013 1.038 1.024 1.179

Figure A-6. Chart gives weight of finned tubes. [Heating, Piping, Air-Conditioning]

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Waste Heat Boiler Dcskbook

IMPORTANT CONCLUSIONS Several important aspects of finned tubes were discussed in this Appendix: They are summarized below for convenience. 1. More surface area does not necessarily mean more energy transfer. So don't purchase a boiler or finned equipment assuming that it will do better simply because it has more surface area. The spread sheet approach to purchasing should be avoided even though the difference in surface areas could be 100 to 200 %. Fin configuration affects U significantly and one should look at the product of U x S and not at S alone. One can transfer more energy with lesser surface area as shown in the examples. 2. Lower the tube side coefficient, as in superheaters or air heaters lower should be the ratio of external to internal surface area. That is, don't use the same high density fin configuration for an evaporator as well as for the superheater. It should be much lower for the superheater. Else, the tube wall and fin tip temperature would be higher as also the gas pressure drop. 3. Staggered arrangement results in lesser surface area requirements for the same energy transfer and same pressure drop. However the impact on ligament or other mechanical considerations should be looked into. 4. Inside fouling factor affects the performance more than the .outside fouling factor; the effects are similar to a low tube side coefficient situation. The tube wall and fin tip temperatures are also influenced.

NOMENCLATURE a-factor defined in gas pressure drop equation Ao, At, Ai - obstruction area, total area, inside tube area, sq ft/ ft AI - area of fins, sq ft/ft b - fin thickness, in. C1 to C6 - factors defined in Table A-I Cp - gas specific heat, Btu/Lb F

Appendix A - Extended Surface Heat Transfer

305

d - tube OD, in. 11- gas viscosity, Lb/ft h E - fin efficiency, fraction F, FI - factors accounting for gas properties f - factor defined in gas pressure drop equation ffi, ffo - fouling factor, inside and outside, sq ft h F /Btu G - gas mass velocity, Lb/sq ft h h - fin height, in. hi - tube side heat transfer coefficient, Btu/ sq ft h F he, hg - convective gas heat transfer coefficient, Btu/sq ft h F h n - non-luminous heat transfer coefficient, Btu/ sq ft h F h 0 - total outside heat transfer coefficient, Btu/ sq ft h F k - gas thermal conductivity, Btu/ ft h F K - tube or fin metal thermal conductivity, Btu/ ft h F L - tube length, ft m - factor defined in equation (A-11) n - fin density, fins/in. N w , Nd - number of tubes wide and deep ~Pg - gas pressure drop, in. wc q - heat flux, Btu/ sq ft h; subscript i stands for inside and 0 for outside. Re - Reynolds number S - Surface area, sq ft s - fin spacing, in. St - transverse pitch of tubes, in. S1 -longitudinal pitch, in. ~ T - log-mean temperature difference, F ti, tb, tf -temperature of fluid, wall and fin tip, F tg - average gas temperature, 1 and 2 refer to inlet and exit conditions. U - overall heat transfer coefficient, Btu/ sq ft h F-subscript 0 refers to outside and i to inside. W g - gas flow, pph ws - width of serration, in. TI - effectiveness of fins 11- viscosity of gas, Lb / ft h Pg- gas density,lb/cu ft

306

Waste Heat Boiler Deskbook

REFERENCES 1. Escoa fin tube manual, published by Escoa Corp. Tulsa, 1979 2. V. Ganapathy, "Applied heat transfer," Pennwell Books, Tulsa, USA 3. V. Ganapathy, "HRSG heat transfer," Heating, Piping, Airconditioning, March 90, Pg 99 4. V. Ganapathy,"Understanding and evaluating extended surfaces," paper presented at the 12th National Industrial Energy Technology conference, June 20, 1990, Houston, Texas. 5. V. Ganapathy, "How fin configuration affects heat transfer," Chemical Engineering, March 90, Pg 147 6. V Ganapathy, "Charts simplify spiral finned tube calculations," Chemical Engineering, Apri125, 1977 7. V. Ganapathy, "Charts help evaluate finned tube alternatives," Oil and Gas Journal, Dec 3, 1979 8. V. Ganapathy, "Estimate surface area of finned tubes," Chemical Engineering, May 27,1985, Pg 156 9. V. Ganapathy, "Chart speeds up fin weight estimation," Heating, Piping, Air-conditioning, March 1988, Pg 107

)

b

AI

---------------~ .. - - - - -

-

Appendix B

Low Temperature Corrosion In Chapter 1 we discussed two main areas of concern in waste heat boiler design, namely those due to high and low temperature corrosion. High temperature corrosion concerns as discussed in Chapter 1 may be alleviated by use of proper materials and design and by keeping the tube surfaces clean so as to prevent the formation of deposits of salts responsible for corrosion. Low temperature corrosion problems may be found in boilers at the cold end, in equipment such as economizers water heaters, and air heaters. They are caused by the condensation of corrosive acids on the surfaces of the tubes or duct work which operate below the acid dew point. A few widely used methods of dealing with this problem will be addressed in this section.

CAUSES AND CURES Whenever fossil fuels containing sulfur are fired in heaters or boilers, sulfur dioxide, and to a small extent sulfur trioxide, are formed in addition to C02 and water vapor. The 503 combines with water vapor in the flue gas to form sulfuric acid and condenses on heat transfer surfaces, which could lead to corrosion and destruction of the surfaces. This condensation occurs on surfaces that are at or below the dew point of the acid gas. Also when cooled below the water vapor dew point, C02 can combine with water vapor to form carbonic acid, which though weak, can attack mild steel. While thermal efficiency of the equipment is increased with reduction in exit gas temperature (or enthalpy), lower temperatures 307

Waste Heat Boiler Deskbook

308

than the acid gas dew point are not advisable for metallic surfaces in contact with the gas. In municipal solid waste fired plants, in addition to sulfuric acid, one has to deal with hydrochloric and hydrobromic acid formation. This article deals with methods for solving cold, or back end corrosion as it is called, with the most commonly used heat recovery equipment, namely economizers or water heaters. These are used to preheat feed water entering the system (Figure B-l) and operate at low metal temperatures, thereby increasing their susceptibility to corrosion by sulfuric, hydrochloric, hydrobromic and carbonic acid.

t Economizer Deaerator

Gas

Feed water Evaporator

Figure B-1. Economizer in a heat recovery boiler system.

Estimating the dew point of these acid gases is the starting point in understanding the problem of back end corrosion. Table B-1 gives the dew points of the various acid gases as a function of their partial pressures in the flue gas. Figure B-2 gives the dew point for sulfuric acid. As an example consider a typical glue gas having the following analysis: CO = 8%, H20 = 12%, N2 = 73%,502 = 0.02%, HCL = 0.015%, 02 = 6%,HBR = 0.01%, all by volume.

309

Appendix B - Low Temperature Corrosion

170r-----------------------------~

6

160

100 7

60

8 9

u Oct

140

10

~

5

130

i

11 r{

Example = 32 ppm

12 13 14

PS03

PH20 = 10% TDP = 150°C

15

16 17

120

18

110

~

________________

~

__________

19 ~20

Figure B-2. Dew Point of sulfuric acid as a function of partial pressure of S0;3 and water vapor. [Hydrocarbon Processing]

To compute the sulfuric acid dew point, one should know the amount of 503 in the flue gases. The formation of 503 is primarily derived from two sources. 1. Reaction of 502 with atomic oxygen in the flame zone. It depends on the excess air used and the sulfur content. 2. Catalytic oxidation of 502 with the oxides of vanadium and iron, which are formed from the vanadium in the fuel oil. It is widely agreed that 1 to 5% of 502 converts to 503. Hence the % volume in our case would be 4 ppm, assuming a 2% conversion. Using these numbers and after proper conversion and substitution in the equations in Table B-1, we have: dew point of sulfuric acid = 267°P, dew point of hydrochloric acid = 128°P, dew point of hydrobromic acid = 134°P and dew point of water vapor = 121°P.

" 310

Waste Heat Boiler Deskbook

HCI, HBr, HN03and S02 correlations were derived from vapor-liquid equilibrium data. 4 The H2S04 correlation is from reference 5. Hydrobromic acid: 1,000/Top = 3.5639 - 0.1350 In (PH20) 0.0398 In (PHB,) + 0.00235 In (PH20) In (PHB,) Hydrochloric acid: 1,000/Top = 3.7368 - 0.1591 In (PH2o) 0.0326 In (PHC1) + 0.00269 In (PH20) In (PHCI) Nitric acid: 1,000/Top = 3.6614 - 0.1446 In (PH20 )0.0827 In (PHNo3) + 0.00756 In (PH20) In (PHN03) Sulfurous acid: 1,000/Top = 3.9526 - 0.1863 In (PH30) + 0.000867 In (PS02) - 0.000913 In (PH20) In (PS02) Sulfuric acid: 1 ,OOO/Top = 2.276 - 0.0294 In (PH20) 0.0858 In (PH3S04) + 0.0062 In (PH20 ) In (PH2S04) Where: Top is dew point temperature (K) and P is partial pressure (mmHg). Compared with published data, the predicted dew points are within about 6K of actual values except for H2S04 which is within about 9K. Table B-lo Dew points of acid gases.

Hence, it is apparent the limiting dew point is that due to sulfuric acid and any heat transfer surface should be above this temperature (267°F) if condensation is to be avoided. There is a misconception even among experienced engineers that the gas temperature dictates the metal temperature of surfaces such as economizers. It is not so. To explain this, an example will be worked to show the metal temperature of an economizer with two different gas temperatures. Figure B - 8 shows this calculation. It can be seen that the water side coefficient is so high that the tube wall temperature runs very close to the water temperature in spite of a large difference in gas temperatures. Thus, the tube wall temperature will be close to the water temperature and the water temperature fixes the wall temperature and hence, the dew point. Some engineers think that by increasing the flue gas temperature the economizer corrosion can be solved; not so! It should

311

Appendix B - Low Temperature Corrosion

be noted also that the maximum corrosion rate occurs at a temperature much below the dew point (Figure B-3),

Peak corrosion Q}

~ c: o 'iii

Dew point

e o o

100

I

I

110 Wall temperature, ·C

130

Figure B-3. Corrosion rate as a function of wall temperature [Hydrocarbon processing]

Steam for preheating To process

jll~[E:conomizer

Feed water

Condensate To drum

Figure B-4. Steam to water exchanger preheats feed water.

Methods of dealing with cold end corrosion:

Basically there are two approaches used by engineers to combat the problem of cold end corrosion:

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312

Waste Heat Boiler Deskbook

A. Avoid it by using protective measures such as maintaining a high cold end temperature so that condensation of any vapor does not occur. B. Permit condensation of acid vapor or both acid and water vapor, thereby increasing the duty of the heat transfer surface, and use corrosion resistant materials such as glass, etc.

Methods of avoiding cold end corrosion: 1. Maintain a reasonably high feed water inlet temperature. If the computed dew point is say 250°F, a feed water of 250°F should keep the minimum tube wall temperature above the dew point. With finned heat transfer surfaces, the wall temperature will be slightly higher than with bare tubes. The simplest way would be to operate the deaerator at a slightly higher pressure, if the feed water enters the economizer from a deaerator (Figure B-l). At 5 psig the saturation is 228°F and at 10 psig it is 240°F. 2. In case the deaerator pressure cannot be raised, a heat exchanger may be used ahead of the economizer (Figure B-4) to increase the feed water temperature. It may be steam or water heated. 3. Figure B-5 shows a method for using an exchanger to pre-heat the water. The same amount of water from the economizer exit preheats the incoming water. By controlling the flow of the hotter water, one can adjust the water temperature to the economizer so that a balance between corrosion criterion and efficiency of operation can be maintained. 4. Hot water from either the economizer exit or the steam drum (Figure B-6) can be recirculated and mixed with the incoming water. The economizer has to handle a higher flow, but the exchanger is eliminated and a pump is added. Note that some engineers have the misconception that bypassing a portion of the economizer (Figure B-7) would solve the problemi not so. While bypassing, the heat transfer surface reduces the duty on the economizer and increases the exit gas temperaturei it does not help to increase the wall temperature of the tubes, which is the most important variable. A higher exit gas temperature probably helps the down stream ductwork and equipment, but not the

-----------------.----------

-----------~-~~-~~

313

Appendix B - Low Temperature Corrosion

economizer. One benefit, however, from bypassing is that steaming possibilities in the economizer are minimized.

Economizer

il!IC

Feed water Water to water heat exchanger

t

Gas

To drum

Figure B-S. Water to water exchanger preheats feed water.

Feed water Economizer

To drum

Recirculating pump

Figure B-6. Recirculation pump mixes hot water with feed water.

Economizer 1

Feed water

Feed water

Economizer 2

t

t

Gas

Economizer 2

TO

drum

A

Gas

To drum B

Figure B-7. Bypass arrangement for economizer. In "B" eco 1 is by-passed. This increases exit gas temperature and avoids steaming but does not solve dew point corrosion in eco 2.

314

Waste Heat Boiler Deskbook

Permitting condensation on surfaces: By using proper materials one can protect the heating surfaces from corrosion attack,if condensation is likely. This concept has now been extended to recovering the sensible and latent heat from the flue gases, thereby increasing the thermal efficiency of the system by several percentage points in what are called condensing heat exchangers. If flue gases contain say 10% by volume water vapor, by condensing even half of it, approximately 30 Btu/lb of flue gas can be recovered. This is nearly equivalent to a 120°F drop in gas temperature if sensible heat alone is transferred. A large amount of sensible and latent heat in the flue gas can be recovered if the gas is cooled below the water dew point. This implies that sulfuric acid, if present in the gas stream, will condense on the heat transfer surfaces as its dew point is much higher than that of water vapor. Borosilicate glass and teflon coated tubes have been widely used as heat transfer surfaces for this service. Glass is suitable for low pressures and temperatures (less than 450°F and 20 to 100 psig). However, presence of fluorides and alkalis is harmful to the glass tubes. One manufacturer of condensing heat exchangers uses teflon coated tubes. A thin film (about 0.015 in.) is extruded onto carbon or alloy steel tubes, and the surface is resistant to corrosion of sulfuric acid. Finned tubes can not be used as teflon cannot be extruded onto these surfaces. Hence, these exchangers will be larger than those with extended surfaces, however, the higher heat transfer rates with condensation process, improves the overall heat transfer coefficients and partly compensates for the lower surface area per linear foot of bare tubes. The high initial investment associated with condensing heat exchangers has to be carefully reviewed along with the energy recovered, fuel costs, etc. If the fuel cost is not high, then the payback period for this type of equipment may be long. Materials such as cast iron and stainless steels probably have better corrosion resistance than carbon steel, but still they are not corrosion proof. It is also felt by some that the higher thickness of cast iron is responsible for the longer life.

Appendix B - Low Temperature Corrosion.

315

The above material outline the importance of the dew point of acid gas and methods for dealing with the problem of condensation on heating surfaces such as economizers. Similar methods could be used for air heaters. The basic difference lies in the fact that the back end temperature is a function of both the gas and air temperatures. Steam air heating or air ~ypassing have been used 0 combat the problem of corrosion. Replaceable matrices and corrosion resistant materials such as enamels have been used at the cold end of regenerative air heaters.

The average wall temperature of a bare tube economizer is given by the simple equation:

tw = 0.5[tl + tg - U(tg - tl) (lIh o - 1!hl») Where:

hi = heat transfer coefficient inside tubes, Btu! ft 2 h of ho = heat transfer coefficient outside tubes, Btu! ft 2 h of tl = temperature of water inside tubes, of tg = temperature of gas outside tubes, of tw = average tube wall temperature, of U = overall heat transfer coefficient, Btulft2h °F

1/u = 1!hl + lIho, neglecting fouling and metal resistance, which are much smaller. Typically hi

= 1,000, ho = 15 and hence U = 14.77

Case 1: Determine tw when tg = 750°F and tl = 250°F tw = 0.5 [250 + 750 - 14.77 (750 - 250) (0.066 - 0.001») = 260°F Case 2: tg = 350°F, ti = 250°F tw = 0.5 [250 + 350 - 14.77 (350 - 250) (0.066 - 0.001») = 252°F Thus, for a variation of 400° F gas temperature, the tube wall temperature hardly changes by 8°F. Thus, the water temperature fixes the tube wall temperature.

Figure B-8. Determining tube wall temperature of economizer. [Hydrocarbon Processing]

CORROSION IN STACKS, DUCTS If a boiler stack or ductwork leading to the stack is uninsulated

then the average wall temperature will be lower than the gas

Waste Heat Boiler Deskbook

316

temperature due to the loss of heat from the casing, which depends upon the ambient conditions and the wind velocity. At lower loads the gas exit temperature from boilers will be lower and hence the problem becomes more acute. By adding external insulation,the heat loss can be minimized and the stack wall temperature can be maintained close to the gas temperature. The following methodology illustrates the procedure to compute the wall temperature of stack or duct. Stacks, ducts and scrubbers are designed so that the inside'wall temperature always remains above the dewpoint of any acid vapors passing through these units. This is because the condensed acids will attack many materials of construction, which would require the equipment to be made from costly alloys, ceramics or plastics. Designing equipment to avoid condensation requires estimating the inside-wall and stack temperatures (the latter being the temperature of the gas exiting the stack), and any heat losses through these walls, whether they are insulated or not. Generally, engineers use rule-of-thumb estimates for these temperatures. However, such guesses are sometimes inaccurate and as a result, structures are built that suffer needless acid corrosion. Here is a simple method to calculate these temperatures, as well as the heat losses, based partly on some equations previously developed by the author. Deriving the equations: The figure shows a typical temperature profile in a wall that is insulated. The heat loss is given by:

The temperature drop across the gas film is: (B-2)

where the gas-film heat-transfer coefficient is given by h = 2.44

WO· 8 e/di l .8

(B-3)

Appendix B - Low Temperature Corrosion

317

Insulation

,

,

, I

Stack or duct wall Figure 8-9. Temperature profile in ducts, stacks.

where: C = (Cp/I1)OA/k o.6

(B-4)

The duct-wall temperature drop is given by: twi - two

= Qdoln(do/di/24Km

(B-5)

The temperature drop across the insulation is: (B-6)

two - tc = QL/Ki

The effect of curvature [the heat-transfer surfaces are curved, since they are part of a cylindrical shape] is neglected, since ducts, stacks another such structures are generally large in diameter. The total heat loss from the duct (or stack or other such structure) is:

Qz =nd o HQ/12

= 3.14 doHQ/12

(B-7)

The temperature of the exiting gas is: tg2 = tgl -.Qe/WCp

(B-8)

318

Waste Heat Boiler Deskbook

Solution method: To solve the above equations: 1. Assume a gas-exit temperature, tg2 and calculate the average gas temperature, tg = 0.5(tgl + tg2) (B-9) 2. Assume an insulation-casing temperature, te. 3. Calculate Q, the heat loss, using Equation (B-V. 4. Calculate h, the film coefficient, and the various temperature drops, using Equations (B-3), (B-2) and (B-4) - (B-6). 5. Assume a value of the casing temperature, te. 6. Set tolerance for the difference between the assumed and calculated values of te. A difference of between 1-2°P will usually be accurate enough. If the tolerance is exceeded, repeat Steps 2-5. 7. Calculate the total heat loss from Equation (B-7) and the exit-gas temperature, tg2, via Equation (B-8) 8. If the calculated and assumed values of tg2 are not within the tolerance, repeat the process, starting from Step 1; otherwise, the calculations are finished. Although this method may seem tedious, the calculations are easily done on a scientific calculator, and the assumed and calculated values quickly converge to within a reasonably accuracy. A few trials are all that is usually needed. A computer program can easily be created to perform the calculations; the author will supply his version upon request. To illustrate the method, here is an example: Example 1; Insulated stack: Flue gases at 110,000 lb/h and 423°P enter a 48-in.-I.D. stack, 50 ft long and 1 in. thick. If the ambient temperature is 67°P and the wind velocity is 125ft/min, determine the wall and casing temperatures if the stack is covered with 2 in. of mineral-fiber insulation. Por flue gases at 420 o P, Cp = 0.266 Btu/(lb)(OF), Jl = 0.058 Ib/(ft)(h), and k = 0.023 Btu/(ft)j(h)/( OF). The gas temperature drop in the insulated stack/ducts is relatively low, typically 2-5°P. Thus, assume that the exit-gas temperature is 420o P: 1. Assume that the casing temperature is 90 o P. Then: Q =0.173 x 0.9(5.5 4 - 5.27 4) + 0.296(550-527)1.25[(125 + 69)/6910.5 =47.3

-

Appendix B - Low Temperature Corrosion

- - - - - - - - - --

------

- - -_ _ _ _ _ _ _

__

~

319

Btu/(ft2)(h), and h= 2.44 x 110,000°.8(0.266/.058)°.40.023°.6/481.8 = 4.74 Btu/(ft2)(h)(OF). 2. The thermal conductivity of the fiber (from manufacture's tables) is 0.30 and 0.42 Btu-in./(ft2 )/(h)/( oF) at 200 and 400 o P, respectively. At 250 oP, the value is 0.33. 3. The temperature drop across the insulation = 47.3 x 2/0.33 = 287°P. The average insulation temperature is 232°P. Por the next trial, the thermal conductivity has to be estimated at this temperature. 4.Stack-wall temperature drop = 47.3 x 50 In(50/48)/24 x 25 = 0.16°F. Note: A thermal conductivity of 25 Btu/(ft)(h)(OF) is used for carbon steel. 5. Gas film drop = 47.3 x (50/48)/4.74 = 1O.4°F. 6. Thus, the corrected casing temperature = 422-10.4-0.16-287 = 124°F. Since 124°P is higher than the assumed value of 90o P, another iteration must be tried. It can be shown that assuming a value of 92°P gives good results. Using this value Q = 52.5 Btuj(ft2)(h) 7.The total heat loss = 3.14 x 50 x 52.5 x 50/12 = 34,300 Btu/h. 8. The gas temperature-drop in the duct = 34,300/110,000/0.266 = 1.2 of. The final results: Temperature drops: gas-film = 11.5 of, duetwall = 0.2 OF, and insulation = 31JOF. Example B-2: Let us now check the stack wall temperature if no insulation is used. Solution: 1. Let the gas temperature drop = 20 0 Pi as the heat loss is higher, the gas temperature drop will also be higher. The average gas temperature is = 413°P. 2. Let the casing temperature = 250o P. This will be checked and corrected later. 3. Heat loss from the casing is: Q = .173 x .9 x [7.1 L 5.27 4J + .296 x (710-527)1.25 x [(125 + 69)/69Jo.5 =610 J3tu/sq ft h. 4. Temperature drop across gas film = 610/4.74 = 129°F. 5. Temperature drop across the stack wall = .0034 x 610 = 2 OF. 6. Hence two = 413-129-2 = 282°F versus 250°F assumed. Hence another round of iteration is required.

7~~~~

_ _...(

320

Waste Heat Boiler Deskbook

It may be shown that a wall temperature of 266°F balances the equations. Heat loss =679 Btu/ sq ft h. The gas temperature drop =3.14 x 50 x 679 X 50/(110000 X .265) = 15°F. This agrees with what we assumed in step 1 and hence no further iterations are necessary. Since insulation calculations particularly those with multiple layers involve iterative calculations, the author has developed a program to solve for the heat loss and the temperature profile across multi-layer insulation. This is presented in Figure B-10, followed by two examples.

