Combined cycle systems for near-zero emission power generation 0857090135, 9780857090133

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Combined cycle systems for near-zero emission power generation

© Woodhead Publishing Limited, 2012

Related titles: Modern gas turbine systems: High efficiency, low emission, fuel flexible power generation (ISBN 978-1-84569-728-0) This book provides a comprehensive review of gas turbine science and engineering. Coverage includes large-scale power generation as well as industrial and smallscale gas turbine applications. This coverage is extended with an examination of operational and maintenance issues, including performance monitoring and control, modelling and optimisation techniques. Finally, advanced operating conditions and concepts are reviewed, including the pursuit of near-zero emissions power plants incorporating carbon dioxide capture technology.

Power plant life management and performance improvement (ISBN 978-1-84569-726-6) Coal- and gas-based power plants currently supply the largest proportion of the world’s power generation capacity. This book critically reviews the fundamental degradation mechanisms that affect conventional power plant systems and components, noting mitigation routes alongside monitoring and assessment methods. Maintenance and replacement routes are further extended through chapters on the management and refurbishment of advanced systems and components.

Advanced power plant materials, design and technology (ISBN 978-1-84569-515-6) Fossil fuel power plants generate the majority of the world’s power, but many plants are ageing and cannot meet rising global energy demands and increasingly stringent emissions criteria. To ensure security and economy of supply, utilities are building a new generation of advanced power plants with increased output and environmental performance. This book initially reviews improved plant designs for efficiency and fuel flexibility, including combined cycle technology and utilisation of lower-grade feedstocks. Coverage extends to advanced material and component use, and the incorporation of alternative energy conversion technology, such as hydrogen production. Environmental and emissions performance issues round off the book. Details of these and other Woodhead Publishing books can be obtained by: • • •

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© Woodhead Publishing Limited, 2012

Woodhead Publishing Series in Energy: Number 32

Combined cycle systems for nearzero emission power generation Edited by Ashok D. Rao

Oxford

Cambridge

Philadelphia

New Delhi

© Woodhead Publishing Limited, 2012

Published by Woodhead Publishing Limited, 80 High Street, Sawston, Cambridge CB22 3HJ, UK www.woodheadpublishing.com www.woodheadpublishingonline.com Woodhead Publishing, 1518 Walnut Street, Suite 1100, Philadelphia, PA 19102–3406, USA Woodhead Publishing India Private Limited, G-2, Vardaan House, 7/28 Ansari Road, Daryaganj, New Delhi – 110002, India www.woodheadpublishingindia.com First published 2012, Woodhead Publishing Limited © Woodhead Publishing Limited, 2012. Chapters 4, 6 and 11 were prepared by US government employees; these chapters are therefore in the public domain and cannot be copyrighted. The authors have asserted their moral rights. This book contains information obtained from authentic and highly regarded sources. Reprinted material is quoted with permission, and sources are indicated. Reasonable efforts have been made to publish reliable data and information, but the authors and the publishers cannot assume responsibility for the validity of all materials. Neither the authors nor the publishers, nor anyone else associated with this publication, shall be liable for any loss, damage or liability directly or indirectly caused or alleged to be caused by this book. Neither this book nor any part may be reproduced or transmitted in any form or by any means, electronic or mechanical, including photocopying, microfilming and recording, or by any information storage or retrieval system, without permission in writing from Woodhead Publishing Limited. The consent of Woodhead Publishing Limited does not extend to copying for general distribution, for promotion, for creating new works, or for resale. Specific permission must be obtained in writing from Woodhead Publishing Limited for such copying. Trademark notice: Product or corporate names may be trademarks or registered trademarks, and are used only for identification and explanation, without intent to infringe. British Library Cataloguing in Publication Data A catalogue record for this book is available from the British Library. Library of Congress Control Number: 2012930197 ISBN 978-0-85709-013-3 (print) ISBN 978-0-85709-618-0 (online) ISSN 2044-9364 Woodhead Publishing Series in Energy (print) ISSN 2044-9372 Woodhead Publishing Series in Energy (online) The publisher’s policy is to use permanent paper from mills that operate a sustainable forestry policy, and which has been manufactured from pulp which is processed using acid-free and elemental chlorine-free practices. Furthermore, the publisher ensures that the text paper and cover board used have met acceptable environmental accreditation standards. Typeset by Newgen Publishing and Data Services

Printed by TJI Digital, Padstow, Cornwall, UK

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Contents

Contributor contact details Woodhead Publishing Series in Energy Preface

ix xi xvii

1

Combined cycle power plants M. P. BOYCE, The Boyce Consultancy Group, LLC, USA

1.1 1.2 1.3 1.4 1.5 1.6 1.7 1.8 1.9 1.10 1.11 1.12

Introduction Typical cycles The Brayton cycle (gas turbine) The Rankine cycle (steam turbine) The Brayton–Rankine cycle (gas turbine and steam turbine) Combined cycle power plant configurations NOx emissions Carbon capture and sequestration Plant operation Availability and reliability Major equipment Sources of further information

1 6 6 12 16 17 19 20 22 24 26 42

2

Advanced industrial gas turbines for power generation M. P. BOYCE, The Boyce Consultancy Group, LLC, USA

44

2.1 2.2 2.3 2.4 2.5

Introduction Gas turbine compressors Gas turbine combustors Gas turbine expander Sources of further information

44 46 61 82 101

3

Natural gas-fired combined cycle (NGCC) systems A. D. RAO, University of California, USA

103

3.1 3.2

Introduction Technology, system design and equipment

103 104

1

v © Woodhead Publishing Limited, 2012

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Contents

3.3 3.4 3.5

Criteria pollutants control Advantages and limitations Future trends for improvements in performance and emissions Sources of further information References

3.6 3.7

116 118 120 126 127

4

Integrated gasification combined cycle (IGCC) systems Y. ZHU, Pacific Northwest National Laboratory, USA and H. C. FREY, North Carolina State University, USA

129

4.1 4.2 4.3 4.4 4.5 4.6 4.7 4.8

Introduction Technology, system design and equipment Prevention and control of pollutant emissions Advantages and limitations Future trends Conclusion Sources of further information References

129 131 142 148 153 155 156 156

5

Novel cycles: humid air cycle systems P. CHIESA, Politecnico di Milano, Italy

162

5.1 5.2 5.3 5.4 5.5 5.6

Introduction Water mixing for power augmentation and NOX control Steam injected gas turbine (STIG) cycles Recuperated water injected (RWI) cycles Evaporative cycles Comparative performance analysis of natural gas-fired humidified air gas turbine cycles Water quality and condensate recovery Further application of humid air turbine (HAT) cycles Conclusions Sources of further information References Appendix: nomenclature

162 163 165 171 174

5.7 5.8 5.9 5.10 5.11 5.12

178 180 181 182 183 183 185

6

Novel cycles: oxy-combustion turbine cycle systems 186 G. RICHARDS, National Energy Technology Laboratory, USA, M. WILLIAMS, URS Corporation, USA and K. CASLETON, National Energy Technology Laboratory, USA

6.1 6.2

Introduction Oxy-fuel power cycle configurations

© Woodhead Publishing Limited, 2012

186 187

Contents 6.3 6.4 6.5 6.6

Component and performance considerations Cycle operation and prospects for coal applications Conclusion References

7

Pressurized fluidized bed combustion (PFBC) combined cycle systems E. KAKARAS, A. K. KOUMANAKOS and A. DOUKELIS, National Technical University of Athens, Greece

7.1 7.2 7.3 7.4 7.5 7.6

vii 201 212 214 215

220

Introduction Fluidized bed combustion: an overview Pressurized fluidized bed combustion Environmental performance Industrial power plants employing PFBC technology Improvements in thermal performance and environmental signature Conclusions References

220 221 221 226 227

8

Externally fired combined cycle (EFCC) systems F. ROBSON, KraftWork Systems Inc., USA and A. D. RAO, University of California, USA

234

8.1 8.2 8.3 8.4 8.5 8.6 8.7 8.8

Introduction Background Early efforts in externally fired systems Large-scale EFCC programs Foster Wheeler high-performance power systems (HIPPS) United Technologies Research Center (UTRC) HIPPS Conclusions References

234 235 237 240 246 253 263 263

9

Hybrid fuel cell gas turbine (FC/GT) combined cycle systems J. BROUWER, University of California, USA

265

Introduction The hybrid FC/GT concept Background Design considerations Cycle configurations Hybrid FC/GT system performance

265 266 267 270 271 272

7.7 7.8

9.1 9.2 9.3 9.4 9.5 9.6

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227 231 232

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Contents

9.7 9.8 9.9 9.10 9.11

Hybrid system dynamic operation potential Commercialization status Conclusion References Appendix: glossary

275 278 279 280 281

10

Integrated solar combined cycle (ISCC) systems J. ZACHARY, Bechtel Corp., USA

283

10.1 10.2 10.3 10.4 10.5 10.6 10.7 10.8 10.9 10.10

Introduction Technology, system design and equipment Example of the evaluation process for an ISCC Additional considerations Advantages and limitations Past and future trends Conclusion Acknowledgment References Appendix: abbreviations

283 284 294 298 301 303 304 305 305 305

11

Techno-economic analysis of combined cycle systems H. C. FREY, North Carolina State University, USA and Y. ZHU, Pacific Northwest National Laboratory, USA

306

11.1 11.2 11.3

Introduction Techno-economic analysis (TEA) methodology Techno-economic analysis of pulverized coal-fired power plants with carbon capture Techno-economic analysis of natural gas-fired gas turbine combined cycle power plants with carbon capture Techno-economic analysis of coal-fired integrated gasification combined cycle power plants with carbon capture Advantages and limitations Summary Sources of further information References

306 308

Index

329

11.4 11.5 11.6 11.7 11.8 11.9

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313 318 321 323 326 326 326

Contributor contact details

(* = main contact)

Editor and chapters 3 and 8 Ashok D. Rao Advanced Power & Energy Program University of California Irvine CA 92697–3550 USA E-mail: [email protected]

Chapters 1 and 2 M. P. Boyce The Boyce Consultancy Group, LLC 2121 Kirby Drive #117 Houston, Texas 77019 USA E-mail: [email protected]

Chapters 4 and 11 Yunhua Zhu* Pacific Northwest National Laboratory P.O. Box 999 Richland, WA 99354 USA

H. Christopher Frey Department of Civil, Construction, and Environmental Engineering Campus Box 7908 North Carolina State University Raleigh, NC 27695-7908 USA E-mail: [email protected]

Chapter 5 Paolo Chiesa Department of Energy Politecnico di Milano Via Lambruschini 4 20156 Milan Italy E-mail: [email protected]

Chapter 6 G. Richards* National Energy Technology Laboratory 3610 Collins Ferry Rd. P.O. Box 880 Morgantown, WV 26507-0880 USA E-mail: [email protected]

E-mail: [email protected] ix © Woodhead Publishing Limited, 2012

x

Contributor contact details

Mark Williams URS Corporation 421 Jefferson Street Morgantown, WV 26501 USA K. Casleton National Energy Technology Laboratory 3610 Collins Ferry Rd. P.O. Box 880 Morgantown, WV 26507-0880 USA

Chapter 7 Emmanuel Kakaras, Antonios K. Koumanakos, Aggelos Doukelis* Mechanical Engineering Department National Technical University of Athens 9, Heroon Polytechniou str. 15780 Athens Greece E-mail: [email protected]; [email protected]; [email protected]

Chapter 8 Fred Robson KraftWork Systems Inc. Amston, CT USA E-mail: [email protected]

Chapter 9 Jack Brouwer National Fuel Cell Research Center University of California Irvine, CA 92697–3550 USA E-mail: [email protected]

Chapter 10 Justin Zachary Bechtel Corp. 5275 Westview Drive Frederick MD 21703-8603 USA E-mail: [email protected]

© Woodhead Publishing Limited, 2012

Woodhead Publishing Series in Energy

1 Generating power at high efficiency: Combined cycle technology for sustainable energy production Eric Jeffs 2 Advanced separation techniques for nuclear fuel reprocessing and radioactive waste treatment Edited by Kenneth L. Nash and Gregg J. Lumetta 3 Bioalcohol production: Biochemical conversion of lignocellulosic biomass Edited by K. W. Waldron 4 Understanding and mitigating ageing in nuclear power plants: Materials and operational aspects of plant life management (PLiM) Edited by Philip G. Tipping 5 Advanced power plant materials, design and technology Edited by Dermot Roddy 6 Stand-alone and hybrid wind energy systems: Technology, energy storage and applications Edited by J. K. Kaldellis 7 Biodiesel science and technology: From soil to oil Jan C. J. Bart, Natale Palmeri and Stefano Cavallaro 8 Developments and innovation in carbon dioxide (CO2) capture and storage technology Volume 1: Carbon dioxide (CO2) capture, transport and industrial applications Edited by M. Mercedes Maroto-Valer 9 Geological repository systems for safe disposal of spent nuclear fuels and radioactive waste Edited by Joonhong Ahn and Michael J. Apted 10 Wind energy systems: Optimising design and construction for safe and reliable operation Edited by John D. Sørensen and Jens N. Sørensen 11 Solid oxide fuel cell technology: Principles, performance and operations Kevin Huang and John Bannister Goodenough xi © Woodhead Publishing Limited, 2012

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12 Handbook of advanced radioactive waste conditioning technologies Edited by Michael I. Ojovan 13 Nuclear safety systems Edited by Dan Gabriel Cacuci 14 Materials for energy efficiency and thermal comfort in buildings Edited by Matthew R. Hall 15 Handbook of biofuels production: Processes and technologies Edited by Rafael Luque, Juan Campelo and James Clark 16 Developments and innovation in carbon dioxide (CO2) capture and storage technology Volume 2: Carbon dioxide (CO2) storage and utilisation Edited by M. Mercedes Maroto-Valer 17 Oxy-fuel combustion for power generation and carbon dioxide (CO2) capture Edited by Ligang Zheng 18 Small and micro combined heat and power (CHP) systems: Advanced design, performance, materials and applications Edited by Robert Beith 19 Advances in clean hydrocarbon fuel processing: Science and technology Edited by M. Rashid Khan 20 Modern gas turbine systems: High efficiency, low emission, fuel flexible power generation Edited by Peter Jansohn 21 Concentrating solar power (CSP) technology: Developments and applications Edited by Keith Lovegrove and Wes Stein 22 Nuclear corrosion science and engineering Edited by Damien Féron 23 Power plant life management and performance improvement Edited by John E. Oakey 24 Direct-drive renewable energy systems Edited by Markus Mueller and Henk Polinder 25 Advanced membrane science and technology for sustainable energy and environmental applications Edited by Angelo Basile and Suzana Pereira Nunes 26 Irradiation embrittlement of reactor pressure vessels (RPVs) in nuclear power plants Edited by Naoki Soneda

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Woodhead Publishing Series in Energy 27 High temperature superconductors (HTS) for energy applications Edited by Ziad Melhem 28 Infrastructure and methodologies for the justification of nuclear power programmes Edited by Agustín Alonso 29 Waste to energy (WtE) conversion technology Edited by Marco Castaldi 30 Polymer electrolyte membrane and direct methanol fuel cell technology Volume 1: Fundamentals and performance of low temperature fuel cells Edited by Christoph Hartnig and Christina Roth 31 Polymer electrolyte membrane and direct methanol fuel cell technology Volume 2: In situ characterisation techniques for low temperature fuel cells Edited by Christoph Hartnig and Christina Roth 32 Combined cycle systems for near-zero emission power generation Edited by Ashok D. Rao 33 Modern earth buildings: Materials, engineering, construction and applications Edited by Matthew R. Hall, Rick Lindsay and Meror Krayenhoff 34 Handbook of metropolitan sustainability: Understanding and improving the urban environment Edited by Frank Zeman 35 Functional materials for energy applications Edited by John Kilner, Stephen Skinner, Stuart Irvine and Peter Edwards 36 Nuclear decommissioning: Planning, execution and international experience Edited by Michele Laraia 37 Nuclear fuel cycle science and engineering Edited by Ian Crossland 38 Electricity transmission, distribution and storage systems Edited by Ziad Melhem 39 Advances in biodiesel production: Processes and technologies Edited by Rafael Luque and Juan Antonio Melero 40 Biomass combustion science, technology and engineering Edited by Lasse Rosendahl 41 Ultra-supercritical coal power plant: Materials, technologies and optimisation Edited by Dongke Zhang

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42 Radionuclide behaviour in the natural environment: Science, impacts and lessons for the nuclear industry Edited by Horst Geckeis and Christophe Poinssot 43 Calcium and chemical looping technology for power generation and carbon dioxide (CO2) capture: Solid oxygen- and CO2-carriers P. Fennell and E. J. Anthony 44 Materials ageing and degradation in light water reactors: Mechanisms, modelling and mitigation Edited by K. L. Murty 45 Structural alloys for power plants: Operational challenges and hightemperature materials Edited by Amir Shirzadi, Rob Wallach and Susan Jackson 46 Biolubricants: Science and technology Jan C. J. Bart, Emanuele Gucciardi and Stefano Cavallaro 47 Wind turbine blade design and materials: Improving reliability, cost and performance Edited by Povl Brøndsted and Rogier Nijssen 48 Radioactive waste management and contaminated site clean-up: Processes, technologies and international experience Edited by William E. Lee, Michael I. Ojovan and Carol M. Jantzen 49 Probabilistic safety assessment for optimum nuclear power plant life management (PLiM) Gennadij V. Arkadov, Alexander F. Getman and Andrei N. Rodionov 50 Coal utilization in industry Edited by D. G. Osborne 51 Coal power plant materials and life assessment: Developments and applications Edited by Ahmed Shibli 52 The biogas handbook: Science, production and applications Edited by Arthur Wellinger and David Baxter 53 Advances in biorefineries: Biomass and waste supply chain exploitation Edited by K. W. Waldron 54 Geoscience of carbon dioxide (CO2) storage Edited by Jon Gluyas and Simon Mathias 55 Handbook of membrane reactors Volume 1: Fundamental materials science, design and optimisation Edited by Angelo Basile 56 Handbook of membrane reactors Volume 2: Industrial applications and economics Edited by Angelo Basile

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57 Alternative fuels and advanced vehicle technologies: Towards zero carbon transportation Edited by Richard Folkson 58 Handbook of microalgal bioprocess engineering Christopher Lan and Bei Wang 59 Fluidized-bed technologies for near-zero emission combustion and gasification Edited by Fabrizio Scala 60 Managing nuclear projects: A comprehensive management resource Edited by Jas Devgun 61 Handbook of process integration: Energy, water, waste and emissions management in processing and power industries Edited by Jiří Klemeš 62 Membranes for clean and renewable power applications Edited by Annarosa Gugliuzza and Angelo Basile

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Preface

Near-zero power generation implies control not only of criteria pollutants such as NOx, CO and volatile organic compounds but also of the greenhouse gas CO2. Approximately, a third of all the CO2 emissions due to human activity come from fossil fuel-based power plants, with each power plant capable of emitting several million tonnes of CO2 annually. Furthermore, the increased demand for electricity is expected to continue well into the middle of this century, with the projected demand on a global basis more than doubling its current value. Improvements in thermal efficiency through the incorporation of more advanced technologies as well as improved plant configurations can certainly decrease the carbon footprint of electricity generation on a per megawatt basis. But to make substantial decreases in the CO2 emissions, capturing and storage of the CO2 is required. Two basic options are available for reducing the carbon footprint of a fossil-fuelled power plant: (1) pre-combustion capture which consists of capture before combustion in the gas turbine in the case of a combined cycle power plant; and (2) post-combustion capture which consists of capture from flue gas before it enters the atmosphere. The separated CO2 may then be sequestered geologically or used for enhanced oil or coal bed methane recovery. The book addresses two types of fossil fuel: natural gas and coal. The book opens with Boyce’s chapter on thermodynamics followed by key turbomachinery design aspects of a combined cycle. Next, Rao introduces the natural gas-fired combined cycle with both pre- and post-combustion carbon capture. The coal-based integrated gasification combined cycle which lends itself to pre-combustion carbon capture is introduced by Zhu and Frey. In addition to coal, the integrated gasification combined cycle may be fuelled or co-fuelled with other ‘dirty fuels’ such as biomass and refinery bottoms. An advantage of co-fuelling with biomass is that it not only uses an essentially carbon neutral energy source but also addresses the seasonal variation in the supply of biomass. Novel cycles such as humid air turbine cycles and oxy-combustion turbine cycles are next discussed by Chiesa and by Richards, Williams and Castleton, respectively, the latter being specifically developed for abatement of CO2 emissions. Coal combustion-based systems follow with Kakaras, xvii © Woodhead Publishing Limited, 2012

xviii

Preface

Koumanakos and Doukelis’s chapter on pressurized coal combustion combined cycles and Robson and Rao’s chapter on externally fired combined cycles. These systems, like the integrated gasification combined cycle, have the advantage of also being able to use the renewable biomass energy. The use of another renewable energy source, solar energy in a fossil-fuelled combined cycle application, is described by Zachary as reducing not only the carbon footprint but also tackling the intermittent nature of solar energy supply on a daily as well as on a seasonal basis. Brouwer discusses a promising technology for ultra-high thermal efficiency, the fuel cell which may be integrated into a combined cycle resulting in a hybrid system. In a fuel cell, where the chemical energy is directly converted into electricity, the intermediate step of conversion into heat as in a gas turbine is eliminated, whereas in a gas turbine, the materials of construction limit the temperature at which the heat is converted into power and thus its thermal efficiency. Finally, the techno-economic analysis and optimisation methodology are presented by Frey and Zhu to assist in the selection of appropriate technology option(s) from among the various alternatives for the energy systems or subsystems. This type of analysis is necessary in order to understand the basis of cost estimates and their strengths and limitations, to establish a framework for dealing with factors that lead to variability and uncertainty in cost estimates and to introduce a framework for making design choices and optimising process technologies. A common theme across the book is that of technology developments to improve the environmental signature of power plants, with special emphasis on reducing the carbon footprint. The various chapters show the range of approaches being pursued to meet this goal. This book, which is aimed at industry practitioners and academic researchers, contains chapters authored by both. The aim throughout the book is to provide a well-referenced assessment of the state-of-the-art technology with suggestions on where more detailed information can be found as well as some indication as to where the future developments might lie. I hope that you will find this book useful and also intellectually gratifying. Ashok D. Rao

© Woodhead Publishing Limited, 2012

1 Combined cycle power plants M. P. BOYCE , The Boyce Consultancy Group, LLC, USA

Abstract: This chapter deals with the theory behind power plants with particular reference to cogeneration and combined cycle power plants (CCPPs). The two key cycles are the Brayton (gas turbine) and Rankine (steam turbine) cycles, within each of which there are further subdivisions. A combination of the Brayton and Rankine cycles, often known as the combined cycle, has been found to obtain efficiencies as high as 60% and consequently is now used extensively worldwide as the main source of power. It is also more commonly being used in combined heat and power (CHP) plants that provide power and other sources of energy such as heat and air conditioning for major complexes or parts of cities. The chapter also deals with availability and reliability of major power plants and with power plant emissions and techniques to contain the emissions from these plants. Key words: gas turbines, steam turbines, Brayton cycle, Rankine cycle, combined cycle.

1.1

Introduction

With today’s concerns about energy and environmental impact, the natural gas combined cycle power plant (CCPP) plays a leading role in the power complexes of the world. CCPPs are the most efficient power plants operating on the power grids throughout the world with an efficiency ranging between 45% and 57%. These power plants come in all sizes. In this chapter we will be emphasizing the larger plants ranging in size from 60 to 1500 MW. These combined cycle plants have as their core the gas turbine, which acts as the topping cycle, and the steam turbine, which acts as the bottoming cycle. Between the gas turbine and the steam turbine is a waste heat recovery steam generator (HRSG), which takes the heat from the exhaust of the gas turbine and generates high-pressure (HP) steam for the steam turbine. The natural gas CCPP is attractive to the power industry because it has the following major characteristics: 1. These plants will displace coal in the power generation sector by 2050, under a model scenario where industrialized nations reduce CO2 emissions by 2050 through carbon emission pricing. Large emerging economies such as Brazil, China, and India will reduce CO2 emissions by 2070. 1 © Woodhead Publishing Limited, 2012

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Combined cycle systems for near-zero emission power generation

2. The plant has efficiency between 50% and 60%, the highest efficiency in the marketplace for a power plant type. 3. The steam turbine plant fired by coal or oil is the most widely used plant in the world and has an efficiency of about 30–35%. 4. The plant has a very compact footprint compared with other major types of power plants, and thus is environmentally friendly to the ecosystem. 5. The construction time of a major CCPP of 200 MW and above could be less than 24 months. 6. The emissions of unburnt hydrocarbons (UHCs), CO and oxides of nitrogen (NOx) (less than 9 ppm) are the lowest of any thermal power plant. 7. For plants larger than 200 MW, the cost of the plant ranges between $450 and $650/kW; for plants below 200 and 5 MW, the cost of the plant can vary, respectively, from $650/kW to about $1200/kW. CCPPs associated with electrical power plants use waste heat from the prime mover for the production of steam and consequently the steam is used in a steam turbine for the production of additional power. This is usually a combination of the Brayton cycle (gas turbine) as the topping cycle, and the Rankine cycle (steam turbine) as the bottoming cycle. However, technically the term ‘combined cycle’ can be used for any combination of two or more cycles. Many small plants use the diesel cycle as the topping cycle with the Rankine cycle as the bottoming cycle; there is a small number of plants that use the Brayton cycle as both the topping and the bottoming cycles. Figure 1.1 shows a large CCPP. The fossil power plants of the 1990s and the early part of the new millennium are the CCPPs, with the gas turbine being the centerpiece of the plant. It is estimated that between 1997 and 2010 an additional 147.7 GW of combined cycle power has been built. These plants have replaced the large steam turbine plants that were the main fossil power plants through most of the 1900s. The CCPP is not new in concept, since some have been in operation since the 1950s. These plants came into their own with the new high-capacity and high-efficiency gas turbines, with their new materials and advanced combustion systems. The new energy conversion marketplace will find many novel concepts in CCPPs. Figure 1.2 shows the heat rates of present and future plants, and Fig. 1.3 shows their efficiencies. The plants referenced are the simple cycle gas turbine (SCGT) plants with firing temperatures of 2400°F (1315°C); recuperative gas turbine (RGT) plants, where the exhaust gases from the turbine are used to heat the incoming air to the combustion chamber; the steam turbine plants; the CCPPs; the advanced combined cycle power plants (ACCPs), such as CCPPs using advanced gas turbine cycles; and finally the hybrid power plants (HPPs), combined cycle plants with fuel cells.

