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The research presented in this thesis was carried out at the Laboratory of Thermal Engineering of the University of Twente and at the Netherlands Energy Research Foundation ECN. Part of the work was realized within the framework of the New Energy Conversion Technologies (NECT) programme by the Netherlands Agency for Energy and Environment Novem. The financial support of Novem is gratefully acknowledged.
Cover by Anja Astapova, Amsterdam, © 1998 Printed by Febodruk BV, Enschede
New and Advanced Energy Conversion Technologies. Analysis of Cogeneration, Combined and Integrated Cycles / M.A. Korobitsyn.
ISBN 90 365 1107 0 Copyright © 1998 by M.A. Korobitsyn
ter verkrijging van de graad van doctor aan de Universiteit Twente, op gezag van de rector magnificus, Prof.dr. F.A. van Vught, volgens het besluit van het College voor Promoties in het openbaar te verdedigen op vrijdag 3 april 1998 om 13.15 uur
Mikhail Aleksandrovich Korobitsyn geboren op 6 oktober 1963 in Sosnogorsk (Rusland)
Hirs .dr.ir.G.Dit proefschrift is goedgekeurd door de promotor Prof. G.
Ìîèì ðîäèòåëÿì .
Basic cycles 1.3 Kalina cycle 1.5 Otto and Diesel cycle 1.10 Conclusions References Chapter 2.1 Introduction 2.2 Rankine cycle 1.1 Advanced and combined cycles 0.3 Exergy 0.4 Joule-Brayton cycle 1.9 Heat pumps 1.2 Rankine/Rankine 2.1 Carnot cycle 1.Table of contents Summary Samenvatting Êðàòêîå ñîäåðæàíèå Preface Introduction 0.8 Fuel cells 1.4 Scope of the thesis References Chapter 1.2 The New Energy Conversion Technologies programme 0.7 Ericsson cycle 1.3 Rankine/Kalina vii xi xiii xv xvii 1 1 2 4 4 5 7 7 8 9 11 13 14 15 15 17 19 20 23 23 26 26 .6 Stirling cycle 1. Combined cycles 2.
5 Semi-closed gas turbine 3. Advanced cycles 3.17 Fuel cell/Otto(Diesel) 2.13 Otto(Diesel)/Joule 2.2 Water flashing 188.8.131.52 Steam recompression 3.3 Advanced Rankine cycles 3.1 Water and steam injection 3.7 Chemically-recuperated gas turbine 3.4 Humid Air Turbine cycle 3.18 Fuel cell/Stirling 2.4 Advanced Joule cycles 3.5 Joule/Rankine 2.2 Background 3.5.6 Joule/Kalina 2.16 Fuel cell/Joule 2.11 Otto(Diesel)/Kalina 2.4.1 Multi-pressure steam boiler 3.6 Discussion and conclusions References Chapter 4.3.9 CO2 gas turbine 3.8 Joule/Otto(Diesel) 2.12 Otto(Diesel)/Stirling 2.4.4 Steam flow spitting 3.4.viii COGENERATION. COMBINED AND INTEGRATED CYCLES 2.6 Water recovery 3.3 Wet compression 3.5 Other cycles 3.3.10 Otto(Diesel)/Rankine 2.14 Fuel cell/Rankine 2.15 Fuel cell/Kalina 2.2 Supercharging and evaporative cooling 3.3.1 Introduction 3.19 Fuel cell/Fuel cell 2. Industrial cogeneration 4.2 Performance criteria 27 28 31 32 33 34 34 35 35 35 36 36 36 38 38 38 39 39 43 43 44 45 45 45 46 46 47 47 50 51 51 53 53 54 55 57 57 57 58 59 65 65 66 .4.1 Introduction 4.1 Ejector topping power cycle 3.4.9 Joule/Stirling 2.7 Joule/Joule 2.8 Gas turbines with multi-stage combustion 3.4 Rankine/Stirling 126.96.36.199.20 Conclusions References Chapter 3.
Integration of gas turbine and waste incinerator 6. Air bottoming cycle 7. Partial oxidation gas turbine 8.5 Exergy analysis 6.6 Conclusions References Chapter 5. Externally-fired combined cycle 5.4 Performance analysis 8.5 Exergy analysis 8.4 Configurations 7.1 Introduction 6.2 Gas turbine integration 6.4 Discussion 5.6 Conclusions References Chapter 7.1 Introduction 7. Nomenclature 70 73 80 81 81 83 84 87 88 89 91 92 93 95 95 97 98 100 102 104 105 107 107 108 110 111 112 114 116 116 117 119 119 120 123 126 127 130 130 133 .6 Cogeneration 7.3 Thermodynamic considerations 7.1 Introduction 5.7 Applications 7.8 Conclusions References Chapter 8.5 Performance analysis 7.1 Introduction 8.4 Performance analysis 4.5 Exergy analysis 5.TABLE OF CONTENTS ix 4.3 Configurations 4.2 Background 8.2 Background 7.3 Configurations 8.3 Plant configurations 5.6 Conclusions References Appendix 1.5 Exergy analysis 4.2 Background 5.3 Configurations 6.4 Performance analysis 6.6 Conclusions References Chapter 6.