HEAT -LOSS CALCULATIONS THROUGH MULTI-LAYERED INSULATION The heat loss from the casing of any surface can be found from

Q = 0.173iTa4) + 0.296 (ts_ta)1.25 [(V + 69)169Jo.5 This equation is obviously difficult to handle manually, but the Basic program presented here provides a quick calculation of heat loss from refractory or insulation-lined pipes or flat surfaces. It also calculates intermediate temperatures and casing outer temperature. The program can handle any number of layers, with the heat loss between layers being given by Q

= t1T/L

(B-l0)

For flat surfaces, L is the insulation or refractory thickness. For cylindrical surfaces, L = 0.5Do In(Do/Di), where Do= outer diameter of the layer and Di = inner diameter of the layer. Thermal conductivity K for each layer is calculated at its mean temperature. K is input at two given temperatures, and the program interpolates for the actual mean K using a linear relationship. Quick converging logic is used to compute the final results. A trial value of casing temperature is assumed, and intermediate temperature and heat loss are calculated. The intermediate

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---------

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Appendix B - Low Temperature Corrosion

321

CLS:KEY OFF:CLEAR DIM A(64) ,A$(IS) ,L(IS) L=O:DO=O 10 PRINT"INSULATION PERFORMANCE-PROGRAM ••• FLAT SURFACES OR PIPES":PRINT" "

20 PRINT"This program can handle any number of layers-information on conductivit y of each layer at any two temeperatures is needed":PRINT" " 30 INPUT"NAlIE OF PROJECT-OATE(up to 30 characters)~";PROJ$:PRINT" " 40 INPUT"pipe outer dia in inch-for flat surface input 0="/00: IF 00=0 THEN B$=" flat surface"ELSE B$=" pipe" 50 00(1)=00 60 PRINT" " 70 INPUT"number of layers,hot face temp, ambient temp, wind vel in fpm,surf emissa: ";P(A(1) ,H,V,J:HF=A(l) 80 PRINT" " 90 PRINT"Input Insulation data starting from the hot surface-thermal conduotivit y in BTU in/ft2 h":PRINT" " 100 FOR 1=1 TO P 110 INPUT"Name(up to 10 characters) ,thickness-in=" ;A$,L 120 A$(I)=A$:L(I-l)=L 130 PRINT" " 140 INPUT"templ,cond l,temp2,cond 2~"/M(I) ,K(I) ,N(I) ,0(1) 150 IF 00=0 THEN A(lS+I)=L :GOTO 170 160 A(lS+I)=.S*LOG( (DO+2*L) /00) :00=D0+2*L 170 A(4S+I)=(K(I)-Q(I» / (M(I)-N(I» :A(30+I)=K(I)-A(4S+I)*M(I) 180 PRINT" ":NEXT I 190 A(P+l)=H+200:U=A(1) 200 Z=A(P+l)-H 210 X=(U-A(P+l) )/p 220 W=.01*(A(P+l)+460) :Y~.01*(H+460) 230 Q=.173*J* (WA4-Y A4)+93.6*(W-Y)A1.2S*SQR( (V+69) /69) 240 XX-.Ol* (460+HF) :QMAX-.173*. 9* (XX A4_Y A4 )+93.6* (XX-Y) A1. 2S*SQR( (V+69) /69) 250 FOR 1=1 TO P 260 A(P+l-I)=A(P+2-I)+X 270 T=.S* (A(P+I-I)+A(P+2-I» :G=A(3l+P-I)+A(46+P-I)*T 280 IF 00=0 THEN FA=l ELSE FA=OO 290 300 300 310 320 330 340 350 360 370 380 390 400 410

R=A(P+2-I)+Q*FA*A(16+P-I) /G IF ABS(R-A(P+I-I»<3 GOTO 320 IF ABS(R-A(P+I-I) )<3 GOTO 320 A(P+l-I)=.S*(R+A(P+I-I»:GOTO 270 NEXT I IF ABS(A(1)-U)<3 GOTO 360 Z=.S*Z:A(P+l)=A(P+l)+SGN(U-A(l) )*Z.GOTO 210 PRINT" " CLS PRINT STRING$(80,20S) PRINT" RESULTS-INSULATION PERFORMANCE- ";B$:PRINT" " PRINT" " PRINT" Project. ";PROJ$ PRINT"--:-"-------'

420 PRINT"Name thick-in Temp-F TEMPl CONDl TEMP2 COND 2". PRINT" " 430 A$ (P+l)~"Casing" 440 FOR 1=1 TO P+l 4S0 PRINT A$ (P+2-I) ,TJSING"######.##" ;L(P+I-I) ;A(P+2-I) /M(P+2-I) /K(P+2-I) ;N(P+2-I ) ;0(P+2-I) 460 NEXT I 470 PRINT" " 480 PRINT"Heat loss -BTU/ft2h·";Q/"Number of layers of insulation=";P 490 PRINT" "

500 PRINT"Amb temp=" ;H; "Wind Vel-fpm=" IV; "Ernias=" ;Jl "Max Loss-BTU/ft2h=" ;QMAX: PR INT" " 510 IF 00(1»0 THEN PRINT"Pipe outer dia-in =" ;DO(l) 520 END

Figure B-10. Listing of program for multi-layered insulation design. [Machine Design]

322

Waste Heat Boiler Deskbook

temperatures are corrected during each iteration, and the final casing temperature is calculated and compared with the assumed temperature. If the two values do not agree, iteration continues. Example B-3: A pipe with an outer diameter of 6.625 in., is layered with 3-in. thick mineral fiber insulation. At 200°F, K = 0.3 Btu-in./ftLh-OF; at 400°F, K = 0.42 Btu-in./ftLh-oF. The pipe hot face is at 650°F, ambient temperature is 70°F, and wind velocity is 150 fpm. Find the outer casing temperature and heat loss per unit surface area, assuming a casing emissivity of 0.9. Figure B-11lists the results of the iteration. Casing temperature is 96°F, and heat loss is 55 Btu/ft2_h. Example B-4: A heater has two layers of refractory, each 4-in. thick. Kast 30 is facing the hot gases and has a K of 3.31 Btu-in./ft2 /hF at 1200°F and 3.11 at 800°F. The second layer is castable block mix with a K of 0.42 Btu-in./ft2-h-F at 400°F and 0.61 Btu-in./ft2-h-F at 800°F. The hot face temperature is 1,800°F, ambient temperature is 80°F, wind velocity is 0 fpm, and casing emissivity is 0.9. Find the heat loss and casing intermediate temperature. For this calculation, pipe diameter is 0, and K values are input in order, starting from the refractory closest to the hot surface. Figure B-11 lists the results: Heat loss is 220 Btu/ft2_h and intermediate temperature is 1,551°F.

HOT CASING DESIGN The casing for boilers could be of two types; membrane wall, Figure B-12, which is fully water cooled and hence maintained at the saturation temperature of steam. Since saturation steam temperature at say 200 psig is usually higher than the dew point of most corrosive gas, low temperature corrosion is not a concern with membrane wall units. The boiler pressure could also be raised if required to alleviate any concerns of low tube wall temperature.

--~-~-------------------

t..

--------

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--~~-~~~~~

323

Appendix B - Low Temperature Corrosion

RESULTS-INSULATION PERFORHANCE-

_ _ _ _ _ _ _Project: Name Casing

min fb

thick-in 0.00 3.00

pipe

example

Temp-F 95.39 651.77

CONnl

TEMPl

TEMP2

0.00' 0.30

0.00 200.00

CONO 2 0.00 0.42

0.00 400.00

Heat loss -BTU/ft2h= S5. 34366 Number of layers of insulation"'" 1 1unb temp= 70 Wind Vel-fpm= 150 Emiss" .9 Max Loss-BTU/ft2h"", 3741.703

Pipe outer dia-in = 6.625

RESULTS-INSULATION PERFORHANCE-

Project: Name

thick-in

Casing CBM KAST30

0.00 4.00 4.00

flat surface

EXAMPLE

Temp-F

TEMPl

CONOl

179.22 1551.40 1801.63

0.00 400.00 1200.00

0.00 0.42 3.31

TEMP2

CONO 2

0.00 BOO.OO 800. 00

0.00 0.61 3.11

Heat loss -BTU/ft2h= 220.2423 Number of layers of insulation= 2

Figure B-l1. Results from program for Examples 3 and 4.

-------------_._------._.__..

_._---

gus flow

.O. -.--(/~. - ·

---.. . .\-__./----)__~.r\ ". ,r-: :-( ) . \ (I

/'0

00 " -

D

(I

......

0>

00 (I III I> (I ~ (I I) .... >(I _ . _ _ •• _. _ _ _ _ _" _ _ _ _ _ _ _ _ _ _ _ _

~,,)

~

d
I)

(I

(l

o!:I-

(to

- - -_ _ _ _ _ _0__. _ -

Figure B-12. Membrane wall casing.

The other type of casing is the one lined internally with refractory, Figure B-13. In this design the casing is protected from the hot gas stream by two or more layers of refractory. The casing temperature could run from 140 to 250°F depending on the gas temperature, ambient conditions and the thickness and combination of refractories used, as seen in example B-3. One concern with this design is that there is a possibility of corrosive gases seeping through the refractory and attacking the

....

324

Waste Heat Boiler Deskbook

casing, portions of which could be blow the acid dew point of the corrosive gases. The casing is usually painted with a corrosion inhibiting paint or coating; however if the casing temperature were raised by adding external fiber insulation, dew point corrosion could be eliminated. Some engineers however do not prefer this solution as they feel that the casing is hidden from the observer and signs of hot spots due to the possibility of the refractory falling off cannot be spotted. Adding a layer of miner fiber insulation externally increases the casing temperature; by using appropriate thickness of mineral fiber and a combination of internal refractory, one can maintain the casing at say 300 to 450°F if required. The outermost casing over the mineral fiber however will be at a low temperature thus minimizing the heat losses to the atmosphere.

._-----------------

9a.5 flow

insula. tlon Figure B-13. Refractory lined casing.

Example B-5: A boiler casing has two layers of refractory namely 4 in. of KS4(AP Green) and 2 in. of Castable Block Mix (CBM). Determine at average gas temperatures of 1000°F and 600°F, the external insulation to be used to maintain a casing temperature of 300 to 350°F. Ambient temperature = 60°F and wind velocity = 100 fpm. thermal conductivity data are shown in Figure B-14, which also gives the results.

325

Appendix B - Low Temperature Corrosion

RESULTS-INSULATION PERFORMANCEProject: NAME Casing MINFB CBM KS4

THICK-IN

flat surface

EXAMPLE

TEMP-F 85.00 327.30 550.49 601.49

0.00 1.50 2.00 4.00

TEMPI 0.00 200.00 400.00 1200.00

CONDI

TEMP2

0.00 0.30 0.42 4.29

0.00 400.00 800.00 800.00

COND2 0.00 0.42 0.61 4.12

HEAT LOSS -BTU/ft2h= 49.4179 Number of layers of insulation= 3 AMB TEMP= 60 WIND VEL-fpm= 100 EMISS= .9 MAX LOSS-BTU/FT2H= 3057.666

eeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeee eeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeee RESULTS-INSULATION PERFORMANCEProject: NAME Casing MINFB CBM KS4

THICK-IN 0.00 0.50 2.00 4.00

flat surface

EXAMPLE

TEMP-F 11B.98 337.30 865.84 999.36

TEMPI

COND1

TEMP2

0.00 200.00 400.00 1200.00

0.00 0.30 0.42 4.29

0.00 400.00 800.00 800.00

'COND2 0.00 0.42 0.61 4.12

HEAT LOSS -BTU/ft2h= 136.846 Number of layers of insulation= 3 AHB TEMP= 60 WIND VEL-fpm= 100 EMISS= .9 MAX LOSS-BTU/FT2H= 3057.666

eeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeeee Figure B-14. Results from program for Example 5.

Solution: As discussed earlier a trial and error procedure is required to solve for the heat loss and casing temperature. The program shown in Figure B-lO was used. Figure B-14 gives the results for the various cases. It may be seen that at the boiler exit where the gas temperature is lower, a higher thickness of insulation is required, 1.5 in. versus .5 in. at the middle portion of the boiler where the average gas temperature is lOOO°F. The analysis could be carried out at

326

Waste Heat Boiler Deskbook

various portions along the boiler length and appropriate external insulation could be selected. Also note the low casing heat loss to the atmosphere.

NOMENCLATURE Cp C d

- gas specific heat, Btu/lb F - factor used in Equation (B-4) - duct diameter, in.; subscript i and 0 refer to inside and outside e,E - emissivity of casing h - gas heat transfer coefficient, Btu/ sq ft h F H - length of duct or stack, ft k - gas thermal conductivity, Btu/ft h F Km - metal thermal conductivity, Btu/ft h F K - insulation conductivity, Btu in./sq ft h F L - thickness of insulation, in. Q - heat loss, Btu/ sq ft h Qz - total heat loss, Btu/h tc,Tc - temperature of casing, F and R ta,Ta- ambient temperature, F and R tg - average gas temperature, F; 1 and 2 refer to inlet and exit V - gas velocity, fpm

REFERENCES 1. V. Ganapathy, "Cold end corrosion: causes and cures," Hydrocarbon Processing, Jan 89 2. V. Ganapathy, "Preventing corrosion in stacks and scrubbers," Chemical engineering, Jan 89 3. V. Ganapathy, "Basic programs for steam plant engineers," Marcel Dekker, New York, 1984 4. Kiang, Yen Hsiung, "Predicting dewpoints of acid gases," Chemical engineering, Feb 9, 1981, Pg 127 V. Ganapathy, "Applied heat transfer," Pennwell books, 5. Tulsa, 1982 6. V. Ganapathy, "Simplified heat loss calculations," Machine Design, Sep 8, 1988.

Appendix C

Heat Transfer Equipment Vibration

Tube bundles in heat exchangers, boilers, superheaters and heaters are often subject to vibration and noise problems. Vibration can lead to tube thinning and wear, resulting in tube failures. Excessive noise can be a problem to plant operating personnel. Large gas pressure drop across the equipment is also a side effect, which results in large operating costs. 1,2 With the design checks presented here, one can predict during design if problems associated with noise and vibration are likely to occur. Vibration causes. Vibration and noise problems are caused when air or flue gases flow over tube bundles, which may be arranged inline or staggered (Figure C-l). Vortices are formed and shed beyond the wake of the tubes, resulting in harmonically varying forces on the tubes perpendicular to the flow direction. It is a self-excited vibration. If the frequency of vibration of the Von-Karman vortices, as they are called, coincide with the natural frequency of vibration of the tube bank, resonance occurs which leads to tube vibration. Another phenomenon that occurs with vortex shedding is acoustic vibration, leading to noise and high gas pressure drop. The duct or the bundle enclosure vibrates when the acoustic oscillation frequency coincides with the vortex shedding frequency.6 The acoustic oscillation is normal to both the direction of gas flow and tube length. 327

Waste Heat Boiler Deskbook

328

To process

rv I

I

Gas In lIP

F 'om economizer or deaeralor

...

D

U

Gas oul

I

Boller-waler-Iube design

Gas in

I

J

-

Air out

l

-Air in

,

Gas oul

Air healer

J Il

, Gas oul Superhealer

Figure Col. Cross flow of gas over tube bundles.

329

Appendix C - Heat Transfer Equipment Vibration

Design methods to check vibration and noise. The first step in the analysis for possible vibration or noise is the estimation of the vortex shedding frequency, Ie. Vortex shedding is prevalent in the Reynolds number range of 300 to 100,000, which is the operating range of many boilers, heaters and exchangers.The vortex shedding frequency may be estimated once the Strouhl number, 5, is known1,2 which is given by the expression:

5=

Ie

d/(12V)

(C-1)

Here d is the tube outer diameter, V is the average gas velocity and 5 is a function of tube geometry. Figures (C -2 to C-5) give typical values of 5. The natural frequency of vibration of the tubes is then determined. For a uniform beam supported at each end, In is given by the expression8: (C-2)

C is a constant depending on end conditions and is given in Table C-l. The tube length in feet is 1 and Me is the total weight of the tube, which includes the contribution of the fluid weight inside and outside the tubes. For carbon steel tubes he above equation may be simplified and written as 8: (C-3)

The next step is estimation of acoustic frequency, I a

la = V s/).

(C-4)

Vs is the sonic velocity of the gas and), is the wave length.

). =2w/n where w is the width of the duct in feet and n is the mode of vibration. For air or flue gases, Vs is approximately 49 vT where Tis the gas temperature in degrees, R. For a cylindrical duct

la =N

Vs/D

(C-5)

Waste Heat Boiler Deskbook

330

N is a constant = 0.5681 for mode 1, 0.9722 for mode 2 and 1.337 for mode 3. Checks and analysis for vibration and noise. To analyze for possible vibration or noise in the tube bundles caused by flow of gases across tube banks, the following calculations are performed: 1. Calculate In for different modes and load conditions. Compute Ie. If Ie and In are within 20% of each other, vibration is almost certain to occur. 2. Estimate Ia at different loads. Compare I a with Ie. If they are within 20% of each other, excessive noise is likely. The first mode of vibration is the most critical one as the amplitude of vibrations is large.

Table C-l Values ofC Mode of vibration End support conditions Both ends clamped One clamped, one hinged Both hinged

1 22.37 15.42 9.87

2 61.67 49.97 39.48

3 120.9 104.2 88.8

2 91 54 112.2 224 336

3 179 54 168.3 336.6 504

Table C-2 Summary of results

n

fn fe fa fa fa

(no baffles) (1 baffle)

(2 baffles)

1 33.1 54 56.1 112.2 168.3

Ed

331

Appendix C - Heat Transfer Equipment Vibration

0.5

_ -

_

-1-V.J.-I-'-I-+t--t-lH-+-t-Hrl--+--I-"+-I-I-H-.. H

...-

Sl/d=1.251- - -i-H-+-t-t--,I--\-+--I-I-H-+-+-H

I--

=- =: ~ .~. ~~ ~ :~ =:= ~: .... _

1_1_

....... - - -1-1- -

1.5 !-

~-_++-I-+-+-+r·-H_. '-+-+1_-+1---1

- .. '''1/ _....... - ..... - _. - ." .. --1-.. +_. ·-1- -!-+"1I--1---I-4I-+-1 =1=11 ~ =- =- =. =- =- =- -::1= - 1:= ~'li. -::1---{"--"HI--+I--lII-_+--I--+"-~~~I~l""~C::!

0.1. -

w

L....

+-I-+-+-'!- I-

1:'+i"--I--I--+--I-+-t-t-+-I-t-+-1

. . _- _./

~ =~

Z

.

=V' . -- . . ;: -.. . . . . -:IlL _.... =- 3=" =I= . . ~. =.; ~.~ -~I-- +-~-J.-I-H+H+-I-+-H ..+"...j.'..

v~ 2.0

- _ . f - ....I-I--+-I--I-,I--+-\A-.. t--l:-r-..-+-I-H-++-t-i

-I

Z

-11--I-H·-+-"hl/~I-···+-I+·.j-H-+d4-++-I-t-+I--.-+--H·-1--1-1/

0.3

. --.. . It -1-'

............ -

~

I/"

-1-

- -II/hf-il-I-I-I-II-I--I-I-H-H-f-f-·-!-

- _..

1/

. - I =~

-'1-1-1-f·--+-Hf---I--h..f44-i1-+-I-+-I---I-+-++--I--I _ .... -I-1-17f- - -f-+..-iI-l--hlf4--I--I--I---I-H--I-H+H-+--I

-I

« :I:

iH-~R+~"~~~/'44+++-+-~

::>

o ....a::

~-

0.2-

1/

VI

.1.1

1/

. i/ ~ .....

1/

17' . . _. - . - -

=. := -

-~I7~·-l-""'.. -I-+~-I-+-I-+-++-I4-H

J . -It ~ ~ -HI--I--I-l---I-I-~-I-I- --+++--1-+--1 1/. -) -1-" i/I,.t' _. -- 1- -HH--t-L-L-L..1.....L-..1...l-./.... 1--

.r

FLOW

0.1-

-I/. . f- tzlt. -1-1-1-+-+++--1-+-+--1--1 ~ 1

tr

.. ~.. -;

J

·11 i

-I--

t'~t]Sl =~

-I--

---'1

·w

.. 1-

I .. .

f..-I--

-I-

d-I-

ST-t

"-1----I-

....

··1-- .......... -'''1-1- .

Hf-i-+-t---/--t-t---/- - ·.. ·1- ....-t-I ..+--t-+--lH---I-H·-t-..H--I-·-f-H

-I--II ..........+-I--I-..... -~-I--

I--I·-t- -1-+-'- -

-t-~-++-1-+-+-I

o,0 +..L....J......1....~-'-L....1....J-.J-1-...L..~...L....JL-J......L...JL-..J-..J.....J-L..L-I.....I-.L.J.-1-I 2

3

Figure C..2. Strouhl number from inline bank of tubes.. Chen. [Hydrocarbon Processing]

.

332

Waste Heat Boiler Deskbook

Eliminating noise and vibration problems. By changing the tube span, tube pitch, or end conditions, the natural frequency may be a1t~red keepingjn and fe apart to avoid vibration problems. Gas velocity can also be changed so that fe is altered. This may be done by changing the tube length and number of tubes wide. Primary correction devices for noise are baffles. Baffles divide the gas column into smaller channels or ducts and thereby increase the acoustic frequency, moving it away from the vortex shedding frequency. If the gas temperature is high, the materials for baffles must be chosen with care. Acoustic vibrations usually lie in the range of 40 to 100Hz. Example problem. A tubular air heater 1l.7ft wide, 17.5 ft deep and 10 ft long is used in a plant. Carbon steel tubes of 2 in. OD and 0.08 in. thick are arranged inline with a transverse and longitudinal pitch of 3.5 in. The bundle is 40 tubes wide and 60 tubes deep. Air flows over the tubes, while flue gas flows inside. Air flow is 300,000 lb/h at an average temperature of 260°F. The tubes are fixed at each end in tube sheets. Analyze the bundle for possible noise and vibration problems. Solution: Estimate fe. For st/d = slId = 3.5/2 = 1.75, from Figure C-3, S = 0.3. From Figure C-5 we see that S = 0.31. Calculate the air velocity, V. Air density = 0.081(492)/(460 + 260) = 0.55 lb/ft.3 V = 300,000(12)1[3600(0.055)40(3.5 - 2) 10] = 30 ft/s. Hence fe = 12SV/d = 12(30)0.30/2 = 54 Hz. Estimate fn using Eq~ation (C-3).l = 10, d = 2, di = 1.84, Mt = 1.67 lb/ft = Me, (neglecting weight of air/gas). For the first three modes, Cl = 22.37, C2 = 61.67 and C3 = 120.9, from Table C-l. Then, fnl = 33.1, fn2 = 91 and fn3 = 179 Hz, using Equation (C-3). Let us compute the acoustic frequencies, fa. Sonic velocity, Vs = 49(460 + 260)°·5 =1.315 ft/s. Width, W = 11.7 ft and A. = 2(11.7)/n, tal = Vs/A. = 56.1, fa2 =112.2, fa3 = 168 Hz .