© Woodhead Publishing Limited, 2012

Combined cycle power plants

3

1.1 A typical combined cycle facility showing two plants of four gas turbines each. 12000

Heat rate

10000 8000 Btu/kWhr

6000

kJ/kWhr

4000 2000

Si m ga ple s tu cyc rb le in e R eg e ga n s era tu ti rb ve in e St ea m tu rb in C e om bi po ne w dc er y pl cle tu A an t cy rbi dv a cl ne n e ce c po om d w b ga er in s pl ed an t H yb rid po pl wer an t

0

Types of power plants

1.2 Typical heat rates of various types of plants.

The steam turbine power plants have an efficiency of about 30% to 35% compared with CCPPs, which have an efficiency from 47% to 58%. Newer gas turbine technology will make combined cycle efficiencies to range between 60% and 65%. As a rule of thumb, a 1% increase in efficiency could mean a saving of 3.3% in capital expenditure for the same size plant. However, one must be careful that the increase in efficiency does not lead to a decrease in availability. From 1996 to 2000 we have seen a growth in efficiency of about 10% and a loss in availability of about 10%. This trend must

© Woodhead Publishing Limited, 2012

4

Combined cycle systems for near-zero emission power generation 80 70

Efficiency

60 50 40 30 20 10

C

po pl wer an t

rid yb H

om

b po ine w dc er y pl cle an t tu A cy rbi dv a n cl e n e c po com ed w b ga er in s pl ed an t

ne tu rb i m St ea

R eg ga en s era tu ti rb ve in e

Si m ga ple s tu cyc rb le in e

0

Types of plants

1.3 Typical efficiencies of various types of plants.

be turned around since many analyses show that a 1% drop in availability needs an increase in efficiency of about 2–3% to offset the loss. The time taken to design, construct and commission a steam plant from conception to production is about 42–60 months as compared to 22–36 months for a natural gas CCPP. The actual construction time is about 18 months, while environmental permits in many cases take 12 months and engineering 6–12 months. The time taken to bring the plant on line affects the economics of the plant: the longer the capital is employed without return, the greater is the interest, insurance and taxes that accumulate. CCPPs and cogeneration plants are very similar in their thermodynamic properties. Cogeneration is the production of two or more forms of energy from a single plant. The most common application of the term is for the production of electrical power and steam for use in process applications. This does not mean that other types of cogeneration plants are not being designed and used. Cogeneration plants are used to produce power and use the direct exhaust gases from the prime movers for preheating air in furnaces or in absorption cooling systems, or for heating various types of fluids in different process applications. During the timeframe 1996–2006, cogeneration plants accounted for 3% of the new power being generated worldwide, amounting to 19.6 GW of electrical power. A typical cogeneration plant uses the waste gases from the gas turbine to produce steam in an HRSG or a waste heat boiler (WHB); these terms are often used interchangeably in various chemical processes. The steam could be used directly in an absorption chiller to produce refrigerated air, or in a steam turbine to drive a cooling system or to produce more power. Water is heated to provide hot water for all types of usage.

© Woodhead Publishing Limited, 2012

Combined cycle power plants

5

Cogeneration systems are also used in petrochemical plants where the prime movers drive compressors to compress process gases. The exhaust heat is used to generate steam for process use, or to operate an extraction, condensing, or back pressure steam turbine to drive a compressor or pump and the extracted steam is then used for a process application. A typical cogeneration plant in a refinery or a chemical plant generates HP steam which then is used in an extracting condensing steam turbine. It usually extracts part of the steam at a lower pressure for use in various chemical processes; the rest of the steam goes through the second part of the steam turbine and then to the condenser. The steam turbines usually drive separate generators; however, the system can be designed in such a way that both the gas turbine and the steam turbine drive the same generator. In most chemical plant applications, the gas turbine is a single-shaft unit if it is used for power generation; a twin shaft is used for mechanical drives such as driving a compressor or pump. The gases from the turbine exhaust are normally between 900°F (482°C) and 1100°F (593°C), depending on the turbine efficiency and the turbine firing temperature. The hot gases (approximately 90 lb/s (40.8 kg/s) for a 15 MW turbine to about 1400 lb/s (636 kg/s) for a 200 MW turbine) are piped into a boiler where steam is generated for use as process heat or for use in an extraction or back pressure steam turbine. Cogeneration plants are also known as combined heat power (CHP) plants and are widely used in Europe where they provide steam heating as well as power to large city centers. Figure 1.4 is a photograph of this kind of plant where the steam is bled after the intermediate-pressure (IP) stage of

Steam bled after IP stage of steam turbine for town center heating purposes

1.4 Steam turbine combined heat and power (CHP) plant.

© Woodhead Publishing Limited, 2012

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Combined cycle systems for near-zero emission power generation

a three-stage steam turbine, which consists of an HP stage, an IP stage and a low-pressure (LP) stage.

1.2

Typical cycles

The CCPPs discussed in this chapter use the Brayton cycle as the topping cycle (gas turbine) and the Rankine cycle (steam turbine) as the bottoming cycle. This is by far the most widely used configuration of CCPP in the power industry.

1.3

The Brayton cycle (gas turbine)

The gas turbine is a power plant which produces a great deal of energy for its size and weight. The gas turbine has found increasing service in the past 40 years in the power industry throughout the world both in utilities and merchant plants and in the petrochemical industry. Its compactness, low weight and multiple fuel application make it a natural power plant for offshore platforms. Today there are gas turbines that run on natural gas, diesel fuel, naphtha, methane, crude, low-Btu gases, vaporized fuel oils and biomass gases. The last 20 years has seen extensive growth in gas turbine technology. The growth has been spearheaded by the growth of new high-temperature materials technology, new coatings and new cooling schemes for the hot section of the gas turbine. This, in conjunction with an increase in compressor pressure ratio, has increased the gas turbine thermal efficiency from about 15% in the 1950s to over 45% in the mid-1990s. The Brayton cycle, which is the base cycle of the gas turbine, consists in its ideal form of two isobaric processes and two isentropic processes. The two isobaric processes are the combustor system of the gas turbine and the heat rejection side of the gas turbine and the heat gas side of the HRSG. The two isentropic processes are represented by the compression (Compressor) and the expansion (Turbine Expander) processes in the gas turbine.

1.3.1 Inlet cooling effect on the Brayton cycle – gas turbine The power in a gas turbine is greatly reduced by an increase in ambient temperature. A 10°F/5.6°C rise in temperature decreases the power output by about 4%. There are several techniques that are used for cooling the turbine compressor inlet, from the simple evaporative cooling to the more complex and costly refrigerated inlet cooling: •

Evaporative systems – Either conventional evaporative coolers or direct water fogging.

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Combined cycle power plants

7

•

Refrigerated inlet cooling systems – Utilizing absorption or mechanical refrigeration. • Combination of evaporative and refrigerated inlet systems – The use of evaporative cooler to assist the chiller system to attain lower temperatures of the inlet air. • Thermal energy storage systems – These are intermittent use systems where the cold is produced off-peak and then used to chill the inlet air during the hot hours of the day.

Evaporative and fogging systems for the Brayton cycle – gas turbines Traditional evaporative coolers that use media for evaporation of the water have been widely used in the gas turbine industry over the years, especially in hot climates with low humidity areas. The low capital cost and the low installation and operating costs make it attractive for many turbine operating scenarios. Evaporation coolers consist of demineralized water being sprayed over the media blocks, which are made of fibrous corrugated material. The airflow through these media blocks evaporates the water; as water evaporates, it consumes about 1059 Btu (1117 kJ) (latent heat of vaporization) at 60°F (15°C). This reduces the air temperature entering the compressor to below that of the ambient air. This technique is very effective in low humidity regions. The work required to drive the turbine compressor is reduced by lowering the compressor inlet temperature, thus increasing the output work of the turbine. The inlet temperature is lowered by about 18°F (10°C), if the outside temperature is around 90°F (32°C). An evaporative cooling system runs at a cost of around $50/kW. Direct inlet fogging is a type of evaporative cooling method where demineralized water is converted into a fog by means of HP nozzles operating at 1000–3000 psi (67–207 Bar). This fog then provides cooling when it evaporates in the air inlet duct of the gas turbine. The air can attain 100% relative humidity at the compressor inlet, thereby giving the lowest temperature possible without refrigeration (the web bulb temperature). Direct HP inlet fogging can also be used to create a compressor intercooling effect by allowing excess fog into the compressor, thus lowering temperatures in the first few stages of compression and further boosting the power output. Refrigeration systems for the Brayton cycle – gas turbines The refrigerated inlets are more effective than the previous evaporative cooling systems as they can lower the temperatures by about 45–55°F (25–30°C). Two techniques for refrigerating the inlet of a gas turbine are

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8

Combined cycle systems for near-zero emission power generation

vapor compression (mechanical refrigeration) and absorption refrigeration. In a mechanical refrigeration system the refrigerant vapor is compressed by means of a centrifugal, screw or reciprocating compressor. Centrifugal compressors are typically used for large systems in excess of 1000 tons (12.4 × 106 Btu/13.082 × 106 kJ) and would be driven by an electric motor. Mechanical refrigeration has significantly higher auxiliary power consumption for the compressor driver and pumps required for the cooling water circuit. After compression, the vapor passes through a condenser where it is condensed. The condensed vapor is then expanded in an expansion valve and provides a cooling effect. The evaporator chills cooling water that is circulated to the gas turbine inlet chilling coils in the air stream. Absorption systems typically employ lithium bromide (LiBr) and water, with the LiBr being the absorber and the water acting as the refrigerant. Such systems can cool the inlet air to 50°F (10°C). The heat for the absorption chiller can be provided by gas, steam or gas turbine exhaust. Absorption systems can be designed to be either single or double effect. A single-effect system will have a coefficient of performance (COP) of 0.7–0.9, and a double-effect unit, a COP of 1.15. Part-load performance of an absorption system is relatively good, and efficiency does not drop off at part load as it does with mechanical refrigeration systems. The costs of these systems are much higher than the evaporative cooling system; however, refrigerated inlet cooling systems in hot humid climates are more effective due to the limitation of the evaporative cooling system in high humidity climates. Combination of evaporative and refrigerated inlet systems for gas turbines Depending on the specifics of the project, location, climatic conditions, engine type and economic factors, a hybrid system utilizing a combination of the above technologies may be the best as shown in Fig. 1.5. The possibility of using an evaporative system ahead of the mechanical inlet refrigeration system should be considered. This may not always be intuitive, since evaporative cooling is an adiabatic process that occurs at constant enthalpy. When water is evaporated into an air stream, any reduction in sensible heat is accompanied by an increase in the latent heat of the air stream (the heat in the air stream being used to effect a phase change in the water from liquid to the vapor phase). If fog is applied in front of a chilling coil, the temperature will be decreased when the fog evaporates, but since the chiller coil will have to work harder to remove the evaporated water from the air stream, the result would yield no thermodynamic advantage. To maximize the effect, the chiller must be designed in such a manner that in combination with evaporative cooling, the maximum reduction in temperature is achieved. This can be accomplished by designing a slightly

© Woodhead Publishing Limited, 2012

Combined cycle power plants Steam generator

Y

1

Air

Compressor

W

Turbine

φ= 60

30

W B

W B

65

100

Ev

ap

60 (h) tion

80

co

oli

ng

20

60

4D evia

φ=

–.0

20 15

X

30

ve

W B

40

cu tf

35

13.0

45

ati

cu tf

50

40

W B

or

14.0

55

120

Refrigeration

25

Drain

140

Sensible heat ratio

4 Combustor

160

–0.2 D eviati on (h φ= ) 40

3

2

180

Specific humidity, grains per lb dry air

En tha lpy at 35 We sa tur t-b ati 70 ulb o an 40 n, B dd tu ew pe r lb p 75 oin t te 45 dry a mp ir era t 80 φ ure =1 00 –0.1 Dev iatio φ= n (h 80 )

Water pump

Evaporative cooling

Absorption chiller

9

20

40

50

60 70 Dry-bulb temperature

80

90

100

1.5 Evaporative and refrigerated inlet systems.

under-sized chiller which is not capable of bringing the air temperature down to the ambient dew point temperature, but in conjunction with evaporative cooling the same effect can be achieved. In this manner the system is taking advantage of evaporative cooling to reduce the load of refrigeration. Mid-compressor flashing of water in gas turbines Another manner of intercooling is the injection of demineralized water into the mid-stages of the compressor to cool the air and approach an isothermal compression process as shown in Fig. 1.6. The water is usually mechanically atomized so that very fine droplets are injected into the air. The water is evaporated as it comes in contact with the high-pressure and high-temperature air stream. As water evaporates, it consumes about 1058 Btu (1117 kJ) (latent heat of vaporization) at the higher pressure and temperature, resulting in lowering the temperature of the air stream entering the next stage. This reduces the work required to drive the compressor. The steam or water injected for cooling purposes also increases the mass flow through the system and therefore increases the power output of the turbine. The steam to be injected can be obtained from the use of an LP single-stage HRSG, and is injected at the first compressor section discharge and/or injection in the combustor if steam is being used for controlling the NOx output in the combustor.

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10

Combined cycle systems for near-zero emission power generation Water pump Exhaust 2

3 4

Combustor Compressor

Turbine

W

1

Air Water metering valve

Air manifold Water manifold

Blank-off/spacer 24 equally Feeder spaced water tubes spray nozzles

Water spray

8th stage bleed air piping

1.6 Mid-compressor cooling showing a schematic as well as an actual application in a GE LM 6000 engine. (Source: Courtesy GE Corporation.)

1.3.2 Recuperative/regeneration effect for Brayton cycle – gas turbines In a simple gas turbine cycle the turbine exit temperature is nearly always appreciably higher than the temperature of the air leaving the compressor. Obviously, the fuel requirement can be reduced by the use of a regenerator in which the hot turbine exhaust gas preheats the air between the compressor and the combustion chamber. Figure 1.7 shows a schematic of the regenerative cycle and its performance in the T-S diagram. In an ideal case the flow through the regenerator is at constant pressure. The regenerator effectiveness is given by the following relationship:

ηreg =

T3 − T2 T5 − T2

[1.1]

Thus the overall efficiency for this system’s cycle can be written as

ηRCYC =

(

−

) − (T2 − T1 ) ( − )

© Woodhead Publishing Limited, 2012

[1.2]

Combined cycle power plants 6

11

5 2

3

4

Turbine w

Combustor Compressor 1 T

4

5 3

6

2 1 S

1.7 The regenerative gas turbine cycle.

1.3.3 Reheat effect for gas turbines The reheat cycle increases the turbine work and consequently the net work of the cycle. This is accomplished not by changing the compressor work or the turbine inlet temperature but instead by dividing the turbine expansion into two or more parts with constant pressure heating before each expansion. This cycle modification, as shown in Fig. 1.8, is known as reheating. By reasoning similar to that used in connection with intercooling, it can be seen that the thermal efficiency of a simple cycle is lowered by the addition of reheating, whereas the work output is increased. However, a combination of regenerator and reheater can increase the thermal efficiency.

1.3.4 The intercooled regenerative-reheat gas turbine The Carnot cycle is the optimum cycle and all cycles attempt to reach this optimum. Maximum thermal efficiency is achieved by approaching the isothermal compression and expansion of the Carnot cycle, or by intercooling in compression and reheating in the expansion process. Figure 1.9 shows the intercooled regenerative-reheat cycle, which approaches this optimum cycle in a practical fashion. This cycle achieves the maximum efficiency and work output of any of the modified Brayton cycles. With the insertion of an intercooler in the compressor, the pressure ratio for maximum efficiency becomes much higher.

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12

Combined cycle systems for near-zero emission power generation Q

Q T

2

3

5

4

3

5

6

1

4

6

2 Compressor

Turbine

Turbine

1 S

1.8 Reheat cycle and T-S diagram. 8 Q a′

1

2′

b

3

Q

Q 4

5′

6

7′ w

Compressor

Turbine 4

T

5

6 5′ 7′ 7

3 2′ 2 a a′ b

8

1

S

1.9 The intercooled regenerative-reheat split-shaft gas turbine cycle.

1.4

The Rankine cycle (steam turbine)

The Rankine cycle, employing water/steam as the working fluid, is the most common thermodynamic cycle utilized in the production of electrical power. It is the cycle utilized by a steam turbine. A schematic of a steam power plant is shown in Fig. 1.10. Water enters the boiler feedwater pump at point 1 and is pumped isentropically into the boiler. The compressed liquid at 2 is heated until it becomes saturated at 2a, after which it is evaporated

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Combined cycle power plants

13

2 2a 2b

3

Turbine

G

Boiler 4

1 P

Condenser

1.10 Schematic of a steam turbine power plant.

to steam at 2b and then superheated to 3. The steam leaves the boiler at 3, expands isentropically in the ideal engine to 4, and passes to the condenser. Circulating water condenses the steam to a saturated liquid at 1, from which state the cycle repeats itself. The thermodynamic diagrams corresponding to the steam power plant in Fig. 1.10 showing the thermodynamic states are shown in the pressure volume (PV) diagram in Fig. 1.11 and the temperature-entropy (T-S) diagram in Fig. 1.12.

1.4.1 The regenerative-reheat Rankine cycle steam turbine It is evident from the Rankine cycle shown in Fig. 1.10 and 1.12 that a considerable amount of heat is required to raise the temperature of the water from 2 to 2a. The Rankine cycle has the disadvantage that the fluid temperature at the pump discharge is much lower than the fluid temperature at the turbine inlet. One way of overcoming this disadvantage is to use the internal system heat rather than the external heat to minimize this difference in the temperatures. This concept is called regenerative heating. In gas turbines, regenerative heating is accomplished using the high-temperature exhaust gases. In steam turbines, IP steam rather than exhaust steam is used for heating the feedwater. As the HP steam expands through the steam turbine, the steam gets very wet in the LP section of the steam turbine. This wet steam is detrimental for a turbine; it results in reduction of efficiency and also in nozzle and blade erosion. The reheat cycle involves heating of the steam withdrawn after partial expansion. This idea, combined with regenerative heating for improved thermal efficiency, is common practice in central power plants.

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14

Combined cycle systems for near-zero emission power generation 2a

2

2b

3

P

1

4

V

1.11 Pressure–volume diagram of a typical steam turbine power plant. 3 Superheater

T 2a

Evaporator

Economizer P = Constant (2f to 3)

2

2b Steam turbine

Pump 4a

Condenser 4

1

S

1.12 Temperature entropy diagram of a typical steam turbine power plant.

A simplified concept of the regenerative-reheat steam cycle is depicted in Fig. 1.13 and the thermodynamic cycle of the same is shown in Fig. 1.14. The water enters the first pump at point 1, from where it enters the feedwater heater at point 2. In the LP economizer/feedwater heater, the pressurized condensate is heated by part of the steam extracted from the HP turbine at an IP, point 6. The rest of the extracted steam is reheated in the reheater and enters the turbine at point 7. The heated water enters the second pump at point 3, from where it enters the boiler at point 4. The compressed liquid at 4 is heated until it becomes saturated at 4a, after which it is evaporated

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Combined cycle power plants

G

Turbine 4

6 4a 4b

7

5

3

8

Reheater

Boiler

P

15

Feed water heater 2 1

Condenser

P

1.13 Schematic of a regenerative-reheat steam turbine power plant.

LP Reheater/superheater Critical point

T HP economizer

4a

7

5 HP superheater

HP evaporator

4b

HP steam turbine

IP steam turbine

4 Pump LP economizer 3 2

6a LP evaporator

6

8a

Pump 8

Condenser 1

S

1.14 Temperature entropy diagram of a regenerative-reheat steam turbine power plant.

to steam at 4b and then superheated to 5. The steam leaves the boiler at 5, expands isentropically in the ideal engine to 6, and in the real case to 6a, where it is extracted for regeneration and reheating. The superheated steam now leaves the reheater at 7, expands isentropically in the ideal engine to 8 and in the real case to 8a, where it passes to the condenser. Circulating water condenses the steam to a saturated liquid at 1, from which state the cycle repeats itself.

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16

Combined cycle systems for near-zero emission power generation

1.5

The Brayton–Rankine cycle (gas turbine and steam turbine)

The combination of the gas turbine with the steam turbine is an attractive proposal, especially for electric utilities and process industries where steam is being used. The schematic of this cycle is shown in Fig. 1.15. The hot gases from the gas turbine exhaust are used in a steam recovery steam generator which may be supplementary fired to produce superheated steam at high temperatures for a steam turbine. Figure 1.16 shows the distribution of the energy in a Brayton–Rankine combined cycle. About 40% of the energy is converted to power by the gas turbine; the remaining 60% of the energy is collected in the HRSG and is used to power a steam turbine which produces about 20% of the energy as power. The computations of the combined Brayton and Rankine cycle are divided into three parts: 1. The work of the Brayton cycle (gas turbine) is the same as already outlined. 2. The work of the Rankine cycle (steam turbine) is also the same as already shown for that cycle. 3. The computation of the exchange of the heat from the gas turbine to the steam turbine needs to be further examined. In the HRSG – also known as the WHB – heat is exchanged from the gas turbine exhaust to the steam condensate. Experience worldwide in CCPPs indicates high levels of availability, reliability and thermal efficiency in most of these plants. Exhaust 4s Steam generator 3a

2a

4a 1s

Combustor w1

Steam turbine

w2 2s

1a

Compressor

Gas turbine

Condenser

Air

3s Pump

1.15 Schematic of the combined cycle.

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Combined cycle power plants

17

Fuel input 100%

Gas turbine output 38%

Radiation losses 0.5%

Radiation losses 0.3%

Energy in exhaust 61.5%

Stack 10%

Steam turbine output 21%

Radiation losses 0.2%

Condenser 30%

1.16 Energy flow diagram of CCPP.

1.6

Combined cycle power plant configurations

Large CCPPs are combinations of single-shaft CCPPs or multiple-shaft CCPPs. The plant arrangements shown in Fig. 1.17a are categorized as multishaft 1:1:1, that is, one gas turbine, one HRSG and one steam turbine.The plant arrangements shown in Fig. 1.17b and 1.17c are categorized as single-shaft 1:1:1 and single shaft without and with clutch respectively. If the desired plant output is higher than can be produced by a single gas turbine plant, other possible arrangements in multiple-shaft CCPP configurations are 2:2:1 (two gas turbines, two HRSGs and one steam turbine), which is the most common in plants above 300 MW, as shown in Fig. 1.18a, or 3:3:1 (three gas turbines, three HRSGs and one steam turbine), as shown in Fig. 1.19a. Corresponding arrangements for the single-shaft CCPP configurations are 2 × (1:1:1) or 3 × (1:1:1), that is, two or three trains of single-shaft arrangement, as shown in Fig. 1.18b and 1.18c or Fig. 1.19b and 1.19c respectively. Arrangements for the single-shaft CCPP configurations are just multiple of trains. For example two 1500 MW plants in Europe have five single-shaft CCPP trains, and a major 2800 MW plant in China has eight single-shaft CCPP trains. The primary advantage of a single-shaft CCPP over a multiple-shaft CCPP is that combining the gas turbine and steam turbine generators into

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18

Combined cycle systems for near-zero emission power generation ST

GEN.

GT

GEN.

COND. HRSG (a) 1:1:1 multi-shaft arrangement GT

ST

GEN.

HRSG

COND. (b) 1:1:1 single-shaft arrangement without clutch GT

GEN.