Of the combined cycles considered. which results in a combined-cycle efficiency of up to 45%. defines the structure of this thesis. Combustion with gas turbine exhaust allows exergy losses to be reduced by more than 9%. Since the air turbine exhaust is a hot sterile air. gas turbine integration adds 12–15% points to the basic plant gross efficiency. These achievements. the basic thermodynamic cycles and their specific features are described. however.Summary At the current status of development in power technology large combined-cycle power plants are exceeding 70% of the maximum theoretical (Carnot) efficiency. heat pumps) and operating modes. combined and advanced cycles are addressed. so industrial cogeneration is assessed in various configurations (steam boiler.or low-temperature cycles for bottoming. improvements within a particular cycle are discussed in the chapter on advanced cycles. The scope of the NECT programme covers power and heat production. In comparison with a steam bottoming plant. and medium. being realized by the Netherlands Agency for Energy and Environment (Novem). At the beginning. can be obtained only at a large scale for power plants above 300 MW. The bottoming cycle adds 20–30% to the power output. Enhancement of direct-fired power plants with the use of gas turbine technology is investigated for a waste incineration plant and a solid fuel fired steam boiler. One of the configurations is the Joule/Joule combined cycle. xi . overall efficiency. which is dedicated to the development of new and improved combinations of existing energy conversion technologies. combination of existing technologies – these are the possible ways to improve performance of small. since the heat rejected by the topping cycle is utilized by the bottoming one. Because no single cycle can offer high efficiency due to the intrinsic limitations and the impossibility to operate within a broad temperature range. and the specific boiler surface area. and better performance can be obtained. The part of the programme. are analyzed in detail. In addition to the combined cycles. The basic cycles are combined according to their temperature level: high-temperature cycles are good candidates for the topping application. which consists of an existing gas turbine and an air bottoming turbine. Advances within power cycles. The hot windbox configuration shows the best match between natural gas consumption. each cycle is outlined and its schematic diagram is given. which are selected for further development within the NECT programme. Identification and development of new energy conversion technologies and systems for distributed power generation applications are the objectives of the New Energy Conversion Technologies (NECT) programme. gas turbine. several technologies. On a smaller scale. the implementation of the cycle in the food processing industries is considered.and medium-scale power technology. innovative designs are required to reach the same level of performance at feasible costs. Combined cycles do not suffer from the drawbacks of the single cycles. integration of cycles. Subsequently. In the case of the waste incinerator. the air bottoming cycle requires no steam/water equipment that enables an operation of small units at high efficiency and low costs without the complexity of the steam cycle.
which substitutes the combustion chamber. in view of its good exergetic efficiency. COMBINED AND INTEGRATED CYCLES In the case of the coal-fired steam boiler. Performance of these cycles was found comparable. One project is aimed at the development and market introduction of several combined and advanced cycles. UT. The studies will evaluate the potential future application of the selected cycles and will include optimization. but some specific features of the POGT deserve further considerations. First. an implementation of the cycle for combined production of power and synthesis gas is proposed. experimental work. Second. This externallyfired combined cycle shows an overall efficiency of 47%. outlined in this thesis. A development team of ECN. The work has resulted in two feasibility projects. parametric and comparative analysis. is considered. . and a cost estimate. manufacturers and end users will be established to introduce proposed systems into the application market. the low level of air excess reduces NOX forming. The use of supplementary firing is employed to increase the turbine inlet temperature and to reduce the costly surface of the high-temperature heat exchanger An advanced gas turbine cycle with partial oxidation (POGT).xii COGENERATION. an engineering company. which is another subject of the NECT programme. The other project assesses the technical feasibility of the gas turbine with partial oxidation. a gas turbine cycle with an external heat exchanger. is compared with the reheat gas turbine. which are financially supported by Novem and being carried out by the Energy Research Foundation ECN and the University of Twente.
warmtepompen) en bedrijfscondities beoordeeld. omdat het met een standaard cyclus niet mogelijk is een hoog rendement te halen vanwege intrinsieke beperkingen en de onmogelijkheid om binnen een breed temperatuurbereik te functioneren. Daarom wordt ook industriële warmte/kracht koppeling in verschillende configuraties (stoom. Diesel en Stirling cycli. Verbetering van direct-gestookte ketels met gebruik van gasturbine technologie wordt onderzocht voor een afvalverbrandingsinstallatie (AVI) en voor een kolengestookte stoomketel. omdat de restwarmte van de topping cyclus nog eens in de bottoming cyclus benut wordt. Om te beginnen worden de standaard thermodynamische cycli en hun specifieke eigenschappen beschreven. De luchtturbine levert 20–30% extra vermogen wat resulteert in overall efficiency van maximaal 45%. Elk van de beschouwde cycli wordt in hoofdlijnen beschreven en in een schema weergegeven. Een ander voordeel van een dergelijke cyclus is de afwezigheid van een water-stoom systeem. die geselecteerd zijn voor verdere uitwerking binnen het NECT programma.Samenvatting Met de huidige stand van de energietechnologie overschrijden grote combined-cycle elektriciteits-centrales al 70% van het maximum theoretisch (Carnot) rendement. combinaties van bestaande technologiën. Gecombineerde en geavanceerde cycli worden beschouwd. Deze prestaties echter worden behaald alleen bij grote eenheden van meer dan 300 MW. Binnen het kader van het NECT programma valt niet alleen elektrische energieopwekking. in detail geanalyseerd. Verbeteringen binnen de cyclus zelf.en gasturbines. Een van de configuraties is de Joule/Joule cyclus. Op kleinere schaal zijn er innovatieve cycli nodig om hetzelfde prestatieniveau tegen aanvaardbare kosten te bereiken. Deze cyclus kan ook warme lucht leveren. Vervolgens worden de cycli. midden en lage temperatuur cycli voor bottoming. en ook brandstofcellen. Verbeteringen binnen standaard cycli woorden in een apart hoofdstuk besproken. dat door de Nederlandse Onderneming voor Energie en Milieu (Novem) wordt gerealiseerd. Identificatie en ontwikkeling van nieuwe energie-conversie technologien en systemen voor decentrale energieopwekking zijn de doeleinden van het Nieuwe EnergieConversie Technologiën (NECT) programma. Otto. Deze maakt exploitatie van kleine units met hoog rendement en lage kosten mogelijk zonder de complexiteit van de stoombottoming cyclus. De structuur van dit proefschrift wordt bepaalt door het deel van het programma. dat zich op ontwikkeling van nieuwe en verbeterde combinaties van bestaande energie-conversie technologiën richt. Joule-Brayton. Een gecombineerde cyclus heeft een beter rendement. integratie van thermodynamische cycli – het zijn de mogelijke methoden om efficiency van klein. In het geval van AVI voegt de gasturbine integratie 12–15% xiii . De standaard cycli zijn te combineren aan de hand van hun temperatuurniveau: hoge temperatuur cycli voor topping toepassingen. Een implementatie van deze cyclus bij de voedselindustrie wordt in beschouwing genomen. bestaande uit een standaard gasturbine en een nageschakelde luchtturbine. ook thermische energie produktie komt aan bod. terwijl hetzelfde prestatieniveau behaald wordt als bij een cyclus met stoombottoming.en middenschalige energieopwekking te verbeteren. Dit deel omvat de Rankine.