Appendix C - Heat Transfer Equipment Vibration

333

AS

-UNFINNtD TVBE DANK -- -FINNED IVBE 8ANK

o.o··Fo-----.----r----.-I

Figure C-3. Strouhl number for staggered bank of tubes-Chen. [Hydrocarbon Processing] S FOR

STAGGERED ARRAYS

0.27

u

ill.

stj. l.1

1.0

0.1'

ill U

I.e

D.ll l.1

U

U

,.,

Figure C-4. Strouhl number for staggered bank of tubes-Fitzhugh. [Hydrocarbon processing]

Waste Heat Boiler Deskbook

334

S FOR IN LINE ARRAYS '.2 -

J.80.12

0.20

0.16

0.21

0.20

J.' .~ J.~

Slid

2.6 0.20 2.2'

1.8-

u-

0.'8

O,M O.U

1.0 0.8

1.2

JlJ.I 1.6

2.0

U

2.8

3.2

U

0.11 '.0

,.,

SLId

Figure C-S. Strouhl number for inline bank of tubes-Fitzhugh. [Hydrocarbon processing]

The summary or results is shown in Table C-2, which also shows the la data with one and two baffles (w being 11.712 = 5.85 ft and 11.713 =3.9 ft). Note that la and Ie are very close to each other in the very first mode. Hence, acoustic vibration leading to noise is likely. If one baffle is used, la and Ie are kept well apart in all the modes. Also, Ie and In are well apart in all modes, and tube vibrations are unlikely.

CONCLUSION The above calculations show how one can check a tube bundle design for possible vibration or noise problem. A simple approach

Appendix C - Heat Transfer Equipment Vibration

335

was discussed. For elaborate analysis, one would use the methods discussed in literature. 9 However, noise and vibration problems are better predicted based on field operating experience of similar sized units. Performing the above calculations and modifying a design to keep the forcing frequencies well apart may not avoid noise/vibrations in all cases, as vibration and noise phenomenon are inexplicable at times. Damping effect of finned tubes, presence of ash in flue gases, manufacturing tolerances used and effect of end connections are variables that cannot be quantified. Hence, field experience coupled with analysis would be the ideal way to deal with the problem of noise and vibration.

NOMENCLATURE C d di E

-

Constant used in Equation (C-3) Tube outer diameter, in. Tube inner diameter, in. Youngs modulus of elasticity, psi I a - Acoustic frequency, hertz Ie - Vortex shedding frequency, hertz In - Natural frequency of vibration of tubes, hertz I - Moment of inertia of tube I - Tube length, ft Me - Total weight of tube per foot, lb n - Mode of vibration 5 - Strouhal number SI - Longitudinal pitch, in. St - Transverse pitch, in. T - Gas temperature, R V - Gas velocity, ft! s V s - Sonic velocity, ft/ s w - Width of duct, ft A - Wave length, ft

336

Waste Heat Boiler Deskbook

REFERENCES 1. Chen, Y.N., "Flow induced vibration and noise in tube bank heat exchangers due to Von Karman Streets," Trans ASME, Jour. Of Engg for Industry, Vol 1, 1968, pp. 134-146. 2. Rogers, J.D., et al., "Vibration prevention in boiler banks of industrial boilers," American Power Conference, 1977. 3. Fitzhugh, J.S., "Flow induced vibration in heat exchangers," Symposium on vibration problems in industry, UK, April 1973. 4. Rogers, J.D. and Peterson, c.A. "Predicting sonic vibration in cross flow heat exchangers-experience of model testing," ASME 1977 WA/DE28. 5. Barrington, E.,A., "Acoustic vibrations in tubular exchangers, " Chemical Engineering Process," vol 69, No 7, July 1973. 6. Putnam, A.A., "Flow induced noise in heat exchangers," Trans ASME, Jour. of Engg for power, Oct 1959, pg. 417. 7. Deane, W.J., and Cohan, L.J., "Baffle plates cure boiler vibration," Power, Feb. 66, Pg 82. 8. Ganapathy, V., "Applied Heat Transfer, " Pennwell Books, Tulsa, Okla. 1982, pp. 650-658. 9. Symposium on Flow Induced Vibrations, Vol 3, Vibration in heat exchangers, ASME, 1984, pp. 87-101.

Appendix D

Gas Turbine Data As discussed in Chapters 4 and 5, in order to evaluate the design and performance of gas turbine HRSGs, the basic gas parameters of the gas turbine should be available. Since the gas parameters such as gas flow, analysis and exhaust temperature vary with load, type of fuel used and its analysis and ambient conditions, the parameters at all of the operating conditions should be made available to the HRSG designer. In addition, if there are special requirements such as steam or water injection, the exhaust gas conditions should be evaluated by the concerned system engineers and furnished to the HRSG designers. In order to obtain an idea of the size of the various machines available in the industry, one can use the data in the following pages. However, in order to design an HRSG for a specific application, the supplier of gas turbine must be contacted to obtain the gas parameters as discussed above. These data are reproduced with permission from "Deisel and

Gas Turbine Worldwide Catalog . "

337

-~~-----~--------"¥ii·0~·m

Waste Heat Boiler Deskbook

338

Gas Turbine Engines 0

j

j :I ASEA BROWN BOVERI

ABB Power Generation Ltd

liJ

HJ

I

I iIl

II.

I

!

567 Type GT 13E EGiMD Type GT 13E EGiMD Type GT 13 EGiMD EGiMD

Type GT llN EGiMD Type GT llN EGiMD Type GT 11 EGiMD Type GT 11 EGiMD Type GT 8

EGiMD

Type GT 8 Type GT 9

EGiMD

Type GT 9 Type GT 10 Type GT 10

EGiMD EGiMD EG MD

GT 35 Jupiter EG

GT 35 Jupiter MO GT 35 Jupiter MD Mars

EG

Mars

MD

AEG- KANIS

1

ALLISON GAS TURBINE DIV.,

Steam Inj. Steam InJ.

7240 7310

10,243

7750

L G

132,100 111,710

98,500 83,300

7830 7700

10,967 11,077

L G

109,560 99,200

L

97,200

G L

65,400 64,100

G

48,400

47,800 36,100

L G

47,500 29,260

35,400 21,820

*Power Tumlne Inlet Temperature

ALSTHOM

JOHN BROWN ENGINEERING LTD,

G

4330 5263

3329 3925 3807

5106 7966

5943 5860 4877

316

8698

6300

359

510 510

4500 4500

13.8 13,8

173 173

79

510 510

1500·1800 7700

12.0 12,0

202

92 92

374

92

389 466

3000·3600 3600 .6000

10,600 11,240 10,920

11.663 11,769 12,522 13,094 11,663

12,0

202 202

16,0 16,0

84 84

8775 6447

11,769

9,3 9.3 9,4 9,3 9.3 9,3

405· 523

34.6 34,6 34.4 33.9 33,3 33,3

15,7

1035

532

14,200

15.7 15,7

1035 1035

532

14,200 13,820

15.4 15,1 15,1

982

534

1035

533

13,820 14,200 14,200 14,200

6500

533 525

14,200

41.0 41,0 44,2

18,6 20,0

'803 '803

565 565 534

11,500 11,500

44,2

20,0

'803

534

11,500

10,745

35,000 35,300

26,080 26,300

8830 8820

12,490

8,6

272

12,470

10,2

51,460 156,900

38,340

8100

116,900 212,200

7690 7450

11.8 12.1 13,5

270 301

284,830

11,460 10,880 10,542

13,700· 284,832

10,200· 212,200

6990·

9880· 13,500

7.1· 13,5

9544

3000·

525

7510 7595

G G

943· 412· 1224 543

982 '803

5738

G G

9500

variable

1035 982

5760

G

1500·1800

17,1

37,8 37,8

12.7

466

17.1 18,6

9,3 6642 9,3 8654 12,167 12,1 12,167 12.1 10,625 12,7

6350 7725 7694

4735

374

38 38

885

L G L

79

115·

7855

8600 8600

3000 3600

515 515

3000 3000

163 163

L G

6540

489

9,0

9255 6698

6456

3000

12.461 10,980

11.0 11,0

6850· 9693· 7.1· 9936 14,060 30,0

8775 8851

516 516 489

3600 6300

10,890

3936

697 697

81

523

7700

3925 3807

406 406

H I-

011

180

8840

5278

895 895

12.5 12,4 12,4

J

.!

397 359

9400

5263 5106

501 501

{!.

16.3 12,337 9,0

12,600

L

1,105 1,105

12,5

! i!

~

16,3

11,940 11,590

10,000· 111,500

13.9 13,9

11,105

8440 8190

13,410·

l

0.-

{!.

3600 3600

15,840

149,525

~

J I

U'

520

21,240 11 ,850

G G

I

J

520 523

G

EG EG

EGiMD

8810

:I

290 180

7940 7720

G L

EG

8720

J;

316 290

16,900 17,400

EGiMD EGiMD

PG 9281 (F)

7770 7850

P

I

693 397

11,358 10,995

22,600

MN

MD 534 MS 5352 B PG 5371 (PA) EG PG 6541 (B) EG PG 9161 (E) EG

48,800

~

693

11,245

22,660 23,340

501·KF 501·KH

EGiMD/MN

7950 8030

30,310

G L

EGiMD

7780

10,901 11,010

G

G

571·K

81,700 74,000 72,500

10,345

G G

MD

570·K 571·K

201,290

7760 7490

EG EG

EGiMD EGiMDiMN

I

150,100 147,200

501·Ke5

570·K

I

100,500

550 501·KB5 501·KB5

501·KH

(

197.400 134,800

EGiMDiMN GiL

501·KH 501·KH

'li:

!.

G

G L

TURBINENFABRIK GmbH

GENERAL MOTORS

I

88

z

Type GT 13

ABB STAL AB

JI

I.

123 120

889 1350

137 404 610

114.9·

52.15·

1300

600

11,500

926 957

490 483

4670 5100

1104 1104

540

5100

529

1260

583

3000 3000

927· 483· 1260 583

Type GG: Gas Generator; EG: Electric Generator Drive; MD: Mechanical Drive; MN: Marine Propulsion Fuel: L: Liquid (Distillate); G: Gaseous (Natural Gas); R: ReSidual Fuel 'This engine builder is not represented in this 1990 Edition of the Catalog with a section description of its products,

3000· 6500

---------------------~--~-----------------

---

---------

Appendix D - Gas Turbine Data

339

Gas Turbine Engines

I

I :i

J

tj

B

.~ j

j~.

Us

I I..

!

i

~

20,240 19,330 35,000 33,432

15,093 14,420 26,100 24,940

6570 6971 6625 7142

EG

LJG

590· 1860

440· 1400

11,230· 15,660· 9.0· 11,660 16,760 9.4

LJG LJG LJG LJG LJG LJG LJG LJG LJG LJG LJG LJG LJG

21.5 261 516 619 1366 1426 1729 1794 2676 3345 3622 4626 7376

16 195 365 625 1050 1090 1320 1365 2030 2545 2650 3660 5520

24.623 15,641 15,645 13,000 13,706 10,706 12,951 10,461 13,625 13,196

EG EG EG EG EG EG EG EG EG EG EG EG EG EG EG EG

LJG

LJG LJG LJG LJG

2050 2615 1676 2026 2240 2535 3017 4167 5900 16,000 21,000 30,600 30,400 34,600 46,210 46,150

MD MD MD MD MD MD

G G G G G G

5000 5900 16,000 21,000 30,600 34,800

3730 4400 13,420 15,660 22.620 25,950

DAIHATSU DIESEL MFG. CO., LTD.

566

DEUTZ MWM MWM DIESEL UND GASTECHNIK GMBH

527 TA 90 KA 215 KT 215 KA 123 KAl34 KA 1134 KA 334 KA 1334 KT 134 KT 334 KT 1334 CA 139 CA 159

T009 SlA·02 S1T-02 SZA-Ol M1A-Ol M1A-M M1A-03 M1A-13 M1T-Ol M1T-03 M1T-13 501 KB5 571 KA

557 KG2-3C KG2-3E KG2-3R KG2-S1 KG2-S2 KG2-S3 KG2-S4 KG5 DR-99 DR-6G DR-l0 DR-6 DR-30 DR-20 DR-40 DR-40 541 DR-22 DR-990 DR-60 DR-160 DR·61 DR-290

DRESSER-RAND TURBO PRODUCTS DIV.

LJG LJG LJG LJG LJG LJG LJG LJG LJG LJG

LJG

I

l

!

,s e

G G G G

!

514 2000 2000 6000 6000

G"

l

I.

MD EG MD EG

COOPER ROLLS, INC. -COBERRA

DRESSER-RAND POWER

I

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z

AVON AVON RB211 RB211

J

.&

-;:a

i i.

12,669 9.2:1 9653 20:1 10.102 20:1

171 171 196 196

76 76 90 90

G.6· 18.1

J

""

E

!1 U

J

!

H

441 441 465 465

5500 5500 4600 4600

3.0· 8.2

460· 520

1500· 1800

0.465 3.97 7.94 10.45 17.64 12.32 20.30 16.76 35.65 40,56 31.74 34.61 44.09

0.22 1.6 3.6 4.74 6.09 5.59 9.21 7.16 16.17 16.4 14.4 15.7 20.0

600 516 516 496 513 540 545 532 509 540 530 561 534

3600 1500-1600 1500-1600 1500-1600 1500-1600 1500-1600 1500-1600 1500-1600 1500-1600 1500-1600 1500·1600 1500-1600 1500-1600

15,490 21,915 3.9 1530 1950 14,610 20,660 4.6 1400 9350 13.220 3.9 15,440 21,649 3.9 1512 15,076 21.335 3.9 1670 1690 15.071 21,326 4.1 14,360 20,319 4.6 2250 11,605 16.415 6.5 3110 4400 6350 11.610 12.6 13.420 7100 10.050 21.5 15,660 8240 11,660 6.7 22,760 6620 9650 16.1 22,670 6636 9676 16.1 25,950 6670 9720 19.3 34,456 6620 9651 26.6 34,414 6932 9609 26.6

26,4 32.9 27.6 26.9 26.9 30.2 32.4 46.5 45 100 170 150 150 196 270 276

12.9 14.9 12.6 13.1 13.1 13.7 14.7 20 45 77 66 66 90 122.5 125.9

625 630 625 615 625 670 910 670 1050 1216 675 1231 1231 1162 1155 1155

565 546 300 554 604 640 609 500 461 502 430 532 526 464 445 443

16000 16600 16000 16000 16000 16600 16600 12000 7200 6800 5500 5500 3600 5400 3600 3000

9.7 12.6 21.5 6.7 16.1 19.3

35 45 100 170 150 196

16 20 45 77 66 90'

1035 1050 1216 675 1231 1162

561 461 502 430 532 464

13,620 7200 6600 5500 5500 5400

9420 6136

6650 6350 7100 6240 6620 6670

12,124 9.2:1

35,100 22,117 22,404 16,363 19,364 15,136 16,313 14,620 19,549 16,663

9.0 9.0 9.0 6.0 9.2 6.0 9.2

13,320 9.3 11,504 12,0

12.520 11,620 10.050 11,660 9650 9720

21,1

676 676 1164 1164

Type GG: Gas Generator; EG: Electric Generator Drive: MD: Mechanical Drive; MN: Marine Propulsion Fuel: L: Liquid (Distillate); G: Gaseous (Natural Gas); R: Residual Fuel

*This engine builder is not represented in this 1990 Edition of the Catalog with a section description of its products.

Waste Heat Boiler Deskbook

340

Gas Turbine Engines

e

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S

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lJ.

888

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~

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!

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!

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d

521 530 523 513 516 506

7000 3600 3600 3600 3000 3000

I

s

:;....

i

I... 743 610 606 794 799 787

1f~

FIAT AVIO

523 LM500 LM2500PE LM2500PE LM2500PE LM2500PE LM2500PE

MN MN EGlMD EGlMD EGlMD EGlMD

L L L G L G

5450 29,500 29.500 29,500 28.500 26.500

4060 21.998 21.996 21.998 21.252 21.252

8096 6920 6920 6875 7125 7080

11.456 9792 9792 9728 10.082 10.018

14.3 16.8 16.7 18.7 18.7 16.7

34.9 146.0 147.0 147.0 148.0 148.0

15.8 67.1 66.7 66.7 67.1 67.1

FIAT AVIO DIVISIONE A GAS (formerly FIAT TTG)

560 TG.16 TG.20 TG.20 TG.50 TG.50

EG EG EG EG EG

G G G G G

24.400 18,200 50.750 37.850 57.000 42,500 138.100 103.000 172.100 126.340

9825 8640 7900 8195 7515

13.900 12,225 11.180 11.595 10.635

7 11 14 12 14

259 351 346 851 977

117 159 157 386 443

409 520 502 535 495

4850 4920 5400 3000 3000

GARRETT AUXILIARY POWER DIV. ALLIED·SIGNAL AEROSPACE CO.

516 IM631·600

7.9

3.6

496

1500·1800

TUR~INE

735

548

11.425 16,166 11.0

GENERAL ELECTRIC CO 508 PG5371 (PA) POWER GENERATION PG6541 (B) PG7111(EA) Power Generation PG7191(F) Machines PG9161(E) Firing temperature is the PG9281 (F) average total temperature LM2500(PE) at the first slg. stator exit LM2500(PE) ·includes generator LM5000PC efficiency M3142(J) M3142R(J) Mechanical drives M5261 (RA) Ratings given for gas tuel; M5352(B) capability to operate on M5382(C) most fuels exists. M5352R(C)

EG

G

35,269

26.300

6619

12.470

10.2

270

123

957

463

5100

EG EG EG EG EG EG EG MD MD EGlMD MD MD MD MD MD MD

G G G G G G L G L G G G G G G G

51.415 111.975 201.153 156.765 264.565 30.400 30.400 30.400 30.400 44.386 14.600 14.000 26,000 35,000 38,000 35.600

38.340 63.500 150.000 116.900 212.200 22.220· 22,220· 22,670 22.670 33,100 10,900 10,400 19.400 26.100 28.300 26.500

8102 7616 7368 7666 7453 7065 7150 7085 7150 7248 9530 7410 9380 8630 8700 6990

11.450 11.060 10.420 10,860 10,643 10,020 10,110 10,020 10.110 10.253 13.463 10.464 13.271 12.493 12.309 9689

11.6 12.4 13.6 12.1 13.5 16.7 18.7 18.7 16.7 30 7,1 7,3 7.5 8.6 8.9 8.7

301 641 916 669 1323 150.4 150.2 150,4 160.2 271 115 115 202 268 272 263

137 291 417 404 601 66.2 68.1 66,2 66,1 123 52 52 92 122 124 120

1104 1104 1260 1104 1260 613 826 813 626 1202

539 530 563 529 563 541 553 541 653 449 526 353 531 491 516 367

5100 3500 3600 3000 3000 3600 3600 3600 3600 3600 6500 6500 4660 4670 4670 4670

GENERAL ELECTRIC 529 LM500 MARINE & INDUSTRIAL LM500 ENGINES AND SERVICE LM500 DIV. LM1600 LM1600 LM1600 LM2500PE LM2500PE LM2500PE (Continues) LM2500PE

EGlMD EGlMD MN MN EGlMD EGlMD EGlMD EGlMD EGlMD EGIMD

G L L L G L L G L G

5600 5600 6000 20,000 18,750 16.750 30.400 30.400 29.500 29,500

4176 4176 4474 14.915 13,960 13,960 22.670 22.670 21.898 21.996

6180 8220 8151 6845 6847 6660 6932 8672 7111 7052

11,575 11,631 11,646 9686 9679 9735 9809 9724 10,062 9979

14.4 14.4 14.7 22.0 22.0 22.0 16.4 18.4 16.6 18.6

35.0 35.0 35.9 102.0 100.0 100.0 148.0 148.0 149.0 150.0

15.9 15.9 16.0 46.0 44.9 44.9 67.2 67.2 67.6 68.1

768 779 806 780 749 761 626 813 816 604

540 550 565 512 462 492 540 526 541 529

7000 7000 7000 7000 7000 7000 3600 3600 3000 3000

Type GG: Gas Generator; EG: Electric Generator Drive; MD: Mechanical Drive; MN: Marine Propulsion Fuel: L: Liquid (Distillate); G: Gaseous (Natural Gas); R: Residual Fuel ""This engine builder is not represented in this 1990 Edition of the Catalog with a section description of its products.

341

Appendix D - Gas Turbine Data

Gas Turbine Engines I~oj.

j ~

1

U~

I

I I

I

z

" I

II.

!

.l!

!

I I

II

151.0 151.0

66.5 68.5

656 802

556 504

3600 3600

20.0

151.0

68.5

773

494

3000

9728 9455 9670 9671 9813 7966

25.3 27,3 26.0 25.3 26.0 32.0

271.0 292.0 276.0 271.0 276.0 312.0

122.9 132.5 126.1 122.9 126.1 141.6

702 638 693 692 683 676

454 476 451 446 443 407

3600 3600 3000 3600 3000 3600

5650

6276

32.0

313.0

142.0

657

404

3000

4850

6855

32.0

309.0

140.2

680

5050· 3760· 6990- 9890· 6.6· 156,912 116,900 10,570 14,950 12.4

45· 889

20· 404

352· 550

3000· 10,290

UG

1459· 69.600

1088· 51,938

12.1' 305

5.5· 138.3

709· 433· 1160 565

3000 10,290

L G

32,400 36,900

24,161 27,517

6900 6220

9764 8800

16.2 20.0

EG

L

34,300

25,576

6396

9050

EGlMD MN EGlMD EGIMD EGlMD EG

L L L G G G

46.200 55,000 46,200 46,200 46,200 72,100

34,450 41,014 34,450 34,450 34,450 53,765

6675 6687 6975 6830 6930 5650

EG

G

67,000

49,960

GG

G

83,700

62,415

HITACHI LTD.

EGlMD

G

ISHIKAWAJIMA·HARIMA HEAVY INDUSTRIES CO" LTD.

EGlMD

KAB

549 G028 G029

EGlMD EGlMD

L L

255 325

190 240

KAWASAKI HEAVY INDUSTRIES, LTD.

564 SlA·02 SlA·02 SlT·02 S1T·02 S2A·Ol S2A·Ol M1A·Ol M1A·Ol M1A·03 M1A·03 M1T·Ol M1T·Ol M1T'03 M1T·03

EGIMD EGlMD EGlMD EGlMD EGlMD EGlMD EGIMD EGlMD EGlMD EGlMD EG/MD EGlMD EGlMD EGlMD EGlMD EGIMD EGlMD EG/MD EGlMD EGlMD

G L G L G L G L G L G l G L G L G L G L

283 276 557 543 936 913 1569 1531 1965 1916 3032 2960 3783 3692 1703 1667 2105 2060 3299 3226

211 206 415 405 696 661 1170 1142 1465 1430 2261 2207 2821 2753 1270 1243 1570 1536 2460 2406

M1A·l1 M1A·13 M1A·13 M1T-ll (Continues)

M1T·11

!

.l!

1

MN EG

M1A·11

G'

i!

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!.

GENERAL ELECTRIC 529 LM2500PF MARINE & INDUSTRIAL LM2500PH ENGINES AND SERVICE STIG LM2500PH DIV. (Continued) STIG LM5000PC LM5000 LM5000PC LM5000PC LM5000PC LM5000PD STIG LM5000PD STIG LM5000GE STIG

• IJ !... H 1

:I

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!

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p

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5740· 6120· 9.3· 9959 14,090 33.0

!