CL

ST

HRSG COND. (c) 1:1:1 single-shaft arrangement with clutch

1.17 (a–c) The single gas turbine plant arrangements (1:1:1). (Source: The Single Shaft Combined Cycle Myth by Ram G. Narula ASME Paper No. 2000-GT-0594 IGTI-Munich 2000.)

a single larger generator results in one less generator, main transformer and associated electrical subsystems. However, as the overall plant size increases, requiring multiple gas turbines, the advantage of one less generator/transformer grows proportionally smaller. As an example, the saving in electrical costs (on a $/kW basis) of going from four to three generators is not as great as going from two generators to one. The primary disadvantage of a single-shaft CCPP is that the number of steam turbines, condensers and condensate systems – and perhaps the cooling towers and circulating water systems – increases to match the number of gas turbines. For multiple-shaft CCPPs, there is only one steam turbine, condenser and the rest of the heat sink for up to three gas turbines; only their size increases. The design having only one larger steam turbine and heat sink results in lower cost because of economies of scale. Further, a larger steam turbine also allows use of a higher pressure and more efficient steam cycle. Thus the overall plant size and the associated number of gas turbines required have a major impact on whether single-shaft CCPP or multipleshaft CCPP is more economical. The total plant gross output and the individual gross outputs of the gas turbine generator and steam turbine generator, the overall plant efficiency and the heat rate for typical plants is shown in Table 1.1 for both the configurations. The numbers shown in these tables are for F technology gas

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Combined cycle power plants ST

GEN.

GT

GEN.

GT

GEN.

19

COND. HRSG HRSG (a) 2:2:1 multi-shaft arrangement GT

ST

GT

ST

GEN.

HRSG COND. GEN.

HRSG COND. (b) 2x(1:1:1) single-shaft arrangement without clutch GT

GEN.

CL

ST

GT

GEN.

CL

ST

HRSG COND.

HRSG COND. (c) 2x(1:1:1) single-shaft arrangement with clutch

1.18 (a–c) The dual gas turbine plant arrangements. (Source: The Single Shaft Combined Cycle Myth by Ram G. Narula ASME Paper No. 2000GT-0594 IGTI-Munich 2000.)

turbines and represent the best of the F technology gas turbines available from the major suppliers. The main point here is that from a performance point of view it does not matter whether the turbine configuration is a single-shaft or a multiple-shaft CCPP.

1.7

NOx emissions

CCPPs with the advanced new gas turbines have had a very positive impact on the environment as compared with other types of power plants. The new low NOx combustors have reduced NOx levels below 10 ppm. Figure 1.20 shows how in the past 30 years the use first of steam (wet combustors) injection in the combustors and then, in the 1990s, the dry low NOx combustors have greatly reduced the NOx output. New units under development aim to reduce NOx levels below 9 ppm. Catalytic converters have also been used in conjunction with both these types of combustors to reduce the NOx emissions even further.

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20

Combined cycle systems for near-zero emission power generation

COND.

ST

GEN.

GT

GEN.

GT

GEN.

GT

GEN.

HRSG HRSG HRSG (a) 3:3:1 multi-shaft arrangement GT

ST

GT

ST

GT

ST

GEN.

HRSG COND. GEN.

HRSG COND.

HRSG

GEN.

COND. (b) 3x(1:1:1) single-shaft arrangement without clutch GT

GEN.

CL

ST

GEN.

CL

ST

GEN.

CL

ST

HRSG GT HRSG GT

COND. COND.

HRSG COND. (c) 3x(1:1:1) single-shaft arrangement with clutch

1.19 (a-c) The triple gas turbine plant arrangements. (Source: The Single Shaft Combined Cycle Myth by Ram G. Narula ASME Paper No. 2000GT-0594 IGTI-Munich 2000.)

New research in combustors, such as catalytic combustion, has great promise, and values as low as 2 ppm can be attainable in the future. Catalytic combustors are already being used in some engines under the US Department of Energy’s (DOE) Advanced Gas Turbine Program, and have obtained very encouraging results.

1.8

Carbon capture and sequestration

Carbon capture and/or sequestration of the exhaust gases in a CCPP are very costly processes, both in initial costs – amounting to an additional cost of around $350/kW – and performance-based costs where the plant output could be reduced by about 7% and the plant efficiency reduced by about

© Woodhead Publishing Limited, 2012

© Woodhead Publishing Limited, 2012

156

67.3

243

480

309

384

143

107

164

52.7

54.7

52.9

52

67.6

121.6

Multiple-shaft CCPPs (1 × 1 × 1) Multiple-shaft CCPPs (2 × 2 × 1) Multiple-shaft CCPPs (3 × 3 × 1) Single-shaft CCPPs (1 × 1 × 1)

189.2

Steam turbine Overall output plant MW eff. %

50 Hz

Single gas Total plant turbine output output MW Plant arrangement MW

Table 1.1 Data on CCPP

6474 (6840)

6238 (6580)

6450 (6805)

6561 (6935)

Overall plant heat rate Btu/kW-h (kJ/kW-h)

271

309

529

262

Total plant output MW

172

67.3

170

170

Single gas turbine output MW

60

99

107

189

92

Steam turbine output MW

57.6

54.7

56.5

56

Overall plant eff. %

5923 (6250)

6238 (6580)

6038 (6375)

6093 (6425)

Overall plant heat rate Btu/kW-h (kJ/kW-h)

Combined cycle systems for near-zero emission power generation

NOx emissions (ppm)

22

200 180 160 140 120 100 80 60 40 20 0 1970

Water injection Dry low NOx combustor 1975

1980

1985

1990 Years

1995

2000

Catalytic combustor 2005

2010

1.20 Control of gas turbine NOx emissions over the years.

6%. Thus for a typical 630 MW plant, power loss could be between 45 and 50 MW and the plant overall efficiency reduced by 3–5%. Before CO2 gas can be sequestered from power plants and other point sources, it must be captured as a relatively pure gas. On a mass basis, CO2 is the nineteenth largest commodity chemical in the United States. Existing capture technologies, however, are not cost-effective when considered in the context of sequestering CO2 from power plants. Most power plants and other large point sources use air-fired combustors, a process that exhausts CO2 diluted with nitrogen. Flue gas from coal-fired power plants contains 10–12% CO2 by volume, whereas flue gas from natural gas combined cycle plants contains only 3–6% CO2. For effective carbon sequestration, the CO2 in these exhaust gases must be separated and concentrated. The US DOE’s Carbon Sequestration Atlas of the United States and Canada documents more than 3500 billion metric tons of CO2 storage potential in oil and gas reservoirs, coal seams and saline formations. Preliminary estimates suggest the availability of more than 1100 years of CO2 storage for the United States and Canada in these geologic formations. Currently projected and historical carbon emissions are shown in Fig. 1.21. The trends in part reflect increasing growth worldwide in the use of coal and fossil fuels for generating electricity. The replacement of existing plants and the use in new plants of more efficient power cycles will mitigate part of the projected increase.

1.9

Plant operation

1.9.1 Base or peaking CCPPs are not, as were originally planned, base-loaded plants. It is not uncommon for the plant to be cycled from 40% to 100% load in a single day, every

© Woodhead Publishing Limited, 2012

Combined cycle power plants

14

23

Billions of tons of carbon emitted per year

ath

dp

te jec

tly

pro

en urr

C

7

Historical emissions No Kyoto cap: Emerging & dev. Kyoto cap not ratified: USA, AUS Kyoto cap ratified: EU25, RUS, UKR, CAN, JAP.

1954

2004

Flat path Kyoto reduction commitment (cap 2008-2012)

Enforced (ratified) Kyoto reduction commitment (cap 2008-2012)

2054

2104

1.21 Estimated carbon emissions, historic and future paths.

day of the year. The intended plant operation mode has some bearing on the choice of multi-shaft or single-shaft CCPP, especially for the larger plants with multiple gas turbines. A larger multi-shaft CCPP is inherently more efficient than the corresponding single-shaft CCPP because of the larger steam turbines and, as such, is more suitable as a base-load plant. However, if the plant has to be operated as a cycling plant with multiple starts and with extended hours at part-loads as low as 40%, then a single-shaft CCPP with multiple smaller turbines may be more desirable. Further, if the plant must operate in simple cycle mode because of phased construction or when the steam turbine is down, then a multi-shaft CCPP or single-shaft CCPP with a clutch is more desirable than a single-shaft CCPP without a clutch.

1.9.2 Start-up techniques The start-up of a gas turbine is done by the use of electrical motors, diesel motors, and in plants where there is an independent source of steam, by a steam turbine. New turbines use the generator as a motor for start-up. After combustion occurs and the turbine reaches a certain speed, the motor declutches and becomes a generator. Use of a synchronous clutch between two rotating pieces of equipment is not new. It is very common in use with start-up equipment. In the case of single-shaft CCPPs a synchronous clutch can be used to connect the steam turbine to the gas turbine. However, use of a clutch in transmitting over 100 MW of power is new and has not found

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24

Combined cycle systems for near-zero emission power generation

unequivocal customer acceptance. While use of a synchronous clutch leads to additional space requirements, additional capital and O&M costs, and potentially reduced availability, it does offer the tangible benefit of easy and fast plant start-up. A major drawback of a single-shaft CCPP with a clutch is that the generator installation and maintenance and power evacuation are more complex and costly because the generator is located in the middle.

1.10

Availability and reliability

The availability of a power plant is the percentage of time the plant is available to generate power in any given period at its acceptance load. The acceptance load or the net established capacity would be the net generating capacity of the power plant, at design or reference conditions, established as a result of the performance tests conducted for acceptance of the plant. The actual power produced by the plant would be corrected to the design or reference conditions and is the actual net available capacity of the power plant. Thus it is necessary to calculate the effective forced outage hours which are based on the maximum load the plant can produce in a given time interval when the plant is unable to produce the power required of it. The effective forced outage hours is based on the following relationship: EFH

HO x

(MW Wd − MWa ) MW Wd

[1.3]

where • MWd = Desired output corrected to the design or reference conditions. This must be equal to or less than the plant load measured and corrected to the design or reference conditions at the acceptance test. • MWa = Actual maximum acceptance test produced and corrected to the design or reference conditions. • HO = Hours of operation at reduced load. The availability of a plant can now be calculated by the following relationship, which takes into account stoppages due to both forced and planned outages, as well as the forced effective outage hours: A=

(

−

− PT

−

)

where • PT = Time period (8760 h/year) • PM = Planned maintenance hours

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[1.4]

Combined cycle power plants • •

25

FO = Forced outage hours EFH = Equivalent forced outage hours.

Reliability of a plant depends on many parameters, such as the type of fuel, the preventive maintenance programs, the operating mode, the control systems and the firing temperatures. The reliability of the plant is the percentage of time between planned overhauls and is defined as: R=

(

−

− PT

)

[1.5]

Availability and Reliability have a major impact on the economics of the plant. Reliability is essential, in that when the power is needed it must be there. When the power is not available it must be generated or purchased and can be very costly in the operation of a plant. Planned outages are scheduled for non-peak periods. Peak periods are when the majority of the income is generated, usually there are various tiers of pricing depending on the demand. Many power purchase agreements have clauses containing capacity payments, thus making plant availability critical in the economics of the plant. A 1% reduction in plant availability could cost $500 000 in income on a 100 MW plant. Starting reliability (SR) is another very important factor in a plant. This reliability is a clear understanding of the successful starts that have taken place and is given by the following relationship: =

Number of starting successes (Number of starting successes Numbe b r of starting failures)

[1.6]

The insurance industry concerns itself with the risks of equipment failure. For advanced gas turbine CCPPs, the frequency and severity of failures are major concerns. In engineering terms, however, risk is better defined as: Risk = Probability of failure × Consequences of failure

[1.7]

where the consequences of failure include the repair/replacement costs and the lost revenue from the downtime to correct the failure. Actions taken to reduce the probability and/or consequences of failure tend to reduce risk and generally enhance insurability. Because of the high risks associated with insuring advanced gas turbines, demonstrated successful operation is important to the underwriting process. Operating and maintenance costs are usually broken up into two categories, variable costs and fixed costs. Variable operating costs include

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Combined cycle systems for near-zero emission power generation

consumables and spare parts, whereas fixed operating costs include staff, insurance, taxes and interest. The variable cost of a combined cycle is lower than a gas turbine, due to the fact that these costs are driven by the gas turbine spares, which can be distributed over a larger output of a combined cycle. As these CCPPs move towards coal-based fuels and the move towards higher turbine inlet temperatures and pressures continues, combined cycles will become more important. There is also significant interest in converting existing gas turbines in simple cycle operation to combined cycle operation, and existing steam plants to modified CCPPs. Results indicate that for a base-loaded plant running 5000 full load equivalent hours per year, the production costs are between 20% and 25% less for a combined cycle plant at a fuel cost of $30/bbl. At fuel costs of $120/bbl these costs are about 60–70% less. A commonly used rule of thumb is that 50–60% of the gas turbine output can be added by using the Rankine cycle (steam turbine) as the bottoming without any further fuel usage.

1.11

Major equipment

In a CCPP application there are four major categories which make up the plant system: 1. 2. 3. 4.

Advanced gas turbines HRSGs Steam turbines Condensers.

In the traditional combined cycle plant, air enters the gas turbine where it is initially compressed and then enters the combustor where it undergoes a very rapid increase in temperature at constant pressure. The high-temperature and HP air then enters the expander section where it is expanded to nearly atmospheric conditions. This expansion creates a large amount of energy, which is used to drive the compressor used in compressing the air, plus the generator where power is produced. The compressor in the gas turbine uses about 50–60% of the power generated by the expander. The air, on leaving the gas turbine, is essentially at atmospheric pressure conditions and at a temperature between 950°F and 1200°F (510–650°C). This air enters the HRSG, where the energy is transferred to the water to produce steam. There are many different HRSG units. Most HRSG units are divided into the same number of sections as the steam turbine. In most cases each section of the HRSG has a pre-heater, an economizer and feedwater, and then a superheater. The steam entering the steam turbine is superheated.

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1.11.1 Gas turbines Gas turbines used in major CCPP applications can be divided into two major types: industrial heavy-duty gas turbines and aeroderivative gas turbines. Industrial heavy-duty gas turbines These gas turbines were designed shortly after World War II and introduced to the market in the early 1950s. The early heavy-duty gas turbine design was largely an extension of steam turbine design. Restrictions of weight and space were not important factors for these ground-based units, and so the design characteristics included heavy wall casings split on horizontal centerlines, sleeve bearings, large-diameter combustors, thick airfoil sections for blades and stators and large frontal areas. The overall pressure ratio of these units varied from 5:1 for the earlier units to 35:1 for the units in present-day service. Turbine inlet temperatures have been increased and run as high as 2500°F (1371°C) on some of these units; this makes the gas turbine one of the most efficient prime movers on the market today, reaching efficiencies of 50%. Projected temperatures approach 3000°F (1649°C) and, if achieved, would make the gas turbine an even more efficient unit. The Advanced Gas Turbine Program sponsored by the US DOE has these high temperatures as one of its goals. To achieve these high temperatures, steam cooling is being used in the latest designs to achieve the goals of maintaining blade metal temperatures below 1300°F (704°C) and to prevent hot corrosion problems. The industrial heavy-duty gas turbines employ axial-flow compressors and turbines. The industrial turbine consists of a 15- to 17-stage axial-flow compressor with multiple can-annular combustors, each connected to the other by crossover tubes. The crossover tubes help propagate the flames from one combustor can to all the other chambers and also assures an equalization of the pressure between each combustor chamber. The earlier industrial European designs have single-stage side combustors; most of the newer European designs do not use the side combustor. The newer European designs have can-annular or annular combustors since side (silo type) combustors had a tendency to distort the casing. Figure 1.22 is a cross-sectional representation of the GE industrial type gas turbine, with can-annular combustors, and Fig. 1.23 is a cross-sectional representation of the Siemens silo type combustor gas turbine. The turbine expander consists of a 2- to 4-stage axial-flow turbine, which drives both the axial-flow compressor and the generator. The large frontal areas of these units reduce the inlet velocities, thus reducing air noise. The pressure increase in each compressor stage is reduced, creating a large, stable operating zone.

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Combined cycle systems for near-zero emission power generation

1.22 A frame type gas turbine with can-annular combustors. (Courtesy GE Turbines.) Generator drive shaft with hydraulic turning gear

Combustion chamber with 6 burners Combined thrust and journal bearing

Turbine admission chamber

17-stage compressor

4-stage turbine

3 Compressor blade carries

Fixed turbine Adjustable compressor Turbine inlet guide vane outer casing support

Turbine inner casing

Builtup rotor with single center tie bolt and disk “Hirth” serration

Journal bearing

Flexible turbine support on 3 rods

Longitudinal section of the Siemens/KWU gas turbine

1.23 Frame type gas turbine with silo type combustors. (Courtesy Siemens.)

The auxiliary modules used on most of these units have gone through considerable hours of testing and are heavy-duty pumps and motors. The advantages of the heavy-duty gas turbines are their long life, high availability, and slightly higher overall efficiencies. The noise level from this type of turbine is considerably less than an aircraft type turbine. The heavyduty gas turbine’s largest customers are the electrical utilities, and independent power producers. Since the 1990s, the industrial turbines have been the bulwarks of most CCPPs.

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Aircraft-derivative gas turbines Jet gas turbines consist of two basic components: an aircraft-derivative gas generator and a free-power turbine. The gas generator serves as a producer of gas energy or gas horsepower and is derived from an aircraft engine modified to burn industrial fuels. Design innovations are usually incorporated to ensure the required long-life characteristics in the ground-based environment. In the case of fan jet designs the fan is removed and a couple of stages of compression are added in front of the existing LP compressor. The axialflow compressor in many cases is divided into two sections: an LP compressor followed by an HP compressor. In these cases there is usually an HP turbine and an LP turbine that drive the corresponding sections of the compressor. The shafts are usually concentric; thus the speeds of the HP and LP sections can be optimized. The power turbine is separate and is not mechanically coupled – the only connection is via an aerodynamic coupling. Therefore in a case where there is an LP and HP compressor with a power turbine, the gas turbine has three shafts, all operating at independent speeds. The gas generator serves to raise combustion gas products to conditions of around 45–75 psi (3–5 Bar) and temperatures of 1300–1700°F (704–927°C) at the exhaust flange. Figure 1.24 shows a cross section of an aeroderivative engine. The power industry uses these units in a combined cycle mode for power generation where the power requirements are less than 100 MW. They are used due to their compactness and fast turnaround at major overhauls, the gas generator, which is the usual source of problems, being easy to replace with a spare gas generator while the problem generator is sent to the manufacturer or authorized repair facility. The benefits of the aeroderivative gas turbines are the following: 1. Favorable installation cost: The equipment involved is of such a size and weight that it can be packaged and tested as a complete unit within the manufacturer’s plant. Generally, the package will include either a genLP and power turbine LP compressor

HP compressor HP turbine

Est. weight 4500 lbs

1.24 A cross section of an aeroderivative gas turbine engine.

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Combined cycle systems for near-zero emission power generation

erator or a driven pipeline compressor and all auxiliaries and control panels specified by the user. Immediate installation at the job site is facilitated by factory matching and debugging. 2. Adaptation to remote control: Users strive to reduce operating costs by automation of their systems. Jet gas turbine equipment lends itself to automatic control; as auxiliary systems are not complex, water cooling is not required (cooling by oil-to-air exchanges), and the starting device (gas expansion motor) requires little energy and is reliable. Safety devices and instrumentation adapt readily for purposes of remote control and monitoring the performance of the equipment. 3. Maintenance concept: The off-site maintenance plan fits in well with these systems where minimum operating personnel and unattended stations are the objectives. Technicians conduct minor running adjustments and perform instrument calibrations. Otherwise, the aeroderivative gas turbine runs without inspection until monitoring equipment indicates distress or sudden performance change. This plan calls for the gasifier section (the aero-engine) to be removed and sent back to the factory for repair while another unit is installed. The power turbine does not usually have problems since its inlet temperature is much lower. Downtime due to the removal and replacement of the gasifier turbine is about 8 h.

1.11.2 Heat recovery steam generators The HRSG is located between the gas turbine and the steam turbine, where the condensate (water) is heated to superheated steam by the gas turbine exhaust. Properties of the gas in the gas turbine exhaust The exhaust from gas turbines in the simple cycle or regenerative mode represents a large amount of waste heat. The gas consists of a large amount of oxygen, since gas turbines use only about 10% of the air in the combustion process. As gas turbines operate with large amounts of excess air, about 18% oxygen is available in the exhaust and this allows supplemental firing. Mass flow rates can vary, depending on the size of the gas turbine, from 4 lb/s (1.8 kg/s) for a 180 kW unit to a tremendous 1100 lb/s (500 kg/s) for a large 143 MW gas turbine. In most cases, exhaust temperatures will be between 900°F (482°C) and 1100°F (593°C). Gas turbines utilizing a regenerator may have exhaust gas temperatures of around 950°F (510°C). Temperatures after supplemental firing can be as high as 1600°F (871°C), and it is often possible to double the steam production by supplemental firing. Variation in the exhaust gas specific heat (cp) for combustion is linear with temperature, varying from 0.259 Btu/lb/°R (1.084 kJ/kg/°K) at 800°F (427°C) to about 0.265 Btu/lb/°R (1.1094 kJ/kg/°K) at 900°F (482°C).

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When computing exhaust heat usable for a cogeneration project, capability variations in the gas turbine performance must be taken into account. These include ambient temperature variations, altitude effects and humidity. The effect of back pressure on the gas turbine capability is also important. Back pressure from heat recovery equipment should not exceed 10 inches of water (gage). The effect of hot ambient temperatures on gas turbine performance can be significant from a cogeneration perspective, as both electrical output and steam generation may drop off. Typically, in a combined cycle plant there is about a 0.5% drop in power output for every 1°F rise in temperature. This drop has led to a number of schemes, such as evaporative cooling and chilling of the inlet air, to overcome the problem. Most gas turbine manufacturers have a set of curves for their different gas turbine models showing steam-generating capabilities under different conditions. If curves are not available, a relatively simple calculation with regard to the HRSG/WHB can be done to get a ballpark estimate. In order to compute the amount of steam generated, the following items are required: • • • •

Gas turbine mass flow rate Exhaust gas temperature Steam conditions pressure and temperature required Feedwater temperature.

Some general rules of thumb may be stated, as they are implicit in the computation. First, the steam outlet temperature will be 50°F (28°C) or more below the exhaust gas temperature. The exhaust gas leaving the boiler evaporator must be 40°F (22°C) greater than the saturation temperature. The pinch point is taken to be not less than 40°F (22°C) and the water entering the evaporator must be 15°F (8°C) below the saturation temperature. Based on a simple heat balance the steam flow rate is: msteam =

mair × cp × (T2 T1 ) × K

( h2

h1 )

[1.8]

where msteam = Steam flow rate (lb/h) (kg/h) mair = Gas turbine mass flow rate (lb/h) (kg/h) T2 = Exhaust gas temperature T1 = Stack exhaust temperature (saturation + 40°F (22°C); it also will vary depending on the type of fuel – fuel with low sulfur or no sulfur can have lower exhaust temperatures) cp = Specific heat at constant pressure (average between T1 and T2) h2 = Steam final enthalpy (leaving HRSG)

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Combined cycle systems for near-zero emission power generation h1 = Enthalpy of water entering evaporator K = Radiation loss factor (approximately 0.985).