toegepast. Deze projecten evalueren de potentiële toepassingen van de geselecteerde cycli en bevatten optimalisatie. . Een geavanceerde gasturbine cyclus met partiële oxidatie (POGT). fabrikanten en eindgebruikers zal ontstaan om voorgestelde systemen bij de toepassingsmarkt te introduceren. De hot windbox configuratie toont een goede overeenkomst tussen aardgas verbruik.xiv COGENERATION.en synthese gas productie mogelijk. gezien het goed exergetisch rendement. Verbranding met het uitlaatgas reduceert de exergieverliezen met meer dan 9%. overall rendement en het specifieke verwarmingsoppervlak van de boiler. wordt vergeleken met de reheat gasturbine. die in dit proefschrift worden beschouwd. Echter sommige eigenschappen van de POGT cyclus verdienen meer aandacht: ten eerste. Een development team bestaande uit het ECN. een ingenieursbureau. Een project richt zich op ontwikkeling en marktintroductie van een aantal gecombineerde en geavanceerde cycli. Dit werk heeft geresulteerd in twee haalbaarheidsprojecten die in opdracht van Novem worden uitgevoerd door het Energieonderzoek Centrum Nederland ECN en de Universiteit Twente. UT. Deze externally-fired combined cycle haalde een rendement van 47%. experimentele werk en een kostenbegroting. De prestaties van de POGT en de reheat gasturbine zijn vergelijkbaar bevonden. Het andere project beoordeelt de technische haalbaarheid van de gasturbine met partiële oxidatie. COMBINED AND INTEGRATED CYCLES punten bij de bruto efficiency. parametrische en vergelijkende analyses. In het geval van een kolengestookte ketel wordt een gasturbine met externe warmtewisselaar. Er kan bijgestokt worden om de turbine entreetemperatuur te verhogen en het kostbare warmtewissellingsoppervlak te verkleinen. is een uitvoering van deze cyclus voor gecombineerde elektriciteit. ten tweede. die ook bij het NECT programma hoort. die verbrandingskamer vervangt. een lage luchtovermaat resulteert in NOX reductie.
ï. Ïîñêîëüêó âûõëîï âîçäóøíîé òóðáèíû ïðåäñòàâëÿåò ñîáîé ñòåðèëüíûé.ä.ï.. è Ñòèðëèíãà) è òîïëèâíûå ýëåìåíòû. ÷òî ïîçâîëÿåò äîñòè÷ü ê. è ïîýòîìó ÷åòâåðòàÿ ãëàâà ïîñâÿùåíà ïðîìûøëåííîé òåïëîôèêàöèè â ðàçëè÷íûõ êîíôèãóðàöèÿõ (ñ òåïëîâûìè íàñîñàìè. â ñâÿçè ñ òåì.ï.. ÷òî ñáðàñûâàåìîå îò âåðõíåãî öèêëà òåïëî óòèëèçèðóåòñÿ â íèæíåì öèêëå. Îïðåäåëåíèå è ðàçðàáîòêà òåõíîëîãèé äëÿ äåöåíòðàëèçîâàííîãî ïðîèçâîäñòâà ýëåêòðîýíåðãèè îñóùåñòâëÿåòñÿ â Íèäåðëàíäàõ Àãåíñòâîì ïî Ýíåðãåòèêå è Îêðóæàþùåé Ñðåäå Novem â ðàìêàõ ïðîãðàììû Íîâûå Òåõíîëîãèè Ïðåîáðàçîâàíèÿ Ýíåðãèè (NECT). à òàêæå êîìáèíèðîâàíèå ñóùåñòâóþùèõ òåõíîëîãèé. Äèçåëÿ. Âîçäóøíàÿ òóðáèíà óâåëè÷èâàåò ìîùíîñòü óñòàíîâêè íà 2030%.  ïîñëåäóþùèõ ãëàâàõ áîëåå ïîäðîáíî àíàëèçèðóþòñÿ êîìáèíèðîâàííûå è óñîâåðøåñòâîâàííûå öèêëû. Îäíîé èç ñõåì ÿâëÿåòñÿ êîìáèíèðîâàííûé öèêë òèïà Äæîóëü/Äæîóëü.