~

ill

1500 1500 15,480 15,680 15,750 15,950 11,650 11,720 12,170 12,340 11,650 11,640 12,470 12,640 11,920 12,080 10,100

21,890 22,170 22,270 22,550 16,470 16,570 17.220 17,460 16.490 16,740 17,640 17,880 16,660 17.090 14,290 10,200 14.430 9830 13,910 9950 14,060 10,430 14,750 10,570 14.950

9.0 9.0 9.0 9.0 9.0 9.0 8.0 8.0 9.0 9.0 8.0 8.0 9.0 9.0 8.8 8.8 8.7 8.7 8.8 8.8

3.9 3.9 7.7 7.7 10.4 10.4 17.5 17.5 19.9 19.9 35.0 35.0 39.9 39.9 16.5 16.5 16.3 16.3 33.1

33.1

1.8 1.8 3.5 3.5 4.7 4.7 7.9 7.9 9.0 9.0 15.9 15.9 16.1 18.1

7.5 7.5 7.4 7.4 15.0 15.0

930 930 930 930 930 930 900 900 960 960 900 900 960 960 900 900 1030 1030 900 900

Type GG: Gas Generator; EG: Electric Generator Drive; MD: Mechanical Drive; MN: Marine Propulsion Fuel: L: liquid (Dislillate); G: Gaseous (Natural Gas); R: Residual Fuel 'This engine builder is not represented in this 1990 Edition of the Catalog with a section description of its products.

520 520 520 520 495 495 515 515 545 545 510 510 540 540 465 465 555 555 465 465

1500·1800 1500·1800 1500·1800 1500·1800 1500·1800 1500·1800 1500·1600 1500·1800 1500·1800 1500·1800 1500·1800 1500·1800 1500·1800 1500·1800 1500-1800 1500·1600 1500·1800 1500·1800 1500-1600 1500-1800

Waste Heat Boiler Deskbook

342

Gas Turbine Engines

I~j

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KAWASAKI HEAVY INDUSTRIES. LTD. (Continued)

KHD LUFTFAHRTIECHNIK

I:i!

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564 MH-13 MH-13 MIA-13CC MIA-13CC MIA-13CC Steam inj. MIA-13CC Steam inj.

EGiMO EGiMD EGIMD EGiMD EGiMD

G L G L G

4090 3999 1602 1762 3339

3050 2962 1344 1314 2490

10.150 10.260 11,660 11.610 7400

14.350 14.540 16,500 16.710 10,470

6.7 6.7 7.6 7.6 9.2

32.6 32.6 16.6 16.6 15.7

14.6 14.6 7.5 7.5 7.1

1030 1030 1010 1010 1010

555 555 575 575 575

EGiMD

L

3265

2435

7500

10,600

9.2

15.6

7.1

1010 575 1500-1600

526 T312 T117 T009

turbo-KHD

L L L

32

105 1040 24

5.05 5.5 3.6

!

turbojet turboshaft

'li:

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0.672 1.56 0.24

657

1500-1600 1500-1600 1500-1600 1500-1600 1500-1600

6000 360012,000

KOBE STEEL. LTD.

EG

GIL

12407725

9255760

7510- 10.657- 4.215.600 22.000 12.7

14.244.2

6.4520.0

603 5321035 610

150014.200

MAN. GUTEHOFFNUNGSHOTIE

EGIMD

UG

724012,400

54009250

7169- 10,140- 7.610.604 15,000 10

7599

3445

905- 4961000 515

76006000

20-200

15-150

77905610- 7310- 10,340- 11199,540 146.600 9060 12.640 IS

57914

26414

497567

360013.600

7693- 10,665- 6.69575 13.546 13.9

10.6230.1

4.6104.4

927- 4661100 522

5070 26,600

16,962 15,095 16,962 15,095 15,095 13.666 15.095 13,963 13,666 10,692 10.567 10,007

5.3 5.3 10.6 10.6 17 17 27.1 27.1 33.9 37.9 37.9 37.2

2.4 2.4 4.6 4.6 7.7 7.7 12.3 12.3 15.4 17.2 17.2 16.9

690 950 690 950 670 935 670 935 935 910 940 1060

MICROTURBO MITSUBISHI HEAVY INDUSTRIES. LTD. POWER SYSTEMS HEADQUARTERS

EG

UG

EG

L

151532,412

113024,170

EG EG EG EG EG EG EG EG EG EG EG EG

L L L L L L L L L L L L

320 400 640 600 1100 1300 1720 2100 2700 3200 3670 4760

240 300 460 600 600 1000 1260 1600 2000 2400 2660 3560

NQTE: ALL RATINGS ON NATURAL GAS FUEL AT GENERATOR TERMINALS MITSUI ENGINEERING & SHIPBUILDING NIIGATA ENGINEERING CO" LTD. Stand-by DUly

(Continues)

SB

197 CNT-300E CNT-375E CNT-600E CNT-750E CNT-IOOOE CNT-1250E CNT-1600E CNT-2000E CNT-2500E CNT-3000E CNT-4500E CNT-5000E

24.021 7.5 21.352 7.5 24,021 7.5 21.352 7$ 21.352 7.5 19,644 7.5 21.352 7.3 19.751 7.3 19.644 7.5 15.124 9.0 14,976 9.0 14.155 10.2

Type GG: Gas Generator; EG: Electric Generator Drive; MD: Mechanical Drive; MN: Ma(ine Propulsion Fuel: L: Liquid (Distillate): G: Gaseous (Natural Gas): R: Residual Fuel

*This engine builder is not represented in this 1990 Edition of the Catalog with a section description of its products.

510 1500-1600 550 1500-1600 510 1500-1600 550 1500-1600 495 1500-1600 535 1500-1600 495 1500-1600 535 1500-1600 535 1500-1600 415 1500-1600 465 1500-1600 565 1500-1800

343

Appendix D - Gas Turbine Data

Gas Turbine Engines

I

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NIGATA ENGINEERING CO., LTD. (Continued) CONTINUOUS DUTY

. U~ I~11.§

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PRATT & WHITNEY CANADA, INC.

(Continues)

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197 CNT-10000E

EG

L

9800

7300

9099

12,870

CNT-1500C CNT-4500C CNT-5000C CNT-10000C

EG EG EG EG

UG UG UG UG

1450 4200 5200 11,900

1080 3130 3880 8640

11,000 9656 9100 8190

15,600 6.1 13,658 9.0 12,872 10.2 11.585 16

14.3 39.2 39.0 83.9

6.5 17.8 17.7 38.1

60340

45253

20,394- 28,650- 3.726,863 38,040 5.8

1.45.0

0.6· 2.3

550· 500

30008000

2800 6800 6500 14,000 13,500 14,600 14,000 18,600 29,500 29,500 38,000 35,000 46,200

2100 5070 4850 10,440 10,030 10,890 10,440 13,870 22,000 22,000 28,350 26,110 34,450 2000 4820 4610 9980 9620 10.450 10,000 13,170 21,000 21,690 26,630 33,760 38,980 83,500 118,690

9750 9670 7440 7485 7090 9530 7410 6850 6880 6880 6470 7040 6820

13,710 13,660 10,520 10,590 10,030 13.480 10,480 9690 9730 9730 11,980 9960 9660 14.400 14,390 11,070 11,250 10,590 14,060 10,940 10,200 10,190 10,230 12.340 6960 11,380 11,050 10,800

12 8.2 8.3 14 14.2 7,1 7.5 21.5 18 18 8.2 8.4 30 12 8.2 8.3 14 14.2 7.1 7.5 21.5 18 18 10.2 30 11.5 12.4 1'.6

10 24.6 24.6 41.2 41.2 52.3 52.3 45 66.7 66.7 123 115 122 10 24.6 24.6 41.2 41.2 52.3 52.3 45.0 66.7 66.7 123.7 122 137.9 642 405.8

538 529 357' 462 415' 526 353' 470 513 513 490 353' 450 538 529 357' 462 415' 526 353' 470 513 513 479 450 537 530 524

22,500 10,290 10,290 7900 7900 6500 6500 7900 6500 3600 4670 4670 3600 22,500 10,290 10,290 7900 7900 6500 6500 7900 6500 3600 5100 3600 5100 3600 3000

550 550 654 862 1049 1440

410 410 487 643 782 1074

12,144 12,144 11,408 10,782 10,359 11,960

17.184 17,184 16,143 15,257 14,658 16,924

5.8:1 5.8:1 7.1:1 7.5:1 8.5:1 6.8:1

2.70 2.70 2.81 3.22 3.83 5.76

525 6188 525 2200 33,000 550 610 33,00011900 577 33,00011700 574 6.600

i

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NOEL PENNY TURBINES LTD. NUOVO PIGNONE

I I

538 PGT 2 MS1000 MS1000R PGT10. PGT10R MS3002 MS3002R PGT16 PGT25 LM2500/30 MS5002 MS5002R LM5000 PGT2 MS1000 MS1000R PGT10 PGT10R MS3002 MS3002R PGT16 PGT25 LM2500/30 MS5001 LM5000 MS6001 MD7001 MS9001

MD MD MD MD MD MD MD MD MD MD MD MD MD EG EG EG EG EG EG EG EG EG EG EG EG EG EG EG

545 ST6K-77 ST6J-77 ST6L·77 ST6L-79 ST6L-81 ST6T-76 Twin Pac

MDIMN MDIMN MDIMN MDIMN MDIMN MD/MN

L L L L L L

"li:

11

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16

82.8

37.6

1077 485 1500-1800 871 904 1010 1057

5.95 5.95 6.20 7.10 8.45 12.7

500 1500-1600 449 1500-1800 516 1500-1800 465 1500-1800

Type GG: Gas Generator; EG: Electric Generator Drive; MD: Mechanical Drive; MN: Marine Propulsion Fuel: L: Liquid (Dislillate): G: Gaseous (Natural Gas); R: Residual Fuel

*This engine builder is not represented in this 1990 Edition of the Catalog with a section description of its products.

344

Waste Heat Boiler Deskbook

Gas Turbine Engines

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PRATT & WHITNEY CANADA, INC, (Continued)

545 SPW901/1 SPW124·2

MDIMN MDIMN

ROLLS·ROYCE PLC

589 Avon RB211 SM1C SM2C SM3C OLYMPUS TM3B TYNE RM1C

GG GG MN MN MN MN

I

j

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[

L L

1765 2471

1317 1843

23,800 39,000 24,140 24,140 24,140 21,500

17,750 29,080 18,000 18,000 18,000 16,033

7230 6099 6862 6862 6862 9438

10,230 8625 9710 9710 9710 13,355

5340

3982

8683

MN

!

L

!

UG UG

G'

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10,966 15,518 7,2:1 9,292 13,148 13.7:1

n

H

15,5 17.0

7,03 7.71

171 198 147 147 147 235

78 90 66.8 66.8 66.8 107

870 1161 710 710 710 585

610 717 457 457 457 427

5500 5500 5500 5660

12,287 12.5

46

21

629

435

3425

900 910

9.2 20.0 21.9 21,9 21.9 10.5

527 532

24,625 20,0001 1200

RUSTON GAS TURBINES 534 TB5000 TB5000 TYPHOON TORNADO TORNADO

EG EG EG EG EG

L G LG L G

5200 5400 5521 8350 8768

3879 4027 4119 6229 6540

9600 9370 8003 8156 8033

13,577 7.0 13,259 7.1 11,318 12.8 11,535 12.1 11,367 12.2

46.3 46.8 37.5 62.2 62.2

21.1 21.2 17.0 28.2 28.2

495 486 496 1000 458 1025 468

1500 1500 1500 1500 1500

TB 5000 TB 5000 TYPHOON TORNADO

MD MD MD MD

L G LG LG

5200 5400

3879 4027

9600 9370

13,577 13,259

7.0 7.1

46.5 46.8

21.1 21.2

900 910

495 486

7950 7950

8500

6341

8156

11,535

12

RLM RLM RLM RLM RLM RLM RLM RLM

1600 1600 2500 2500 2500 2500 5000 5000

MD MD MD MD MD MD MO MO

G L G L G L G G

18,870 18,750 30,410 30,441 29,493 29,512 46,319 46,319

14,070 13,982 22,677 22,700 21,993 22,007 34,540 34,540

6841 6878 6871 6931 7052 7111 6831 6931

9679 21.5 9731 21.5 9722 18.4 9806 18.4 9978 18.6 10,061 18.5 8564 25.4 9806 26.1

RLM 1600 RLM 1600 RLM 2500 RLM 2500 STIG RLM 2500 RLM 2500 RLM 2500 STIG RLM 2500 RLM 5000 RLM 5000 STiG RLM 5000 RLM 5000 STIG

EG EG EG EG

G L G G

18,870 18,750 30,464 37,185

14,070 13,982 22,717 27,729

6841 6878 6872 6169

8679 8731 8722 8728

EG EG EG

L G G

30,477 29,523 35,187

22,727 22,015 26,239

6928 7045 6346

9801 8967 8978

EG EG EG

L G G

29,486 46,319 71,232

21,995 34,540 53,118

7104 6820 5567

EG EG

G G

46,319 67,675

54,540 50,465

6915 5771

60.8

27,6

1000 470

10000

100.8 100.4 150,3 150,2 151.5 151.4 269.4 276

45,7 45.6 68.2 68.1 68.7 68.7 122.2 125.6

482 493 528 540 529 541 446 444

7000 7000 3600 3600 3000 3000 3600 3000

21.5 21.5 18.4 20.0

100.8 100,4 150.3 184.7

45.7 45.6 68.2 74.7

482 493 528 503

7000 7000 3600 3600

18.4 18.6 20.0

150.2 151.5 165.1

68.1 68.7 74.9

540 529 503

3600 3000 3000

10,050 18.5 8849 25.4 7876 32.0

151.4 269.4 350.6

68.7 122.2 159,0

541 446 392

3000 3600 3600

276.9 353.4

125.6 160.3

444 393

3000 3000

9783 8165

26.1 32.1

Type GG: Gas Generator; EG: Electric Generator Drive; MD: Mechanicel Drive; MN: Marine Propulsion Fuel: L: Liquid (Distillate); G: Gaseous (Natural Gas); R: Residual Fuel -This engine builder is not represented in this 1990 Edition at the Catalog with a section description of its products.

345

Appendix D - Gas Turbine Data

Gas Turbine Engines

.•

I:II

SIEMENS KWU °base-Ioad performance referred to mechanical output at turbine coupling

a 2) 111

j

HI

t Iz f I

"

V64/84/94

I II.

!

l

'ii:

!

j

II

p

I8

JI

tJ.

J !... I a 'I

It

:I

n

n

~

i

10.435 '10.610.610 15.8

3751100

170500

535645

10.452 8436 10.933 7825 8589

14.785 11.933 15.470 11.072 12.149

6.7 6.5 6.5 6.4 8.3

14.1 13.9 14.1 13.8 38.9

6.4 6.3 6.4 6.3 17.6

499 1500-1800 296 1500-1800 477 22.300 296 22.300 313 1500-1800

3130 3880

9656 9099

13.658 12,870

9.4 9,3

38.7 38,5

17.6 17.5

449 1500-1800 515 1500-1800

4370

9099

12.870 11.2

46.0

20.9

497 1500-1800

t

EG

UG

EG

G G G G G

1450 1200 1340 1280 3420

1080 890 1000 955 2550

G G

4200 5200

G

5870

1

54.300152.650

i!

J .:!

~

30003600

when fired with Iuel gas (LHV - 50.056 kJlkg)

SOLAR TURBINES INCORPORATED

581 Saturn

Saturn (Recupl EG MD Saturn Saturn (Recup) MD Centaur EG IRecup) Centaur EG Centaur EG Type H Centaur EG Taurus Centaur MD (Recupl MD Centaur Centaur MD Type H MD Centaur Taurus EG Mars Mars MD

STEWART & STEVENSON SERVICES. INC. For 'EG' sets, kW measured at Generator Terminals For "MD~ and uMN sets, Bhp and kW

are measured at power turbine shaft.

(Continues)

337 IM-831-8oo TG501-KB TG501-KB5 TG570-K TG571-K TG1600 TG2500-33 TG2500STIG 40 TG5000 PC TG5000STiG 80 TG5000STIG 120 TMD501K5SS TMD570·K TMD571-K TMD1600

G

3680

2745

7370

10,427

7.9

36.4

16.5

298

14,758

G G

4390 5500

3275 4100

9100 8655

12.875 9.6 12,246 10.2

37.6 38.6

17,1 17.5

451 516

15.500 16.520

G

6500

4845

8210

11,617

G G

11,900 12.600

8840 9400

8190 7700

11,585 15.7 10,894 15.7

82.6 82.6

37.5 37.5

465 1500-1800 465 8790

EG EG EG EG EG EG EG EG

G G G G G G G G

384 4328 4957 6160 7501 18,023 29,846 35,525

515 12,097 17.111 11.0 3224 9420 13,325 9.3 3693 8965 12,681 9.3 4589 9145 12,936 12.1 5588 7945 11,238 12.7 13.427 7126 10,080 22.5 22.236 7012 9919 18.7 26,466 6365 9003 20.0

7.9 34.6 34.6 41.8 44.2 101 150.4 163.3

3,6 15.7 15.65 18.96 20.05 45.81 68.22 73,94

499 524 533 546 533 482 528 504

EG EG

G G

45,318 62.940

33.762 46,890

6973 6098

9864 8625

28.8 28.8

269.0 328.8

122 149

446 3000-3600 413 3000-3800

EG

G

69.289

51.620

5898

8343

28.8

346.5

157

407 3000-3600

MD

G

5278

3936

8851

12.524

9.3

34.6

15.65

533

14,200

MD MD MD

G G G

6540 7925 18,750

4877 5909 13,982

8600 7510 6840

12,169 12.1 10.627 12.7 9677 22.5

41.8 44.2 101

18.96 20.05 45.81

546 533 482

11.500 11.500 7000

Type GG: Gas Generalor; EG: Electric Generator Drive; MD: Mechanical Drive; MN: Marine Propulsion Fuel: L: Liquid (Distillate); G: Gaseous (Natural Gas); R; Residual Fuel "This engine builder is not represented in this 1990 Edition of the Catalog with a section description of its products.

1500-1800 1500-1800 1500-1800 1500-1800 1500-1800 1500-1800 3000-3600 3000-3600

Waste Heat Boiler Deskbook

346

Gas Turbine Engines

I• i

STEWART & STEVENSON SERVICES, INC_ (Continued)

SULZER ESCHER WYSS LTD.

TEXTRON LYCOMING DIVISION

II ,

.~ j

I

u-

~

I

z

B )

I I

1

ie I

J~

[

!.

:I .&

';"

11

I

n

n

I

f

~

I

I...

337 TMD2500-33 TMD5000PC TMP501K5SS TMP570-K TMP571-K TMP1600 TMP2500-33

MD MD

G G

30,400 46,200

22,669 34,451

6872 6831

9723 9666

18.7 28.8

150.4 269

68.22 122

528 3000-3600 446 3000-3600

MN

L

5106

3808

8850

12,523

9.3

34

15.42

536

14,200

MN MN MN MN

L L L L

6350 7694 18,750 32,000

4735 5737 13,890 23,862

8593 7507 6880 6882

12,159 10,622 9753 9756

12.1 12.7 22.5 19.1

41.8 44,2

101 151

18.96 20.0 46 68.5

546 533 492 549

11,500 11.500 7000 3600

561 Type 3 Type R3 Type 53

EG EG MD

G G G

8715 8100 8715

6500 6040 6500

8960 7855 8960

12,678 10.0 11,115 9.9 12,678 10.0

73.4 73.4 73.4

33.3 33.3 33.3

970 970 970

464 365 464

Type SR3

MD

G

8100

6040

7855

11,115

9.9

73.4

33.3

970

365

Type 7 Type R7 Type 57

EG EG MD

G G G

14,750 14,215 14,615

11,000 10,262 14,520 10,600 8054 11.396 10,900 10,430 14,758

7.6 7.7 7.6

143.3 142.9 143.3

65.0 64_8 65.0

925 925 925

491 342 493

Type SR7

MD

G

14,010

10,450

7.7

142.9

64.8

925

342

10,600 10,600 450011,130 450011,130 6400 6400 32007040 32007040

TF Models

EG/MDIMN

15004000

11032940

9108-. 12,842- 6.5 13,034 18,378 13.3

11_528.0

5.212.7

968- 4791060 610

300015.400

THOMASSEN INTERNATIONAL TURBOMECA INDUSTRIAL MARINE DIV.

U]

s

IJ.

585 Oredon IV Astazou IV

!

8155

11.540

l

I

hi

EG/MD

G

14,00045,950

10.44034,265

6869- 9720- 6.99936 14,060 30.0

101893

46405

353543

30007000

EG EG/MN

L

UG

134 442

100 330

16,678 23,600 3.75 12,815 18,130 5.6

1.94 5.62

0.88 2.55

575 490

8000 15001800 15001800 15001800 15001800 15001800 5500 15005500

Bastan VI

EG

UG

818

610

12,270 17,360

5.5

9.7

4.4

480

Bastan VII

EG

UG

1167

870

10,947 15.490

7.2

13.2

6.0

450

Bi·Bastan VI

EG

UG

1548

1155

12,487 17,670

5.5

19.4

8.80

480

Bi·Baslan VII

EG

UG

2206

1645

11 ,283 15,965

7_2

26.4

12,0

450

Turmo III Makila T1

MD MDIMN

UG

UG

1005 1648

750 1230

13,250 18.750 9133 12,918

5.0 9.9

11.9 12.1

5.40 5.5

510 525

Type GG: Gas Generator; EG: Electric Generator Drive: MD: Mechanical Drive: MN: Marine Propulsion Fuel: L: Liquid (Distillale): G: Gaseous (Natural Gas): R: Residual Fuel

*This engine builder is not represented in this 1990 Edition of the Catalog with a section description of its products.

-

..

- - . - - - - - - - - - - - - ----

Appendix D - Gas Turbine Data

347

Gas Turbine Engines

j I

i

:I

I

f

I~j

I:i!

Fully steam-injected

Fully stesm-injected

WESTINGHOUSE CANADA, INC.

,I io!!

!

!