More details on heat recovery unit design are presented in the next section. It must be noted that any process fluid may be utilized in the heat recovery unit. Characteristics of heat recovery steam generators The waste heat recovery system is a critically important subsystem of a cogeneration system. In the past it was viewed as a separate ‘add-on’ item. This view is being changed with the realization that good performance, both thermodynamically and in terms of reliability, grows out of designing the heat recovery system as an integral part of the overall system. Some important points and observations relating to gas turbine waste heat recovery are the following: Multipressure steam generators – These are becoming increasingly popular. With a single-pressure boiler there is a limit to the heat recovery because the exhaust gas temperature cannot be reduced below the steam saturation temperature. This problem is avoided by the use of multipressure level steam generators. Pinch point – This is defined as the difference between the exhaust gas temperature leaving the evaporator section and the saturation temperature of the steam. Ideally, the lower the pinch point, the more the heat recovered, but this calls for more surface area and consequently increases the back pressure and cost. Also, excessively low pinch points can mean inadequate steam production if the exhaust gas is low in energy (low mass flow or low exhaust gas temperature). General guidelines call for a pinch point of 15–40°F (8–22°C). The final choice is obviously based on economic considerations. Approach temperature – This is defined as the difference between the saturation temperatures of the steam and the inlet water. Lowering the approach temperature can result in increased steam production, but at increased cost. Conservatively high approach temperatures ensure that no steam generation takes place in the economizer. Typically, approach temperatures are in the 10–20°F (5.5–11°C) range. Figure 1.25 is the temperature energy diagram for a system and also indicates the approach and pinch points in the system. Off-design performance – This is an important consideration for waste heat recovery boilers. Gas turbine performance is affected by load, ambient conditions and gas turbine health (fouling, etc.). This can affect the exhaust gas temperature and the air flow rate. Adequate consideration must be given to bow steam flows (LP and HP), and superheat temperatures vary with changes in the gas turbine operation. Economizers – These are used to heat the condensate entering the HRSG from the condenser. In the economizer, the condensate (water) is maintained

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Exhaust gas

Temperature

Super heater

Pinch point

Approach temperature Economizer

Energy transfer

1.25 Energy transfer diagram in an HRSG of a CCPP.

as a liquid while the temperature is increased. The condensate is in the tubes that usually have fin-tube design. Evaporators – These usually utilize a fin-tube design. Spirally finned tubes of 1.25–2 in outer diameter (OD) with three to six fins per inch are common. In the case of unfired designs, carbon steel construction can be used and boilers can run dry. As heavier fuels are used, a smaller number of fins per inch should be utilized to avoid fouling problems. Forced circulation system – Using forced circulation in a waste heat recovery system allows the use of smaller tube sizes with inherent increased heat transfer coefficients. Flow stability considerations must be addressed. The recirculating pump is a critical component from a reliability standpoint and standby (redundant) pumps must be considered. In any event, great care must go into preparing specifications for this pump. Backpressure considerations (gas side) – These are important, as excessively high backpressures create performance drops in gas turbines. Very low pressure drops would require a very large heat exchanger and more expense. Typical pressure drops are 8–10 inches of water. Types of heat recovery steam generator There are many factors in the selection of the HRSG, but essentially the choice depends on initial cost and the efficiency of the overall plant. A 1%

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Combined cycle systems for near-zero emission power generation

105 104 103 102 101 100 99 98

Relative price Relative net output Relative net efficiency

ng l H e-p R re SG s s pl ure an Si t ng on lep H ce res R th s SG ro ur e pl ugh an t D ua H l-p R re SG s s pl ure an t Tr ip l H e-p R re SG s s pl ure an Tr t H ip R le SG -p r w ess .re u h re D ua pla eat n l t H -hi R gh SG pr w es .r s eh ur ea e t

97 96 95

Si

Relative cost, power, and efficiency

increase in efficiency can support a 3–4% increase in cost; this is due to the fact that the initial cost contributes 10%, maintenance cost contributes 15% and fuel cost contributes about 75%, respectively, to the life-cycle cost. The typical HRSG units used are the drum type HRSG. The drum type HRSG is composed of drums, separate economizers, separate superheaters, circulation systems, generating tubes and blowdown systems. In the past decade once-through heat recovery steam generators (OTSGs) have evolved into a cost-competitive and technologically advanced HRSG. Water tube drum units were developed to prevent scaling, corrosion and control of the steam-generating process. Today’s technology in the areas of materials, water treatment systems and modern control technology does not require the traditional boiler components. This enables the OTSG to be more competitive both in initial costs and in erection and installation costs. The most important effect on performance of the combined cycle as a result of the HRSGs is whether the HRSG is a three-pressure or dual-pressure steam cycle, whether or not reheat is used. Based on manufacturer data of the F turbines, we note that the dual pressure is about 1–2% lower than the triple pressure and about 5–7% for reheat systems, as shown in Fig. 1.26. However, in some cases much higher pressure dual-pressure HRSGs are used, making up for the loss of production in the IP steam turbine, and actually resulting in higher output and efficiency. Figure 1.23 also introduces the OTSG. These steam generators have a once-through flow path on the water/ steam side, therefore all water that enters leaves as steam without any recirculation, thus getting rid of the traditional steam drums. These OTSGs are in many cases faster, cheaper and have smaller footprints than the traditional

Types of CCPPs

1.26 Comparison of cost, output power and cycle efficiency of various types of CCPPs.

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HRSGs. They are used mostly in smaller plants below 70 MW. An important point to remember is that an increase in cycle efficiency of 1–2% can justify an increase in initial cost of about 3–5%. The increase in cost and efficiency between single-pressure and triplepressure systems as well as the increase in power is shown in Table 1.2. Duct combustors There are several reasons for supplemental firing of a waste heat recovery unit. Probably the most common is to enable the system to track demand (i.e., produce more steam when the load swings upward than the unfired unit can produce). This may enable the gas turbine to be sized to meet the baseload demand with supplemental firing taking care of higher load swings. Raising the inlet temperature at the WHB allows a significant reduction in the heat transfer area and, consequently, the cost. Typically, as the gas turbine exhaust has ample oxygen, duct burners can be conveniently used. An advantage of supplemental firing is the increase in heat recovery capability (recovery ratio). A 50% increase in heat input to the system increases the output 94%, with the recovery ratio increasing by 59%. Some important design guidelines to ensure success include the following: • • • •

Special alloys may be needed in the superheater and evaporator to withstand the elevated temperatures. The inlet duct must be of sufficient length to ensure complete combustion and avoid direct flame contact on the heat transfer surfaces. If natural circulation is utilized, an adequate number of risers and feeders must be provided as the heat flux at entry is increased. Insulation thickness on the duct section must be increased.

1.11.3 Steam turbines The steam turbine consists of four basic parts: the rotor that carries the blades or buckets; the stator consisting of cylinder and casing, which are Table 1.2 Typical cost and efficiency increases in CCPPs

Type of process

Increase in cost

Increase in efficiency

Increase in power

Single-pressure HRSG Single-pressure OTSG Dual-pressure HRSG Triple-pressure HRSG Dual-pressure reheat HRSG Triple-pressure reheat HRSG

1.0 0.98 1.025 1.03 1.04 1.035

1 1 1.015 1.02 1.035 1.02.7

1 1 1.015 1.02 1.035 1.027

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Combined cycle systems for near-zero emission power generation

often combined and within which the rotor turns; the nozzles or flow passages for the steam, which are generally fixed to the inside of the cylinder; and the frame or base for supporting the stator and rotor. In small steam turbines, the cylinder casing and frame are often combined. Several other systems, such as the lubrication systems, steam piping systems and a condensing system, make up the rest of the turbine. There are two major types of flow characteristics in steam turbines, the impulse turbine and the reaction turbine. The steam volume increases whenever the pressure decreases, but the resulting velocity changes depend on the type of turbine. These velocity changes are distinguishing characteristics of the different types of turbines. The degree of reaction (R) in an axial-flow turbine is defined as the ratio of the change of enthalpy drop in the rotor to the change in total enthalpy drop across the stage:

R=

H rotor Hstage

[1.9]

By definition, the impulse turbine has a degree of reaction equal to zero. This degree of reaction means that the entire enthalpy drop is taken in the nozzle, and the exit velocity from the nozzle is very high. In practice, there must be a pressure drop across the rotating blades to generate flow. Since there is no change in enthalpy in the rotor, the relative velocity entering the rotor equals the relative velocity exiting from the rotor blade. Most steam turbine HP stages are typically impulse stages by design but average a 5% reaction at full load. For a symmetric flow (50% reaction) the enthalpy drop in the rotor is equal to the drop in the stationary part of the turbine. This also leads to equal pressure drop across the stationary and rotating parts. Due to the difference in the turbine blade diameters at the tip and the root of the blade, the reaction percentages are different to counteract the centrifugal forces acting on the steam flow. If this is not done too much flow would migrate to the blade tips. In IP and LP turbines the basic design is reaction; however, these turbines are about 10% reaction at the root, while at the tip going from about 60% reaction to about 70% reaction in case of the LP turbine. The impulse turbine produces about twice the power output of the 50% reaction type turbine; the reaction turbine, on the other hand, is more efficient. Thus the combination of HP stages being impulse and the later stages being about 50% reaction produces a high-power and-efficiency turbine. In most large plants the axial-flow steam turbine consists of three sections: an HP turbine stage with pressures between 1450 and 4500 psia (100– 310 Bar), with temperatures as high as 1212°F (656°C); an IP turbine stage

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with pressures between 300 and 600 psia (20.6–41.7 Bar); and an LP turbine stage with pressures between 90 and 60 psia (6.1–3.1 Bar). Before it enters the IP stage, the steam leaving the HP stage is usually reheated to about the same temperature as to the HP stage inlet temperature. On leaving the IP stage, the steam is mixed with steam from the LP superheater and then enters the LP stage. The steam from the LP turbine enters the condenser. The condenser is maintained at a vacuum of between 0.13 and 0.033 Bar. The increase in back pressure in the condenser will reduce the power produced. Care must be taken to ensure that the steam leaving the LP stage blades does not have a high liquid content. High liquid content in the steam leads to erosion of the leading edge of the last stages of the LP turbine blading. Steam turbines used in modern CCPPs are simple machines. The following are the important requirements for a modern combined cycle steam turbine: 1. 2. 3. 4. 5. 6.

Ability to operate over a wide range of steam flows High efficiency over a large operating range Reheat possibilities Fast start-up Short installation time Floor mounted installations.

The plants operate over a wide range of steam flows as the plants are now often cycled between base load and 50% of the base load in a 24-h period. This requires a high efficiency over a wide operating range. CCPPs operate at many pressure levels. It is not uncommon for the same manufacturer’s plant using the same gas turbine to operate at two- or three-pressure levels. The combined cycle steam turbine has fewer or even no bleed points as compared to anywhere between 4 and 8 bleed points for feedwater heaters in fired boiler power plants. Rapid start-up is very often important since many of these plants are started up on a daily basis. Great care must be exercised in the design due to rapid increase in temperatures. This does not allow for rotor wheels, which use a shrink fit on to the shaft. There have been cases during rapid start-up where ‘walking’ of the rotor wheel on the shaft occurred due to the different growth rates between the shaft and the rotor wheel. Steam reheat is used in many of today’s steam turbines, improving overall combined cycle efficiency. There are five basic types of steam turbine shaft and casing arrangement: single casing, tandem-compound, cross-compound, double-flow and extraction steam turbines. Figure 1.27 shows schematics of various steam turbine arrangements.

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Single-flow single casing extraction steam turbines In a single-flow turbine the steam enters at one end, flows once through the blading in a direction approximately parallel to the axis, emerges at the other end and enters the condenser. In turbines with a single casing, all sections are contained within one casing and the steam path flows from throttle to exhaust through that single casing. Figure 1.27a shows a simple path where steam enters a turbine and is exhausted to the atmosphere or a condenser. It also shows an extraction for cogeneration purposes. This is the most common arrangement in small and moderately large turbines. The extraction flow steam turbine is the term applied to a turbine where part of the flow is extracted for various reasons such as steam for the plant or for absorption type chillers or for any other plant process. These turbines may be back pressure or condensing, depending on the application, and are used most commonly in a cogeneration application. Compound-flow or tandem-compound turbine Compound-flow or tandem-compound turbine is the term applied to a machine in which the steam passes in sequence through two or more separate units, expanding in each. The two units arranged in a tandem-compound design have both casings on a single shaft and driving the same electrical unit (G). In most cases the HP exhaust is returned for reheating before

HP

G

HP

IP

G

G

Extraction (b) Compound flow

(a) Single casing HP

IP

LP

Boiler

HP

RHB

G

IP

G DFLP

(c) Compound turbine with reheat and double flow LP turbine

G

(d) Cross-compound steam turbine with reheat and double flow LP turbine

1.27 (a–d) A schematic of arrangements of various types of steam turbines.

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entering the IP turbine. Most often the HP and the IP portions are in one casing and the LP portion in another. The IP exhausts into crossover piping and on to one or more LP turbines. Figure 1.27b shows two sections being used, an HP and an LP section. Figure 1.27c indicates a condition where split steam is used in a double-flow, low-pressure (DFLP) section. LP turbines are typically characterized by the number of parallel paths available to the steam. The steam path through the LP turbines is split into parallel flows because of steam conditions and practical limitations on blade length. Typically, several LP flows in parallel are required to handle the large volume of flow rates. The steam enters at the center and divides, the two portions passing axially away from each other through separate sets of blading on the same rotor. This type of unit is completely balanced against end thrust and gives a large area of flow through two sets of LP blading. Cross-compound turbine A cross-compound design typically has two or more casings, coupled in series on two shafts, with each shaft connected to a generator. In cross-compound arrangements the rotors can rotate at different speeds but cannot operate independently as they are aerodynamically coupled. The cross-compound design is inherently more expensive than the tandem-compound design, but has a better heat rate, so that the choice between the two is one of economics. Figure 1.27d shows a turbine setup where there are three casings, each with their own generator, and a reheat boiler (RHB) between the HP and the IP section. Figure 1.28 is a schematic of a three-section steam turbine in two casings. The HP and IP turbine section is in one casing and the LP turbine is a double-flow turbine in another casing. Figure 1.29 is a photograph of the lower half of a steam turbine casing housing the HP and IP turbines. Figure 1.30 is a close-up of the regulating (first stage) nozzle and the regulating stage rotor followed by the other stages, nozzles and rotors. Steam turbine characteristics To understand the effect of the steam operating conditions on efficiency and load in a steam turbine at optimum operating conditions, two types of steam turbine, condensing and back pressure, must be considered. The significant operating parameters of a steam turbine are the steam inlet temperature and pressure, and the exhaust back pressure or the vacuum drawn for condensing steam turbines. The variations in these parameters affect steam consumption and efficiency. In a 100-MW steam turbine at a pressure of about 600 psia (41.4 Bar) and 660°F (350°C), a 1% reduction in steam consumption can cost between $500 000/year and about $900 000/year. This is based on a boiler efficiency of about 85% and lower heating value (LHV) of fuel at 18 900 Btu/lb (10 500 Kcal/kg).

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Combined cycle systems for near-zero emission power generation Low-pressure stages

Crossover piping HP turbine inlet Nozzle Thrust and High box journal pressure bearings stages

Journal bearing Journal bearings

Rotor

Front Pedestal

Pedestal To reheater

Extractions IP turbine inlet

Intermediate pressure stages

To condenser

1.28 A typical compound turbine.

1.29 Lower half of an HP-IP turbine casing with the rotor in the casing.

Turbine steam inlet pressure is a major parameter affecting turbine performance. To retain the design efficiency the steam inlet pressure should be maintained. Lowering steam inlet pressure reduces turbine efficiency and increases steam consumption. A 10% increase in steam pressure will reduce the steam consumption by about 1% in a condensing steam turbine and

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Regulating stage rotor

Rotors

Diaphragm/nozzles Regulating stage nozzle

1.30 HP control and subsequent stages showing nozzles/diaphragms and rotors.

by about 4% in a back pressure steam turbine. The effect on efficiency for 10% increase in pressure is about 1.5% for a condensing steam turbine and 0.45% for a back pressure steam turbine. Turbine steam inlet temperature is another major parameter affecting turbine performance. Reducing steam inlet temperature reduces the enthalpy, which is a function of both the inlet temperature and pressure. At higher steam inlet temperatures, heat extraction by the turbine will also be increased. An increase of about 100°F (55°C) will reduce the steam consumption by about 6.6% in a condensing steam turbine and 8.8% in a back pressure turbine. At this temperature, efficiency will increase by 0.6% for a condensing steam turbine and 0.65% for a back pressure turbine. It should be noted that the overall efficiency in most cases for a condensing steam turbine (30–35%), is about twice that of a back pressure turbine (18–20%). In condensing or exhaust back pressure steam turbines, the increase of this back pressure will reduce the efficiency and increase the steam consumption where all other operating parameters remain constant. In condensing steam turbines, the condenser vacuum temperature will also increase if the removal of heat from the condenser is reduced. Thus in a water-cooled condenser, if the temperature of the inlet water is increased, the power produced by the turbine is decreased because the back pressure will be increased. In summary, condensing steam turbines are more efficient and produce more power than back pressure steam turbines. The cost of a condensing steam turbine is about $25/kW more than a back pressure turbine.

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Steam turbines traditionally have down (or bottom) exhaust, with the condenser located below the steam turbine. In such an arrangement, the steam pedestal is raised about 30 ft (10 m) above the base slab where the condenser sits. Care must be taken that none of the condensate splashes onto the turbine casing, otherwise casing distortion and cracks can occur. For smaller steam turbines up to 250 MW, axial exhaust is available from some steam turbine suppliers. Side exhaust is also available from some suppliers for steam turbines up to 450 MW. In the axial or side exhaust arrangement, the condenser is located side by side with the steam turbine. As a result, the steam turbine pedestal can be lowered by 15–21 ft (5–7 m); this reduces civil/structural costs. Further, if the steam turbine is enclosed in a building, then the building height is also reduced, further lowering civil/ structural costs. It should be noted that some of these savings are partially offset by the increased cost of main steam and cold/hot reheat piping, since lowering the pedestal increases the vertical length of this piping. With axial or side exhaust machines, proper attention must be paid to the moisture drainage and bearing lubrication aspects.

1.11.4 Types of condenser Large steam turbines are usually condensing type steam turbines. Condensers can be water-cooled or air-cooled units. Water-cooled condensers are more common and also more effective. The cooling effectiveness of the condenser is very important to the performance of the LP steam turbine. If the cooling effectiveness is reduced, then the back pressure of the LP steam turbine is increased, and the power output of the steam turbine is reduced. The use of multishaft CCPPs has one major benefit: the gas turbine and the steam turbine can be from two different manufacturers. A major limitation of the single-shaft CCPP is that it does not allow the flexibility of choosing the best available steam turbine for a given plant.

1.12 Sources of further information Boyce, M. P., Cogeneration and Combined Cycle Power Plants. Second Edition, April 2010; First Edition, January 2002. ASME Press, New York. Boyce, M. P., Perry’s Chemical Engineering Handbook. Eighth Edition, 2008, McGraw Hill. Chapter on Transport and Storage of Fluids. Boyce, M. P., ‘Combined Cycle Power Plants’, Power and Energy International, pp. 34– 35, Nov/Dec 2006, London. Boyce, M. P., Gas Turbine Engineering Handbook. Third Edition, February 2006, Elsevier Publisher; Second Edition, December 2001, Butterworth-Heinemann Publisher; First Edition 1985. Boyce, M. P., Operation of Combined Cycle Power Plants. June 2004, ASME, Vienna, Austria.

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Boyce, M. P., ‘Increase in Availability and Reliability of Combined Cycle Power Plants’, ASM International Conference on Combustion Turbines and Hot Section Components, Conference Proceedings, Pittsburgh, PA, October 2003. Boyce, M. P., ‘Advanced Condition Monitoring for Advanced Gas Turbines’, EPRI Conference, International Conference on Advances in Life Assessment and Optimization of Fossil Power Plants, Orlando, FL, March 2002. Boyce, M. P., Assessments of Trends in Condition Monitoring Systems and Life Cycle Cost Management for Combined Cycle Power Plants, DOE Morgantown December 2001. Boyce, M. P., ‘Reliability, Availability, and Maintainability Program for Advanced Gas Turbines in Combined Cycle Applications’, US DOE Next Generation Turbine Stakeholder Meeting, 9–10 April 2001. Boyce, M. P. ‘Performance Characteristics of a Steam Turbine in a Combined Cycle Power Plant’. Proceedings of the 6th EPRI Steam Turbine Generator Workshop, EPRI, Palo Alto, CA. August 1999.

© Woodhead Publishing Limited, 2012

2 Advanced industrial gas turbines for power generation M. P. BOYCE, The Boyce Consultancy Group, LLC, USA

Abstract: The chapter explores new trends in combined cycle power plants. The effects of major parameters such as temperature and pressure ratio on plant efficiency, power and operational ability are fully investigated. The characteristics, relationships and effects, both practical and theoretical, of each of the plant’s major components are discussed. The major equations governing most turbomachines, including equation of state, momentum equation, energy equation and conservation of mass are explained. Compressor issues and diffusion and dry low NOx type combustors are fully investigated. Impulse and reaction type turbines are discussed, together with blade cooling, coatings for adequate blade life and the Larson–Miller parameter, which computes the blade life for various materials and temperatures. Key words: combined cycle power plants (CCPPs), plant efficiency, compressors, combustors, impulse and reaction type turbines, turbine blades.

2.1

Introduction

The new trends for power generation around the world are the combined cycle power plants based on the new advanced technology gas turbines. The new advanced gas turbines are pushing the limits of technology in the areas of material science due to the very high firing temperatures, and of aerodynamics due to the very high-pressure ratios developed in the compressors. The dry low NOx combustors are also pushing the technology in the areas of combustion and flame stability. The concept that combined cycle power plants must be totally base-load operated is now a myth, as plant loads must vary during the day. This type of operation requires online total condition monitoring to ensure that the entire plant is operated at its optimum efficiency through the large operating range. The last 20 years have seen extensive growth in gas turbine technology. The growth is spearheaded by the growth of materials technology, new coatings, new cooling schemes and the growth of combined cycle power plants. This, in conjunction with the increase in compressor pressure ratio from 7:1 to as high as 45:1 and firing temperatures from 1400ºF to 2700ºF (760– 1482ºC), has increased the simple cycle gas turbine thermal efficiency from about 15% to over 45%. 44 © Woodhead Publishing Limited, 2012

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Figures 2.1 and 2.2 show the growth of the pressure ratio and firing temperature. They have grown in parallel, as both are necessary to achieve the optimum thermal efficiency. The increase in pressure ratio increases the gas turbine thermal efficiency when accompanied by an increase in turbine firing temperature. Figure 2.3 shows the effect on the overall cycle efficiency of the increasing pressure ratio and firing temperature. The increase in the pressure ratio increases the overall efficiency at a given temperature; however, increasing the pressure ratio beyond a certain value at any given firing temperature can actually result in lowering the overall cycle efficiency. In the past the gas turbine was perceived as a relatively inefficient power source when compared to other power sources. Its efficiencies were as low as 15% in the early 1950s; today, its efficiencies are in the 45–50% range, which translates to a heat rate of 7582 (8000 kJ/kW-h) to 6824 Btu/kW-h (7199 kJ kW-h). The limiting factor for most gas turbines has been the turbine inlet temperature. With new cooling schemes using steam or conditioned air 45 40 Pressure ratio

35 30 25

Pressure ratio aircraft Pressure ratio industrial

20 15 10 5 0 1940

1950

1960

1970 1980 Year

1990

2000

2010

2.1 Development of engine pressure ratio during 1950–2000. 1600 Development of single crystal blades

Temperature (°C)

1400 1200 1000

Temp aircraft

800 600

Temp industrial

400 200 0 1940

1950

1960

1970 1980 Year

1990

2000

2.2 Trend in improvement in firing temperature.

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2010

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Combined cycle systems for near-zero emission power generation

Thermal cycle efficiency (x)

Overall cycle efficiency Temp = 15°C Eff. comp. = 87% Eff. turb. = 92% 70 60 50

Overall Eff.@ 800°C Overall Eff.@1000°C Overall Eff.@1200°C Overall Eff.@ 1300°C Overall Eff.@ 1350°C Overall Eff.@1400°C Ideal cycle

40 30 20 10 0

0

5

10

15 20 25 Pressure ratio

30

35

40

2.3 Overall gas turbine cycle efficiency.

and breakthroughs in blade metallurgy, higher turbine temperatures have been achieved. The new gas turbines have fired inlet temperatures as high as 2600°F (1427°C), and pressure ratios of 40:1 with efficiencies of 45% and above. Advanced gas turbines have developed very high efficiencies of between 40% and 45% due to high-pressure ratio (30:1 for frame engines) as shown in Fig. 2.1, and high firing temperatures (2400°F, 1315°C) as shown in Fig. 2.2. The advantages of advanced gas turbines have been eclipsed by the following major problems experienced in their operation: 1. 2. 3. 4.

Lower availability (by up to 10%) Reduced life of nozzles and blades (averaging 15 000 h) Higher degradation rate (5–7% in first 10 000 h of operation) Instability of low NOx combustors.

The gas turbine consists of three major components: 1. Gas turbine compressor 2. Combustor 3. Gas expander, often called the turbine section of the gas turbine, consisting of (a) Gasifier turbine (b) Power turbine.

2.2

Gas turbine compressors

A compressor is a device that pressurizes a working fluid. In gas turbines the axial-flow compressor, which is a continuous flow compressor, is generally used for compressing the air. The pressure ratio per stage of the industrial axial-flow compressor is low because the operating range needs to be large. The operating range is the

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range between the surge point and the choke point. Figure 2.4 shows the operating characteristics of a compressor. The surge point is the point when the flow is reversed in the compressor. The choke point is the point at which the flow has reached Mach 1.0, that is, the point where no more flow can get through the unit, a ‘stone wall’ condition. When surge occurs, in addition to the flow being reversed, all the thrust forces thrust acting on the compressor are also reversed, especially the thrust forces, which can lead to total destruction of the compressor. Thus surge is a region that must be avoided. Choke conditions cause a large drop in efficiency, but do not lead to destruction of the unit. It is important to note that with the increase in pressure ratio and the number of stages, the operating range is narrowed. The compressors transfer energy by dynamic means from a rotating member to the continuously flowing fluid. Nearly all gas turbines producing over 5 MW that are used in the power industry use axial-flow compressors.

2.2.1 Thermodynamic and fluid mechanic relationships of a compressor The following relationships that govern the flow in an axial-flow compressor will be employed later in discussing the aerodynamic characteristics of the compressor: Equation of state P

ργ

= Const.