ï. ïîñêîëüêó îòäåëüíî âçÿòûé öèêë íå ìîæåò îáåñïå÷èòü âûñîêîãî ê. äî 45%. óòèëèçèðóþùåé îòõîäÿùåå òåïëî âåðõíåãî öèêëà. Îòòî. îïðåäåëèëà ñòðóêòóðó íàñòîÿùåé äèññåðòàöèè. Êîìáèíèðîâàííûå öèêëû èìåþò áîëåå âûñîêèé ê. â ñâÿçè ñ ïðèñóùèìè åìó îãðàíè÷åíèÿìè è íåâîçìîæíîñòüþ ðàáîòû â øèðîêîì äèàïàçîíå òåìïåðàòóð. ïîñâÿùåííàÿ ðàçðàáîòêå íîâûõ è óñîâåðøåíñòâîâàííûõ êîìáèíàöèé ñóùåñòâóþùèõ òåõíîëîãèé. Âîçìîæíûìè ïóòÿìè óëó÷øåíèÿ ýíåðãåòè÷åñêèõ ïîêàçàòåëåé ñòàíöèé ñðåäíåé è ìàëîé ìîùíîñòè ÿâëÿþòñÿ ìîäåðíèçàöèÿ è èíòåãðàöèÿ öèêëîâ. çà ñ÷åò êîìáèíèðîâàíèÿ è óñîâåðøåíñòâîâàíèÿ ðàññìàòðèâàåìûõ öèêëîâ. Îäíàêî òàêèå âûñîêèå ïîêàçàòåëè âîçìîæíû òîëüêî äëÿ êðóïíûõ áëîêîâ êîìáèíèðîâàííîãî öèêëà ìîùíîñòüþ ñâûøå 300 ÌÂò. Ïîäîáíûé óðîâåíü ýêîíîìè÷íîñòè íà óñòàíîâêàõ ìåíüøåé ìîùíîñòè âîçìîæåí çà ñ÷åò óñîâåðøåíñòâîâàíèÿ öèêëîâ. Ïðèâîäÿòñÿ âîçìîæíûå êîìáèíàöèè öèêëîâ. îïðåäåëåííûå èñõîäÿ èç èõ ðàáî÷èõ òåìïåðàòóð. äîñòèãàþùèì 70% îò òåîðåòè÷åñêîãî ìàêñèìóìà Êàðíî. ïàðîâûìè è ãàçîâûìè òóðáèíàìè) è ðåæèìàõ. Ðàáî÷èå õàðàêòåðèñòêè ýòîãî öèêëà ñîïîñòàâèìû ñ ïàðîãàçîâûì öèêëîì. êîòîðûé ñîñòîèò èç òèïîâîé ãàçîâîé òóðáèíû è âîçäóøíîé òóðáèíû. ñ óêàçàíèåì èõ õàðàêòåðíûõ îñîáåííîñòåé è ðàáî÷èõ ïàðàìåòðîâ.  îòäåëüíîé ãëàâå îáñóæäàþòñÿ óñîâåðøåíñòâîâàííûå öèêëû.ä.ï. xv . ãîðÿ÷èé âîçäóõ. Âòîðàÿ è òðåòüÿ ãëàâà ðàññìàòðèâàþò âîçìîæíîñòè óëó÷øåíèÿ ê. Äàåòñÿ êðàòêîå îïèñàíèå è ñõåìà ðàññìàòðèâàåìûõ êîìáèíèðîâàííûõ öèêëîâ. Îòñóòñòâèå ïàðîâîäÿíîé ñèñòåìû ïîçâîëÿåò óïðîñòèòü êîíñòðóêöèþ è îñóùåñòâèòü ýêñïëóàòàöèþ óñòàíîâîê ìàëîé ìîùíîñòè ñ âûñîêèì ê.ï. ×àñòü ïðîãðàììû. âûáðàííûå äëÿ ðàçðàáîòêè íà ñëåäóþùåì ýòàïå ïðîãðàììû NECT.  ïåðâîé ãëàâå îïèñûâàþòñÿ îñíîâíûå òåðìîäèíàìè÷åñêèå öèêëû (Ðåíêèíà. ðàññìàòðèâàåòñÿ ðàáîòà öèêëà â òåïëîôèêàöèîííîì ðåæèìå.ä.ä. Ïðîãðàììà NECT ðàññìàòðèâàåò ïðîèçâîäñòâî íå òîëüêî ýëåêòðè÷åñêîé.ä. Äæîóëÿ-Áðýéòîíà. ïðèãîäíûé äëÿ òåõíîëîãè÷åñêèõ íóæä.Êðàòêîå ñîäåðæàíèå Ñîâðåìåííûå òåõíîëîãèè ïîçâîëÿþò ïðåîáðàçîâûâàòü õèìè÷åñêóþ ýíåðãèþ òîïëèâà â ýëåêòðè÷åñêóþ ñ ê.ä. íî è òåïëîâîé ýíåðãèè.
êîòîðûå ïî çàêàçó Novem.ï. èíæåíåðíîé êîìïàíèè è ïðîèçâîäèòåëåé îáîðóäîâàíèÿ. COMBINED AND INTEGRATED CYCLES  øåñòîé è ñåäüìîé ãëàâàõ àíàëèçèðóþòñÿ äâà êîìèíèðîâàííûõ öèêëà òèïà Äæîóëü/Ðåíêèí. Äîïîëíèòåëüíîå ñæèãàíèå ïðèðîäíîãî ãàçà äëÿ ïîäíÿòèÿ âåðõíåé òåìïåðàòóðû ãàçîòóðáèííîãî öèêëà ïîçâîëÿåò óìåíüøèòü ïîâåðõíîñòü äîðîãîñòîÿùåãî âûñîêîòåìïåðàòóðíîãî òåïëîîáìåííèêà.  âîñüìîé ãëàâå öèêë ãàçîâîé òóðáèíû ñ ÷àñòè÷íîé îêñèäàöèåé òîïëèâà. Íàñòîÿùàÿ ðàáîòà âûðàçèëàñü â äâóõ ïðîåêòàõ.ï. âêëþ÷åííûé â ïðîãðàììó NECT. îêîëî 47%.xvi COGENERATION.ï. ïàðàìåòðè÷åñêèé è ñðàâíèòåëüíûé àíàëèç. Ïðîåêòû âêëþ÷àþò â ñåáÿ îïòèìèçàöèþ. áðóòòî îò 25% äî 3741%. óíèâåðñèòåòà Òâåíòå. Èñïîëüçîâàíèå îòõîäÿùèõ ãàçîâ òóðáèíû â êà÷åñòâå ïîäîãðåòîãî âîçäóõà ãîðåíèÿ óìåíüøèëî ïîòåðè ýêñåðãèè â ïðîöåññå ñæèãàíèÿ íà 9%. âîçìîæíà êîìáèíèðîâàííàÿ âûðàáîòêà ýëåêòðè÷åñòâà è ñèíòåç-ãàçà ñ âûñîêèì ýêñåðãåòè÷åñêèì ê. . Ðàñ÷åòû êîìáèíèðîâàííîãî öèêëà. Ïåðâûé ïðîåêò íàïðàâëåí íà ðàçðàáîòêó è âíåäðåíèå íåñêîëüêèõ êîìáèíèðîâàííûõ è óñîâåðøåíñòâîâàííûõ öèêëîâ. îäíàêî íåêîòîðûå îñîáåííîñòè öèêëà ñ ÷àñòè÷íîé îêñèäàöèåé çàñëóæèâàþò äîïîëíèòåëüíîãî âíèìàíèÿ: âî-ïåðâûõ. Ñõåìà ïðåäâêëþ÷åíèÿ ãàçîâîé òóðáèíû ñî ñáðîñîì îòõîäÿùèõ ãàçîâ â òîïêó ìóñîðîñæèãàòåëÿ è âíåøíèì ïàðîïåðåãðåâîì ïîêàçàëà íàèëó÷øåå ñîîòíîøåíèå ìåæäó ïîòðåáëåíèåì ïðèðîäíîãî ãàçà. ñîñòîÿùàÿ èç ECN. ýêïåðèìåíòàëüíóþ ÷àñòü è ïðåäâàðèòåëüíûé òåõíèêî-ýêîíîìè÷åñêèå ðàñ÷åòû. ïîêàçàëè ê. îáùèì ê.