83.9 79.4 167.8

450 ~000-36oo 440 3600 450 3000-3600

10.400 15.8 lo.s65 10.6 9945 15.8

375 770 857

170.1 349.3 389.5

530 540 560

5400 3600 3600

16.1

10.2 17.4 16.4 19.8 16.2 34.8 32.9 39.7 32.5 34.1 34.1 41.0 43.3 15.7 34.8 105.7 122.9 105.7

4.6 7.9 7.4 9.0 7.3 15.8 14.9 18.0 14.7 15.5 15.5 18.6 19.6 7.1 15.8 47.9 55.7 47.9

930 959 869 1020 1031 948 869 1006 1031 976.6 1043.9 1048.8 992.8 1056 924,2 1013 982 1027

22,000 22,000 22,000 22,000 22,000 22,000 22,000 22,000 22,000 14,200 14,200 11,500 11,500 22,000 14,600 9660 9660 9660

7.3 8.2 9.3 5.9

222 248 282 172

100 112 127 78

565 540 520 350

5000 5000 5000 5000

9480 7.3 9710 8.5 14,120 7.5 11.130 15.3

213 249 270 382

97 113 123 173

365 375 412 510

5000 5000 4912 5426

10,570 14.2 10,270 14.2

813 935

369 424

519 570

3600 3600

4.71 7.89 11.3 15.8 22.69 34.03

2.14 3.58 5.15 7.16 10.30 15.45

480 510 505 510 505 505

1500-1800 1500-1800 1500-1800 1500-1800 1500-1800 1500-1800

[ 26,125 22,982 52,250

6495 6720 6495

EG EG EG

G G G

72,530" 54,085" 140,940 105,100 194,310 144,900

7350 7466 7030

620 UST700 UST1100 UST1200 UST1400 UST1500 UST2200 UST2400 UST2700 UST3OO0 UST3400 UST3800 UST4600 UST5700 UST2500CC UST5600CC UST12000 UST15000 UST18000

EG EG EG EG EG EG EG EG EG EG EG EG EG EG EG EG EG EG

G G G G G G G G G G G G G G G G G G

933 1569 1702 1964 2104 3032 3298 3783 4090 4689 5262 8539 7823 3338 7875 17,820 20,570 23,570

661 112 1206 1392 1491 2148 2337 2680 2898 3307 3711 4612 5588 2365 5554 12,760 14,730 16,880

11,682 12,173 10,105 11,659 9838 12.467 10,430 11,921 10,146 8934 8996 8601 7511 7403 6307 8027 7863 7129

16,530 17,225 14,299 16.497 13,921 17,641 14,758 16,868 14,357 12,642 12,729 12,170 10,628 10.475 8924 11,358 11,126 10,088

9.0 8.0 8.0 9,0 8.7 8.0 8.8 9.0 8.7 9.3 9.3 12.1 12.7 9.4 17.1 12.8

519 CW352MM CW352MA CW352MB CW352 RMM

MD MD MD MD

G G G G

30,100 35,000 39,800 20,800

22.400 26,100 29,700 15,500

9460 8900 8600 6750

13,380 12.590 12,170 9550

CW352RI.\A CW352RMB CW191PG CW251PG

MD MD EG EG

G G G G

27,900 33,100 23,758 62.416

20,800 24,700 17,700 46,500

6700 6860 9982 7869

519 WD501D 501F

EG EG

G G

YANMAR

578 AT36C AT60C AT90C AT120C AT180C AT270

EG EG EG EG EG EG

UG UG UG UG UG UG

107,850 150,000

380 652 1087 1304 2174 3263

280 480 800 960 1600 2400

12,790 12,760 11.400 12,760 11.400 11.400

n

!

185 175 370

t

35,035 30,820' 70.070

WESTINGHOUSE ELECTRIC CORP. COMBUSTION TURBINE DEPARTMENT

H

20.0 18.7 20.0

..

G L G

!

.f !

I l

EGlMD MN EG

I

0-

:&

~

1

575 FT8 FT8 ·Constant to Tamb-80'F u@Coupling FT8 Twin Pac V64.3 NOTE: EXCLUDES INLET V84.2 & EXHAUST DUCTING V64.41 AND GENERATOR USE

rI !...,.

ii

:8

'it

TURBO POWER

U.S. TURBINE CORPORATION

H!

0-

I

I I

9190 9509 9190

18.090 18,050 16,120 18,050 16,120 16,120

14,7

8.1 8.1 8.1 8.1 8.1 8.1

:!

'it

1724 1670 1652 1778 1888 1670 1652 1760 1886 1800 1895 1477 1477 1850 1800 2282 2282 2282

900 900 900 900 900 900

Type GG: Gas Generator: EG: Electric Generator Drive; MD: MeChanical Drive; MN: Marine Propulsion Fuel: L: liquid (Dislllla1e); G: Gaseous (Natural Gas); R: Residual Fuel

·This engine builder is not represented in this 1990 Edition of the Catalog with a section description of its products.

Appendix E

Gas and Steam Properties

SPECIFIC HEAT, VISCOSITY AND THERMAL CONDUCTIVITY OF GASES The specific heat, viscosity and thermal conductivity data for a few common gases at atmospheric pressure are given below. From heat transfer consideration, pressure effect becomes significant above 250 psig and at gas.temperatures below 400°F. See references 1 and 2 for more information on pressure effect. In Tables E-1 to E-9, the following units are used: temp - F, Cp - gas specific heat-Btu/Lb F, 11- viscosity-Lb/ft h, k - thermal conductivity-Btu/ft h F.

Table E-1. Carbon dioxide k temp Cp Jl

Table E-2. Water vapor k temp Cp Jl

200 400 600 800 1000 1200 1400 1600 1800 2000

200 400 600 800 1000 1200 1400 1600 1800 2000

.2162 .2369 .2543 .2688 .2807 .2903 .2980 .3041 .3090 .3129

.0438 .0544 .0645 .0749 .0829· .0913 .0991 .1064 .1130 .1191

.0125 .0177 .0227 .0274 .0319 .0360 .0400 .0435 .0468 .0500 349

.4532 .4663 .4812 .4975 .5147 .5325 .5506 .5684 .5857 .6019

.0315 .0411 .0506 .0597 .0687 .0773 .0858 .0939 .0119 .1095

.0134 .0197 .0261 .0326 .0393 .0462 .0532 .0604 .0678 .0753

- - -

350

.

-~~~~--

Waste Heat Boiler Deskbook

-,

v'-

Table E-3. Nitrogen Cp temp Jl 200 400 600 800 1000 1200 1400 1600 1800 2000

.2495 .2530 .2574 .2624 .2678 .2734 .2791 .2846 .2897 .2942

.0518 .0608 .0694 .0776 .0854 .0927 .0996 .1061 .1122 .1178

k .0189 .0219 .0249 .0279 .0309 .0339 .0369 .0399 .0429 .0459

Table E-S. Sulfur dioxide k Cp temp Jl 200 400 600 800 1000 1200 1400 1600 1800 2000

.1578 .1704 .1806 .1887 .1950 .1997 .2030 .2054 .2069 .2079

.0386 .0493 .0595 .0692 .0784 .0871 .0954 .1030 .1103 .1170

.0074 .0109 .0143 .0175 .0205 .0234 .0261 .0286 .0310 .0332

Table E-4. Oxygen Cp temp Jl 200 400 600 800 1000 1200 1400 1600 1800 2000

.2250 .2332 .2404 .2468 .2523 .2570 .2611 .2647 .2678 .2705

.0604 .0716 .0823 .0924 .1021 .1111 .1197 .1278 .1353 .1423

{

k .0186 .0229 .0272 .0313 .0352 .0389 .0425 .0460 .0492 .0523

Table E-6. Hydrogen chloride k Cp temp Jl 200 400 600 800 1000 1200 1400 1600 1800 2000

.1907 .1916 .1936 .1965 .2002 .2043 .2086 .2128 .2168 .2203

.0412 .0534 .0655 .0774 .0892 .1009 .1124 .1239 .1351 .1463

.0113 .0143 .0175 .0209 .0245 .0283 .0327 .0364 .0407 .0452

Table E-S. Flue gas

Table E-7. Air (dry)

(% vol C0:2 = 12, H20 = 12, N2 = 70, 0:2 = 6)

temp

Cp

200 400 800 1200 1600 2000

.2439 .2485 .2587 .2696 .2800 .2887

Jl .0537 .0632 .0809 .0968 .1109 .1232

k .0188 .0221 .0287 .0350 .0412 .0473

temp

Cp

200 400 800 1200 1600 2000

.2570 .2647 .2800 .2947 .3080 .3190

Jl .0492 .0587 .0763 .0922 .1063 .1188

k .0174 .0211 .0286 .0358 .0429 .0499

-- -

351

Appendix E - Gas and Steam Properties

Table E-9. Gas turbine exhaust gases (% vol C02 = 3, H20 = 7, N2 =75, 02 =15)

temp 200 400 600 800 1000

f.l

Cp .2529 .2584 .2643 .2705 .2767

.0517 .0612 .0702 .0789 .0870

k .0182 .0218 .0253 .0287 .0321

ENTHALPY OF GASES The following table gives the enthalpy of a few gases with reference to 60 0 P in Btu/lb. They were computed using data for individual gases from reference 3. Table E-10. Enthalpy of gases gas type

% vol

A - gas turbine exhaust 8 - sulfur combustion C - flue gas D - dry air

temp, F 200 400 600 800 1000 1400 1800

A 34.98 86.19 138.70 192.49 247.56

CO2 3

H2 O 7

12

12

B 31.85 78.57 126.57 175.77 226.20 330.15 437.86

N2 75 81 70 79

C 35.52 87.83 141.79 197.35 254.47 (~!f9~ 496.42

02 15 10 6 21

80 2 9

D 33.74 83.00 133.42 184.91 237.52 345.77 457.82

ESTIMATING FLUE GAS PROPERTIES Plu gas properties are required for performing heat transfer calculations. Since they consist of several components, a quick estimate of specific heat Cp , viscosity f.l and thermal conductivity k may be obtained for low pressure gases as follows.

l

04

352

Waste Heat Boiler Deskbook

(E-l)

km

=.I:Yiki VMWi YiVMWi

(E-2)

(E-3)

where MW is the molecular weight, Y the volume fraction of any constituent and m stands for mixture property. Example 1: Determine C p , J!, k for gas turbine exhaust gases at 600 oP; the gas analysis is: % volume C02 = 3, H20 = 7, N2 = 75 and 02 = 15. Data for individual gases may be obtained from tables E-l to E-6.

I

/

Solution: 1 IYiMWi = .03 x 44 + .07 x 18 + .75 x 28 + .15 x 32 =8.38 .I:Yi VMWi

= 0.3 x Y44 + .07 sym

.75 x f18 + .15

X

3

X

rn + .15

X

+

fIT = 5.31

.I:Yi VMWi = 0.3 X V44

.75

m

V32

+ .07 X ill +

= 3.038

The gas data at 600 0 P are as follows:

Cp J! k

C02

H2 0

N2

02

.2543 .0645 .0227

.4810 .0506 .0261

.2574 .0694 .0249

.2404 .0823 .0272

units Btu/lb F Lblft h Btultt h F

substituting in Equations (E-1) to (E-3):

d

353

Appendix E - Gas and Steam Properties

Cfrn

(.fix .2543x44 +.D7x .481x 18 + .75x .2574x28 + .15x.2404x32) 28.38 =.2643

(.fix .0645x {44 + .D7x .0000x;i18 + .75x .0694x.fiB+ .15x .0823x .f32) 5.31

= JJ7CJ2

(.03 x .CJ227 xVii + JJ7x .CJ2fJJ7x

km

V18 + .75x .CJ249x V28 + .15x .CJ272x V32) 3.fi8

=.CJ2524

EFFECT OF PRESSURE ON HEAT TRANSFER The effect of pressure on heat transfer becomes more significant above 20 atmospheres. Reference 1 describes the methodology for computing pressure effect on gas mixture properties. Figures E-1 to E-12 show the effect of pressure on some common gases for flow inside tubes and Figures E-13 to E-24, the effect of pressure for flow outside tubes. Two examples illustrate the use of the charts. Example 2: 200 pph of air at 400°F and 100 atmospheres flow inside a tube of inner diameter 1.7 in. determine the tube side heat transfer coefficient. Solution: Using the Equation (2-7) of Chapter 2, hi =2.44 w .8C/ di 1.8i factor C is obtained from Figures E-1 to E12. C = .18 for air at 400°F and 100 atmospheres. hi =2.44 x 200. 8 x .18/1.71.8 =11.71 Btu/sq ft h F If air had been at atmospheric pressure at 400°F, C hi =11.12 Btu/sq ft h F.

= .171 and hence

\

,,"

.

--

----------------

Waste Heat Boiler Deskbook

354

Example 3: Determine the heat transfer coefficient ho when air at 400°F and 100 atmospheres flows over a bank of 2-in tubes. Gas mass velocity G =5000 lb/sq ft h.

Solution: Using Equation (3-23) of Chapter 3 ho = .9 G·6 F/do.4 .F is obtained from Figures E-13 to E-24 for' flow outside tubes. In this case it is .109. ho = .9 x 5000.6 x .109/2.4 =12.32 Btu/sq ft h F If air had been at atmospheric pressure, F = .1025 and hence ho = 11.58 Btu/sq ft h F. Another example is given below to show the effect of gas analysis on gas properties. The data are from Cheng cycle system and the comparison will be made between the dry and injection modes of the gas turbine. Example 4: In the Cheng cycle system, the gas analyses in the dry and injection modes are as given below in Table E-11. Calculate the gas properties at 1000°, 600° and 400°F. Solution: Using the above methodology, the gas properties were obtained and are shown below. Table E-11. Gas data for dry and injection modes

1. 2. 3. 4.

\

C02 (%vol) H2O N2 02

dry 2.9 7.1 76.0 14.0

injection 2.7 25.1 61.0 11.2

temp

Cp

J.L

k

Cp

1000 700 400

.2771 .2677 .2588

.0868 .0744 .0611

.0321 .0269 .0217

.3086 .2971 .2862

J.L

.0839 .0713 .0578

k

.0333 .0273 .0214

l

----=-- - - - " ----"-,,-=---

355

Appendix E - Gas and Stearn Properties

Factor F at 700°F is .1355 for dry and .144 for injection modes, showing that a significant difference in heat transfer coefficients can arise if the effect of gas analysis is neglected.

CONVERTING FLUE GAS ANALYSIS FROM WEIGHT TO VOLUME BASIS One of the frequent calculations performed by heat transfer engineers is the conversion from weight to volume basis and vice versa. The following example shows how this is done. Example 5: A gas contains 3% C02, 6% H20, 74% N2 and 17% 02 by weight. Determine the gas analysis in % volume. Solution: gas % weight CO2 3 H2 O N2 02

6 74 17

MW 44 18 28 32 total

moles .06818 .3333 2.6429 .5312 3.57563

% volume 1.91 9.32 73.91 14.86 100

Moles of a gas are obtained by dividing the weight by the molecular weight; moles of C02 = 3/44 = .06818. The volume of each gas, then, is the mole fraction x 100. Percent volume of 02 = (.5312/3.57563) x 100 = 14.86 and so on. One can work in reverse and convert from volume (or mole) basis to weight basis.

PROPERTIES OF STEAM Tables E-12, E-13 and E-14 give the properties of saturated and superheated steam and compressed water. Tables E-15 and E-16 give the equations for saturated and superheated steam properties.

~

- - - - - - - - - - - - - - - - - - - - - - - - - -

356

Waste Heat Boiler Deskbook

The author has developed a program which computes steam properties using the equations presented in Tables E-15 and E-16 for saturated and superheated steam and tabular data for compressed water. The results of the program are shown below for three cases. Example 6: Determine the following: 1. Properties of 85% quality and saturated steam at 1000 psia. 2. Superheated steam properties at 450 psia and 750°F. 3. Compressed water properties at 1500 psia and 300°F. Solution: For saturated or wet steam, the inputs are pressure and quality; temperature is inputted as zero and the program automatically computes the saturated steam properties. Results are shown in Figure E-25a. For superheated steam the inputs are pressure and temperature. The program determines whether the temperature is below or above saturation temperature and selects the appropriate subroutine, Figure E-25c.

REFERENCES 1. V. Ganapathy, "Applied Heat Transfer," Penwell Books, Tulsa. 2. R.c. Reid and T.K. Sherwood, "The properties of gases and liquids," second edition, McGraw Hill, New York, 1966. 3. North American Combustion Handbook, Cleveland, 1978. 4. V. Ganapathy, "Program computes turbine steam rates and properties," Hydrocarbon processing, November 1988, p 105. 5. V. Ganapathy, "To get heat transfer coefficients," Hydrocarbon Processing, Nov and Dec 1977. 6. V. Ganapathy, "Basic programs for steam plant engineers," Marcel Dekker, New York, 1984. 7. American Society of Mechanical Engineers, "Correlations for Superheated Steam Properties," 1967. 8. Gonzales-Pozo, V., "Formulas estimate properties for dry saturated steam," Chemical Engineering. May 12, 1986, p. 123.

r

(Conclusion - Table E-12)

110 120 130

334.77 341.25 347.32

0.01782 0.01789 0.01796.

4.049 3.728 3.455

305.7 312.4 318.8

883.2 877.9 872.9

1188.9 1190.4 1191.7

140 145 150

2.8886 3.281 3.718

0.01629 0.01632 0.01634

123.01 109.15 97.07

107.9 112.9 117.9

1014.1 1011.2 1008.2

1122.0 1124.1 1126.1

140 150 160 170 180

353.02 358.42 363.53 368.41 373.06

0.01802 0.01809 0.01815 0.01822 0.01827

3.220 3.015 2.834 2.675 2.532

324.8 330.5 335.9 341.1 346.1

868.2 893.6 859.2 854.9 85M

1193.0 1194.1 1195.1 1196.0 1196.9

155 160 165 170 175

4.203 4.741 S.g35 5.92 6.715

0.01637 0.01639 0.01642 0.01645 /0.01648

86.52 77.29 69.19 62.06 55.78

122.9 127.9 132.9 137.9 142.9

1005.2 1002.3 999.3 996.3 993.3

1128.1 1130.2 1132.2 1134.2 1136.2

190 200 250 300 350 400' 450

37·7.51., 381.79 " 400.95 ) 417:33 431.72

0.01833 0.01839 0.01885 0.01890 0.01913

2.404 2.288 1.8438 1.5433 1.3260

350.8

11~~

393.8 409.7

846.8 843.0 825.1 809.0 794.2

/1198.4'\ \.12.l!1.v 1202.8 1203.9

180 185 190 200 212

7.510 8.383. 9.339 11.526 14.696

0.01651 0.01654 0.01657 0.01663 0.01672

50.23 45.31 40.96 33.64 26.80

147.9 152.9 157.9 168.0 180.0

990.2 987.2 984.1 977.9 970.4

1138.1 1140.1 1142.0 1145.9 1150.4

444.59 456.28 467.01 488.21 503.10

0.0193 0.0195 '0:0197 0.0201 0.0205

1.1613 1.0320 . 0.9278 0.7698 0.6554

424.0 :-437:2 '; 449:4 471.6 491.5

780.5 767.4 755.0 731.6 709.7

1204.5

220 240 260 280 300

17.186 24.969 35.429 49.203 67.013

0.01677 0.01692 0.01709 0.01726 0.01745

23.15 16.323 11.763 8.645 6.466

188.1 208.3 228.6 249.1 269.6

965.2 952.2 938.7 924.7 910.1

1153.4 1160.5 1167.3 1173.8 1179.7

518.23 531.98 544.61 561.22 596.23

0.0209 0.0212 0.0216 0.0223 0.0235

0.5687 0.5006 0.4456 0.3619 0.2760

509.7 526.6 542.4 571.7 611.6

688.9 668.8 649.4 611.1 556.3

321.6 375.0 430.1 487.8 549.3

870.7 826.0 774.5

1192.3 1201.0 1204.6 1201.7 1190.0

$00 ....

iloo 700 800

900 1000 1200 1500

't

(~ 3Z§

\1204.61 1204.4 1203.2 1201.2 . 1198.6 \1 1195.4-r-

1191.8 1183.4 1167.9

350 400 450 500 550

134.63 247.31 422.6 680.8 1045.2

0.01799 0.01864 0.0194 0.0204 0.0218

3.342 1.8633 1.0993 0.6749 0.4240

113.9

640.8

~

'"

ii)

::r: (I)

!! I:I:l 0

.

r::: (I)

J':"''---

,'-

0

1&

g 0

"'"

i::s'"

e: )(

tt1

Table E-13. Properties of Superheated Steam.

I CJ 0:0

Vl

0:0

::s

Abe Pr..aure, psi (Sat lemp, F) 15 (213.03)

h

20 (227.96)

h

40 (267.25)

v v v h

v

60 (292.71)

h

80 (312.03)

h

100 (327.81)

h

150 (358.42)

h

200 (381.79)

h

250 (400.95)

h

v v v v v

~

Temperature 01 ateam, F

Sal liquid

Sal YBper.

400 33.97 1239.9

600

700

600

900

1000

1200

26.29 1150.8

300 29.91 1192.8

500

0.02 181.1

37.99 1287.1

41.99 1334.8

45.98 1383.1

49.97 1432.3

53.95 1482.3

57.93 1533.1

65.89 1637.5

0.02 196.2

20.09 1156.3

22.36 1191.6

25.43 1239.2

28.46 1286.6

31.47 1334.4

34.47 1382.9

37.46 1432.1

40.45 1482.1

43.44 1533.0

49.41 1637.4

0.017 236.0

10.498 1169.7

11.040 1186.8

12.628 1236.5

14.168 1284.8

15.688 1333.1

17.198 1381.9

18.702 1431.3

20.20 1481.4

21.70 1532.4

24.69 1637.0

0.017 262.1

7.175 1177.6

7.259 1181.6

8.357 1233.6

9.403 1283.0

10.427 1331.8

11.441 1380.9

12.449 1430.5

13.452 1480.8

14.454 1531.9

16.451 1636.6

0.018 282.0

5.472 1183.1

6.220 1230.7

7.020 1281.1

7.797 1330.5

8.562 1379.9

9.322 1429.7

10.077 1480.1

10.830 1531.3

12.332 1636.2

0.018 298.4

4.432 1187.2

4.937 1227.6

5.589 1279.1

6.218 1329.1

6.835 1378.9

7.446 1428.9

8.052 1479.5

8.656 1530.8

9.860 1635.7

0.018 330.5

3.015 1194.1

3.223 1219.4

3.681 1274.1

4.113 1325.7

4.532 1376.3

4.944 1426.9

5.352 1477.8

5.758 1529.4

6.564 1634.7

0.018 355.4

2.288 1198.4

2.361 1210.3

2.726 1268.9

3.060 1322.1

3.380 1373.6

3.693 1424.8

4.002 1476.2

4.309 1528.0

4.917 1633.7

0.0187 376.0

1.8438 1201.1

2.151 1263.4

2.427 1318.5

2.688 1371.0

2.942 1422.7

3.192 1474.5

3.439 1526.6

3.928 1632.7

fJl

it 0:0

S

~

0

l::r. '" Vl

(Continued) (jJ

V1

\0

'"

w ~

(Conclusion - Table E-13) 0 v

300 (417.33)

h

400 (444.59)

h

500 (467.01)

h

600 (486.21)

h

800 (518.23)

v v v ~v

h

v

1000 (544.61)

h

1500 (596.23)

h

2000 (635.82)

h

2500 (668.13)

h

3000 (695.36)

h

3206.2 (705.40)

h

v v v v v

,/1

0.0189 393.8

1.5433 1202.8

1.7675 1257.6

2.005 1314.7

2.227 1368.3

2.442 1420.6

2.652 1472.8

2.859 1525.2-

3.269 1631.7

0.0193 424.0

1.1613 1204,5

1.2851 1245.1

1.4770 1306.9

1.6508 1362.7

1.8161 1416.4

1.9767 1469.4

2.1t.! 1522.4

2.445 1629.6

0.0197 449.4

0.9278 1204.4

0.9227 1231.3

1.1591 1298.6

1.3044 1357.0

1.4405 1412.1

1.5715 1466.0

1.6996 1519.6

1.9504 1627.6

0.0201 471.6

0.7698 1203.2

0.7947 1215.7

0.9463 1289.9

1.3013 1462.5

1.4096 1516.7

1.6208 1625.5

0.0209 509.7

0.5687 1198.6

0.6779 1270.7

l-013~N.2"~ /1.1899 ~~V'! ;f>.11@."u';; 0.1.833 0.8763

1338.61j[..~.2@9~::S~··

0.9633 1455.4

1.0470 1511.0

1:2088 1621.4

0.0216 542.4

0.4456 1191.8

0.5140 1248.8

0.6084 1325.3

0.6878 1389.2

0.7604 1448.2

0.8294 1505.1

0.9615 1617.3

0.0235 611.6

0.2765 1167.9

0.2815 1174.5

0.3719 1287.2

0.4352 1363.8

0.4893 1429.3

0.5390 1490.1

0.6318 1606.8

0.0257 671.7

0.1878 1135.1

0.2489 1240.0

0.3074 1335.5

0.3532 1409.2

0.3935 1474.5

0.4668 1596.1

0.0287 730.6

0.1307 1091.1

0.1686 1176.8

0.2294 1303.6

0.2710 1387.8

0.3061 1458.4

0.3678 1585.3

0.0346 802.5

0.0858 1020.3

0.0984 1060.7

0.1760 1267.2

0.2159 1365.0

0.2476 1441.8

0.3018 1574.3

::t

0.0503 902.7

0.0503 902.7

0.1583 1250.5

0.1981 1355.2

0.2288 1434.7

0.2806 1569.8

I:l::I 0

NOTE: v = specific volume. cu ft/lb; h = enthalpy. BtU/lb

~

(Jl

iii /1)

!!:.