[2.1]

Surge line

Pressure ratio

Speed lines

Operational range

Choke point

Flow rate

2.4 Operating characteristics of a high-pressure compressor.

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Combined cycle systems for near-zero emission power generation

where P = pressure ρ = density γ = for an isentropic adiabatic process γ = cp/cV; where cp and cv are the specific heats of the gas at constant pressure and volume, respectively, and can be written as p

cv = R

[2.2]

where cp =

γR γ−

and cv =

R γ −1

[2.3]

Energy equation h1 +

V12 V2 + 1Q 2 h2 + 2 + 1W2 2 gc J 2 gc J

[2.4]

where h = enthalpy V = absolute velocity W = work Q = heat rejection J = mech. equiv. of heat gc = gravitational constant and Conservation of mass/continuity equation . m = A1 V1 ρ1 = A2 V2 ρ2

[2.5]

where A = area V = velocity ρ = density . m = Mass flow. The flow per unit cross-sectional area can be written as follows: ɺ m = A

γ P R T 1+ ( −

(

M (

)M2 )

+

)(

−

)

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[2.6]

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49

where P = total pressure T = total temperature γ = ratio of specific heats cp/cv. The Mach number (M) is defined as M=

V a

[2.7]

It is important to note that the Mach number is based on static temperature. Static conditions are the conditions of flow in a moving stream and total conditions occur when the flow is brought to rest in a reversible adiabatic manner. The acoustic velocity (a) in a gas is given by the following relationship: a2 =

∂P ∂ρ

[2.8] s c

for an adiabatic process (s = entropy = constant) the acoustic speed can be written as follows: a=

γ gc RT Ts MW

[2.9]

where Ts = static temperature. It is important to note that the pressure measured can be either total or static; however, only total temperature can be measured. For the total conditions of pressure and temperature to change, energy must be added to or extracted from the fluid stream. The relationship between total and static conditions for pressure and temperature is as follows: T0

Ts +

V2 2cp

[2.10]

where T0 = total temperature Ts = static temperature V = gas stream velocity and P0

Ps + ρ

V2 2 gc

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[2.11]

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Combined cycle systems for near-zero emission power generation

where P0 = total pressure Ps = static pressure. Equations [2.10] and [2.11] can be written in terms of the Mach number as follows: T0 ⎛ γ − 1 2⎞ = ⎜1+ M ⎟ ⎝ ⎠ Ts 2 and

[2.12]

γ

P0 ⎡ γ − 1 2 ⎤ γ −1 = 1+ M ⎥ Ps ⎣ 2 ⎦

[2.13]

2.2.2 Axial-flow compressors An axial-flow compressor compresses its working fluid by first accelerating the fluid and then diffusing it to obtain a pressure increase. The fluid is accelerated by a row of rotating airfoils or blades (the rotor) and diffused by a row of stationary blades (the stator). The diffusion in the stator converts the velocity increase gained in the rotor to a pressure increase. One rotor and one stator make up a stage in a compressor. A compressor usually consists of multiple stages. One additional row of fixed blades (inlet guide vanes) is frequently used at the compressor inlet to ensure that air enters the firststage rotors at the desired angle. In addition to the stators, an additional diffuser at the exit of the compressor further diffuses the fluid and controls its velocity when entering the combustors. In an axial compressor, air passes from one stage to the next with each stage raising the pressure slightly. By producing low-pressure increases of the order of 1.1:1–1.4:1, very high efficiencies can be obtained. The use of multiple stages permits overall pressure increases up to 40:1. Figure 2.5 shows multistage high-pressure axial-flow turbine rotor. The turbine rotor depicted in this figure has a low-pressure compressor followed by a high-pressure compressor. There are also two turbine sections. The reason for the large space between the two turbine sections is that this is a reheat turbine and the second set of combustors are located between the high-pressure and the low-pressure turbine sections. The compressor produces 30:1 pressure in 22 stages. As with other types of rotating machinery, an axial compressor can be described by a cylindrical coordinate system. The Z-axis is taken as running the length of the compressor shaft, the radius r is measured outward from the shaft, and the angle of rotation is the angle turned by the blades in Fig. 2.6.

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LP axial flow turbine

HP axial flow turbine HP axial flow compressor LP axial flow compressor

2.5 A high-pressure ratio turbine rotor. (Source: Courtesy ALSTOM.)

z

r θ

ω

2.6 Coordinate system for axial-flow compressor.

This coordinate system will be used throughout this discussion of axial-flow compressors. Figure 2.7 shows the pressure, velocity, and total enthalpy variation for flow through several stages of an axial compressor. It is important to note here that the changes in the total conditions for pressure, temperature and enthalpy occur only in the rotating component where energy is input into the system. As the air passes from one stage to the next with each stage raising the pressure and temperature slightly to the order of 1.1:1–1.4:1, very high efficiencies can be obtained. The use of multiple stages permits overall pressure increases

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Stator

Rotor

Rotor

Stator

Combined cycle systems for near-zero emission power generation

IGV

52

CL Po,To, Total pressure and temperature Ps,Ts, Static pressure and temperature

V Absolute velocity

2.7 Variation of flow and thermodynamic properties in an axial-flow compressor.

up to 40:1. As can be seen in this figure, the length of the blades and the annulus area, which is the area between the shaft and shroud, decrease through the length of the compressor. This reduction in flow area compensates for the increase in fluid density as it is compressed, permitting a constant axial velocity. In most preliminary calculations used in the design of a compressor, the average blade height is used as the blade height for the stage. The rule of thumb for a multiple-stage gas turbine compressor would be that the energy rise per stage would be constant, rather than the commonly held perception that the pressure rise per stage is constant. The energy rise per stage can be written as ΔH =

[ H 2 H1 ] NS

[2.14]

where H1, H2 = total inlet and exit enthalpy Btu/lbm (kJ/kg), respectively Ns = number of stages. Assuming that the gas is thermally and calorically perfect (cp, and γ are constant), Equation [2.14] can be rewritten in terms of total temperatures as shown below:

ΔT Tstage

( Tin ⎡( P2 P1 ) ⎣ = Ns

−

)

γ

− 1⎤ ⎦⎥

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[2.15]

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where Tin = total inlet temperature (°F, °C) P1, P2 = total inlet and exit pressure (psia, bar). To understand the flow in a turbomachine, the concepts of absolute and relative velocity must be grasped. Absolute velocity (V) is gas velocity with respect to a stationary coordinate system. Relative velocity (W) is the velocity relative to the rotor. In turbomachinery the air entering the rotor will have a relative velocity component parallel to the rotor blade and an absolute velocity component parallel to the stationary blades. Mathematically, this relationship is written as    [2.16] V U +W where the absolute velocity (V) is the vector addition of the relative velocity (W) and the linear rotor velocity (U). The absolute velocity can be resolved into its components, the axial velocity (VZ) and the tangential component Vθ. As stated earlier, an axial-flow compressor operates on the principle of putting work into the incoming air by acceleration and diffusion. Air enters the rotor as shown in Fig. 2.5 with an absolute velocity (V1) and an air angle α1, which combines vectorially with the tangential velocity of the blade (U1) to produce the resultant relative velocity W1 at a discharge air angle α2. The relative velocities should be parallel to the blade. Air flowing through the passages formed by the rotor blades has a relative velocity W2 at an angle α4, which is less than α2 because of the camber (turning) of the blades. The relative velocity at the exit W2 is less than that at W1, resulting from an increase in the passage width as the blades become thinner towards the trailing edges. Therefore, some diffusion will take place in the rotor section of the stage. The combination of the relative exit velocity and blade velocity produces an absolute velocity V2 at the exit of the rotor. The air then passes through the stator, where it is turned through an angle so that the air is directed into the rotor of the next stage with a minimum incidence angle. The incidence angle (i) is the angle between the actual air angle (α) and the blade angle (β). The air entering the rotor has an axial component of the absolute velocity VZ1 and a tangential component Vθ1. Figure 2.8 also shows the movement of air in the compressor blades and the resultant velocity triangles in an axial-flow compressor. The following relationships are obtained: V12

Vθ21 + VZ21

V2 2

Vθ22 + VZ22

W12

(U 1 − Vθ1 )2 VZ21

W2 2

(U 2 − Vθ 2 )2 + VZ22 (U

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[2.17]

54

Combined cycle systems for near-zero emission power generation

Inlet guide vane

α1 α2

V1

Vz Vθ

U

Rotor

Direction of rotation

W2 α4

α3

V2

Vz U

Vθ Stator

2.8 A typical velocity diagram for an axial-flow compressor.

The Euler turbine equation, a simplification of the momentum equation depicting the total energy transfer, is given as follows: E=

ɺ m (U 1Vθ1 − U 2Vθ 2 ) (U gc

[2.18]

The total energy (E) when written for a unit mass flow is the head produced (Hstage) in a stage per unit mass: Hstage =

1 (U 1Vθ1 − U 2Vθ 2 ) (U gc

[2.19]

By placing relationships defined in Equation [2.17] into Equation [2.19], the following relationship is obtained: Hstage =

(

1 ⎡ 2 VZ 2 gc ⎣

)

VZ22 + (U 12 U 22 ) + (W22 W12 )⎤ ⎦

[2.20]

The change in diameter between the inlet and exit diameter of a stage is negligible; therefore the blade speeds (U) and the axial velocity (Vz) remain constant, reducing Equation [2.20] to: stage

=

UV Vz (tan α 1 − tan t n gc

3)

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[2.21]

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55

and the pressure ratio across the stage can be written as: ⎫⎪ Vz P2 ⎧⎪ UV =⎨ ⎡⎣tan α 2 − tan α 4 ⎤⎦ + 1⎬ P1 ⎩⎪ gc pTin ⎭⎪

γ γ +1

[2.22]

Degree of reaction (compressor) The degree of reaction in an axial-flow compressor is defined as the ratio of the change of static head in the rotor to the head generated in the stage: R=

H rotor Hstage

[2.23]

The change in static head in the rotor is equal to the change in relative kinetic energy: H rotor =

1 (W2 2 (W 2 gc

W12 )

[2.24]

Hrotor is negative since work is being put into the system and W12

Vz21 + (Vz1 tan α 1 )2

[2.25]

W2 2

Vz22 + (Vz2 tan α 3 )2

[2.26]

Thus, the reaction of the stage can be written as: R=

Vz 2U

(tan α 2 + tan t n

3)

[2.27]

The 50% reaction stage is widely used, since an adverse pressure rise on either the rotor or stator blade surfaces is minimized for a given stage pressure rise. When designing a compressor with this type of blading, the first stage must be preceded by inlet guide vanes to provide pre-whirl and the correct velocity entrance angle to the first-stage rotor. With a high tangential velocity component maintained by each succeeding stationary row, the magnitude of W1 is decreased. Thus, higher blade speeds and axial-velocity components are possible without exceeding the limiting value of 0.70–0.75 for the relative inlet Mach number. Higher blade speeds result in compressors of smaller diameter and less weight. Another advantage of the symmetrical stage comes from the equality of static pressure rises in the stationary and moving blades, resulting in a maximum static pressure rise for the stage. Therefore, a given pressure ratio can

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Combined cycle systems for near-zero emission power generation

be achieved with a minimum number of stages, a factor in the lightness of this type of compressor. A disadvantage of the symmetrical stage is the high exit loss resulting from the high axial-velocity component. In stationary turbine applications, where weight and frontal area are of lesser importance, one of the other stage types is used. The term ‘asymmetrical stage’ is applied to stages with reaction other than 50%. The axial-inflow stage is a special case of an asymmetrical stage where the entering absolute velocity is in the axial direction. The moving blades impart whirl to the velocity of the leaving flow, which is removed by the following stator. From this whirl the major part of the stage pressure rise occurs in the moving row of blades with the degree of reaction varying from 60% to 90%. The stage is designed for constant energy transfer and axial velocity at all radii so that the vortex flow condition is maintained in the space between blade rows. Many of the gas turbine axial-flow compressors designed before 1990 have blade profiles which are based on the NACA 65 series blades (NASA SP-36, Aerodynamic Design of Axial-Flow Compressors) and the double circular arc blades. The new high-pressure and high-loading compressors require a higher Mach number to increase their capacity and efficiency. The diffusion bladings increase the loading at the tips and tend to distribute the loading equally between the rotor and the tip. Efficiencies in the later stages of multiple-stage axial-flow compressors are lower than in the earlier stages due to the distortions of the radial flow.

2.2.3 Compressor operation characteristics A compressor operates over a large range of flow and speed delivering a stable head/pressure ratio. During start-up, the compressor must be designed to operate in a stable condition at low rotational speeds. There is an unstable limit of operation known as ‘surging’, and it is shown on the performance map as the surge line. The surge point in a compressor occurs when the compressor back pressure is high and the compressor cannot pump against this high head, causing the flow to separate and reverse its direction. Surge is a reversal of flow and is a complete breakdown of the continuous steady flow through the whole compressor. It results in mechanical damage to the compressor due to the large fluctuations of flow which produce changes in direction of the thrust forces on the rotor, creating damage to the blades and the thrust bearings. The phenomenon of surging should not be confused with the stalling of a compressor stage. Stalling is the breakaway of the flow from the suction side of the blade aerofoil, causing an aerodynamic stall. A multistage compressor may operate stably in the unsurged region with one or more of the stages stalled and the rest of the stages unstalled.

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Compressor surge Compressor surge is a phenomenon of considerable interest, yet it is not fully understood. It is a form of unstable operation and should be avoided. Unfortunately it occurs frequently, sometimes with damaging results. Surge has been traditionally defined as the lower limit of stable operation in a compressor, and it involves the reversal of flow. This reversal of flow occurs because of some kind of aerodynamic instability within the system. Usually a part of the compressor is the cause of the aerodynamic instability, although it is possible for the system arrangement to be capable of augmenting this instability. Compressors are usually operated at a working line, separated by some safety margin from the surge line. Surge has been extensively investigated. Poor quantitative universality or aerodynamic loading capacities of different blades and stators and an inexact knowledge of boundary layer behavior make the exact prediction of flow in the compressor at the offdesign stage difficult. A decrease in the mass flow rate, an increase in the rotational speed of the impeller or both can cause the compressor to surge. Whether surge is caused by a decrease in flow velocity or an increase in rotational speeds, the blades or the stators can stall. One should note that operating at higher efficiency implies operation closer to surge, and that total pressure increases occur only in the rotational part of the compressor, the blades. To make the curve general, the concept of aerodynamic speeds and corrected mass flow rates has been used in the performance maps. The surge line slope on multistage compressors can range from a simple single parabolic relationship to a complex curve containing several break points or even ‘notches’. The complexity of the surge line shape depends on whether or not the flow limiting stage changes with operating speed from one compression stage to another; in particular, very closely matched stage combinations frequently exhibit complex surge lines. In the case of compressors with variable inlet guide vanes, the surge line tends to bend more at higher flows than with units which are speed controlled. Surge is usually linked with excessive vibration and an audible sound, but there have been cases where a surge not accompanied by an audible sound has caused failures. Usually, operation in surge and, often, near surge is accompanied by several indications, including general and pulsating noise level increases, axial shaft position changes, discharge temperature excursions, compressor differential pressure fluctuations and lateral vibration amplitude increases. With high-pressure compressors, operation in the incipient surge range is frequently accompanied by the emergence of a low frequency, asynchronous vibration signal which can reach predominant amplitudes, as well as excitation of various harmonics of blade passing frequencies. Extended operation in surge causes thrust and journal bearing

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Combined cycle systems for near-zero emission power generation

failures. Failures of blades and stators are also experienced due to axial movement of the shaft causing contact of blades and stators. Due to large flow instabilities experienced, severe aerodynamic stimulation of one of the blade’s natural response frequencies is caused, leading to blade failure. The performance map of an axial-flow compressor displays the variation of total pressure ratio across a compressor as a function of corrected mass flow (usually expressed as percent of design value), at a series of constant corrected speed lines (Nc). The axial-flow compressor adiabatic efficiencies (ηc) are shown as islands on the performance map and can also be depicted aginst corrected mass flow as shown for a representative multistage compressor in Fig. 2.9. On a given corrected speed line, as the corrected mass flow is reduced the pressure ratio (usually) increases until it reaches a limiting value on the surge line. For an operating point at or near the surge line the ‘orderly’ (i.e., nearly axisymmetric) flow in the compressor tends to break down (flow becomes asymmetric with rotating stall) and can become ‘violently’ unsteady. Thus the surge line is a locus of unstable compressor operating points and is to be avoided. To cope with this, one specifies the surge margin (SM), defined as:

SM =

(PR

surge

− PR working

)

[2.28]

PR working

In Equation [2.28], PRsurge/working denotes the pressure ratio on the surge/ working line at the same corrected mass flow rate; thus the corrected speed 140

Compressor pressure ratio, percent design

C

Constantefficiency lines

120

A

100 80 Stall-limit line

60

Rotating-stall 40 region 20 0

B (o) 40

60

70

80

90

100 110

N/√θ, percent design

90 100 110 50 60 70 80 Equivalent weight flow, percent design

2.9 Multistage axial-flow compressor map.

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would be higher for operating points on the surge line. For operation on the constant corrected speed line an alternative definition for surge margin in terms of corrected mass flow on the working line and on surge line at the same corrected speed would be preferable. For stable operation of multistage compressor a surge margin is specified. Compressors are designed to operate at a condition referred to as the design point. At design point the various stages mounted on the same shaft are matched aerodynamically, that is, the inlet flow to each stage is such that the stage is at the design point and this occurs for only one combination of corrected speed and mass flow (for this reason the design point is also known as match point). While the design point is one at which the compressor will operate most of the time, there are situations of low-speed operation during the starting of gas turbines where the compressor must also provide adequate pressure rise and efficiency. For compressor operations at corrected speed and corrected mass flow different from those at design, difficulties arise due to the requirements of matching the inlet flow of a stage to the outlet flow from the upstream stage. To illustrate this, consider changes along constant corrected speed line. The effect of reduction in mass flow relative to the working line results in a higher pressure rise and therefore a greater increase in density in the first stage than was predicted at design. The greater increase in density means the second stage has an even lower value of flow coefficient than the first stage, with an even greater increase in density. The effect is cumulative, so that the last stage approaches stall while the front stage is only slightly altered. Conversely, increasing the mass flow relative to working line would result in a lower pressure rise and therefore a smaller increase in density. The smaller increase in density means the second stage has an even higher value of flow coefficient than the first stage, with an even smaller increase in density. The consequence is that the last stage approaches stalling at negative incidence with low efficiency performance. Similarly, one can also show that reducing the rotational speed along the working line through the design point can lead to stalling of front stages and windmilling of rear stages. Methods for coping with low-speed difficulties include use of compressor air bleed at intermediate stage, use of variable geometry compressor and use of multi-spool compressors, or combinations of the above. Compressor choke The compressor choke point is when the flow in the compressor reaches Mach 1 at the blade throat, a point where no more flow can pass through the compressor. This phenomenon is often known in the industry as ‘stone walling’. The more the stages, the higher the pressure ratio and the smaller the operational margin between surge and choke regions of the compressor, as shown previously in Fig. 2.4 and 2.9.

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Compressor stall There are three distinct stall phenomena. Individual blade stall and rotating stall are aerodynamic phenomena. Stall flutter is an aeroelastic phenomenon. Individual blade stall This type of stall occurs when all the blades around the compressor annulus stall simultaneously without the occurrence of a stall propagation mechanism. The circumstances under which individual blade stall is established are unknown at present. It appears that the stalling of a blade row generally manifests itself in some type of propagating stall and that individual blade stall is an exception. Rotating stall Rotating, or propagating stall, was first observed by Whittle and his team on the inducer vanes of a centrifugal compressor. Rotating stall (propagating stall) consists of large stall zones covering several blade passages. It propagates in the direction of the rotation and at some fraction of rotor speed. The number of stall zones and the propagating rates vary considerably. Rotating stall is the most prevalent type of stall phenomenon. The propagation mechanism can be described by considering the blade row to be a cascade of blades as shown in Fig. 2.10. A flow perturbation causes blade 2 to reach a stalled condition before the other blades. This stalled blade does not produce a sufficient pressure rise to maintain the flow around it, and an effective flow blockage or a zone of reduced flow develops. This retarded flow diverts the flow around it so that the angle of attack increases on blade 3 and decreases on blade 1. In this way, a stall ‘cell’ may move along the cascade in the direction of the lift on the blades. The stall propagates downward relative to the blade row at a rate about half the rotational speed; the diverted flow stalls the blades below the retarded-flow zone and unstalls the blades above it. The retarded flow or stall zone moves from the pressure side to the suction side of each blade in the opposite direction to that of rotation. The stall zone may cover several blade passages. The relative speed of propagation has been observed from compressor tests to be less than the rotor speed. Observed from an absolute frame of reference, the stall zones appear to be moving in the direction of rotor rotation. The radial extent of the stall zone may vary from just the tip to the whole blade length. Stall flutter This phenomenon is caused by self-excitation of the blade. It must be distinguished from classic flutter, since classic flutter is a coupled torsional-flexural

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Direction of rotation

1 Retarded flow 2

Direction of propagation

3

2.10 Propagating stall in a blade cascade.

vibration that occurs when the free-stream velocity over a wing or airfoil section reaches a certain critical velocity. Stall flutter, on the other hand, is a phenomenon that occurs due to the stalling of the flow around a blade. Blade stall causes Karman vortices in the airfoil wake. Whenever the frequency of these vortices coincides with the natural frequency of the airfoil, flutter will occur. Stall flutter is a major cause of compressor blade failure.

2.3

Gas turbine combustors

All gas turbine combustors perform the same function: they increase the temperature of the high-pressure gas. Combustor inlet temperature depends on engine pressure ratio, load and engine type, and whether or not the turbine is regenerative or non-regenerative, especially at the low-pressure ratios. The new industrial turbine pressure ratios are between 17:1 and 35:1, which means that the combustor inlet temperatures range from 850°F (454°C) to 1200°F (649°C). Combustor exit temperatures range from 1700°F (927°C) to 2900°F (1593°C). The new aircraft engines have pressure ratios in excess of 40:1. Combustor performance is measured by efficiency, the pressure decrease encountered in the combustor and the evenness of the outlet temperature profile. Combustion efficiency is a measure of combustion completeness. Combustion completeness affects fuel consumption directly, since the heating value of any unburned fuel is not used to increase the turbine inlet temperature. Normal combustion temperatures range from 3400°F (1871°C) to 3500°F (1927°C). At this temperature, the volume of nitric oxide in the

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combustion gas is about 0.01%. If the combustion temperature is lowered, the amount of nitric oxide is substantially reduced. The use of natural gas and the new dry low NOx combustors has reduced NOx levels below 10 ppm. Figure 2.11 shows how, in the past 30 years, the reduction of NOx by first the use of steam (wet combustors) injection in the combustors, and then in the 1990s, the dry low NOx combustors has greatly reduced the NOx output. New units under development have goals of reducing NOx levels below 9 ppm. Catalytic converters have also been used in conjunction with both these types of combustors to even further reduce the NOx emissions. New research in combustors, such as catalytic combustion, shows great promise, and values of as low as 2 ppm can be attainable in the future. Catalytic combustors are already being used in some engines under the US Department of Energy’s (DOE), Advanced Gas Turbine Program, and have obtained very encouraging results.

2.3.1 Typical combustor arrangements There are different methods to arrange combustors on a gas turbine. Designs fall into three main categories: 1. Can-annular 2. Annular 3. Silo-type combustor. Can-annular and annular combustors

NOx emissions (ppm)

Most American large gas turbines have can-annular combustors, as shown in Fig. 2.11. Figure 2.12 shows a set of can-annular combustors on a frame type gas turbine. There are 10–16 cans in an annular arrangement on a single gas turbine. The can-annular combustors are easy to maintain as each can be 200 180 160 140 120 100 80 60 40 20 0 1970

Water injection Dry low NOx combustor 1975

1980

1985

1990 Years

1995

2000

Catalytic combustor 2005

2.11 Control of gas turbine NOx emissions during 1975–2000.

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2.12 GE frame 6 gas turbine with DLE can-annular combustors. (Source: Courtesy GE.)

removed easily and worked on independently. In most can-annular designs each can is connected with the can next to it through a ‘cross-over tube’. The cross-over tubes are used to equalize the pressure in each can; they are also used during start-up to allow the flame to travel from the two igniter cans to all the other cans. This ensures start-up reliability. Can-annular combustors can be of the straight-through or reverse-flow design. If can-annular cans are used in aircraft, the straight-through design is used, while a reverse-flow design may be used on industrial engines. The annular designs are used on European frame type gas turbines. Figure 2.13 is a typical frame type annular combustor used in large gas turbines. Annular combustors are especially popular in new aircraft designs; however, the can-annular design is still used because of the developmental difficulties associated with annular designs. Annular combustor popularity increases with higher temperatures or low-Btu gases, since the amount of cooling air required is much less than in can-annular designs due to a much smaller surface area. The amount of cooling air required becomes an important consideration in low-Btu gas applications, since most of the air is used up in the primary zone and little is left for film cooling. Annular combustors are almost always straight-through flow designs. Development of a can-annular design requires experiments with only one can, whereas the annular combustor must be treated as a unit and requires much more hardware and a large amount of compressor flow.

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2.13 A typical industrial type annular combustor. (Source: Courtesy Siemens Corp.)

2.14 Silo-type combustor.

Silo-type combustors and side combustors These designs are found on large industrial turbines, especially European designs. Figure 2.14 shows a large frame type gas turbine with two silotype side combustors. Smaller side combustors and some small vehicular gas turbines have a combustor as shown in Fig. 2.15. They offer the advantages of simplicity of design, ease of maintenance and long life due to

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Air/gas duct combustor arrangement Burner assy.