ä. êàìåðà ñãîðàíèÿ êîòîðîé çàìåíåíà âûñîêîòåìïåðàòóðíûì òåïëîîáìåííèêîì. âûïîëíÿþòñÿ Íèäåðëàíäñêèì Öåíòðîì ïî Ýíåðãåòè÷åñêèì Èññëåäîâàíèÿì ECN è óíèâåñèòåòîì Òâåíòå. Äëÿ ðåàëèçàöèè ýòèõ ïðîåêòîâ áóäåò ñîçäàíà ðàáî÷àÿ ãðóïïà. è óäåëüíîé ïîâåðõíîñòüþ òåïëîîáìåíà. Ðàññìàòðèâàþòñÿ ñõåìû èíòåãðàöèè ãàçîâîé òóðáèíû ñ ìóñîðîñæèãàòåëüíîé óñòàíîâêîé è ñõåìà ãàçîâîé òóðáèíû.ä. ïðåäëàãàåìûå äëÿ ìîäåðíèçàöèè ïàðîâûõ êîòëîâ ñ ïðÿìûì ñæèãàíèåì òîïëèâà. èñïîëüçóþùåãî ãàçîâóþ òóðáèíó ñ âíåøíèì òåïëîîáìåííèêîì. âî-âòîðûõ.ä. ñðàâíèâàåòñÿ ñ öèêëîì ÃÒÓ ñ ïðîìåæóòî÷íûì íàãðåâîì. Âòîðîé ïðîåêò äîëæåí äàòü òåõíè÷åñêîå îáîñíîâàíèå ãàçîòóðáèííîãî öèêëà ñ ÷àñòè÷íîé îêñèäàöåé. Èíòåãðàöèÿ ìóñîðîñæèãàòåëüíîé óñòàíîâêè ñ ãàçîâîé òóðáèíîé ïðèâåëà ê ðîñòó ê.ï. ðàññìîòðåííûõ âî äàííîé äèññåðòàöèè. â ñâÿçè ñ ìàëîé âåëè÷èíîé êîýôôèöèåíòà èçáûòêà âîçäóõà óìåíüøàåòñÿ âûáðîñ NOX.ä. Îáíàðóæåíû ñõîäíûå ïîêàçàòåëè ýòèõ öèêëîâ. ðàñïîëîæåííûì â òîïêå êîòëà.
Second. Also my thanks go to all staff of Laboratory of Thermal Engineering. Leshiner. I wish to thank Anja for her attention and patience. Pieter Jellema. My appreciation is due to Iris Mayer for her editorial support. Geert-Jan Wiechers. G. xvii .H. I wish to thank my promotor Prof. my teacher who stimulated my interest in thermodynamics and power technology with his excellent knowledge of the subject and a remarkable erudition. A. In spite of my rather short and intermittent stay in Enschede. which is not limited only by exergy analysis. is acknowledged for permitting to publish parts of the report done on their behalf. in the person of Dr. The management of the business unit ECN Fuels.I. Novem. First. I thank my roommate. but also in engineering and industrial communities in this country and abroad. I would like to thank my former colleagues at Stork Engineers and Contractors (then Comprimo) in Amsterdam: Edwin Goudappel and Bas Kerkhof (currently with Tebodin BV) for supporting my exergy studies. and her professional help with graphic and layout matters. Iantovski is thanked for numerous suggestions he made to this thesis. and Patrick Kers for their enthusiasm and devotion. Kipperman. R. Finally.Preface I am grateful to many people in writing this thesis. Much appreciation deserve the graduate students who contributed to this thesis. especially during the last months of writing. He is the driving force behind the penetration of exergy concept not only in academic. I owe a debt to Martijn Wagener. my gratitude is expressed to the late Prof.E. Conversion & Environment is thanked for permitting me an absence leave and the possibility to complete this thesis. Hirs for allowing me to conduct this research at his Energy Technology chair at the University of Twente. I wish to thank my colleagues at the Netherlands Energy Research Foundation ECN for a pleasant and stimulating atmosphere. and Piet Baaijens for his inexhaustible expertise. Prof.M.G. it has been a great pleasure working with you. René Cornelissen. E. for fruitful discussions and his critical approach.