=:

(!l

0

~

~

~

J

.• ,i .'

Appendix E - Gas and Steam Properties

361

Table E-14. Properties of Compressed Water. seo (447.13)

p (l Sal.)

S~c. 31 SO ~OO

150

lQO

.016011 •. 01 ·.01 .016014 /8.06 /8.06 .016110 68.01 68.01 .0/610 117.91 111.91 .01661J 168.0J 168.01

150 ...01100) 300 .0'/1/H 3SO .016000 400 .01.'666, 450 .0/9101

SOO 51G S10 SJO 540

sse 560 510 5,80 S9G

1000 (544.7H

•

h

u

·.00001 .01667 .//961 .11 JQ.( .1UOl

.019748 <M7.70 "9.53 .64904

.021591 538.39 542.)6 .743

.00 1.49 .015994 .OI5~8 11.02 19.50 .016106 61.81 69.36 .016318 111.66 119.11 .016608 161.65 1069.19

.00000 .03599 .12932 .21451 .29341

.015967 .03 2.'9 .015972 17.99 20.94 .016082 61.70 10.68 .016293 111.38 120.40 .016580 161.26 110.32

,

.000 .035 .129' .214 .292'

116.J1 169.6 I 11I.J9 JlI.81 119.96

1I,K:Jl 169.61 11I.J9 111.8J 119.'6

.16117 .11111 .JOJJ9 J6UQ .619111

.016972 .011416 .011954 .018608 .019420

21'1.99268..9'2: 320.11 )73.68 428.40

2-19.56 270.53 322.31 375.40 430.19

.36102 .43641 .50249 .56604 .62198

.016941 .011319 .011909 .018550 .019340

.01060 .01087 .01116 .01118 .01181

186.1 JOO.l Jll.1 J1J.J JJ8.6

IoU. I JOO.l Hl.1 J1J.J JJ6.6

.6919 .1016 .71 71 .1101 .1111

.0101.' .01071 .01100 .Ql110 .01/61

U'J.9 1''1>1.9 HO.I J11.6 JJ1.1

4Rl.8 199.8 J/l.0 11U 117.1

.6896 .1011 .1116

483.8 9;11' 495.6 501.6 511.5 519.9 523.8 532.4 , S36.3

.68; .6!X .1 Ii

.1101

.02036 .02060 .02086 .02114 .02144

.01111 .01161 .011IJ

"1.1 166.1 180.8

1H.1 166.1 180.8

."69 .1707 .7SJ/

.0119.' ,01111 ,011RI .0])11 .0])91

H.!.l 161.0 "6,0 J90.R 606.1

JJO.1 161.0 "8.1 J91,9 60.'.6

.1Hl .7666 .7801 .7916 .8096

.011 11 .01111 .011H .01198 .01119

JO.I JJ8.1 J1/ .6 18J,9 600.6

Jl9.1 J6U "6.0 190.1 601.9

.119 .161 .116 .T69 .801

.01109 .OUSI

616,1 6J1.9

610.6 6Jl.J

.818 .611

.7111

600 61G

p (l Sal.)

1500 (596.39)

1000 (636.00)

•

211.41 268.24 319.83 312.SS 426.89

220.61: .366.

211.46) 0435:

323.15 ,SOl· 315.98 .564: 430.41 .626:

cmP'

.724 .731

1500 (668.31)

h

h

s.c.

.023461 604.91 611.48 .80824

.025649 662.40 611.89 .86221

.028605 111.66 130.89 .9130,

3~

.015939 .05 4.41 .00001 .015946 17.95 22.38 .03584 .016058 67.53 71.99 .12870 .016268 117.10 121.62 .21364 .016554 166.87 171.46 .29221

.015912 .06 5.95 .015920 17.91 23.81 .016034 67.31 73.30 .016244 116.83 122.84 .016527 166.49 172.60

.00008 .03515 .12839 .21318 .29162

.0 I 5885 .08 1.43 .015895 17.88 25.23 .016010 67.20 74,61 .016220 116's6 124.07 .016501 166.11 173.75

.016880 .011308 .011822 .018439 .019191

216.46 266.93 318.15 310.38 424.04

222.10 213.33 324.14 317.21 431.14

.36482 .43316 .49929 .56216 .62313

.016851 .011214 .011180 .018386 .019120

215.96 266.29 311.33 369.34 422.68

.68S3 .6914 .1096 .1219 .7343

.02014 .02036 .02060 .02085 .02112

419.8 491.4 503.1 514.9 521.0

481.3 498.9 510.1 522.6 534.8

.6832 .6953 .1013 .7311

.02004 .02025 .02048 ,02012 .02098

418.0 489.4 501.0 512.6 524.5

481.3 498.8 510.4 522.2 534.2

.7469 .7596 .1125 .1851 .1993

.02141 .02112 .02206 .02243 .02284

539.2 551.8 564.6 517.8 591.3

541.2 559.8

.021lS .02154 .02186 .02221 .02258

536.6 548.9 561.4 514.3 581.4

546.4 558.8

586.1 599.8

.1440 .1565 .1691 .1820 .1951

584.5 597.9

.7413 .75J6 .7659 .7185 .7913

.8111 ,8181 .801 ,8609

.02330 .02382 .02443 .02514 .0160)

605.4 620.0 635.4 651.9 669.8

614.0 628,8 644.5 661.2 679.1

.8086 .8225 .8371 .8525 .8691

.02300 .02346 .02399 .02459 .02530

601.0 615.0 629.6 644.9 661.2

611.6 625.9 640.1 656.3 672.9

.8041 .8171 .8J1S .8459 .8610

.01114

690.1

700.1

.8UI

.02616 .02729 .0189J

618.7 698.4 711.1

690.8 111.0 7JJ.J

.8773 .8954 .9/11

50 100

ISO 100 2SO

.016910 .011343 .011865 .018493 .019264

216.96 261.58 318.98 311.45 425.44

510 SlO 530 540

.02024 .02048 .Q2012 .02099 .02127

481.8 493.4 50D 51D 529.6

550 560 510 580 590

.02158 .02191 .02228 .02269 .02314

542.1 554.9 568.0 581.6 595.7

600 610 610 630

.01166 .01416 ,ON .01 J90

610,1 61J .•' 61},J

)00

350 400 450

sao

6~0

650 660 670

9.'

~60,8

~.36554 ~2.39

.43463 3B. 4 .50034 376.59 .56343 430.19 .62470

Cl8l]l 499.1 SIl.O 523.1 535.5 548.1 S61.0 574.2

S87.9 602.1 616,9 ~J1.6

619.1 668.0

S12.8

.1I9S

223.15 214.28 J25.56 317 .84

.0000' .0356· .1280: .2127: .2910· .36411

043291

.49821 .5609: 431.52 .62161

S11.S

.681:

.693; .1051 .1111 .7292

Table E-15. Formulas to estimate properties of dry, saturated steam.

Equation used: Y = Ax + Blx + Cxll2 + Dlnx + Ex2 + Fil + G y = property A to G are constants as given below in Table x = pressure, psia TABLE-Coefficients to predict saturated-steam properties Property Temperature, OF liquid specific volume, ft3/lb Vapor specific volume, ft3/lb 1 to 200 psia 200 to 1,500 psia liquid enthalpy, Btu/lb Vaporization enthalpy, Btu/lb Vapor enthalpy, Btu/lb Liquid entropy, Btul(lb)(OR) Vaporization entropy, Btu/(lb)(OR) Vapor entropy, Btu/(lb)(OR) Liquid internal energy, Btu/lb Vapor internal energy, Btullb

A

o

E F G 8.762969 x 10-5 - 2.78794 X 10-8 86.594 14 10 8.408856 x 101.86401 X 100.01596

- 0.17724 - 5.280126 x 10- 7

B 3.83986 2.99461 x 10-5

C 11.48345 1.521874 x 10-4

31.1311 6.62512 x 10-5

- 0.48799 2.662 x 10- 3 - 0.15115567 0.008676153 -0.14129 -1.6m2 x 10-4

304.717614 457.5802 3.671404 -1.3049844 2.258225 4.272688 x 10-3

9.8299035 - 0.176959 11.622558 - 8.2137368 3.4014802 0.01048048

-16.455274 0.826862 30.832667 -16.37649 14.438078 0.05801509

9.474745 x 10-4 - 4.601876 x 10-7 8.74117 x 10-5 - 4.3043 x 10-5 4.222624 x 10-5 9.101291 x 10-8

-1.363366 X 10-6 6.3181 X 10- 11 - 2.62306 X 10-8 9.763 X 10- 9 - 1.569916 X 10-8 -2.7592 x 10- 11

19.53953 - 2.3928 54.55 1,045.81 1,100.5 0.11801

3.454439 x 10-5 -1.476933 X 10-4 - 0.1549439 - 0.0993951

- 2.75287 X 10-3 1.2617946 X 10-3 3.662121 1.93961

- 7.33044 X 10-3 3.44201 X 10-3 11.632628 2.428354

- 0.14263733 - 0.08494128 30.82137 10.9818864

- 3.49366 X 10-8 6.89138 X 10-8 8.76248 X 10-5 2.737201 x 10-5

7.433711 X 10- 12 - 2.4941 x 10- 11 - 2.646533 x 10-8 -1.057475 x 10-8

1.85565 1.97364 54.56 1,040.03

Appendix E - Gas and Steam Properties

Table E-16. Correlations for superheated steam properties.

CI C2 C3

C4 Cs

= 80,8701T2 = ( - 2641.6211') X lOCI

= 1.89 + Cz = =

C3(P2ITl) 2 + (372420IT2)

Co = CSC2 Ci = 1.89 + Co Cs = 0.21878T - (126,970IT) Cg = 2CSC7 - (C3IT)(126,970) CtO = 82.546 - (162,460IT) CII = 2ClO C7 - (C3IT)(162.460) Using the variables CI through Cll , u, H. and S have been evaluated. u = {[(CS C4 C3 + ClO)(C/P) + l]C3 + 4.55504

H

=

(TIP)}0.016018 775.596 + 0.63296T + 0.000162467T2 + 47.3635 log T + 0.043557{C7P + 0.5C4 [C II + C3(C lO

+ c9 QJ) 1IT{[(CSC3 - 2C9 )C3C/2 - Cll ]C/2 + (C3 - C7 )P} ( - 0.0241983) - 0.355579 - 11.42761T + 0.00018052T - 0.253801 log P + 0.809691 log T Where P = pressure (atm) T = temperature (K) u = specific volume (ft3lIb) -: H = enthalpy (Btullb) S = entropy (BtullbOF)

S

=

363

TO GET HEAT TRANSFER COEFFICIENTS 0.21

----,---"""T----r---.,...-----.

w ~

0.24 ......

0.20

0.19 0.211----.:tr-_+-_ _ _+-_ _ _+ - -_ _+-_ _--I

~

0.18

~

":

§ 0 •

0.17

~

0

0.16

0.15

t-+-----,f-i--,---+---+---+------1

0.17 I-----+-=""'----,,<'I------+----+_

:E 0> rn nr

0.161--_---,;L+-_ _ _+-_ _ _+-_ _ _+-_ _---l

0.14 t - - + - - - t - - - - t - - - - t - - - - t - - - - - j

::r: (I)

~ O:l 0

=:

... (I)

0.15 L...._~--J_ 200 o

___L_--L_

400

__'__

_ ' _ _........_

600

Temperature, of

......._..I.____J

600

1,000

0.13 "--___"'-----'_--L_--'-_ o 200 400

__'__........_-'-_-'--_.l..----l

600

Temperature. of

BOD

1,000

0(I)

g

"'

~

Figure E-l. Air

Figure E-2. Oxygen

0.22 ...-..--_......,_ _ _

~

_ _ _-.-_ _ _-.-_ _---,

2 . 4 , . . - - - - , - - - - . -_ _r--_ _-.._ _----,

0.21 1-_+_-+_ _ _--1f-_ _..:.P-=a"-lra~m..:.e"'t.::.er...:s.:..:.:..P",re=+ss..:.u:::re..::.:..,:::al:::m..:.s"""-l' 2.3 t----I----+---+---+--~

0.20 Hr--~d----+----+----~----..,."e:::2.2r-...:p...:a=r:::am~m=efrs=:..:.p..:.re=s=s=u=re~·fa=tm=s=.-~~~~-_+-_ _~

0.19 1-\-_-+_+-_ _ _+-_ __

2.1 t----h~h'~---+---+----I 0.18 1---~"+__-::..--+-_r__4-----1---___I

0.17 f----+--+-+----+----+-----1

0.18 '--_-'-_.1.-_-'--_.1.-_.1.-_-'--_-'--_.1.-_.1.-----1 1,000 o 200 800 400 600

2.0

I-T-I-.H4-----f---+---I-----I

1.9 E..-----I_......L_...L_L---1._-L_.L_L-L_-' o 200 400 600 BOO 1,000

Temperature, of

Temperature, of

Figure E-3. Nitrogen

Figure E-4. Hydrogen

O~~----~------~----~~-----r------' 1.0 0.9 0.8 0:;

-

0.22 H . . - - - / - - - - + - - - + - - - - + - - - - I

-

0.21 1 - - - - \ - - ; - - - _ + - - - - + - - - - + - - - - 1 Parameters: Pre8lllure, atms.

0.6

0.20 I---rl---_+---+----+----:::;~

0.5

... 0 ...0~

0.19 1----1--"'...::--1--0.4

Q.

()

'"

0

0.3

:.!

i\\

\\

II

0

0.2

0.17 1:-------II----:::;."......I--7.c:------i~---;---_'_I

~

\~ r---..... i-

0.1

0.18 1-~--;---_+-7""'--7I""'_7'c---+---_I Parameters: Pressure, alms.

-

V

a

V\

""

'~ ~

I

200

-

250

I

400

I I

600

I----"...'-----+-----+----+------l

0.151----1----1----1-------11----;

I

800

0.18

1,OqO

o

200

400

600

600

Temperature, of

Temperature, of

FiQ"ure E-S. Carbon dioxide

Figure E-6. Carbon monoxide

1,000

4.0 3.0 2.0

----

1.0

\

~

\

0.8

...ci

-l!:.

~

0

Co

'":;;" ""II 0

0.6 0.5

\ \ \ \

1\ \ \

Parameters: Pressure, alms.

0.4 r\ \ \ 0.3

-

\\

0.2

0.1 0.08

--

0.06 0.05 0.04

"'-."

\ '&"-av,~""-

-

-- --

I~ ~ ~...

'

0.30 hf----t----t-------1r--------1------1

-

0.03

o

I

I

200

I

400

I 600

Temperature, OF

Figure B-7. Sulfur dioxide

I 800

0.25

1,000

'--_'--_'--_'----.JL-.--.JL-.--.JL-.--.JL-.---l_---l_---l

o

200

400

600

Temperature, OF

Figure B-S. Methane

800

1,000

/ Charts give heat-transfer coefficients considering pressure effect / 2.5...-_ _--.._ _ _-.-_ _-;r-_ _,-_ _--,

0.36 0.34

1--\----1------+---- -- --- -.--

0.32 0.30

I+--\---\--i----+---=-.-~~~ -~--=t~~~

0.28

~

\..

,;

...

0.26

ci

0.24

Parameters: Pressure, alms .

Il

~ I

Paramelers: Pressure. alms.

~~ ci

0

.K

...~ -'"

0.22

•

()

o

0.40 I--+_ _~ --~+-----t------;0.35 \---''---+- ' - - - - ; . . - - - - t - -

0.20

0.18

0.30

\-----+--........--t--"'....- - t

0.25

I--_ _ _~.

:E I»

0.20 1-----+--'''=---+

fIl

;0.16

::r: It>

0.15 h-----I---:~""__l.

~

O:l

g. 0.14 L---I---L---L_-L_-L_...l.-_..L-_L-_L...---l o 200 400 600 800 1,000 Temperature, of

Figure E-9. Nitric oxide

t!i

o

200

400

600

800

Temperature, ° F

1,000

I ,I

~

~ 0

~

Figure E-IO. Nitrogen peroxide

II

1:1

!

j <\,

<\

5.0 1.0 4.0

0.90

0.80 3.0 0.70

0.60 0.50

0.40

~

11l

0.30

0.20

~ -il

.

1'1 \\

0.35

0.25

2.0

ci

Q.

Parameters: Pressure, alms.

ll\

-\2

0.15

-"

~

•

0

~25

--=::::;:;

\ \

\

0.8

0.8

\

\

""

200

400

_1

I

600

800

Temperalure, OF

Figure E-ll. Nitrous oxide

0.2

1,000

V o

..........250

-

""' -100 40~

,,--

t-

I

""-

~

0.3

...1

---_.-

'\.

~

I-

---

Parameters; Pressure, atms.

\

\ \ ~

0.7

,

.--- ---_._- f - - - -

\\\

0.4

V"'--::"-

I

o

1.0 0.9

~

0.5

"'--10040 ;:::::--

V 0.10

~

Q

-----

~\-

F:::::= ::::::-r---

--

'" I

I 200

...1 400

I

600

Temperature, OF

Figure E-12. Ammonia

I

800

1,000

I

I

TO GET HEAT TRANSFER COEFFICIENTS

0.13 , - - - - - r - - - - r - - - - r - - - - r - - - - - ,

0.13 n - - - - , . - - - - r - - - - , . - - - - , . - - - - - , 0.12 ~---1_---1_---1____:7_"S4rc,...L-""7"'~

0.12 0.11 ~

'"ci

; l ci

[;;

0.111---"'<-_

ci

; ci

0.10 f-:::,..L-------:.IL----7f-------!-----!-----l

~ II lL

C.

,;;;0 ci

x II

lL

Parameters: Pressure, atms. 0.09 I+---~I_---I_---r----r---------l

0.10 I-----i----.'~

0.08 r.--f---t-----t-----t-----t----------l

~

go

;'"

0.09

:::r:: (I) ~ t:I:l

0

....== (I)

o

200

400

600

Temperature, 0 F

0.07 L - _ L - _ L - _ . . . I - _ L - _ L - _ L - _ L - _ L - _ L - - - - I 1000 600 800 400 200 o Temperature, of

0

t 0

l>';'

Figure E-13. Air

Figure E-14. Oxygen

/ Charts give heat-transfer coefficients considering pressure effect / 2.5...-_ _--.._ _ _-.-_ _-;r-_ _,-_ _--,

0.36 0.34

1--\----1------+---- -- --- -.--

0.32 0.30

I+--\---\--i----+---=-.-~~~ -~--=t~~~

0.28

~

\..

,;

...

0.26

ci

0.24

Parameters: Pressure, alms .

Il

~ I

Paramelers: Pressure. alms.

~~ ci

0

.K

...~ -'"

0.22

•

()

o

0.40 I--+_ _~ --~+-----t------;0.35 \---''---+- ' - - - - ; . . - - - - t - -

0.20

0.18

0.30

\-----+--........--t--"'....- - t

0.25

I--_ _ _~.

:E I»

0.20 1-----+--'''=---+

fIl

;0.16

::r: It>

0.15 h-----I---:~""__l.

~

O:l

g. 0.14 L---I---L---L_-L_-L_...l.-_..L-_L-_L...---l o 200 400 600 800 1,000 Temperature, of

Figure E-9. Nitric oxide

t!i

o

200

400

600

800

Temperature, ° F

1,000

I ,I

~

~ 0

~

Figure E-IO. Nitrogen peroxide

II

1:1

!

j <\,

<\

5.0 1.0 4.0

0.90

0.80 3.0 0.70

0.60 0.50

0.40

~

11l

0.30

0.20

~ -il

.

1'1 \\

0.35

0.25

2.0

ci

Q.

Parameters: Pressure, alms.

ll\

-\2

0.15

-"

~

•

0

~25

--=::::;:;

\ \

\

0.8

0.8

\

\

""

200

400

_1

I

600

800

Temperalure, OF

Figure E-ll. Nitrous oxide

0.2

1,000

V o

..........250

-

""' -100 40~

,,--

t-

I

""-

~

0.3

...1

---_.-

'\.

~

I-

---

Parameters; Pressure, atms.

\

\ \ ~

0.7

,

.--- ---_._- f - - - -

\\\

0.4

V"'--::"-

I

o

1.0 0.9

~

0.5

"'--10040 ;:::::--

V 0.10

~

Q

-----

~\-

F:::::= ::::::-r---

--

'" I

I 200

...1 400

I

600

Temperature, OF

Figure E-12. Ammonia

I

800

1,000

I

I

TO GET HEAT TRANSFER COEFFICIENTS

0.13 , - - - - - r - - - - r - - - - r - - - - r - - - - - ,

0.13 n - - - - , . - - - - r - - - - , . - - - - , . - - - - - , 0.12 ~---1_---1_---1____:7_"S4rc,...L-""7"'~

0.12 0.11 ~

'"ci

; l ci

[;;

0.111---"'<-_

ci

; ci

0.10 f-:::,..L-------:.IL----7f-------!-----!-----l

~ II lL

C.

,;;;0 ci

x II

lL

Parameters: Pressure, atms. 0.09 I+---~I_---I_---r----r---------l

0.10 I-----i----.'~

0.08 r.--f---t-----t-----t-----t----------l

~

go

;'"

0.09

:::r:: (I) ~ t:I:l

0

....== (I)

o

200

400

600

Temperature, 0 F

0.07 L - _ L - _ L - _ . . . I - _ L - _ L - _ L - _ L - _ L - _ L - - - - I 1000 600 800 400 200 o Temperature, of

0

t 0

l>';'

Figure E-13. Air

Figure E-14. Oxygen

0.13 1.25

0.12

1.20

1.15

:;; d

0.11

"-

if d

~

;

C.

..."

d

~

.l<

C.

r;;"

II

"-

1.10

':i.:

0.10

II

Parameters: Pressure. atms.

1.05

"-

1.00

Parameters: Pressure. atms.