Cooling air

Primary air Pressure probe Straightening vanes

Metallic tiles

Flame monitor and sight glass Sight glass Hot gas Secondary air to turbine Air from compressor

Hot gas to turbine

Air flow control vanes

2.15 A typical single can side combustor.

low heat release rates. These combustors may be of the straight-through or reverse-flow design. In the reverse-flow design, air enters the annulus between the combustor can and its housing, usually a hot gas pipe to the turbine. Reverse-flow designs have minimal length.

2.3.2 Combustion in combustors There are two types of combustion in combustors: • •

diffusion combustion dry low NOx (DLN) or dry low emission (DLE) combustion.

Gas turbine combustors have seen considerable design changes. The original diffusion type combustors were changed to wet combustors by adding water or steam in the combustion zone to restrict the amounts of NOx produced. Most new turbines have progressed to DLE NOx combustors from the wet diffusion type, which were injected by steam in the primary zone of the combustor. The diffusion combustors have a single nozzle, while most DLE combustors have multiple fuel nozzles for each can.

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The diffusion type combustor This is the most common combustor on the market but it is being displaced by the more complex DLN/DLE combustors. The gas turbine diffusion combustor uses very little of its air (10%) in the combustion process. The rest of the air is used for cooling and mixing. Many of the new combustors circulate steam in the combustor can casing for cooling purposes. The air from the compressor must be diffused before it enters the combustor. The velocity of the air leaving the compressor is about 400–600 ft/s (122–183 m/s) and the velocity in the combustor must be maintained below 50 ft/s (15.2 m/s). Even at these low velocities, care must be taken to avoid the flame being carried on downstream. The combustor is a direct-fired air heater in which fuel is burned almost stoichiometrically with one-third or less of the compressor discharge air. Combustion products are then mixed with the remaining air to arrive at a suitable turbine inlet temperature. Despite the many design differences in combustors, all gas turbine combustion chambers have three features: (1) a recirculation zone, (2) a burning zone (with a recirculation zone which extends to the dilution region) and (3) a dilution zone, as shown in Fig. 2.16. The air entering a combustor is divided so that the flow is distributed between three major regions: (1) primary zone, (2) dilution zone and (3) annular space between the liner and casing. Combustion takes place in the primary zone of a combustor. Combustion of natural gas is a chemical reaction that occurs between carbon (or hydrogen) and oxygen. Heat is given off as the reaction takes place. The products of combustion are carbon dioxide and water. The reaction is stoichiometric, which means that the proportions of the reactants are such that there are exactly enough oxidizer molecules to bring about a complete reaction

Recirculation Rec ecirculation ulation zone

Burning zone

Dilution Zone zone

2.16 A typical diffusion combustor can with straight-through flow.

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to stable molecular forms in the products. The air enters the combustor in a straight-through flow or reverse flow. Most aero-engines have straightthrough flow type combustors. Most of the large frame type units have reverse flow. The function of the recirculation zone is to evaporate, partly burn and prepare the fuel for rapid combustion within the remainder of the burning zone. Ideally, at the end of the burning zone, all fuel should be burnt so that the function of the dilution zone is solely to mix the hot gas with the dilution air. The mixture leaving the chamber should have a temperature and velocity distribution acceptable to the guide vanes and turbine. Generally, the addition of dilution air is so abrupt that if combustion is not complete at the end of the burning zone, chilling occurs which prevents completion. However, there is evidence with some chambers that if the burning zone is run over-rich, some combustion does occur within the dilution region. Figure 2.17 shows the distribution of the air in the various regions of the diffusion type combustor. The theoretical or reference velocity is the flow of combustor inlet air through an area equal to the maximum cross-section of the combustor casing. The flow velocity is 25 fps (7.6 mps) in a reverse-flow combustor, and between 80 (24.4 mps) and 135 fps (41.1 mps) in a straightthrough flow turbojet combustor. Figure 2.18 shows a typical diffusion type combustor used in frame type gas turbines. There may be 6–16 of this type of combustor placed in an annular configuration. Note that the main flow from the compressor in many of these combustors goes up between the combustor can and the liner and flows into the combustor liner at various points. It is therefore known as a reverse-flow combustor. Only about 18% of the flow enters the can at the top through the swirler where it combusts with the fuel. The rest of the flow enters the liner through a series of small holes around the liner diameter, keeping the liner and the can cool. Combustor performance is measured by efficiency, the pressure decrease encountered in the combustor and the evenness of the outlet temperature

Flame tube

10%

8% 82% 10% 18%

Primary zone

Dilution zone

28%

72%

2.17 Air distribution in a typical diffusion combustor.

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Combined cycle systems for near-zero emission power generation Outer combustion case assembly

Duel fuel nozzle assembly

Retractable spark plug Combustion air

Fuel gas

Flow sleeve

Cooling air

Transition piece assembly

Combustion zone

Liquid fuel oil

Dilution air

Atomizing air

Cooling air

Fl ow

Combustion case cover

Crossfire tube connection Slot-cooled liner assembly

Compressor discharge air

2.18 A typical reverse-flow diffusion can-annular combustor.

profile. Combustion efficiency is a measure of combustion completeness. Combustion completeness affects fuel consumption directly, since the heating value of any unburned fuel is not used to increase the turbine inlet temperature. The combustion efficiency is calculated as the ratio of the actual heat increase of the gas to the theoretical heat input of the fuel (lower heating value).

ηcomb =

ɺ a h2 (ma + m f )h3 − m Δhactual = ɺ f LHV Δhtheoretical m

[2.29]

where h2 = enthalpy leaving the compressor section h3 = enthalpy entering the turbine section . ma = mass of air flow . mf = mass of fuel flow LHV = lower heating value of the fuel. The loss of pressure in a combustor is a major problem, since it affects both the fuel consumption on a unit MW basis and power output. Pressure loss occurs in a combustor because of diffusion, friction and momentum. Total pressure loss is usually in the range of 2–4% of static pressure (compressor outlet pressure). The efficiency of the engine will be reduced by an equal percentage. The result is increased fuel consumption and lower power output, which affects the size and weight of the engine. The uniformity of the combustor outlet profile affects the useful level of turbine inlet temperature, since the average gas temperature is limited by

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the peak gas temperature. The profile factor is the ratio between the maximum exit temperature and the average exit temperature. This uniformity assures adequate turbine nozzle life, which depends on operating temperature. The average inlet temperature to the turbine affects both fuel consumption and power output. A large combustor outlet gradient will work to reduce average gas temperature and consequently reduce power output and efficiency. The traverse number is defined as the peak gas temperature minus mean gas temperature divided by the mean temperature rise in the combustor. Thus the traverse number must have a lower value – between 0.05 and 0.15 – in the nozzle. Equally important are the factors which affect satisfactory operation and life of the combustor. To achieve satisfactory operation, the flame must be self-sustaining, and combustion must be stable over a range of fuel-to-air ratios to avoid ignition loss during transient operation. Moderate metal temperatures are necessary to assure long life. Steep temperature gradients which warp and crack the liner must be avoided. Carbon deposits can distort the liner and alter the flow patterns to cause pressure losses. Smoke is environmentally objectionable as well as a fouler of heat exchangers. Minimum carbon deposits and smoke emissions also help ensure satisfactory operations. The ratio of the oxygen content at stoichiometric conditions and actual conditions is called the equivalence ratio. Ф = (oxygen/fuel at stoichiometric condition)/ (oxygen/fuel at actual condition)

[2.30]

The amount of oxygen in the combustion gas is regulated by controlling the ratio of air to fuel in the primary zone. The ideal volumetric ratio of air to methane is 10:1. If less than 10 volumes of air are used with 1 volume of methane, the combustion gas will contain carbon monoxide (CO). The reaction is as follows: CH 4 + 1 5(O2

N2 )

2 H 2 O + CO + 6 N 2

Heat

[2.31]

In gas turbines there is plenty of air, so the CO is very low, but it can be significant from an environmental emissions standpoint. Velocity is used as a criterion in combustor design, especially with respect to flame stabilization. The importance of air velocity in the primary zone is known. A transition zone is often included before the primary zone so that the high-velocity air from the compressor is diffused to a lower velocity and higher pressure and distributed around the combustion liner. The secondary, or dilution, air should only be added after the primary reaction has reached completion. Dilution air is added gradually so as not to quench the reaction in ‘conventional’ combustors. Flame tubes should be designed to

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produce a desirable outlet profile and to last a long time in the combustor environment. Adequate life is assured by film cooling of the liner. The air enters the annular space between the liner and casing and is admitted into the space within the liner through holes and slots because of the pressure difference. The design of these holes and slots divides the liner into distinct zones for flame stabilization, combustion and dilution, and provides film cooling of the liner. The liner experiences a high temperature because of heat radiated by the flame and combustion. To improve the life of the liner, it is necessary to lower the temperature of the liner and use a material which has a high resistance to thermal stress and fatigue. The air-cooling method reduces the temperature both inside and outside the surface of the liner. This reduction is accomplished by fastening a metal ring inside the liner to leave a definite annular clearance. Air is admitted into this clearance space through rows of small holes in the liner and is directed by metal rings as a film of cooling air along the liner inside.

2.3.3 Air pollution problems in a diffusion combustor Smoke In general, it has been found that much visible smoke is formed in small, local, fuel-rich regions. The general approach to eliminating smoke is to develop leaner primary zones with an equivalence ratio between 0.9 and 1.5. Another supplementary way to eliminate smoke is to supply relatively small quantities of air to those exact, local, over-rich zones. Unburnt hydrocarbons and CO Unburnt hydrocarbons and CO are only produced in incomplete combustion typical of idle conditions. It appears probable that idling efficiency can be improved by detailed design to provide better atomization in the case of liquid fuels and higher local temperatures. Oxides of nitrogen The main oxide of nitrogen produced in combustion is NO, with the remaining 10% as NO2. These products are of great concern because of their role in creating harmful particulate matter, ground-level ozone and acid rain in the atmosphere, especially at full load conditions. The formation mechanism of NO can be explained as follows: 1. Fixation of atmospheric oxygen and nitrogen at high-flame temperature. 2. Attack of carbon or hydrocarbon radicals of fuel on nitrogen molecules, resulting in NO formation. 3. Oxidation of the chemically bound nitrogen in fuel.

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In 1977 the Environmental Protection Agency (EPA) in the US published plans to limit the emissions of new, modified and reconstructed gas turbines to • •

75 vppm NOx at 15% oxygen (dry basis) 150 vppm SOx at 15% oxygen (dry basis), controlled by limiting fuel sulfur content to less than 0.8% wt.

These standards applied to simple and regenerative cycle gas turbines, and to the gas turbine portion of combined cycle steam/electric generating systems. The 15% oxygen level was specified to prevent the NOx ppm level being achieved by dilution of the exhaust with air. In 1977 it was recognized that there were a number of ways to control oxides of nitrogen: 1. Use of a rich primary zone in which little NO formed, followed by rapid dilution in the secondary zone. 2. Use of a very lean primary zone to minimize peak flame temperature by dilution. 3. Use of water or steam admitted with the fuel for cooling the small zone downstream from the fuel nozzle. 4. Use of inert exhaust gas recirculated into the reaction zone. 5. Catalytic exhaust cleanup. ‘Wet’ control became the preferred method in the 1980s and most of the 1990s since ‘dry’ controls and catalytic cleanup were both at very early stages of development. Catalytic converters were introduced in the 1980s and are still being widely used; however, the cost of rejuvenating the catalyst is very high. There has been a gradual tightening of the NOx limits over the years from 75 ppm down to 25 ppm, and the new gas turbine goals are as low as 2 ppm. Advances in combustion technology now make it possible to control the levels of NOx production at source, removing the need for ‘wet’ controls. This of course opened up the market for the gas turbine to operate in areas with limited supplies of suitable quality water, for example, deserts or marine platforms. Although water injection is still used, ‘dry’ control combustion technology has become the preferred method for the major players in the industrial power generation market. DLN (dry low NOx) was the first acronym to be coined, but with the requirement to control NOx without increasing CO and unburned hydrocarbons, this has now become DLE. The majority of the NOx produced in the combustion chamber is called ‘thermal NOx’. It is produced by a series of chemical reactions between the nitrogen (N2) and the oxygen (O2) in the air that occur at the elevated

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temperatures and pressures in gas turbine combustors. The reaction rates are highly temperature dependent, and the NOx production rate becomes significant above flame temperatures of about 3300°F (1815°C). Figure 2.19 shows schematically flame temperatures and therefore NOx production zones inside a conventional combustor. This design deliberately burned all of the fuel in a series of zones going from fuel-rich to fuel-lean to provide good stability and combustion efficiency over the entire power range. The great dependence of NOx formation on temperature reveals the direct effect of water or steam injection on NOx reduction. Recent research showed an 85% reduction of NOx with steam or water injection using optimized combustor aerodynamics. NOx prevention Emissions from turbines are a function of temperature and thus a function of the fuel-to-air (F/A) ratio. Figure 2.20 shows that as the temperature is increased, the amount of NOx emissions increases and the CO and unburnt hydrocarbons decrease. The principal mechanism for NOx formation is the oxidation of nitrogen in air when exposed to high temperatures in the combustion process; the amount of NOx is thus dependent on the temperature of the combustion gases and, to a lesser amount, on the time the nitrogen is exposed to these high temperatures. Flame tube

10%

8% 82% 10% 18%

Primary zone

Dilution zone

28%

72%

2500 K 4040°F

1697°F

Nox production zone

1200 K

2.19 Flame temperature in various zones in a combustor.

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Emissions

NOx

Temperature Lean

Fuel/air ratio

Rich

2.20 Effect on emissions with fuel air ratio (temperature) increase.

The challenge in these designs is to lower the NOx without degradation of unit stability. In the combustion of fuels that do not contain nitrogen compounds, NOx compounds (primarily NO) are formed by two main mechanisms: thermal and prompt. In the thermal mechanism, NO is formed by the oxidation of molecular nitrogen through the following reactions: O + N2 ↔ NO + N

[2.32]

N + O2 ↔ NO + O

[2.33]

N + OH ↔ NO + H

[2.34]

As can be seen from the above equations, NOx is primarily formed through high-temperature reaction between nitrogen (N) and oxygen (O2) from the air. Hydrocarbon radicals, predominantly through the reaction, initiate the prompt mechanism CH + N2 → HCN + N

[2.35]

The HCN and N are converted rapidly to NO by reaction with oxygen and hydrogen atoms in the flame. The prompt mechanism predominates at low temperatures under fuel-rich conditions, whereas the thermal mechanism becomes important at temperatures above 2732°F (1500°C). Due to the onset of the thermal mechanism, the formation of NOx in the combustion of fuel/air mixtures increases

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Combined cycle systems for near-zero emission power generation 90 80 70 NOX ppm

60 50 40 30 20 10 0 1200

1300

1400

1500 1600 1700 Flame temperature

1800

1900

2000

2.21 Correlation between adiabatic flame temperature and NOx emission.

rapidly with temperature above 2732°F (1500°C) and also increases with residence time in the combustor. The important parameters in the reduction of NOx are the temperature of the flame, the nitrogen and oxygen content and the residence time of the gases in the combustor. Figure 2.21 is a correlation between the adiabatic flame temperature and the emission of NOx. Reduction of any and all these parameters will reduce the amount of NOx emitted from the turbine.

2.3.4 Dry low emission combustors The DLE approach is to burn most (at least 75%) of the fuel at cool, fuellean conditions to avoid any significant production of NOx. The principal features of this combustion system are the premixing of the fuel and air before the mixture enters the combustion chamber, and the lean mixture required to lower the flame temperature and reduce NOx emission. This action brings the full load operating point down on the flame temperature curve and closer to the lean limit, as shown in Fig. 2.22. Controlling CO emissions can thus be difficult, and rapid engine off-loads bring the problem of avoiding flame extinction, which if it occurs cannot be safely re-established without bringing the engine to rest and going through the restart procedure. Figure 2.23 shows a schematic comparison of a typical DLE NOx combustor and conventional combustors. In both cases a swirler is used to create the required flow conditions in the combustion chamber to stabilize the flame. The DLE fuel injector is much larger because it contains the fuel/air

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NOX emissions

Flame temperature

Conventional combustor

Lean limit Lean pre premixed ed Ultr tra-l a-lean Ultra-lean premixed pre ed Cata Catalyt ytic Catalytic Lean

Rich

Fuel/air ratio

2.22 Effect of fuel/air ratio on flame temperature and NOx emissions.

Dry low emission combustor Premix zone LP stage 2 LP stage 1 Pilot

Lean, cool low NOX Rich stable

Lean, cool low NOX

Main fuel Swirlers Conventional combustor

Main fuel

2.23 A schematic comparison of a typical DLE NOx combustor and a conventional diffusion combustor.

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premixing chamber and the quantity of air being mixed is large, approximately 50–60% of the combustion air flow. The DLE injector has two fuel circuits. The main fuel, approximately 97% of the total, is injected into the air stream immediately downstream of the swirler at the inlet to the premixing chamber. The pilot fuel is injected directly into the combustion chamber with little, if any, premixing. With the flame temperature being much closer to the lean limit than in a conventional combustion system, some action has to be taken when the engine load is reduced to prevent flame out. If no action were taken, flame out would occur since the mixture strength would become too lean to burn. A small proportion of the fuel is always burned richer to provide a stable ‘piloting’ zone, whereas the remainder is burned lean. The LP fuel injector is much larger because it contains the fuel/air premixing chamber and the quantity of air being mixed is large, approximately 50–60% of the combustion air flow. Figure 2.24 shows a schematic of an actual DLE annular NOx combustor in an aero-engine. Note the three concentric rings of swirlers and fuel nozzles. Figure 2.25 shows a ring of fuel nozzles in an annular DLE combustor of a large frame type gas turbine. Can type combustors are arranged annularly on a gas turbine. There are 10–16 of these cans in an annular arrangement on a single gas turbine. Each can-annular combustor has a set of between 3 and 8 fuel nozzles, plus a pilot nozzle in the center, for each individual can on a single gas turbine. Figure 2.26 is a photograph of a typical set of DLN fuel nozzles in a GE frame gas turbine. Figure 2.27 shows a combustor can which houses five fuel

Combustion liner Heat shield 1 Premixer

2

Three rings of fuel nozzles

3

First-stage turbine nozzles

2.24 Aero-engine DLE combustor. Note the three concentric rings of fuel nozzles.

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2.25 A ring of fuel injector in a DLN annular combustor (Siemens V94.3 Gas Turbine).

2.26 A typical set of 5 fuel nozzles and center pilot nozzle for a single DLE combustor can (GE DLN Combustors).

nozzles and a central pilot nozzle. Figure 2.28 shows a set of 8 fuel nozzles, plus a pilot nozzle, for a single can of MHI/Westinghouse Frame type gas turbine. Figure 2.29 shows a combustor can which houses the 8 fuel nozzles and a center pilot nozzle. In a DLE combustor, the flame temperature is much closer to the lean limit than in a conventional combustion system. This requires action to prevent flame out to be initiated as otherwise the mixture strength would become too lean to burn. One method is to close the compressor inlet guide vanes progressively as the load is lowered. This reduces the engine airflow and hence the change in mixture strength that occurs in the combustion chamber. This method, on a

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2.27 A typical GE combustor can which houses the 5 fuel nozzles and a center pilot nozzle.

2.28 A typical set of 8 fuel nozzles and center pilot nozzle for a single DLE combustor can (Mitsubishi/Westinghouse DLN Combustors).

single-shaft engine, generally provides sufficient control to allow low emission operation to be maintained down to 50% engine load. Another method is to deliberately dump air overboard prior to or directly from the combustion section of the engine. This reduces the airflow and also increases the fuel flow required (for any given load) and hence the combustion fuel/air ratio can be held approximately constant at the full load value. This latter method causes the part-load thermal efficiency of the engine to fall off by as much as 20%. Even with these air management systems, the combustion stability range can be insufficient, particularly when the load is rapidly reduced. If the combustor does not feature variable geometry, then it is necessary to turn on the fuel in stages as the engine power is increased. The expected

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Rich

2.29 A typical MHI/Westinghouse liner which houses the 8 fuel nozzles and the center pilot nozzle.

Pilot

Lean

Main fuel Stage 1

Stages 1+2

Power

2.30 The staging of DLE combustor as the turbine is brought to full power.

operating range of the engine will determine the number of stages, but typically at least two or three stages are used, as shown in Fig. 2.30. Some units have very complex staging as the units are started or operated at off-design conditions. Gas turbines often experience problems with these DLE combustors, some of the common problems experienced are • •

Auto-ignition and flashback Combustion instability.

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These problems can result in sudden loss of power because a fault is sensed by the engine control system and the engine is shut down. Auto-ignition is the spontaneous self-ignition of a combustible mixture. For a given fuel mixture at a particular temperature and pressure, there is a finite time before self-ignition will occur. Diesel engines rely on selfignition, but spark-ignition engines must avoid it. DLE combustors have premix modules on the head of the combustor to mix the fuel uniformly with air. To avoid auto-ignition, the length of time the fuel is in the premix tube must be less than the auto-ignition delay time of the fuel. If auto-ignition does occur in the premix module then it is probable that the resulting damage will require repair and/or replacement of parts before the engine is run again at full load. Some operators are experiencing engine shutdowns because of auto-ignition problems. The response of the engine suppliers to rectify the situation has not been encouraging, but the operators feel that the reduced reliability cannot be accepted as the ‘norm’. If auto-ignitions occur, then the design does not have sufficient safety margin between the auto-ignition delay time for the fuel and the residence time of the fuel in the premix duct. Auto-ignition delay times for fuels can be found in the literature, but a search will reveal that there is considerable variation for a given fuel. Reasons for auto-ignition could be classified as follows: • • • •

Long fuel auto-ignition delay time assumed Variations in fuel composition reducing auto-ignition delay time Fuel residence time incorrectly calculated Auto-ignition triggered ‘early’ by ingestion of combustible particles.

Flashback into a premix duct occurs when the local flame speed is faster than the velocity of the fuel/air mixture leaving the duct. Flashback usually happens during unexpected engine transients, for example, compressor surge. The resultant change of air velocity would almost certainly result in flashback. Unfortunately, as soon as the flame front approaches the exit of the premix duct, the flame-front pressure drop will cause a reduction in the velocity of the mixture through the duct. This amplifies the effect of the original disturbance, thus prolonging the occurrence of the flashback. Advanced cooling techniques could be offered to provide some degree of protection during a flashback event caused by engine surge. Flame detection systems coupled with fast-acting fuel control valves could also be designed to minimize the impact of a flashback. The new combustors also have steam cooling provided. High-pressure burners for gas turbines use premixing to enable combustion of lean mixtures. The stoichiometric mixture of air and fuel varies between

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1.4 and 3.0 for gas turbines. The flames become unstable when the mixture exceeds a factor of 3.0, and below 1.4 the flame is too hot and NOx emissions will rise rapidly. The new combustors are, therefore, shortened to reduce the time the gases are in the combustor. The number of nozzles is increased – in most cases increases by a factor of 5–10 – to give better atomization and better mixing of the gases in the combustor, leading to a more complex control system. The trend now is to an evolution towards the can-annular burners. For example, a frame type turbine had one combustion chamber with one burner and a new similar turbine has 12 can-annular combustors and 72 burners. Combustion instability used to be a problem with conventional combustors only at very low engine powers. The phenomenon was called ‘rumble’. It was associated with the fuel-lean zones of a combustor, where the conditions for burning are less attractive. The complex 3D flow structure that exists in a combustor will always have some zones that are susceptible to oscillatory burning. In a conventional combustor, the heat release from these ‘oscillating’ zones was only a significant percentage of the total combustor heat release at low power conditions. With DLE combustors, the aim is to burn most of the fuel very lean to avoid the high combustion temperature zones that produce NOx. So these lean zones that are prone to oscillatory burning are now present from idle to 100% power. Resonance can occur (usually) within the combustor. The pressure amplitude at any given resonant frequency can rapidly build up and cause failure of the combustor. The modes of oscillation can be axial, radial or circumferential, or all three at the same time. The use of a dynamic pressure transducer in the combustor section, especially in the low NOx combustors ensures that each combustor can is burning evenly. This is achieved by controlling the flow in each combustor can till the spectrums obtained from each combustor can match. This technique has been found to be very effective and ensures combustor stability. The calculation of the fuel residence time in the combustor or the premixing tube is not easy. The mixing of the fuel and the air to produce a uniform fuel/ air ratio at the exit of the mixing tube is often achieved by the interaction of flows. These flows are composed of swirl, shear layers and vortex. CFD modeling of the mixing tube aerodynamics is required to ensure the success of the mixing process and to establish that there is a sufficient safety margin for autoignition. By limiting the flame temperature to a maximum of 2650°F (1454°C), singledigit NOx emissions can be achieved. To operate at a maximum flame temperature of 2650°F (1454°C), which is up to 250°F (139°C) lower than the LP system previously described, requires premixing 60–70% of the air flow with the fuel prior to admittance into the combustion chamber. With such a high amount of the available combustion air flow required for flame temperature control, insufficient air remains to be allocated solely for cooling the chamber wall or diluting

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the hot gases down to the turbine inlet temperature. Consequently, some of the air available has to do double duty, being used for both cooling and dilution. In engines using high turbine inlet temperatures, 2400–2600°F (1316–1427°C), even though dilution is hardly necessary there is not enough air left over to cool the chamber walls. In this case the air used in the combustion process itself has to do double duty and be used to cool the chamber walls before entering the injectors for premixing with the fuel. This double-duty requirement means that film or effusion cooling cannot be used for the major portion of the chamber walls. Steam cooling is being investigated in some units. Walls are also coated with thermal barrier coating (TBC), which has low thermal conductivity and hence insulates the metal. This is a ceramic material that is plasma-sprayed onto the combustion chamber during manufacture. The temperature drop across the TBC, of typically 300°F (149°C), means the combustion gases are in contact with a surface that is operating at approximately 2000°F (1094°C), which also helps to prevent the quenching of the CO oxidation.