Switzerland (ABB.Introduction 0. the use of advanced/combined schemes was no longer required for efficient plant operation. first in the form of existing steam plant repowering. In order to improve their performance. gas turbines were inefficient. have become a common power plant configuration. such as reheat. advances within the gas turbine cycle have been reconsidered: in 1948 ABB utilized reheat in a gas turbine power plant built in Beznau. . In the beginning.. and later. Nevertheless. and industrial gas turbines could utilize the achievements obtained in jet engine technology. more developed cycles. 0. after the oil crisis in the 1970s the efficiency of power plants became the top priority. the reheat concept has been implemented in the ABB GT24/26 series turbine. compactness. or recuperation. as specially-designed gas-and-steam turbine plants. bulky. 1994).1 Advanced and combined cycles The development of some technologies often follows an S-curve. intercooling. where growth slows down and progress is achieved only by increasing the unit size (Fig. and combined-cycle plants. The open-cycle gas turbines offered low capital costs. costs drop due to the mass production. which was used in the 1940s and 1950s to increase engine efficiency. the technology enters the saturation phase. but low specific power. were applied. and unreliable engines. and efficiency close to that of the steam plants.1). But as soon as the break-through period was achieved in the 1960s. Another approach was combining the turbines with other. At the moment of introduction the technology performs poorly and cannot compete with other types. is now employed in the Rolls Royce/Westinghouse WR21 turbine (Shepard et al. Recuperation. Then comes a break-through period: performance improves very rapidly. 1995). and more investments are attracted for further research and development. such as the Rankine cycle. Forty years later. Also. The plant with multiple shafts had high efficiency for that time. The development of the gas turbine aptly illustrates this principle. After awhile. modifications.
The best simple-cycle gas turbine offers a value of 42%. In some cases.g. developments in direct-fired steam plants with supercritical parameters have resulted in efficiencies approaching 50%. 1993). based on existing energy conversion technologies.2 COGENERATION. (3) new combinations. 0. The objective of Part 3 is as follows: to identify and to develop to market introduction new and advanced energy conversion technologies which can provide higher efficiency and lower emissions than existing. which has been initiated by the Netherlands Agency for Energy and Environment (Novem) in order to identify and to develop the new energy conversion technologies and systems for distributed power generation applications (Novem. The gains and losses of combination/integration should be evaluated as shown in Fig. and (4) cogeneration.2 The New Energy Conversion Technologies programme Identification and development of advanced and combined cycles is one of the topics of the New Energy Conversion Technologies (NECT) programme. a separate generation scheme. Combined together.1. power and heat). 0. and feasible within 5–10 years. To maintain a performance gain. which is perpendicular to the the critical line. Meanwhile. (2) combustion. The programme's scope includes projects in the field of research and development. combination and integration of two cycles may not always be beneficial. these technologies generate electricity with up to 59% efficiency. COMBINED AND INTEGRATED CYCLES Combined advanced cycle Combined cycle Advanced cycle Simple cycle Time Fig. The programme consists of four parts: (1) scouting study. experiments. However. can provide better performance. or the use of two different fuels. Technology development paths. feasibility studies. competing technologies. 0. especially when it involves the production of two different products (e. and demonstration. developments in combined-cycle technology should proceed in the direction of the advance vector. The realization of this part began with an inventory of basic thermodynamic cycles and their combinations. Of the known energy conversion routes (as outlined by Rogers and .2.
emissions.2.3. solar and atomic energy sources were excluded from consideration. The systems were presented in a two-dimensional matrix. the most important of which are efficiency. 0. where they were combined according to their thermodynamic parameters: the high temperature cycles were regarded as the most suitable for topping applications. Energy resources and energy conversion technologies.3. an LHV efficiency. as shown by the dotted line in Fig. The Earth's energy. 0. coal Uranium Geothermal Biomass Wind. in the case of cogeneration of heat and power. and the low temperature cycles for bottoming. . The combinations were assessed on the basis of several criteria. other performance parameters are required. Mayhew.INTRODUCTION 3 Separate plants Combined/integrated plants Gain Loss Parameter Fig. Combined/integrated plants versus separate plants. oil. Several combined and advanced cycles were selected for a detailed performance analysis. While performance of power production is easily assessed by. for example. DIRECT CHEMICAL FOSSIL FUELS WASTE ATOMIC ENERGY EARTH'S ENERGY INDIRECT SOLAR DIRECT SOLAR Gas. a working area was defined. waves Battery Fuel cell HEAT Photovoltaic Steam Gas turbine Stirling MHD Thermionic Thermoelectric Spark-ignition and Gas turbine compression-ignition (topping cycle) turbine (bottoming cycle) engine engines Fig. and feasibility. A system with the highest efficiency was chosen as the reference. 0. A reference system was set for a comparative analysis. 1991).