0.09 0.95

0.08 L-_'--_'--_'--_'--_'---'-.l-._.l-._.l-._.L.---I 800 1000 400 600 o 200 Temperature. OF

Figure E-15. Nitrogen

0.90 '--_'--_'--_'--_'--_'--_'--_'--_'--_'-----' o 200 400 600 1000 Temperature. of

Figure E-16. Hydrogen

0.6 0.13

0.5

0.4

r..----,-----,-----,----,------,

\-----i-------+----~---

0.12 r-'\---+----t------+--___,,,,c---1':T'-r7"+-l 0.3

0.11 1-\-_ _ _+-_ _ _-+ ~

'"a

; ci

0.2

--- ----------- ---------- - - - - - - - j

a.

:;;0 ci

-'" II

0.10 r",.,----l7"'----:r---t-:r-----+------jr-------j

u..

0.1 Parameters: Pressure. atms.

0.09 1_--\-___+_ Parameters: Pressure. atms.

O.OB

f-----+_

0.07

0.09 1-----¥----~-----+-----1r-------j

10-£....--+----,-----,----1 1 - - - - - .----------- . _ -

0.06 0.08

0.05

o

200

400

600

BOO

Temperature. of

Figure E-17. Carbon dioxide

1000

L-_L-_L-_1-_1-_1-_1-_1-_1-_.L---l

o

200

400

600

BOO

Temperature. OF

Figure E-18. Carbon monoxide

1000

0.8 0.7 0.6 0.5

"-

r\ \

Co

-\ \ \

0.3

-

0.

0.15

0

-"

I rn po

if po

II·

"-

0.1 0.09 0.08

\

0.07

"-

0.04

0.03

V o

/

--

---::..

'-..

.....- .....

0.06 0.05

!3

~

\$~\~

-,

:l Co

0.30

\' ~\

0

tI1

C'l po

\\\

0.2

x'

0.35

0.4

~

r;,0

i;g

1\ \

,... ~ 0

0.40

\

/

t 0

0.

r;,0

~ 0

"0 ~

0.25

g-

rn

o

-"

-

II

"-

--

:::::---

~

Parameters: Pressure, atms.

0.20 f-------,4-~-+-+---_+---__t---__l Parameters: Pressure, atms. 0.19 I---==-----,¥-F---+-----r----r-----j 0.18 I----.,I-.-F-----+----+----+----J 0.17 I__--/----,f---I----~~----.I__-----jl__---I

I

I

200

,

I

I

400

600

0.16

800

1000

IC--F---+-

o

200

---If-----+----+----I

400

600

Temperature, OF

Temperature, OF

Figure E-19. Sulfur dioxide

Figure E-20. Methane

800

1000

/ Charts give heat-transfer coefficients considering pressure effect in crossftow situations / 0.20

4.0

0.19

3.0

0.18 0.17

!\

\

0.16

\

0.15

~

!,f-" ~

0.14 0.13

'1 u.

2.0

0.12

0.11

0.10

0.09

f\

1.00

\

0.8

""

\

---

~

~~

250

~ ~~ ~' --~ V. !'IQ

---

-V

o

0.6

?

\\,\

0.3

~

0.08

400

I

600

Tempereture, of

Figure E-21. Nitric oxide

800

~

1000

'\

~./

I

I

0.04

I

o

'---"

\ r-.....

0.1

I

~

i\ '" r--....

I-

Parameters: Pressure, elms.

I

200

~~\ \\ \

0.4

0.08 0.05

,

0.06

::0.

Parameters: Pressure, alms.

0.2

/'

V

0

...•

toO . . /

'-

I;

~ P\.\i

200

~

:E I»

~

en

Ii)

::r: ~

~ IJ:l 0

I::::

I

400

.

___c·

I

600

III

I

800

Temperature, of

1000

tj ~ en

~

0

~

Figure E-22. Nitrogen peroxide

j

/ft·

0.7

5.0 4.0

0.6

3.0 0.5

2.0 0.4

~

~

;. 0

l

1.

0.2.

........

0.09

0.07

y

V

~ \\

~

s~

0.8 , - - 0.5

1\

00.

0.4

'it

'l.

•

- ~~ ~~

-

0.1

~

" ~

\:

\

0.2

~

\.

1\

0.3

I&-

~

Parameters; Pressure, alms.

~

~.

~ ~ ~~ -

0.1

1.00 0.8

~

I

...

0.08

Paramelers; Pressure, alms.

\

0.3

~

~

V-

/

-='

.............

.....

1/ I

0.08

o

400

800

I

I

I

I

I

200

800

1000

o

I

r

I

400

800

Temperalure, of

Tamperature, of

Figure E-23. Nitrous oxide

Figure E-24. Ammonia

I 800

1000

Waste Heat Boiler Deskbook

376

Figure E-25. Results from steam properties program. Saturated Steam Properties PSIA 1000

TSAT-F 547

HLIQ 545.1

HVAP 1193.0

VLIQ 0.0216

WAP 0.4456

SLIQ SVAP 0.745'84 1.39201

Mixture properties Quality - % = 85 Enth = 1095.848 SP vol = .3819935 Entrpy =1.295087

H - Enthalpy in Btu/lb, V = Sp vol in eu ft/lb, s = entropy in Btu/lb F... VAP refers to vapor and LIQ to water phase ...

a Superheated Steam Properties PSI A TEMP HV AP 450 750 1387.2 HVAP = Enthalpy in Btu/lb, VVAP =SP Vol in eu ft/lb, SVAP = Entropy in Btu/lb F .

VVAP 1.53194

b

Compressed Water'Sp Vol & Enthalpy PRES = 1500 TEMP = 300 ENTH =272.4067 SP VOL =1.734667E-02 e

SVAP 1.648026

Appendix F

Tube Thickness Calculations Determining thickness of tubes is an important aspect of boiler design. Cost, weight and fluid pressure drop (if water tube type boiler) are influenced by size of tubes. Tubes of fire tube boilers are subject to external pressure; tubes of water tube boilers and equipment such as economizers, superheaters are subject to internal pressure. ASME code sections 1 and 8 provide guidelines for estimating the tube thickness. While using the code formula one should keep in mind the fact that the allowable stress values for the particular tube material should be evaluated at the design temperature. If several combinations of tube materials, sizes are used as in, say, superheaters, then it is likely to have different tube thicknesses along its length. Table F-1 gives the allowable stress for a few commonly used boiler tube materials. The latest edition of the code should be referred to for the allowable stress stress values for a given grade of material.

TUBES AND PIPES SUBJECT TO INTERNAL PRESSURE Tubes are specified by the outer diameter and minimum wall thickness, while pipes are specified by the nominal diameter and average wall thickness. Table F-2 gives dimensions of steel pipes. A few examples illustrate the computation procedure. Example 1: Determine the thickness of the tubes required for a hoiler superheater. The material is SA 213 T 11; the metal temperature is 900°F, (see chapter 3 for discussions on tube wall temperature and 377

Waste Heat Boiler Deskbook

378

method of computing the same); and the tube outer diameter is 1.75 in. The design pressure is 1000 psig. Solution: Per ASME Boiler and Pressure Vessel Code, Sec. 1, 1980, Para. pg. 27, the following equation may be used to obtain the thickness or the allowable pressure for tubes.

tw=~+0.005 +e 2S. +p

(F-1)

p _ S{ 2tw - O.Old - 2e ] d - (tw - 0.005d - e)

(F-2)

where = minimum wall thickness, in. P = design pressure, psig d =tube outer diameter, in. e = factor that accounts for compensation in screwed tubes, generally zero Sa = allowable stress, psi tw

From Table F-l, Sa is 13,100. Substituting in Equation (F-1) yields

Table F-l. Allowable Stress Values Ferrous Tubing, 1000 psi Temperatures not exceeding (OF): Material 20 to Specifications 650 800 750 800 850 900 950 10001200 1400 SA 178gr A

10.0 9.7 9.0 7.8 12.8 12.2 11.0 9.2 11.8 11.5 10.6 9.2

6.7 5.5 3/8 2.1 7.4 5.5 3.8 2.1 7.9 6.5 4.5 2.5

SA 1929rC SA 210 gr A1 SA 53 B 15 14.4 13.0 10.8 8.7 6.5 17.5 16.6 14.8 12.0 7.8 5.0 grC

4.5 3.0

2.5 1.5 (Continued)

.,J

379

Appendix F - Tube Thickness Calculations

Table F-1. (Cont'd) Allowable Stress Values Ferrous Tubing, 1000 psi Temperatures not exceeding (OF):

Material 20 to Specifications 650 800 750 800 850 900 950 1000 1200 1400 SA 213 T 11, P 11 15.0 15.0 15.0 15.0 14.4 13.1 11.0 T22, P 22 15.0 15.0 15.0 15.0 14.4 13.1 11.0' T9 13.4 13.1 12.5 12.5 12.0 10.8 SA 213 TP 304H 15.9 15.5 15.2 14.9 14.7 14.4 TP316H 16.3 16.1 15.9 15.7 15.5 15.4 TP321 H 15.8 15.7 15.5 15.4 15.3 15.2 14.7 14.7 14.7 14.7 14.7 14.6 TP347H

7.8 7.8 8.5

1.2 1.6

13.8 15.3 14.0 14.4

6.1 7.4 5.9 7.9

2.3 2.3 1.9 2.5

Source: ASME, Boiler and Pressure Vessel Code, Sec. 1, Power Boilers, 1980. tw::: 1000 x 1.75 + 0.005 x 1.75 = 0.073 in. 2 x 13100 + 1000 The tube with the next higher thickness would be chosen. A corrosion allowance, if required, may be added to two

MAXIMUM ALLOWABLE PRESSURE FOR PIPES Example 2: Determine the maximum pressure that a SA 53 B carbon steel pipe of size 3 in. schedule 80 can be subject to at a metal temperature of 550°F. Use a corrosion allowance of 0.02 in. By the ASME Code, Sec. 1, 1980, pg. 27~ the formula for determining allowable pressures or thickness of pipes, drums and headers is tw = Pd +c (F-3) 2SaE + 0.8P

Waste Heat Boiler Deskbook

380

Table F-2. Dimensions of Steel Pipe (IPS) Nominal

pi~e 6ize,

I S, in. ~

OD, in.

Schedule No.

----40· 0.405

ID, in.

Flowarca per pipe, in.'

;Surrace per lin It, It,IfCt. Outside

Inside

Weight per lin It, 'Ib steel

sot

0.269 0.215

0.058 0.036

0.106

0.070 0.056

0.25 0.32

}1

0.540

40· sot

0.364 0.302

0.104 0.072

0.141

0.095 0.079

0.43 0.54

~

0.675

40· sot

0.493 0.423

0.192 0.141

0.177

0.129 0.111

0.57 0.74

H

0.840

40· sot

0.622 0.546

0.304 0.235

0.220

0.163 0.143

0.85 1.09

X

1.05

40' sot

0.824 0.742

0.534 0.432

0.275

0.216 0.194

1.13 1.48

1

1.32

40· SOt

1.049 0.957

0.864 0.718

0.344

0.274 0.250

1.68 2.17

1}1

1.66

40· sot

1.3SO 1.278

1.50 1.28

0 . 435

0.362 0.335

2.28 3.00

1~

1.90

40· sot

\.610 1.500

2.04 1. 76

0.498

0.422 0.393

2.72 3.64

2

2.38

40· sot

2.067 1.939

3.35 2.95

0.622

0.542 0.508

3.66 5.03

2~

2.88

40· sot

2.469 2.323

4.79 4.23

0.753

0.647 0.609

5.SO 7.67

3

3.50

40· sot

3.068 2.900

7.38 6.61

0.917

0.S04 0.760

7.58 10.3

4

4.50

40' sot

4.026 3.826

12.7 11.5

1.178

1.055 1.002

10.8 15.0

6

6.625

40· sot

6.065 5.761

28.9 26.1

1.734

1. 590 1.510

19.0 28.6

8

8.625

40· sot

7.!l81 7.625

50.0 45.7

2.258

2.090 2.000

28.6 43.4

78.8 74.6

2.814

2.62 2.55

40.5 54.8

3.338 3.665 4.189 4.712 5.236 5.747 6.283

·3.17 3.47 4.00 4.52 6.05 5.56 6.09

43.8 M.6 62.6 72.7 78.6 84.0 94.7

10

10.75

40· 60

10.02 9.75

12 14 16 18 20 22 24

12.75 14.0 16.0 18.0 20.0 22.0 24.0

30

12.09 13.25 15.25 17.25 19.25 21.25 23.25

30

30 20: 20 20: 20

*t Commonly known as

115 138 183 234 291 355 425

standard. Commonly known as extra heavy. t Approximately.

r--------Appendix F - Tube Thickness Calculations

381

Where E is the ligament efficiency; it is 1 for seamless pipes. c is the corrosion allowance. From Table F-2, a 3-in. schedule 80 pipe has an outer diameter of 3.5 in. and a nominal wall thickness of 0.3 in. Considering the manufacturing tolerance of 12.5%, the minimum thickness available is 0.875 x 0.3 = 0.2625 in. Substituting Sa = 15,000 psi (Table F-1) and c = 0.02 in. Equation (F-3), we have 0.2625 = 3.5P + 0.02 2 x 15,000 + 0.8P Solving for P, we have P = 2200psig. For alloy steels, the factor 0.8 in the denominator would be different. The ASME Code may be referred to for details. A simplified approach is seen below. Using standard pipe dimensions and Equ. (F-3), the maximum allowable pressure for each pipe was arrived at. One need not refer to the code for stress values. In addition, specific pipe dimensions (O.D. and thickness) need not be known, as the schedule number and pipe material are all that are necessary. Table F-3 gives the maximum allowable working pressure, in psig, based on an allowable stress of 15,000 psi. To correct for the specific material and temperature, multiply the value obtained from table by the appropriate F factor from Figure F-l. Table F-4 lists the various materials by composition and ASME code specification. Example 3: Estimate the maximum allowable working pressure for a 2-in., Schedule 40 carbon-manganese (SA 53B) pipe at 750°F. Solution: From Table F-3, a value of 1,782 psi is obtained. From the figure, for SA 53B (Curve 1) at 750°F, F is 0.86. Hence, the maximum allowable pressure is 0.86 x 1,782 = 1,532 psig.

DESIGNING VESSELS AND TUBES FOR EXTERNAL PRESSURE The standard procedure for designing pressure vessels and tubes (or pipes) subject to external pressure is tedious and time-consuming,

382

Waste Heat Boiler peskbook

1.0

0.9

0.8

0.7 F

0.6

0.5

0.4

, See Table

for key

650

Temperature, of Figure F-l. Factor F corrects for temperature. [Chemical Engineering]

Table F-3. MaXimum alalowable pressure* Nominal pipe size, in. Schedule 40 %

% 1 1% 2 2% 3 4 5 6 8

4,830 3,750 2,857 2,112 1,782 1,948 1,693 1,435 1,258 1,145 1,006

Schlldule 80 Schedule 160 6,833 5,235 3,947 3,000 2,575 2,702 2,394 2,074 1,857 1,796 1,587

• Based on allowable stress of 15,000 psi

6,928 5,769 4,329 4,225 3,749 3,601 3,370 3,191 3,076 2,970

~~~~~~~~~-~~~-

------------

383

Appendix F - Tube Thickness Calculations

Table F-4. ASME specifications for various steels

Composition Carbon-manganese Carbon-manganese-silicon CarbQn-%Mo (%%molybdenum) 5Cr-%Mo (5% chromium, Y:.% molyl:!denum) 9Cr-1Mo 1%Cr-Y:.Mq 2%Cr.1Mo

Code designation

Curve on figure

SA 538 SA 1068 SA 335 P1

2

SA 335 P5 SA 335 P9 SA335P11 SA 335 P22

3 4 5 5

Source: ASME boiler and pressure vessel code, Sec. 1,,1980, p. 184, Table PG23.1

because it involves a trial-and-error approach, and reference to many charts and tables. However, by using he method explained in this article, and Figures F-2 and F-3, quick estimates of shell or tube thickness, or external design pressure, can be obtained for a wide variety of materials without the need for a complicated iterative procedure. Many chemical process equipment items must be designed for external pressure. For example, process vessels often may be subjected to accidental vacuum. The ASME Boiler and Pressure Vessel Code requires that these be treated as vessels subject to external pressure. Vessels with jackets, depending on the pressure in the jackets, may have to be designed for external pressure. In heat exchangers, evaporators, aftercoolers, etc., where there is high pressure on the shell side, the tubes must be designed for external pressure. To show the advantage of this new method, let us first describe the procedure of the ASME Boiler and Pressure Vessel Code, Sec.VIII, Div.l, Para. UG-2S. ASME Procedure For cylinder$, shells and tubes where Dlt> 10: Step l-Assume a value for t, and determine LID and Dlt.

...... L ___________

~~~~

___

~.

______ _____________ ~

~~~_

384

Waste Heat Boiler Deskbook

Step 2-Enter Figure UGO-28.0 (in the ASME manual) at the value of LID. For LID> 50, use LID =50. Step 3-Move horizontally to the line for the value of Dlt, and from the point of intersection read off A. Step 4-Using A, enter the appropriate chart for material and temperature and read off B (from Figure UCS 28-1 or 28-2 in the code). Step 5-Calculate Pa = 4Bf[3(Dlt)]. Step 6-If Pa is less than the design pressure, P, increase t and repeat Steps 1 to 6.

For cylinders, shells and tubes where Dlt < 10: Step I-Using the same procedure as before, find A and B. For Dlt less than 4, A = 1.11(DI02 . For A > 0.1, use 0.1. Step 2-Calculate Pal and Pa2 . pal

=[2.167 D/t

pa2

0.0833] B

=~-h _-L) D/t\

D/t

Where S is the lesser of: twice the maximum allowable stress value at the design metal temperature (from the applicable table in Subsection C), 0.9 times the tabulated yield strength of the material at the design temperature (values of yield strength are available in Section VIII, Div. 2, Table F-ACS 2).

NOMENCLATURE A,B -

DLP Pa -

Factors from charts in ASME Boiler and Pressure Vessel Code, Section VIII, Div 1 and 2, P 8 UGO-28.0. UGS-28 and UGS28.2 Outer diameter of shell or tube in. Design length of shell or tube in. External design pressure, psi Maximum allowable external pressure on shell or tube psi

~.------~----------~-----------

Appendix F - Tube Thickness Calculations

s-

385

Lesser of twice the maximum allowable stress value at design temperature or 0.9 times the tabulated yield strength at the design temperature, (yalues of tabulated yield strength in Section VIII, Div. 2 may be used in the absence of such values in Div. 1). Minimum thickness of shell or tube, exclusive of corrosion allowance, in.

Table F-S. Classification of materials for use in Figures F-2 & F-3 Type for use ~

i[] Eig E-j

lli2...

Malru:ial

1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12.

SA 36 AandB SA 283 AandB SA 283 CandO SA 285 AandB SA 285 C SA414 AandB SA414 C,O,E and F SA 442 55 and 60 SA 455 All SA515 All SA 516 All SA 537 Up to 2.5 in. incl. CI.1 SA 53 AandB SA 106 A, BandC SA 179 SA 192 SA 333 1.6 SA 334 1.6 SA 524 1 and 11 SA 556 A2 SA 556 82 SA 53 Aand8 SA135 Aand8 SA 178 AandC SA 226 SA 334 1 SA 334 6 SA 557 A SA 557 B SA 587 SA 105 SA 1811 and II SA 216 WCA and WC8 All SA 266

13. 14. 15. 16. 17. 18. 19. 20. 21. 22. 23. 24. 25. 26. 27. 28. 29. 30. 31. 32. 33. 34.

B A B A B A B B B B B B B B A A B B 8 A 8 8 8 8 A A 8 A B 8 8 8 8 8

~

E-2 lli2... 35. 36. 37. 38. 40. 41. 42. 43. 44. 45. 46. 47. 48. 49. 50. 51. 52. 53. 54. 55. 56. 57. 58. 59. 60. 61. 62. 63. 64. 65. 66. 67.

Type for use i[] Eig E-j ~ E-2 LF1 and LF2 B LCB B I B AandO B 2CI. . B 12 cl.1 .B 11 cl.1 B 22 cl.1 B 21 cl.1 B All A All B All B 4,7and3 B 7and3 B All B All B 1 and 2 B T1b A T1, T1a 8 7and3 8 1 and 2 8 WCI 8 cI F22a 8 cI F21a B dF5 8 LF4,LF3 8 LCI B 1 B 1 8 405,41 OS, 410 8 TP 405,TP410 8 405 8

Malru:ial SA 350 SA 352 SA 372 SA 203 SA 387 SA 387 SA 387 SA 387 SA 387 SA 199 SA 209 SA 213 SA 333 SA 334 SA 335 SA 369 SA 423 SA 260 SA 260 SA 334 SA 423 SA217 SA 336 SA 336 SA 336 SA 350 SA 352 SA 508 SA 541 SA 240 SA 268 SA 479

~

Waste Heat Boiler Deskbook

386

Step 3-Choose the smaller of Pal andPa2 for maximum allowable design pressure. This can be compared with P and t, and may be altered if necessary. A quick estimating method Figure F-1 and F-3 have been developed based on detailed calculations, and give a very good estimate of Pa if D, t and L are known. Or, t may be found if the other parameters are known. The two examples that follow are worked out using both methods, to illustrate the versatility of the charts Two categories of materials commonly used are considered in these charts under Type A and Type B. They are categorized in Table F-5. Example 4:

A shell of material SA 515 Grade 70 is to be designed to withstand external pressure under the following conditions: P= 30 psi; D =80 in.; L = 180 in.; design temperature, T, =500°F. Estimate the thickness that is suitable. Method A (using code procedure): Assume the thickness to be 1.5 in. Then:

L/D = 2.25; D/t = 53 From Figure UGO-28, for LID =2.25 and D/t::: 53, find A = 0.0015. , From Figure UCS-28.2, for A = 0.0015 and T = 500 of, find B = 11,000. Then: Pa

= 4Bj[3(D/t)] = (4)(11,000)/(3)(53) = 277 psi

Therefore, we have to try a lower value for t. Try t = 0.5 in. Then, L/D = 2.25; D/t = 160 From Figure UG0-28, A =0.00028 From Figure UCS-28.2, B = 3,800

Pa = (4)(3,800)/(3)(160) = 32 psi Hence, t =0.5 is reasonable.

r

----------------~~-------~

=-'"--------'

~~~~~-~--~-~-~-~~-~-~~~-

"<

Appendix F - Tube Thickness Calculations

387

Method B (using Figure F-2 and F-3); Since D/t> 10, use Figure F-2 from Table SA 515, Grade 70 comes under Type B. On Figure F-2, start from Pa = 30, and go up to cut T = 500 of. Move horizontally to intersect a vertical line from L/D = 2.25 at a D/t of approximately 160. Hence, t = D/160 =0.5 in. Example 5: An SA 335 Pllpipe having a diameter of 5 in., a thickness of 0.8 in. and a length of 40 in. is used at a temperature of 800°F. Determine the maximum external pressure that this pipe can withstand. Method A (using Code procedure);

L/D

= 40/5 = 8; D/t = 5/.08 = 6.25

From Figure UGO-28, A = 0.3 From Figure UCS-28.2, B = 12,000

Pal

= [(2.167/6.25) -0.0833]12,000 = 3,160 psi

To find Pa2 we must first obtain S. For the material specified, the allowable stress (from Table F-UCS-23, in Section VIII, Div. 1) is 15,000 psi, and yield strength (Section VIII, Div. 2) is 22,500 psi. 0.9(22,500)

= ~O,250

Twice the allowable stress is: 2(15,000) = 30,000. We choose the lower of the two (20,250) to use in the equation for Pa2. Pa2 = [(2)(20,500)/6.25][1-(1/6.25)] = 5,443 psi

Hence the maximum allowable pressure is 3,160 psi.