2.4

Gas turbine expander

There are two types of turbines used in gas turbines: the axial-flow type and the radial-inflow type. The axial-flow turbine is used in more than 95% of all applications. Radial turbines are used in smaller turbines. This chapter concentrates on the axial-flow turbine.

2.4.1 Axial-flow turbines The axial-flow turbine, like its counterpart the axial-flow compressor, is the most common type of turbine used in large gas turbines. In an axial-flow turbine, the flow enters and leaves in the axial direction. There are two types of axial-flow turbine blade profile design: the impulse type and the reaction type. In the impulse type turbine blade profile the entire enthalpy drop occurs in the nozzle (reaction = 0); therefore, it has a very high velocity entering the rotor. In the reaction turbine blade profile the enthalpy drop is divided between the nozzle and the rotor. Many blades have a variable reaction from the hub, where the reaction is about zero, to the tip, where the reaction could be more than 50%. Figure 2.31 is a schematic of an axialflow turbine, also depicting the distribution of the pressure, temperature and absolute velocity. Most axial-flow turbines consist of more than 1 stage; in a typical industrial turbine, there are more than 15 stages. The front stages are usually impulse (zero reaction) and the later stages have 50% reaction or higher. The impulse stages produce about twice the output of a comparable 50% reaction stage, although their efficiency is less.

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Combustor

Nozzle Blades

Ho,Po,To Ps,Ts Vabs NB N B N B

N

B

2.31 Schematic of an axial-flow turbine, with pressure, temperature and velocity distribution throughout the turbine section.

Degree of reaction (turbine) The relative proportions of energy transfers obtained by a change of static and dynamic pressure are used to classify turbomachinery. The parameter used to describe this relationship is called the degree of reaction. Reaction is the energy transfer by means of a change in static pressure in a rotor to the total energy transfer in the rotor. (h2 R= ( H01

h3 ) = H04 ) ⎡(V12 ⎣

⎡(U 12 U 22 ) + (W12 W22 )⎤ ⎣ ⎦ V22 ) + (U 12 U 22 ) (W12 − W22 )⎤⎦

[2.36]

where H01 = total enthalpy entering the nozzle H04 = total enthalpy leaving the blade h2 = static enthalpy at the leading edge of the blade h3 = static enthalpy at the trailing edge of the blade U1 = blade velocity at the leading edge of the blade U2 = blade velocity at the trailing edge of the blade V1 = absolute velocity leaving the nozzle entering the blade (parallel to the exit angle of the nozzle) V2 = absolute velocity leaving the blade (parallel to the following nozzle entrance angle)

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Combined cycle systems for near-zero emission power generation W1 = relative velocity entering the blade (parallel to the blade entrance angle) W2 = relative velocity leaving the blade (parallel to the blade exit angle).

In a turbine not all energy supplied can be converted into useful work even with an ideal fluid. There must be some kinetic energy at the exit velocity. The velocity exiting from a turbine is considered unrecoverable. The total head (Had) divided by the total head plus the absolute exit velocity (V4) head is known as the utilization factor (∈). Thus the utilization factor is defined as the ratio of ideal work to the energy supplied. ∈=

Had

Had +

((

)V42 )

[2.37]

The effect of the utilization factor with speed is shown in Fig. 2.32. The figure also shows the difference between an impulse and a 50% reaction turbine. An impulse turbine is a zero-reaction turbine. In addition to the degree of reaction and the utilization factor, another parameter used to determine the blade loading is the work factor. The work factor is defined as the total head developed divided by the rotor head. Γ=

U 1Vθ1 − U 2Vθ 2 U 22

[2.38]

The two conditions that vary the most in a turbine in the power generation mode are the inlet airflow and temperature, since speed is a constant. Two diagrams are needed to show their characteristics. Figure 2.33 is a performance map, which shows the effect of turbine inlet temperature and airflow on the performance of the unit. Figure 2.34 is also a typical performance map of a turbine showing the effect of heat rate and exhaust gas temperature on power. Velocity diagrams An examination of various velocity diagrams for different degrees of reaction is shown in Fig. 2.35. These types of blade arrangements with varying degrees of reaction are all possible; however, they are not all practical. Examining the utilization factor, the discharge velocity head (V42/2g) represents the kinetic energy loss or the unused energy part. For maximum utilization, the exit velocity should be at a minimum and, as the velocity diagrams demonstrate, this minimum is achieved when the exit velocity is axial. This type of velocity diagram is considered to have zero exit swirl. Figure 2.36 shows the various velocity diagrams as a function of the work

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Advanced industrial gas turbines for power generation

Utilization factor (⑀)

1.0

⑀ max

⑀ max

85

Nozzle angle α = 20°

.8 .6 .4

R=0

R = 0.5

U =.47 V1

U =.94 V1

.2 0 0

.2

.4

.6

.8

1.0 1.2 U/V

1.4

1.6

1.8

2.0

2.32 Variation of utilization factor with the speed to absolute velocity ratio (U/V ). Inlet temperature (%) 100% 75%

100%

45% 100%

Shaft power (%)

80% 90%

60%

80% Inlet pressure (%) 70%

60% 80% 100% Air flow (%)

2.33 Turbine performance map showing the effect of flow, inlet pressure and inlet temperature on the power output.

factor and the turbine type. It can be seen from the diagram that zero exit swirl can exist for any type of turbine. Zero exit swirl diagram In many cases the tangential angle of the exit velocity (Vθ4) represents a loss of efficiency; a blade designed for zero exit swirl (Vθ4 = 0) minimizes the exit loss. If the work parameter is less than two, this type of diagram produces the highest static efficiency. Also, the total efficiency is approximately the same as in the other types of diagrams. If the work factor (Г) is greater than 2.0, stage reaction is usually negative, a condition best avoided.

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Combined cycle systems for near-zero emission power generation 120

Maximum continuous speed

100 Power – percent of site rated

Rated firing temperature

Turbine exhaust temperature Compressor surge line

80 Constant heat rate

60

40

20

0 75

80

85 90 95 100 Shaft speed – percent of rated

105

2.34 Performance map showing the variation of heat rate and power as a function of turbine speed. (Courtesy API Standard 616.)

V1

V1

W2 V2

W1

α

V2

W1

U

U

W2

V1

W2

α U

V2

W1

U

U Wθ

Wθ V2 < W1 R V1 R>1

(d)

(e)

2.35 Turbine velocity triangles showing the effect of various degrees of reaction.

Impulse rotor diagram For the impulse rotor, the reaction is zero, so the relative velocity of the gas is constant, or W3 = W4. If the work factor is less than 2.0, the exit swirl is positive, which reduces the stage work. For this reason, an impulse diagram

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Diagram type

Stage work factor Zero exit swirl

Symmetrical

V3

W4 1

Impulse

W2 V4

U3

U4 2

4

2.36 Effect of diagram type and stage work factor on velocity diagram shape.

should be used only if the work factor is 2.0 or greater. This type of diagram is a good choice for the last stage, because for Г greater than 2.0, an impulse rotor has the highest static efficiency. Symmetrical rotor diagram The symmetrical type diagram is constructed so that the entrance and exit diagrams have the same shape V3 = W4 and V4 = W3. This equality means that the reaction is R = 50%. If the work factor Г equals 1.0, then the exit swirl is zero. As the work factor increases, the exit swirl increases. Since the reaction of 50% leads to a high total efficiency, this design is useful if the exit swirl is not counted as a loss as in the initial and intermediate stages.

2.4.2 Impulse turbine The impulse turbine is the simplest type of turbine. It consists of a row of nozzles followed by a row of blades. The gas is expanded in the nozzle, converting the high thermal energy into kinetic energy. This conversion can be represented by the following relationship: V

[2.39]

h

The high-velocity gas impinges on the blade where a large portion of the kinetic energy of the moving gas stream is converted into turbine shaft work. Figure 2.37 shows a diagram of a single-stage impulse turbine. The static pressure decreases in the nozzle with a corresponding increase in the

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Combined cycle systems for near-zero emission power generation Moving blades

Nozzle

Po,To Vabs Ps,Ts

2.37 Schematic of an impulse turbine showing the variation of the thermodynamic and fluid mechanic properties.

absolute velocity. The absolute velocity is then reduced in the rotor, but the static pressure and the relative velocity remain constant. To get the maximum energy transfer, the blades must rotate at about half the velocity of the gas jet velocity. By definition, the impulse turbine has a degree of reaction equal to zero. This degree of reaction means that the entire enthalpy drop is taken in the nozzle, and the exit velocity from the nozzle is very high. Since there is no change in enthalpy in the rotor, the relative velocity entering the rotor equals the relative velocity exiting from the rotor blade. For the maximum utilization factor, the absolute exit velocity must be axial. The power developed by the flow in an impulse turbine is given by the Euler equation: (ma + m f )(U 1Vθ1 U 2Vθ 2 )

(ma + m f )U (Vθ1 Vθ 2 )

[2.40]

The relative velocity W remains unchanged in a pure impulse turbine, except for frictional and turbulence effect. This loss varies from about 20% for very high-velocity turbines (3000 ft/s) to about 8% for low velocity turbines (500 ft/s). Since for maximum utilization, as seen in Fig. 2.38, the blade speed ratio is equal to U/V = (cos α)/2 = 0.47 where α is the turbine nozzle vane angle (α ~ 20°), the energy transferred in an impulse turbine can be written as: P

(ma + m f )U ( U 2

)

2(ma + m f )U 2

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[2.41]

Advanced industrial gas turbines for power generation

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V1 W 1.0

V1 ≈ 2U1 U1 ≈ U2 V2 U2

V2

W

Utilization factor E

U1

E = max

.6 .4

– U = .47 V1

.2 0 U1

α = 20°

.8

0

.2

.4

.6

.8

1.0 1.2

2.38 Effect of velocity and air angle on utilization factor.

2.4.3 Reaction turbine The axial-flow reaction turbine is the most widely used turbine. In a reaction turbine both the nozzles and blades act as expanding nozzles. Therefore, the static pressure decreases in both the fixed and moving blades. The fixed blades act as nozzles and direct the flow to the moving blades at a velocity slightly higher than the moving blade velocity. In the reaction turbine the velocities are usually much lower, and the entering blade relative velocities are nearly axial. Figure 2.39 shows a schematic view of a reaction turbine. In most designs, the reaction of the turbine varies from hub to shroud. The hub is close to an impulse turbine. The impulse turbine is a reaction turbine with a reaction of zero (R = 0). The shroud is usually above a reaction of 50%. The utilization factor for a fixed nozzle angle will increase as the reaction approaches 100%. For R = 1, the utilization factor does not reach unity but reaches some maximum finite value. The 100% reaction turbine is not practical because of the high rotor speed necessary for a good utilization factor. For reaction less than zero, the rotor has a diffusing action. Diffusing action in the rotor is undesirable, since it leads to flow losses. The 50% reaction turbine has been used widely and has special significance. The velocity diagram of a 50% reaction is symmetrical, and, for the maximum utilization factor, the exit velocity (V2) must be axial. Figure 2.40 shows a velocity diagram of a 50% reaction turbine and the effect on the

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or

Combined cycle systems for near-zero emission power generation

Blade

Blade

Nozzle

Nozzle

Co

mb

ust

Nozzle

Moving blades

Nozzle

Moving blades

Exhaust

Wheel Po Total pressure

Ps Static pressure

Labyrinth seals

Vo Absolute velocity

Shaft

2.39 Schematic of a reaction type turbine showing the distribution of the thermodynamic and fluid mechanic properties.

V1 α U1

0.94 V1≈U1 W1 U1 U1≈U2 V2 U2 W2≈V2

W1

V2

W2

U2 1.0 Utilization factor (⑀)

90

E = max

0.8

Nozzle angle = α20°

0.6 0.4 0.2 0

U V1 = 0.94

0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8 2.0 U/V 1

2.40 The effect of exit velocity and air angle on the utilization factor for a 50% reaction axial turbine stage.

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utilization factor. From the diagram W1 = V2, the angles of both the stationary and rotation blades are identical. Therefore, for maximum utilization, U = cos α V1

[2.42]

the 50% reaction turbine has the highest efficiency of all of the various types of turbines. Equation [2.42] shows that the effect on efficiency is relatively small for a wide range of blade speed ratios (0.6–1.3). The power developed by the flow in a reaction turbine is also given by the general Euler equation. This equation can be modified for maximum utilization for a 50% reaction turbine with an axial exit and the Euler equation reduces to P

(ma + m f )U (U

)

(ma + m f )U 2

[2.43]

The work produced in an impulse turbine with a single stage running at the same blade speed is twice that of a reaction turbine. Hence, the cost of a reaction turbine for the same amount of work is much higher, since it requires more stages. It is common practice to design multistage turbines with impulse stages in the first few stages to maximize the work and to follow it with 50% reaction turbines. The reaction turbine has a higher efficiency due to blade suction effects. This type of combination leads to an excellent compromise: an all-impulse turbine would have a very low efficiency and an all-reaction turbine would have an excessive number of stages.

2.4.4 Turbine blade cooling concepts The turbine inlet temperatures of gas turbines have increased considerably in recent years and will continue to do so. This trend has been made possible by advances in materials and technology, and the use of advanced turbine blade cooling techniques. Figure 2.41 shows the history of the various cooling schemes and development of metals throughout the years, increase in firing temperatures having been a major goal of the industry. The cooling air is bled from the compressor and is directed to the stator, the rotor and other parts of the turbine rotor and casing to provide adequate cooling. The effect of the coolant on the aerodynamics depends on the type of cooling involved, the temperature of the coolant compared with the mainstream temperature, the location and direction of coolant injection, and the amount of coolant. In high-temperature gas turbines, cooling systems need to be designed for turbine blades, vanes, endwalls, shroud and other components to meet metal temperature limits. The

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Combined cycle systems for near-zero emission power generation (1538ºC) 2800

Firing temperature ºF (ºC)

2600 (1316ºC) 2400 (1204ºC) 2200

Steam cooling

2000 Advanced air cooling

(982ºC) 1800 1600 (760ºC) 1400 1200

Conventional air cooling Firing temperature

IN 733

U 500 RENE 77 Blade metal temperature

(538ºC) 1000 1950

1960

1970

GTD111

GTD 111 SC

GTD 111 GTD 111 SC DS

1980

1990

2000

2010

Year

2.41 Various materials and cooling schemes used in high-temperature turbine blade cooling.

Hot stream

Film cooling

Hot stream

Steam/ water

Convection cooling

Steam water cooling

Hot stream Hot stream Hot stream

Full coverage film cooling

Impingement cooling

Transpiration cooling

2.42 Various cooling schemes as used in turbine blade cooling.

concepts underlying the following six basic air-cooling schemes are shown in Fig. 2.42: 1. 2. 3. 4. 5. 6.

Film cooling Convection cooling Water/steam cooling Full coverage film cooling Impingement cooling Transpiration cooling.

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Until the late 1960s, convection cooling was the primary means of cooling gas turbine blades; some film cooling was occasionally employed in critical regions. However, in the year 2000 steam cooling was introduced in production frame type engines used in combined cycle applications. The new turbines have very high-pressure ratios and this leads to compressor air leaving at very high temperatures, which affects their cooling capacity. Film cooling This type of cooling is achieved by allowing the working air to form an insulating layer between the hot gas stream and the walls of the blade. This film of cooling air protects an airfoil in the same way as combustor liners are protected from hot gases at very high temperatures. Convection cooling This form of cooling is achieved by designing the cooling air to flow inside the turbine blade or vane and removing heat through the walls. Usually, the airflow is radial, making multiple passes through a serpentine passage from the hub to the blade tip. Convection cooling is the most widely used cooling concept in present-day gas turbines. Steam cooling Steam is passed through a number of tubes embedded in the nozzle vanes. This method keeps blade metal temperatures below 1000°F. The steam is usually taken in a combined cycle plant from the exit of the HP stage of the steam turbine and then sent to the gas turbine transition piece and first-stage turbine nozzles where it is used as a cooling medium, and in that process is reheated and then sent back to the IP stage of the steam turbine as reheated steam. Full coverage film cooling It is a high-intensity form of film cooling which is achieved by allowing the working air to form an insulating layer between the hot gas stream and the walls of the blade from multiple sources in the blade, thus giving full blade coverage by the use of film cooling. Impingement cooling In this high-intensity form of convection cooling, the cooling air is blasted onto the inner surface of the airfoil by high-velocity air jets, allowing an increased amount of heat to be transferred to the cooling air from the metal surface. This cooling method can be restricted to desired sections of the airfoil to maintain even temperatures over the entire surface. For instance, if

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the leading edge of a blade needs to be cooled more than the mid-chord section or trailing edge, the gas is impinged only onto this part. Transpiration cooling Cooling by this method requires the coolant flow to pass through the porous wall of the blade material. The heat transfer is directly between the coolant and the hot gas. Transpiration cooling is effective at very high temperatures, since it covers the entire blade with coolant flow. Multiple small hole cooling design Primary cooling is achieved by film cooling, with the cooling air injected through small holes over the airfoil surface, leading to temperature distribution over the entire surface (see Fig. 2.43). The temperature distribution as seen in the figure shows the uncooled values and the cooled values in brackets over the surface of the blades. The highest temperatures are seen at the leading and trailing edge.

1498 (1088)

1531 (1106)

1599 (1144)

1560 (1122)

1522 (1101) 1570 (1128)

1555 (1119)

1576 (1131) 1500 (1089)

1488 1575 (1082) (1130) 1571 (1128) 1570 (1128) 1624 (1158)

2.43 Multiple small hole transpiration cooled blades.

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Cooled-turbine aerodynamics The injection of coolant air in the turbine rotor or stator causes a slight decrease in turbine efficiency; however, the higher turbine inlet temperature usually makes up for the loss of the turbine component efficiency, giving an overall increase in cycle efficiency. Total pressure surveys have been made downstream of the stators in both the radial and circumferential directions to determine the effect of coolant on stator losses. The wake traces for the stator with discrete holes and the stator with trailing edge slots show that there is a considerable difference in total pressure loss patterns as a function of the type of cooling and the amount of cooling air supplied. As the coolant flow for the porous blades increases, the disturbance to the flow pattern and the wake thickness increases. Consequently, the losses increase. In a blade with trailing edge slots, the loss initially starts to increase with coolant flow as the wake thickens. However, as the coolant flow is increased, it tends to energize the wake and reduce losses. For a higher coolant flow, the coolant pressures must be higher, causing the flow to be energized. It is obvious from a comparison of the various cooling techniques shown in Fig. 2.44 that a blade with trailing edge slots is thermodynamically the most efficient. The porous stator blades decrease the stage efficiency considerably. This efficiency indicates losses in the turbine but does not take into account cooling effectiveness. The porous blades are more effective for cooling.

2.4.5 Materials The development of new materials as well as cooling schemes has seen a rapid growth in the turbine firing temperature, leading to high turbine efficiencies. The stage 1 blade must withstand the most severe combination of temperature, stress and environment; it is generally the limiting component in the machine. Since 1950, turbine bucket material temperature capability has increased by a total of approximately 400°F (205°C), and, this has allowed the firing temperature of the gas turbine to increase by 1100ºF (593ºC) or 20ºF/10ºC per year as shown in Fig. 2.41. The importance of this increase can be appreciated by noting that an increase of 100°F (56°C) in turbine firing temperature can provide a corresponding increase of 8–13% in output and 2–4% improvement in simple cycle efficiency. Advances in alloys and processing, although expensive and time-consuming, provide significant incentives through increased power density and improved efficiency. The composition of the new and conventional alloys used in turbine blades and nozzles are shown in Table 2.1. The increases in blade alloy temperature capability accounted for the majority of the firing temperature increase until air cooling was introduced, which decoupled firing temperature from

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Combined cycle systems for near-zero emission power generation 100 Solid blades Turbine eff., %

Trailing edge slots 90 Leading edge holes Transpiration cooling

80

70

0

2

4 6 Coolant flow, %

Porous discrete holes

8

2.44 The effect of various types of cooling on turbine efficiency.

the blade metal temperature. Also, as the metal temperatures approached the 1600°F (870°C) range, hot corrosion of blades became more life-limiting than strength until the introduction of protective coatings. During the 1980s, emphasis turned towards two major areas: improved materials technology, to achieve greater blade alloy capability without sacrificing alloy corrosion resistance, and advanced highly sophisticated air-cooling technology, to achieve the firing temperature capability required for the new generation of gas turbines. The use of steam cooling to further increase combined cycle efficiencies in combustors was introduced in the mid- to late 1990s. Steam cooling in blades and nozzles was introduced in commercial operation in the year 2002. In the 1980s, IN-738 blades were widely used. IN-738 was the acknowledged corrosion standard for the industry. New alloys, such as GTD-111, were developed and patented by GE in the mid-1970s. It possesses about a 35ºF (20ºC) improvement in rupture strength as compared to IN-738. GTD111 is also superior to IN-738 in low-cycle fatigue (LCF) strength. The design of this alloy was unique in that it utilized phase stability and other predictive techniques to balance the levels of critical elements (Cr, Mo, Co, Al, W and Ta), thereby maintaining the hot corrosion resistance of IN-738 at higher strength levels without compromising phase stability. Most nozzle and blade castings are made by using the conventional equiaxed investment casting process. In this process, the molten metal is poured into a ceramic mold in a vacuum, to prevent the highly reactive elements in the superalloys from reacting with the oxygen and nitrogen in the air. With proper control of metal and mold thermal conditions, the molten metal solidifies from the surface to the center of the mold, creating an equiaxed structure. Table 2.1 shows the composition of high-temperature alloys.

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© Woodhead Publishing Limited, 2012

18.5 15 16 14

25 25 28 21 22.5

23 22 20 22

19 16 1 15 12

12 12 15.5

Turbine blades U500 RENE 77 (U700) IN738 GTD111

Turbine nozzles X40 X45 FSX414 N155 GTD-222

Combustors SS309 HAST X N-263 HA-188

Turbine wheels ALLOY 718 ALLOY 706 Cr-Mo-V A286 M152

Compressor blades AISI 403 AISI 403 + Cb GTD-450

Source: GE Power Systems.

Cr

Component

– – 6.3

BAL BAL 0.5 25 2.5

13 BAL BAL 22

10 10 10 20 BAL

BAL BAL BAL BAL

Ni

Table 2.1 High-temperature alloys

– – –

– – – – –

– 1.5 20 BAL

BAL BAL BAL 20 19

18.5 17 8.3 9.5

Co

BAL BAL BAL

18.5 37.0 BAL BAL BAL

BAL 1.9 0.4 1.5

1 1 1 BAL –

– – 0.2 –

Fe

– – –

– – – – –

– 0.7 – 14.0

8 8 7 2.5 2.0

– – 2.6 3.8

W

–

–

– – 0.8

3.0 – 1.25 1.2 1.7

9 6

3 2.3

– – –

4 5.3 1.75 1.5

Mo

– – –

0.9 1.8 – 2 –

– – 2.1 –

1.2

– – – –

3 3.35 3.4 4.9

Ti

– – –

0.5 – – 0.3 –

– – 0.4 –

– – – – 0.8

3 4.25 3.4 3.0

Al

– 0.2 –

5.1 2.9 – – –

– – – –

– – – – –

– – 0.9 –

Cd – – – –

– – –

– – 0.25 0.25 0.3

– – – –

– – – – 0.10

V

0.11 0.15 0.03

0.03 0.03 0.30 0.08 0.12

0.10 0.07 0.06 0.05

0.50 0.25 0.25 0.20 0.008

0.07 0.07 0.10 0.10

C

– – –

– – – 0.006 –

– 0.005 – 0.01

0.01 0.01 0.01 – 1.00

0.006 0.02 0.001 0.01

B

– – –

– – – – –

– – – –

– – – – –

– – 1.75 2.8

Ta

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Combined cycle systems for near-zero emission power generation

Directional solidification (DS) is also being employed to produce advanced technology nozzles and blades. First used in aircraft engines more than 25 years ago, it was adapted for use in large airfoils in the early 1990s. By exercising careful control over temperature gradients, a planar solidification front is developed in the blade, and the part is solidified by moving this planar front longitudinally through the entire length of the part. The result is a blade with an oriented grain structure that runs parallel to the major axis of the part and contains no transverse grain boundaries, as in ordinary blades. The elimination of these transverse grain boundaries confers additional creep and rupture strength on the alloy, and the orientation of the grain structure provides a favorable modulus of elasticity in the longitudinal direction to enhance fatigue life. The use of directionally solidified blades results in a substantial increase in the creep life, or substantial increase in tolerable stress for a fixed life. This advantage is due to the elimination of transverse grain boundaries from the bucket, the traditional weak link in the microstructure. In addition to improved creep life, the directionally solidified blades possess more than 10 times the strain control or thermal fatigue compared to equiaxed blades. The impact strength of the DS blades is also superior to that of equiaxed, showing an advantage of more than 33%. In the late 1990s, single-crystal blades were introduced in gas turbines. These blades offer additional creep and fatigue benefits through the elimination of grain boundaries. In single-crystal material, all grain boundaries are eliminated from the material structure and a single crystal with controlled orientation is produced in an airfoil shape. By eliminating all grain boundaries and the associated grain boundary strengthening additives, a substantial increase in the melting point of the alloy can be achieved, thus providing a corresponding increase in high-temperature strength. The transverse creep and fatigue strength is increased, compared to equiaxed or DS structures. The advantage of single-crystal alloys compared to equiaxed and DS alloys in LCF life is increased by about 10%. Blade life comparison is provided in the form of the stress required for rupture as a function of a parameter that relates time and temperature (the Larson–Miller parameter). The Larson–Miller parameter is a function of blade metal temperature and the time the blade is exposed to those temperatures. Figure 2.45 shows the comparison of some of the alloys used in blade and nozzle application. The Larson–Miller parameter is one of several important design parameters that must be satisfied to ensure proper performance of the alloy in a gas turbine blade application to ensure long service life. Other material properties such as creep life, high-cycle fatigue, LCF, thermal fatigue, tensile strength and ductility, impact strength, hot corrosion and oxidation resistance to hot corrosion must also be considered.