New methods of exergy analysis have been introduced recently. 1985. Due to the irreversibility of thermal processes. Materials from the following. Combined generation of heat and power is discussed in Chapter 4. . technological chains of processes (Szargut. Exergy method can be applied for the analysis of thermal. unless otherwise stated. Bejan et al. G. developers' evaluations. computer simulations. 1997). Joule-Brayton. 1996). and possible improvements suggested. by analyzing work loss within a system. were used in this thesis. 1987) and exergy load redistribution methods (Sorin and Paris. Otto. and Stirling cycles) and fuel cells. Exergy. also known as availability. The Joule/Rankine combined cycle is addressed in the form of an externally-fired combined cycle (Chapter 5) and in an integrated scheme of a gas turbine and a waste incinerator (Chapter 6). Korobitsyn. The Joule/Joule cycle (a gas turbine with an air bottoming cycle) forms the topic of Chapter 7. 1973. 1995).4 COGENERATION. Hence. such as graphic exergy analysis with the use of the energy-utilization diagram (Ishida. 1996). 1985. Valero et al. on which the width of a flow line is determined by its exergy value. All efficiency values quoted are based on the lower heating value (LHV) of the fuel. Combined with economic analysis. 1995. imperfections can be pinpointed and quantified. 1988).G. In Chapter 2 their possibilities and performance in combination are considered. a number are analyzed in more detail. and metallurgical plants (Brodyanski.. Rankine. or a whole country (Wall.. 1987). 1. chemical. exergy equipartitioning (Tondeur and Kvaalen. the life cycle of a product (Cornelissen. Szargut. 0. Kotas. 1982).4 Scope of the thesis Chapter 1 discusses the basic thermodynamic cycles (Carnot. different sorts of energy can be directly compared in exergetic terms. 1991).. An advanced gas turbine cycle with partial oxidation is discussed in Chapter 8. are used to illustrate the exergy flow across a system. the work obtained is always less than the maximum work. or are estimated. ASME Paper 95-CTP-11. Also. and Hirs. 1986. and to evaluate labour and economic trends (Bandura and Brodyanski. exergy is used in thermoeconomic studies (Tsatsaronis and Winhold. previously published papers. Of the cycles reviewed in Chapters 1–3. is a measure of the maximum useful work that can be obtained when a system is brought to a state of equilibrium with the environment in reversible processes. Grassmann diagrams. "Analysis of Cogeneration Alternatives". Advances within the basic cycles are reviewed in Chapter 3. Diesel. The performance specifications of the analyzed systems are based on manufacturers' claims.3 Exergy An objective criteria for assessment offers exergy efficiency. M.A. COMBINED AND INTEGRATED CYCLES 0.
Petten. The Netherlands. Korobitsyn M. and Jellema P. Switzerland. Bejan. "Energy and Exergy Analysis of a Chemical Process System with Distributed Parameters Based on the Enthalpy-Direction Factor Diagram". 5. 1997. Proceedings of the Power-Gen International '96 Conference. 557–562. Essex. Advanced Cycle System. 6..G. Japan. 4. ABB Power Generation Ltd. Rogers.W. and Mayhew..J. 1985... M. June 2–5. Kers. 1994.A.. Moscow. University of Twente. pp.. vol. 199–210. 51–58. T. "The Use of Supplementary Firing in an ExternallyFired Combined Cycle Power Plant". pp. Stockholm. Korobitsyn. Work and Heat Transfer.A. 690–695. The Use of Exergy Analysis and the Reduction of Irreversibility.A. T.T.. December 4–6. Journal of Engineering for Gas Turbines and Power. 21. 1973. 1–2.. Kotas. "Performance Analysis of the Dual Gas Turbine Combined Cycle". 255–259. Vol.J. 1993. New York. ASME Publications. 1995. New Energy Conversion Technologies.G.A. and Chiprich... 1997. Proceedings of the International Conference on Power Engineering ICOPE '97. and Kawamura. pp. to be presented at the 43rd ASME Gas Turbine and Aeroengine Congress. Y. (in Russian). Shepard. P.L. and Hirs. 5.G.. "Integration of a Gas Turbine and a Municipal Waste Incinerator". Analysis. An In-Depth Review of the NECT Research and Development Programme... pp. 1995. by R. 3. Ishida. Design and Development. Thermodynamics and Sustainable Development.INTRODUCTION 5 2. AES-Vol. Hirs. Korobitsyn. Vol. M. in: Thermodynamics and the Design.C. pp. London.. 7. 8. "Analysis of a Gas Turbine with Partial Oxidation". M.L.. M. Process. J. G. M. and Moran. "Review on New and Improved Combinations of Existing Energy Conversion Technologies". "GT26.B. Also appeared in Power Technology International. John Wiley & Sons. The Netherlands. References 1. 1982. The Netherlands. M. 4. V. July 13–17. Hirs G.R. Engineering Thermodynamics.P. Cornelissen.A. "Design and Development of the WR-21 Intercooled Recuperated (ICR) Marine Gas Turbine". UK. Baden. S.. Utrecht. Bowen. 35. 1997. R. G. 6.M. 1996. Florida. New York. and Korobitsyn. 117.. July. 1998. G. Tokyo. G. PhD Thesis.G. .. the Innovative Answer to Lower the Cost of Electricity". 46–49. Novem (Netherlands Agency for Energy and Environment). and Hirs. A. Final Report ECN-CX-96-060. Brodyanski. Vol. The Exergy Method of Thermal Plant Analysis. and Improvement of Energy Systems (ed. Orlando. Thermal Design and Optimization. Company information.F. Tsatsaronis. Krane)... 1996. K. Exergy Method of Thermodynamic Analysis. G. Longman Scientific & Technical. Wagener. M. Netherlands Energy Research Foundation ECN... 3.A. 1996. Sweden. 1992. 9. Energia. 2. Industrial and Engineering Chemistry.M. pp. 1. Butterworths. ABB.. Enschede. Korobitsyn.
Lozano. 11. by R. 55–73. No. Exergy Analysis of Thermal. Hemisphere Publishing. and Winhold. D. Energy... Krane).ity Criterion for Transfer and Separation Processes". Szargut. Part I – A New General Methodology". Morris. J. M. 14. and Steward. Sciences. 26. Sorin. J. New York.. Vol. Gaggioli). G. 12. . Chemical. 513–521. pp. and Torres. Tsatsaronis. "Exergoeconomic Analysis and Evaluation of Energy Conversion Plants. "Equipartition of Entropy Production. 10. COMBINED AND INTEGRATED CYCLES 10. and Kvaalen. 1986. E. pp. pp. 143–160. ASME Publications. 50–66. and Improvement of Energy Systems (ed. An Optima. "Exergy Losses in the Chains of Technological Processes"... G.. AES-Vol. 13–21. 2–3.. Vol. Valero. 36.R. Vol. 9. "A General Theory of Exergy Savings – Parts I–III. F. 69–94. "Exergy Conversion in the Swedish Society". 16.. M. ASME Publications. in: Thermodynamics and the Design.R. pp. Industrial & Engineering Chemistry Research. 1988. D.. 1. Resources and Energy.A. by R. pp. New York. Tech. Bulletin of the Polish Academy of Sciences. 1995. 13. 35. Wall. pp. J. AES Vol. M.. 1987. and Paris. Tondeur. in: Computer-Aided Engineering of Energy Systems (ed.6 COGENERATION. "Exergy Load Redistribution Method for Multi-Step Process analysis". 15. No. 1988. Analysis.V. 1985. S. Szargut. New York. and Metallurgical Processes.. 7–9. A.