Method B (using Figure F-2 and F-3); Since D/t< 10, use Figure F-3 from Table I, material used falls under Type B. On Figure F-3 for D/t = 6.25 and T = 800°F, read off Pa = 3,200 psi.

til

,j~11

I-

.,'~.:~

~~ '~~

1b. ~

-r-

1\

~ 1\

1\ ~ 1\ 1\ ~

1\

~

r--I--

~t\

J

500

200

100 P•. psi

50

I

~r--~ SO

t-

-r-.

rI"-

~

"'"

" ~~

~

20

~

10

0.2

'"

1\

"-

1\ 1\

D,S

1.0

3 2 LlO-

~~

"

t

I

~

~,

\~ t\r--

~q,

I'~

t--.I'

-....."

1'..
" ~'.

I-

-

t-.....

-~

-

r- I-- r-

~~

20

rr--- r-F-r-. t--. ~r-.

--

r-- I -

I--

r---_~_

l-

~t\~

1,000

r---

I-r-.

CO CO

I I Dlt=1O

~

~ I:C

o

t:::

III •

5

10

20

i

~ o

Figure F-2. Chart used for solving problems concerning shells and tubes subjected to external pressure. Use this chart when D/t >10.

~

>

'1j '1j (!)

::s

&. ><

'Tl

8,000

I

~ cr'

(!)

7,000

;1 ~ ::s

6.000

...

(!)

rn rn (j

['" 'g:"

5,000

0-

::srn

",,'"

1,000 '---'-....,:,~...._~_"'-_.1.......Ii..-~_....._..I.._.,j,.,_~_"-.....1_...,I,_..J._.........J 4 4.5 5 6 7 8.5 8

DIt-

Figure F-3. Chart for shells and tubes subjected to external pressure. Use when Dft <10.

-

------

390

----------

~------

--~--

----

Waste Heat Boiler Deskbook

CONCLUSION These charts are reasonably accurate for engineering purposes. We have worked out over 20 cases, and the maximum error was 6%. For this approach: • No trial and error is involved if the thickness is not known. No "feel" for the value of t is needed. • There is no need to spend time looking up values for yield strength and allowable stress at the appropriate temperature. • There is no need to read values for A and B from the figures in the Code manual (Figure UGO-28.0 and UGO-28.2i this saves further time.) In addition, it should be pointed out that the two figures in the Code (UGO-28.0 and 28.2) have log-log scales, and the accuracy obtained by their use depends in part on the skill of the user. From the foregoing examples, one can see the elegance, simplicity and versatility that make the use of Figures F-2 and F-3 preferable to the Code methods.

Appendix G

Conversion Factors j\merican To Mebip American To American

Metric To American Metric To Metric AREA: 1 mm' = 0.001 55 in.' = 0.000 010 76 ft' 1 em' = 0.155 in.' = 0.001076 ft, 1 m' = 1550in.' = 10.76 ft'

1 in.' = 645.2 mm' = 6.452 cm' = 0.000 645 2 m' 1 ft' = 92903 mm' = 929.03 cm' = 0.0929m' 1 acre = 43 560 ft, 1 circular mil = 0.7854 square mil = 5.067 x to- IO m' = 7.854 x 10"' in.'

DENSITY and SPECIFIC GRAVITY: 1 g/cm' = 0.036 13lb/in.> = 62.43 Iblft' 1000 kg/m' 1 kg/l = 62.43 Ib/ft' = 8.3451b1USgal l",g/m' = 136 grains/ft' (for particulate pollution) 1 kg/m' = 0.062431b1ft'

=

lib/in.' = 27.68 g/cm' = 27680 kg/m' Ilblft' = 0.0160 g/cm' = 16.02 kg/m' = 0.0160 kg/I Specific gravity relative to water (SGW) of 1.00 = 62.43Ib/ft'at4Cor39.2Ft Specific gravity relative to dry air (SGA) of 1.00 = 0.0765Ib/ft':j: = 1.225 kg/m' Ilb/USgal = 7.481lb/ft' = 0.1198 kg/I 1 gIft' = 35.3 x 10" ",g/m' 1lb/l0ooft' = 16 x 10'",g/m'

=

ENERGY. HEAT. and WORK: 1 cal = 0.003 968 Btu 1 kcal = 3.968 Btu = 1000 cal = 4186 J = 0.004 186 MJ 1 J = 0.000948 Btu = 0.239 cal = 1 W's = 1 N'm = 10' erg = 10' dyne ·cm 1 W·h = 660.6 cal

3413 Btu = 1 kW·h 1 Btu = 0.2929 W·h 252.0 cal = 0.252 kcal = 778Ct·lb = 1055 J = 0.001055 MJ tfHb = 0.1383 kg· m = 1.356 J 1 hp ·hr = 1.98 x 10' fHb 1 therm = 1.00 x 10' Btu 1 bhp (boiler horsepower) = 33 475 BtuIhr 8439 kcallh 9.81 kW

=

HEAT CONTENT and SPECIFIC HEAT: 1 callg = 1.80 Btullb = 4187 J/kg 1 callcm' = 112.4 Btulft' 1 kcaUm' = 0.1124 Btulf!' = 4187 JIm' 1 callg·oC = 1 Btullb·oF = 4187 J/kg·oK

t 62.35Iblft' at 60 F. 15.6 C; 8.335 IblUSgal. :I: 0.07631blft' for moist air.

=

1 Btullb = 0.5556 callg = 2326 J/kg 1 Btulft> = 0.008 90 callcm' = 8.899 kcallm' = 0.0373 MJlm' 1 BtulUSgal = 0.666 kcal/l 1 BtuJlb·oF = 1callg· oC = 4187JIkg·ol(_

391

,,------------------------------"

""

""-""~~

------

Waste Heat Boiler Deskbook

392

Metric To American Metric To Metric

American To Metric American To American

HEAT FLOW. POWER: 1 N . mls = 1 W = 1 J/s = 0.001 341 hp = 0.7376 ft ·Ib/see 1 keallh = 1.162 J/s = 1.162 W = 3.966 Btulhr 1000 Jls 3413 BtuIhr 1 kW = 1.341 hp

1 hp = 33 000 ft Ib/mfn = 550 ft Ib/see = 745.7 W = 745.7 J/s = 641.4 keal/ll 1 Btulhr = 0.2522 kcallh = 0.000 393 1 hp = 0.2931 W = 0.2931 J/s

=

=

HEAT FLUX and HEAT TRANSFER COEFFICIENT: 1 BtulftZ . sec = 0.2713 eallem"s 1 callem"s = 3.687 Btu/ft' . sec 1 Btu!ft" hr 0.003 153 kW/m' = 41.87 kW/m' = 2.713 keallm"h 1 eallem" h = 1.082 WIrt' = 11.65 W/m' 1 kW/ft' = 924.2 eal/em'·h .""" lkW/m' = 317.2 Btu!ft'·hr /1Bfii7rt%~P'F";;'-4:89 kcal/m'·h· °C 1 kW/m"oC = 176.2 Btu!ft'·hr·oF

=

'--~~~'-~--~'----

-

LENGTH: 1 mm

= 0.10 em = 0.03937 in. = looem = loo0mm = 39.37in.

1 in. == 25.4 mm = 2.54 em = 0.0254 m 1 ft 304.8 mm 30.48 em 0.3048 m 1 mile = 5280 ft 1 micron = 1!l = 1
=

= 0.003281 ft 1m

= 3.281 ft lkm = 0.6214mile

PRESSURE: 1 N . m' 0.001 kPa 1.00 Pa 1 mm H,O = 0.0098 kPa 1 mm Hg = 0.1333 kPa = 13.60 mm H,O = 1 torr = 0.01933Ib/in.' 1 kg/em' 98.07 kPa 10000 kg/m' 10000 mm 394.1 in. = 735.6 mm Hg = 28.96 in. Hg 227.6oz/in.' 14.221b/in' = 0.9807bar 1 bar 100.0 kPa 1.020 kg/em' 10200 mm H,O = 401.9 in. H,O 750.1 mm Hg = 29.53 in. Hg = 232.1oz/in.' = 14.501b/in.' = 100000 N/m' 1 g/em' = 0.014221b/in.' = 0.2276oz/in.' = 0.3937 in. H,O

=

=

= = =

=

= ao = =

=

ao

=

=

1 in. H,O = 0.2488 kPa = 25.40 mm H,O = 1.866mmHg = 0.00254 kg/em' = 2.54 g/em' 1 in. Hg = 3.386 kPa = 25.40 mm Hg = 345.3 mm H,O = 13.61 in. H,C = 7.858 ozIin.' = 0.491 Ib/in.' = 25.4 torr lIb/in.' = 6.895 kPa = 6895 N/m' 703.1 mm H;O 27.71in. H,O = 51.72 mm Hg = 2.036 in. Hg = 16.00 ozlin.' = 0.0703 kg/em' = 70.31 g/em' = 0.06897 bar 1 oz/in.' = 0.4309 kPa = 43.94 mm = 1.732 in. H,O = 3.232mmHg = 0.004 39 kg/em' = 4.394 g/em'

=

=

ao

Appendix G - Conversion Factors

Metric To American Metric To Metric

393

American To Metric American To American

PRESSURE (cont'd) (For rough calculations. 1 atm = 1 kg/cm' 1 bar = 10 m H,o = 100 kPa)

=

1 atm t = 101.3 kPa = 101 325 N/m' 10330mmH,Q 407.3in.H,Q = 760.0 mm Hg = 29.92 in. Hg '" 235.1OZ/in.' 14.70Ib/in.' 1.033 kg/cm' = 1.013 bar

=

=

=

TEMPERATURE: C = 'I,(F - 32) F = ('I, C) + 32 K = C + 273.15 R = F + 459.67 THERMAL CONDUCTIVITY: 1 W/m· oK 0.5778 Btu· Wft" hr· OF = 6.934 Btu·inJft'·hl'·oF 1 cal'cmlcm"s,oC = 241.9 Btu·ftlft'·hr·oF = 2903 Btu· inJft' . hI' . °F 418.7W/m·oK

=

=

THERMAL DIFFUSIVITY: 1 m'/s = 38 760 ft'ihr 1 m'ih 10.77 ft'/hr

=

VELOCITY: 1 cmls = 0.3937 inJsec = 0.032 81 Wsec '" 10.00 mmls 1.969 ftlmin 1 mls '" 39.37 inJsec = 3.281 Wsec = 196.9 ftlmin = 2.237 mph = 3.600 km/h = 1.944 knot

=

= 1.730 W/m' OK = 1.488kcaUm·h·oK 1 Btu·in.lft"hr·oF = 0.1442W/m·oK

1 Btu· Wft'· hr' ° F

=

1 Btu·ftlft'·hr·oF O.004139ca!·cmlcm'·s·oC 1 Btu·inJft"hl'·oF = 0.000 3445cal·cml cm"s,oC

1 ft'/hl'

= 0.0000258 m'/s = 0.0929 m'ih

= 25.4 mmls = 0.0254 mls = 0.0568mph 1 Wsec 304.8 mmls = 0.3048 mls = 0.6818mph 1 ftlmin = 5.08 mmls = 0.00508 mls = 0.0183 km/h 1 mph = 0.4470 mls 1.609 kmih = 1.467 Wsec 1 knot = 0.5144m1s 1 rpm - 0.1047radians/sec 1 in.lsec

=

=

t Normal atmosphere = i60 torr (mm Hg at 0 C)-not a "technical atmosphere". which is 736 torr or 1 kg/em'. Subtract about 0.5Ib/in.' for each 1000 ft above sea level.

Waste Heat Boiler Deskbook

394

American To Metric American To American

Metric To American Metric To Metric VISCOSITY. absolute.,.,: 0.1 Pa·s = Idyne·s/em' = 360kg/h·m = 1 poise = 100 centipoise = 242.11b masslhr· ft = 0.002 0891b force·seelft' 1 killh· m 0.672 Iblhr· ft 0.002 78 gls' em 0.000005 81!b force·sec/f!'

= =

=

1 lb mass/hr· ft

= 0.000008 634 Ib Corce ·sec/f!'

= 0.413 centipoise = 0.000413 Pa·s lib force 'sec/ft' = 115800 Ib mass/hr· Ct = 47880centipoise = 47.88 Pa's 1 reyn

= 1 Ib force' sec/in.'

= 6.890 x 10' centipoise ,., ofwatert = 1.124centipoise 2.72lb mass/hr· ft 2.349 x lO-'1b· sec/ft'

,., of airt = 0.0180 centipoise = 0.0436 Ib/hr . ft = 3.763 x 10" Ib ·sec/ft

VISCOSITY. kinematic. v: 1 cm'/s = 0.0001 m'ls

• 1 m'ls

= 1 stoke = 100 centis tokes = 0.001 076 ft'/sec = 3.874 ft'ihr = 3600 m'lh = 38736 ft'/hr = 10.76 ft'lsee

v of watert = 1.130centistokes

= 32 SSU

1 ft'lsee = 3600 ft'/hr = 92 900 centistokes = 0.0929 m'ls 1 ft'/h 0.000 278 ft'/sec 25.8 centis tokes = 0.000 025 8 m'/s

=

=

v of airt

= 14.69 centis tokes 1.581 x 10" ft'/sec

= 1.216 x 10'" ft'lsee

VOLUME: 1 em'(ce) = 0.000001 OOm' 0.0610 in.' 0.0338 US fluid oz 1 1 (dm'] = 0.0010 m' = 1000 cm' 61.02 in.' 0.03531 f!' 0.2642 VSgal

1 m'

=

=

= =

=

= 1000 1 = 1 000 000 cm' = 61 020 in.' = 35.31 ft' = 220.0 Brga!

= 6.290bbl 264.2 USgal

= 1.308 yd' t At stp.

1 in.' = 16.39 cm' ::: 0.000 163 9 m' 0.016391 1 ft' 1728 in.' 7.481 USgal = 6.229 Brgal = 28320 cm' = 0.028 32 m' = 28.321 = 62.427 IboC 39.4 F(4C)water = 62.3441b of 60 F(15.6 C] water

= =

1 USgal

=

= 3785 cm' = 0.003 785 m' = 3.7851 = 231.0 in.' = 0.8327 Brga! = 0.1337 ft' = sp gr x 8.335 Ib ::. 8.3351b oC water = ,/ •• barre! (oil)

Index ABMA boiler and feed water guidelines

Cogeneration systems 5 Combined cycle plants 6, 45 Computer programs COGEN228 desuperheating steam 194 expansion of steam 244 insulation performance 321 supplementary firing 25 Condensate heater 233 Convective heat transfer 135-139 (See heat transfer coefficients also) Corrosion in economizer 307-310 high temperature 77 low temperature 80, 307-310 in stacks, ducts 315 superheater 79, 80

268

Acoustic vibration 329 Air heaters 91, 332 Allowable stress values 378-379 Approach point design mode 34, 165,209 off-design mode 227-228 economizer steaming 34,212 fired mode 218 selection of 209-211 ASME boiler and feed water guidelines 266

Auxiliary firing bumers 15, 17 combustion calculations 20, 24, 28 computer program 25 efficiency of system 30 fresh air 19, 31

Dampers 41 leakage through 42-44 Deaerator in HRSG systems 237 steam consumption 238-241 Departure from nucleate boiling, (DNB) 174 Desuperheater 191 Dew points of acid gases 310

Baffles 332 Beam length 104, 141 Blow down 265 Boiler water 266-268 Burners duct 15 register 17

Drum

Carbon Monoxide (CO) 35 catalysts 40 conversion calculations 35-36 Casing design 322-324 Cheng cycle 47-50 Chlorine corrosion 78-80,310 Circulation calculations 174-175 forced 6 natural 6 system 170, 175 COGEN software 228 features 229 optimizing temperature profiles 245

hold up time 272 internals 258 sizing 271

Economizer configura tions196 corrosion in 310-312 design 196 steaming 34 Efficiency fin 278 HRSG system 30, 31, 231 improving 231 395

396 Emissions C035 conversion calculations 36 SCR40 NOx 35 stearn injection 39 water injection 39 specifications 257 Emissivity of gases 142-146 Enhanced oil recovery 51 Enthalpy of gases 351 Eutectics 77 melting point 78 Evaporator (see water tube boiler also) bare tube 154, 286 comparison of bare versus finned 167, 286 configurations 170 design of 155,286 finned tube 286 forced circulation 6 heat flux 157, 286 heat transfer 135-140 inline versus staggered 153, 287 natural circulation 6 performance 162 Expansion of stearn 242-245 Feed water 265, see ABMA, ASME Finned tubes 33, 277 comparison with bare tubes 167, 286 efficiency 278 effectiveness 278 effect of fin configuration 294, 297 effect of tube side coefficient 291 fin tip temperature 280 fouling factors 294 gas pressure drop 278, 295-297 heat transfer 276-280 heat flux 291, 294 inline arrangement 287-289 serrated 279 solid 279 staggered arrangement 287-289 surface area 292,300 weight 302 Fire tube boilers design of 100,103 effect of scale 121 effect of tube size 108

Waste Heat Boiler Deskbook elevated drum 3,97-99 gas bypass calculations 124 guidelines for selection 99 heat loss calculations 125 heat flux 124 . hydrogen plants 52, 55, 123 incineration plants 84,88 multi-pass 99 performance 117 simplified design procedure 109 simplified performance evaluation 119 single pass 84, 97 single shell 84, 97 Fluidised bed combustors 74 Fluidised bed cat cracker 62 Forced circulation boilers 6 Fouling factors effect in finned tubes 292, 295 inside tubes 120, 156, 292 outside tubes 294 scale 121 Furnaces membrane wall 16, 66 refractory lined 66, 70 water cooled 66, 174 Gas analysis effect on gas properties 354 converting % volume to. weight, 355 of heat recovery systems 2 Gas mass velocity 136, 156, 276 Gas properties computing 352 data 349 enthalpy 351 effect of pressure 353 effect of gas analysis 354 specifications 260-262 specific heat 352 thermal conductivity 352 viscosity 352 Gas turbine characteristics 5 cogeneration plants 6 combined cycle plants 5 data 260, 337, 351 exhaust gas analysis 2, 351 HRSG features 5 NOx reduction 39-41

Index stearn injection 39 water injection 39 Hazardous waste incineration 76 Heat flux in finned tube boilers 33, 291-294 in fire tube boilers 124 in water tube evaporators 33, 157 Heat loss calculations 317-320 computer program 320 through multi-layered insulation 320 in fire tube boiler 125 Heat pipes 92 Heat transfer coefficients effect of pressure 353 inside tubes 139 outside bare tubes inline 135-137 outside bare tubes staggered 135-137 outside finned tubes inline 276, 287289 outside finned tubes staggered 276, 287-289 non-luminous 142-146 stearn 139 HRSGs (see waste heat boilers also) cogeneration plants 5 combined cycle plants 6 fired 12 fluidised bed cat cracker 62 forced circulation 6 gas turbine exhaust 5 hydrogen plants 54 improving efficiency 231 incineration systems 80, 86-87 multi-pressure 4, 245, 250 natural circulation 6 performance 197 simulation of performance 205 specifying 257 sulfuric acid plants 57, 60 temperature profiles 207, 213 High temperature corrosion 77 Hydrochloric acid dew point 310 Hydrogen plant 54 Incinerators fluidised bed 67, 74 rnodular70 refractory lined 66 rotary kiln 68, 76

397 waterwall 66 Incineration furne87 gaseous waste 87 hazardous waste 76 liquid waste 87 solid waste (MSW) 62, 84 RDF72,73 V0Cs87,88 Inline arrangement bare tubes 136, 148, 153 finned tubes 287-289 Insulation calculations 316-320 Landfill gas 44 Leakage through dampers 42-44 Life cycle costing 115 Low temperature corrosion 307 Melting point of eutectics 77-78 Multi-pressure HRSG 4, 245, 250 Municipal solid waste 64 analysis 64 boilers 80 corrosion 77 Natural circulation boilers 6, 174 Natural frequency of vibration 329 NOx gas turbine 35 conversion calculations 36 method of reducing 37 SCRs 37-41 stearn injection.39 water injection 39 NTU method 181 NusseltNumber 102, 137 Optimization fire tube boilers 115 pinch and approach points 165 temperature profiles 165, 245 Oxygen for combustion 20-28 (see auxiliary firing also) Pinch point design mode 165, 210-213 fired case 218 selection of 209-211 optimization 165

398 Prandtl Number 102, 137 Pressure drop gas inside tubes 105 gas outside inline bare tube bank 148150 gas outside staggered bare tube bank 148-150 gas outside inline finned tube bundle 278,290 gas outside staggered finned tube bundle 278, 290 cost implications 115,166, 171,295-297 Radiation distribution of 187 non-luminous 140--145 Reciprocating engine 52 Recycling 63 Refuse Derived Fuel (RDF) analysis 73 boilers 75 corrosion 83 preparation of 72-75 Reynolds Number 102, 136, 149,277 Rotary Kiln 76

Scale conductivity 121 effect of 121, 122 Selective Catalytic Reduction Systems (SCRs) 37-39 Simplified procedure fire tube boiler design 109 fire tube boiler performance 119 water tube boiler evaporator 157, 164 superheater /economizer 181 NTU method 181 Simulation of HRSGs 205 Slagging 77, 78 Solid waste, see Municipal solid waste 64 Specific heat computing mixture 352 data 349-351 effect of gas analysis 354 effect of pressure 353 Specifications (HRSG) 257 Staggered arrangement bare tubes 137, 138,148, 153 finned tubes 287-289

Waste Heat Boiler Deskbook Steam deaeration 238-241 desuperheater 191 expansion of 242~245 heat transfer coefficient 139 injection 39 properties 352, 362-363 purity 258 temperature control 191 velocity 140 Steaming economizer 34, 212 Strouhl Number 329 Sulfur condenser 60 Sulfuric acid corrosion 309 dew point 310 Superheater configurations 176 design of 176,179 header arrangement 183 flow in parallel passes 189 high temperature corrosion 77 performance of 181 tube wall temperature 182, 190 Supplementary firing (see auxiliary firing) Surface area evaporator 156, 167 finned tubes 292, 300 fire tube boilers 101, 108 superheater 180

Temperature profile in design mode 213-215 off-design mode 216-222 optimizing 245 fired HRSG 218 from field data 250 multi-pressure HRSG 250 unfired HRSG 213-215 Thermal conductivity of mixture 352 data for gases 349 effect of pressure 353 scale 121 boiler tubes 184 Tube thickness external pressure 381 internal pressure 377

399

Index Tube wall temperature bare tubes 157, 182, 190 finned tubes 280 economizer 315 Velocity gas mass 136, 276 steam 140 Vibration acoustic 329 avoiding 330-332 vortex shedding 329 Viscosity mixture 352

data for gases 349 effect of pressure 353 Vortex sheading frequency 329 Waste heat boiler (see HRSGs) Water tube boiler economizer 196 evaporator 155 forced circulation 6 guidelines for selection 133 multi-pressure 4,245,250 natural circulation 6, 174 superheater 176

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