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Advanced industrial gas turbines for power generation

40

3.0

kg/cm2 × 10–3

Stress 4.0

KSI

60

99

FSX-414 IN-738

GTD-222

2.0

U-500 GTD-111

N-155 Blades Nozzles

2.0

10 Temp. 100 000 hrs Life

500

600 1000

700 °C 1200

800 1400

1600

°F 40

42

44

46

48

50

Larson–Miller parameter PLM = T(20+log t )×10–3

2.45 Larson–Miller parameter for various blade materials.

2.4.6 Coatings Blade coatings are required to protect the blade from corrosion, oxidation and mechanical property degradation. As superalloys have become more complex, it has been increasingly difficult to obtain both the higher strength levels that are required and a satisfactory level of corrosion and oxidation resistance without the use of coatings. Thus, the trend towards higher firing temperatures increases the need for coatings. The function of all coatings is to provide a surface reservoir of elements that will form very protective and adherent oxide layers, thus protecting the underlying base material from oxidation and corrosion attack and degradation. The main requirements of a coating are to protect blades against oxidation and/or corrosion problems. Other benefits of coatings include thermal fatigue from cyclic operation, surface smoothness and erosion in compressor coatings, and heat flux loading when one is considering thermal barriers. A secondary consideration, but perhaps rather more relevant to thermal barriers, is their ability to tolerate damage from light impacts without spalling to an unacceptable extent because of the resulting rise in the local metal temperatures. Coatings extend life of the blades by providing protection from hostile operational conditions. The coating can be attacked but the base metal is protected and can be coated over again. This greatly reduces the cost of overhauling blades as the base metal is not usually damaged nor its strength compromised.

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Life of coatings depends on composition, thickness and the standard of evenness to which it has been deposited. Coatings help extend the life of bladings by protecting them against oxidation, corrosion, thermal fatigue, temperature and foreign object damage. Oxidation is a prime consideration in a ‘clean fuel’ regime, whereas corrosion is due to higher metal temperatures and is emphasized in ‘dirtier’ fuels. For a given combination of loadings, coating life is governed by: 1. Composition of the coating which includes environmental mechanical properties such as thermal fatigue. 2. Coating thickness which provides a greater protective reservoir if thicker; however, thicker coatings may have lower thermal fatigue resistance. 3. Standard of deposition such as thickness uniformity, or defined thickness variation and coating defects. There are three basic types of coating: TBCs, diffusion coatings and plasma-sprayed coatings. The advances in coating have also been essential in ensuring that the blade base metal is protected at these high temperatures. Coatings ensure that the life of the blades is extended and in many cases they are used as a sacrificial layer which can be stripped and recoated. The general types of coatings used today differ little from the coatings used 10–15 years ago. These include various types of diffusion coatings, such as aluminide coatings originally developed nearly 40 years ago. The thickness required is between 25 and 75 µm; these coatings consist of a combination of Ni/Co and contain about 30% Al. The new aluminide coatings with Pt increase the oxidation resistance and also the corrosion resistance. Coatings developed some 30–35 years ago, commonly known as MCrAlY, have a wide range of composition tailored to the type of performance required and are Ni/Co-based as shown in these three common types of coatings: 1. Ni, 18% Cr, 12% Al, 0.3% Y 2. Co, 29% Cr, 3% Al, 0.3% Y 3. Co, 25% Ni, 20% Cr, 8% Al, 0.3% Y. These coatings are usually 75–500 µm thick and sometimes have other minor elements added to improve environmental resistance, such as Pt, Hf, Ta and Zr. If carefully chosen, these coatings can give very good performance. The TBCs have an insulation layer of 100–300 µm, and are based on ZrO2-Y2O3; they can reduce metal temperatures by 90–270°F (50–150°C). This type of coating is used in combustion cans, transition pieces, nozzle guide vanes and also blade platforms, as shown in Fig. 2.46. The interesting point to note is that some of the major manufacturers are switching away from corrosion protection-based coatings to coatings which

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Cross-section of cooled airfoil

TBC TBC blade

Conventional blade Cooling gas

Cooling gas

ΔT Temp. Hot gas

Hot gas Effect of TBC Base metal Bond layer Ceramic layer

2.46 Airfoil cross-section with TBC.

are oxidation resistant not only at high gas temperatures but also at higher metal temperatures. TBCs are being used on the first few stages in all the advanced technology units. The use of internal coatings is becoming popular due to the high temperature of the compressor discharge, which results in oxidation of the internal surfaces. Most of these coatings are aluminide type coatings. The choice is restricted, due to access problems, to slurry-based, or gas phase/chemical vapor deposition. Care must be taken in production; otherwise, internal passages may be blocked. The use of pyrometer technology on some of the advanced turbines has located blades with internal passages blocked causing those blades to operate at metal temperatures of 50–100°F (28–56°C) higher than the neighboring blades.

2.5

Sources of further information

Boyce, M. P., Cogeneration and Combined Cycle Power Plants. Second Edition, April 2010; First Edition, January 2002. New York: ASME Press. Boyce, M. P., ‘Maintaining Combustors – New Designs Reduce Emission but Require Additional Care’, Turbomachinery International, Pg 34–5, Nov/Dec 2006. Boyce, M. P., ‘Advanced Axial Flow Turbines’, March/April 2007 Turbomachinery International, Pg 40–41, July/August 2006. Boyce, M. P., ‘Recovering Performance Compressor Operation and Maintenance is Crucial to Regaining Losses in Gas Turbine Efficiency and Output’, Turbomachinery International, Pg 6–7, July/August 2006.

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Boyce, M. P., Gas Turbine Engineering Handbook. Third Edition, Elsevier Publisher, February 2006; Second Edition, Butterworth-Heinemann Publisher, December 2001; First Edition 1985. Boyce, M. P., ‘They Don’t Make ’Em Like That Anymore: Reliability, Availability, Maintainability of Advanced Gas Turbines’, Turbomachinery International, September/ October 2004, 45 (5) Pg. 16. Boyce, M. P., ‘Reliability, Availability, and Maintainability Program for Advanced Gas Turbines in Combined Cycle Applications’, US DOE. Next Generation Turbine Stakeholder Meeting, 9–10 April 2001. Boyce, M. P. and Francisco Gonzales, ‘A Study of On-Line and Off-Line Turbine Washing to Optimize the Operation of a Gas Turbine’, ASME – GT-2005–69126, Reno, Nevada. June 2005. Also published in the Journal of Turbomachinery June 2006 GTP-05–1136. Boyce, M. P., Jeffery N. Phillips, Jay Grandmont and Leonard Angelo, ‘Simplified On-Line Performance Monitoring of Simple Cycle Gas Turbines using Spreadsheet based Calculation’, ASME – GT-2004–54123, Vienna, Austria, June 2004. Davis, L. B., ‘DLN Combustion Systems for HDGTs’, GER3568G OCT 2000. The Gas Turbine Handbook. US Department of Energy Office of Fossil Energy National Energy Technology Laboratory, Axial Flow Compressors Section 2.0, 2006. DOENETL-2006/1230.

© Woodhead Publishing Limited, 2012

3 Natural gas-fired combined cycle (NGCC) systems A. D. RAO, University of California, USA

Abstract: A combined cycle combines two power cycles in series to obtain a high overall thermal efficiency, significantly higher than the individual efficiencies. This chapter discusses utilization of a Brayton cycle or gas turbine for the topping cycle and a steam Rankine cycle for the bottoming cycle. The size of a combined cycle may range from less than 10 MW to in excess of 500 MW while using a single gas turbine. In addition to having thermal efficiencies exceeding 60% natural gas lower heating value (LHV) basis with a ‘J class’ gas turbine with a nominal 1600°C turbine inlet temperature (Ito et al., 2010), outstanding environmental performance, easy start-up and shutdown and low cooling water requirements, combined cycles require significantly lower staffing, capital cost and construction time than boiler-based power plants. However, the clean fuels required, such as natural gas, synthesis gas or distillate, are significantly more expensive than fuels such as coal and biomass that can be directly combusted in a boiler. This chapter focuses on control technologies for criteria pollutants and CO2 emissions. Future trends for improvements in performance and emissions are also discussed. Key words: combined cycle, efficiency, gas turbine, HRSG, steam turbine, topping cycle, bottoming cycle, syngas, reforming, partial oxidation, CO2 separation, CO2 capture, carbon capture, emissions control, precombustion control, post-combustion control, SCR.

3.1

Introduction

A combined cycle consists of combining two power cycles in series to obtain a high overall thermal efficiency, significantly higher than the individual efficiencies of the two cycles making up the combined cycle. In the combined cycle discussed in this chapter, a Brayton cycle or gas turbine is utilized for the topping cycle and a steam Rankine cycle for the bottoming cycle. Combined cycles come in a variety of sizes depending on the size and number of gas turbines utilized. Combined cycle sizes may range from less than 10 MW to in excess of 500 MW when using a single General Electric or Siemens Energy ‘H class’ 50-cycle gas turbine. A combined cycle with a thermal efficiency as high as 61% natural gas lower heating value (LHV) basis is being commercialized by Mitsubishi Heavy Industries using a ‘J class’ gas turbine which has a nominal turbine inlet temperature of 1600°C. High 103 © Woodhead Publishing Limited, 2012

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thermal efficiency not only reduces operating cost (primarily by reducing the fuel cost, and capital cost (by increasing the specific power output of the equipment), but also the CO2 emissions per unit of net power generated by the plant for a given percentage CO2 capture. Increased firing temperatures, however, require materials capable of withstanding the higher temperatures within the combustor and the turbine of a gas turbine. Furthermore, control of NOx emissions, a criteria pollutant, becomes more challenging.

3.2

Technology, system design and equipment

There are various options being considered for control of CO2 emissions from a natural gas-fired combined cycle. The various approaches may be broadly classified as: • •

Pre-combustion capture Post-combustion capture.

Before some of the approaches under each of these categories are described, a basic combined cycle plant, that is, without any CO2 emissions control, is described.

3.2.1 Combined cycle without CO2 capture Figure 3.1 depicts a state-of-the-art combined cycle plant utilizing an H class steam-cooled gas turbine (Smith, D., 2004) integrated with a triple pressure reheat steam cycle with wet cooling towers and without CO2 capture. Most gas turbines are air-cooled with the cooling air being provided by the gas turbine compressor; however, many of the advanced gas turbines with their significantly higher turbine inlet temperatures for higher efficiency are utilizing steam cooling, the cooling steam being provided by the bottoming Rankine cycle. The gas turbine air compressor draws in ambient air via a filter to remove airborne particulates, especially those that are larger than 10 µm. Fuel and compressed air are mixed and combusted. Hot gas turbine exhaust flows through a heat recovery steam generator (HRSG). Demineralized makeup boiler feed water (BFW) is sprayed directly into the surface condenser which condenses steam leaving low-pressure (LP) section of a steam turbine at a vacuum. This negative operating pressure of the condenser is set by the temperature of the cooling medium used in the surface condenser. The combined stream of cold vacuum condensate and makeup BFW is drawn from the surface condenser by the vacuum condensate pump and is heated in an economizer within the HRSG and then supplied to an integral deaerator that also generates LP steam (at about 460 kPa). The deaerator removes

© Woodhead Publishing Limited, 2012

© Woodhead Publishing Limited, 2012

Filter

Gas turbine cooling steam supply

Gas turbine

TC

BFW TC

Superheater

Intermittent blowdown drum

Heat recovery steam generator

TC

PC

M

BFW

Vent

TC

Superheater LC

LC

HP steam drum

Condenser

Cooling water supply / return

LC

Continuous blowdown drum

Continuous Blowdown

Boiler

LC

BFW from natural gas heater

IP steam drum

FC

Vent

LC

Demineralizer

Deaerator

BFW to natural gas heater

To desuperheating & attemperation

Superheater

Boiler

FC

To stack

Cold vacuum condensate

Blowdown

Cems

Treated makeup water

3.1 Combined-cycle with steam cooled gas turbine and triple pressure reheat steam cycle along with major steam cycle controls.

To disposal

Quench water

Auxiliary

Air

Gas turbine cooling steam return

BFW return / supply

Gas turbine cooling steam return

Superheater

Natural gas

BFW

Reheater

M

Reheater

Gas turbine cooling steam supply

PC

Economizer

TC

Superheater

PC

Economizer

M

Boiler

Vacuum system

Economizer

Steam turbine

Economizer FC

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Combined cycle systems for near-zero emission power generation

dissolved gases such as O2 and CO2 in the feed water which can cause corrosion. Chemicals are also injected into the water to scavenge the small amounts of remaining O2. Excess steam generated in the deaerator after superheating is fed to the LP section of the steam turbine. Superheating in addition to increasing cycle efficiency also avoids condensation of steam into droplets. Intermediate-pressure (IP) BFW is extracted from the main BFW pump and flows through the IP economizer in the HRSG. Saturated IP steam generated (at about 2850 kPa) in the HRSG is combined with steam leaving the high-pressure (HP) section of the steam turbine before it is reheated (to about 570°C, depending on the gas turbine exhaust temperature) and fed back to the IP section of the steam turbine. Saturated HP steam generated (at about 17 400 kPa) in the HRSG is superheated (to about 570°C, again depending on the gas turbine exhaust temperature) and fed to the HP section of the steam turbine. The optimum steam cycle conditions depend on the exhaust temperature of the gas turbine corresponding to the ambient temperature chosen, while its exhaust flow rate dictates the type of steam cycle to be used. The BFW pump supplies water to the attemperators for temperature control of the superheated and reheated steam. In an attemperator, the steam comes into direct contact with water when it is cooled through the evaporation of the water. Cooling steam required by the gas turbine is provided from the HP steam turbine exhaust. Steam returning from this closedcircuit cooling of the gas turbine is also combined with the IP steam before it is reheated within the HRSG in parallel with the superheater coils. Steam drums of the HRSG are continuously purged to control the build-up of dissolved solids. The continuous blowdown is cascaded from HP steam drum to IP steam drum and blowdown from the IP steam drum is routed to a drum where LP steam is recovered. Water discharging from this drum is fed to a second lower pressure drum and flash steam produced is vented to atmosphere.

3.2.2 Combined cycle with pre-combustion CO2 capture In pre-combustion capture, as depicted in Fig. 3.2, natural gas is converted to a syngas which is primarily a mixture of CO and H2 by reaction with steam, and O2, depending on the type of process. In order to avoid carbon formation, deposition on the catalyst leading to catalyst deactivation, and metal dusting, the amount of steam used is far in excess of that required for stoichiometric conversion. Metal dusting is a catastrophic corrosion phenomenon which leads to disintegration of structural metals and alloys into dust consisting of fine particles of the metal, alloy and carbon. The tendency for metal dusting increases with the system operating pressure and can thus take away some of the efficiency gain that may be realized by operating

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107

Steam

Gas turbine Syngas generation Syngas

Combustor Booster compressor

Air

Extraction air

HRSG

Reheated decarbonized syngas

Stack gas

Hightemperature heat recovery

Shift conversion and heat recovery Cooled shifted syngas

STEAM TURBINE Steam turbine

CO2 removal

To condenser Reheat Decarbonized syngas

CO2 to sequestration

CO2 compression

3.2 Combined-cycle with pre-combustion CO2 capture.

the syngas generation step at increased pressures. The next step is catalytic shifting of the CO to CO2 by the reaction CO + H2O → H2 + CO2, which is followed by heat recovery, syngas cooling and separation of CO2 from the syngas for sequestration, utilizing an absorber column and a stripper column with a suitable solvent circulating between the two columns. About 85–90% of the CO2 present in the syngas is absorbed into the solution in the absorber although higher levels of CO2 removal are achievable as in ammonia (NH3) synthesis plants, where a high-purity H2 stream is produced from the syngas to react with N2 to form NH3. Solvent loaded with CO2 exiting the absorber is regenerated in the stripper using steam, while a high-purity CO2 stream is released. With some solvents, it may be possible to release the CO2 from solution by simply depressurizing the CO2-loaded solvent. Solvent circulation rate, size of the columns and energy penalty all increase as the level of CO2 removal is increased. The pressure at which the CO2 is released depends on the type of solvent used. Remaining gas (decarbonized syngas) leaving the absorber which is now mostly H2 is combusted in gas turbines of the combined cycle with reduced CO2 emissions to the atmosphere (Rao et al., 1999). The steam bottoming cycle of the combined cycle is integrated

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with the steam system of the decarbonized syngas generation section of the plant. An advantage of this pre-combustion CO2 capture scheme as compared to post-combustion capture is that the CO2 present in the syngas is available at a high partial pressure, thereby lowering the energy penalty of separating and pressurizing the captured CO2 stream. On the other hand, since the shift reaction is exothermic, a significant amount of the chemically bound energy of the syngas is degraded to heat and it enters directly into the bottoming steam cycle of the combined cycle while bypassing the gas turbine, thereby not fully taking advantage of the higher conversion efficiency of a combined cycle. Furthermore, there are inefficiencies associated with the syngas generation process itself. Syngas generation Various options are available for the conversion of natural gas to syngas, such as steam-methane reforming, partial oxidation or a combination of the two. In steam-methane reforming, saturated hydrocarbons present in natural gas are reacted with steam in the presence of a nickel-based catalyst at 500– 900°C to produce syngas. The major overall reactions are the following: CH4 + H2O → CO + 3H2 (reforming reaction) and CO + H2O → H2 + CO2 (water gas shift reaction) with the overall heat of reaction being endothermic. The natural gas has to be purified prior to reforming to remove catalyst poisons typically consisting of compounds of sulphur (odorants added to natural gas to render it detectable by smell are organic compounds of sulphur called mercaptans) and chlorine in some cases. Desulphurization may be accomplished near ambient temperatures by passing the natural gas through a bed containing activated carbon or at high temperatures (around 400°C) through a bed containing zinc oxide (ZnO) to form zinc sulphide (ZnS) or alternating beds of ZnO and a COS hydrolysis catalyst sandwiched between them to convert COS if present to H2S by the reaction: COS + H2O → H2S + CO2. ZnO containing a COS hydrolysis catalyst such as alumina (Al2O3) is also available. Chlorine compounds if present in the feed gas may be removed by an additional chemical absorbent installed upstream of the desulphurization unit. Chlorine compounds are mainly a problem in natural gas spiked with landfill gas or certain refinery off-gases as long as the steam itself does not introduce any chlorides. Unsaturated hydrocarbons if present in the natural gas may be catalytically hydrogenated by reaction with recycled H2 in the presence of a cobalt or nickel molybdenum catalyst to avoid or minimize carbon deposition in the reforming catalyst. The operating temperature is typically limited to 400°C. This hydrogenation process also converts the COS and mercaptans to H2S which may then be removed by the ZnO bed. Pre-reforming may be used upstream of a reformer for converting any higher hydrocarbons present in natural gas to result in stable and mild operating

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conditions for the downstream reformer, ensuring reliable operation of the reformer. Higher hydrocarbons are converted into a mixture of CO, CO2, H2 and CH4 approaching equilibrium based on the reforming and water gas shift reactions. It also allows operation at lower steam to carbon ratios, thereby reducing the overall energy consumption and increasing the lifetime of the downstream reformer and the shift catalysts, as any sulphur remaining in the hydrocarbon feed is absorbed by the pre-reforming catalyst. In partial oxidation, CH4 and other hydrocarbons in natural gas are reacted with a limited amount of O2 that is not enough to completely oxidize the hydrocarbons to CO2. The major overall reactions are the following: CH4 + ½O2 → CO + 2H2, CH4 + H2O → CO + 3H2 and CO + H2O → H2 + CO2 with the overall heat of reaction being exothermic. Steam is added to limit soot formation while O2 is provided by an air separation unit. In autothermal reforming, again the major overall reactions are CH4 + ½O2 → CO + 2H2, CH4 + H2O → CO + 3H2 and CO + H2O → H2 + CO2 but the process is carried out in the presence of a catalyst. The operating temperature of the autothermal reforming process can be lower (reactor exit temperature of about 900–950°C versus about 1150°C for non-catalytic partial oxidation) since the reactions are carried out in the presence of a catalyst, making it more thermally efficient than the non-catalytic partial oxidation process, since in the former a lesser fraction of the chemically bound energy of the natural gas is degraded into heat. Furthermore, the amount of O2 required is also lower, thereby reducing the efficiency and cost penalty of O2 supply. Air instead of O2 may be utilized to eliminate the need for an air separation unit but the thermal efficiency of the reforming process itself is reduced (although to a lesser degree than in the non-catalytic partial oxidation process which operates at a higher temperature). With today’s cryogenic air separation technology, however, the overall trade-off is typically in favour of using air instead of O2. The amount of steam required for the autothermal reforming process is greater than that required by the non-catalytic partial oxidation process and tends to limit its thermal efficiency advantage, and the reactor required for carrying out the process is more complex. A previous study (Rao et al., 1999) has shown that the air-blown autothermal reforming scheme is more efficient than a configuration utilizing partial oxidation in pre-combustion CO2 capture plants, with the heat rate for the partial oxidation option being about 8% higher while utilizing cryogenic air separation for producing O2. Compared with a plant without CO2 capture, where natural gas is directly fired in gas turbines, both plant efficiency and cost are significantly compromised with pre-combustion capture of CO2: an increase of more than 30% in heat rate, and a more than doubling of the plant cost on a per kW basis may be expected (Rhudy, 2005). A synergistic combination of the autothermal reformer with the steammethane reformer consists of utilizing the heat generated in the exothermic

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autothermal reformer in the endothermic steam-methane reformer to reduce the thermal inefficiency of generating the syngas. Shift reaction The shift reactors suitable for this application are adiabatic reactors such that the reactor exit temperature may be maximized (but constrained by limits imposed by the catalyst) to generate steam at as high a pressure as possible, thereby maximizing the work extracted from the steam, the steam being supplied to the bottoming cycle of the combined cycle. CO2 recovery Options currently available for separating CO2 from the shifted syngas include scrubbing the syngas with physical solvents – such as propylene carbonate (Fluor SolventSM), N-methyl-2-pyrrolidone (Purisol®) and mixtures of the dimethyl ethers of polyethylene glycol (Selexol™) – and chemical solvents – such as aqueous monoethanolamine and diethanolamine solutions with proprietary additives. The suitability of a particular type of solvent depends on the syngas pressure; chemical solvents are favoured at low to moderate CO2 partial pressures typical of steam-methane as well as autothermal reforming, while physical solvents are more suitable for higher pressures. An advantage with some of the physical solvents is that the CO2 may be released at higher pressures, thereby decreasing the parasitic power consumption of the CO2 pressurization step. On the other hand, refrigeration of the physical solvent may be required in order to limit its circulation rate. The CO2 thus separated may be pressurized by intercooled compression either to the final pressure or to an IP where it forms a liquid or a supercritical fluid upon cooling, and then pumping the fluid to the final pressure. A significant saving in power can be realized by this latter approach but for this configuration to be practical, significant variations in the concentration of impurities in the CO2 stream may not be tolerated since the impurities can affect the phase diagram of the CO2 stream. The final pressure which depends on the disposal method is presented in Table 3.1 along with other specifications (Woods et al., 2007). Typically, dehydration of the CO2 stream Table 3.1 CO2 pipeline specification Parameter units Inlet pressure MPa (psia) Outlet pressure MPa (psia) Inlet temperature °C (°F) N2 concentration ppmv O2 concentration ppmv Ar concentration ppmv

Parameter value 15.3(2215) 10.4(1515) 26(79)