except for the fuel cells. and a low-temperature reservoir (sink). and two isentropic processes.1 Carnot cycle All the cycles reviewed in this chapter are heat engines. A heat engine is an apparatus that produces work operating between a high-temperature reservoir (source). 1. while more advanced developments are covered in Chapter 3. The vapour expands isentropically to state 3. In the heating process. liquid from state 1 is brought to saturated vapour state 2. 1.1a) heat is added at a constant upper temperature and rejected at a constant lower temperature. The cycle can be performed on the working fluid within a two-phase region under the saturation line. At state 4. and their advantages and disadvantages are discussed.1) . The operating parameters. Using the general expression for efficiency. 2–3 and 4–1. A corollary of the Second Law of thermodynamics proves that a heat engine possesses the highest efficiency when operated in the Carnot cycle. It consists of two isothermal processes. isentropic compression begins. In the Carnot ideal cycle (Fig. Ways of improving performance are also mentioned. The overall thermal efficiency of the engine is the proportion of the heat supplied which is converted into mechanical work. the Carnot cycle efficiency can be described as η CARNOT = W Q ADDED = Q ADDED − Q REJECTED T MAX ∆S − T MIN ∆S T MAX − T MIN = = Q ADDED T MAX ∆S T MAX (1.Chapter 1 Basic Cycles Abstract This chapter gives an overview of basic thermodynamic cycles and fuel cells. efficiency and power range are outlined for each cycle. and the fluid is brought back to state 1. 1–2 and 3–4. and condensation with heat rejection occurs along line 3–4.
A further improvement can achieved by adding a reheat stage.8 COGENERATION. Superheat can be introduced into the cycle in order to raise the maximum cycle temperature. The cycle also gains in performance when the boiler feed water is preheated by the steam extracted from the steam turbine (line c–d in Fig.2c). 1. This arrangement covers even larger area on the T–s diagram.1. and. 1. where the working fluid is condensed completely (line 5–1 in Fig. The Carnot and Rankine cycle.2a).032 to 0.2b). efficiency is increased.16 bar. this results in a practical condensing temperature of 25–55 °C at the condensing pressure of 0. Using water as the working fluid. .1b) and brought to the boiler pressure by a feed water pump (line 1–2). The ideal efficiency that a Rankine plant with reheat can reach is about 55%. where steam expands to an intermediate pressure (line 5–a). By this means. The average temperature of the heat addition is then higher than in the conventional Rankine cycle. 1.3–0. the efficiency of a Carnot plant with a boiler pressure of 40 bar will lie in the range of 37–43%.2 Rankine cycle Practical difficulties with the compression of a wet vapour in the Carnot cycle (high volumetric flow. The temperature difference in the condenser can vary from 10–15 K for a water-cooled condenser and 25–40 K for an air-cooled unit. such as the critical temperature and pressure. thus improving the efficiency of the cycle (line 4–5 in Fig. therefore. It is evident that the greater the temperature difference the more efficient the cycle. non-homogeneous medium) lead to the Rankine cycle. The maximum temperature has thermodynamic constraints. and then is passed back to the boiler to reach the initial temperature (line a–b in Fig. 1. The minimum temperature in a real process depends on the heat sink temperature and the heat transfer characteristics of the condenser unit. A natural heat sink possesses an ambient temperature which is assumed to be 15 °C. The work done by the pump on the working fluid is much less than that of the compressor: at low pressures it amounts to 0.5% of the total cycle work and can often be disregarded. COMBINED AND INTEGRATED CYCLES T 2 3 3 4 2 1 4 1 5 s a) b) Fig. The highest steam temperature in the wet region is the critical temperature of 374 °C. depending on the maximum temperature. meaning a higher work output. 1. 1.
Further developments of the cycle are expected by raising the metallurgic limit with the use of high-temperature materials. since it is an inexpensive. Advanced Rankine cycles: (a) with superheat. 1. At the medium-temperature level. based on chlorofluorocarbons (CFC) or other fluids with low critical temperatures. non-toxic and well-known medium. Extremely irreversible heat exchange caused by this large temperature difference indicates an imperfection of this cycle. steam remains the most suitable fluid. The organic Rankine cycle.BASIC CYCLES T 9 5 3 4 2 1 6 2 1 3 5 4 a b 3 d c 2 6 1 5 s 4 a) b) Fig. 1997). 1992). various media other than steam have been proposed for the Rankine cycle. 1994). where a mixed working . The organic cycle can be used for waste heat recovery. c) The temperature of the flue gases in the steam boiler approaches 1800 °C. A mixture of fluids has been proposed to substitute steam in a Rankine cycle. or by choosing other working fluids. The choice of a medium depends on the temperature level.3 Kalina cycle The use of the Rankine cycle in bottoming applications has a serious drawback due to the constant temperature of vaporization. while that of the steam is typically just 500–600 °C. unless otherwise stated. proves to be a workable one at a lower temperature. At high temperatures (above 500 °C) a fluid with a high critical temperature is required. such as mercury or potassium vapour. Modern supercritical Rankine plants with reheat and regeneration that operate at temperatures of 600 °C and pressures of 350 bar obtain an efficiency not higher than 48–49% (Rogers and Mayhew. readily available. as for example. other refrigerants are preferred for use in the organic Rankine cycle. Further in the text steam will be regarded as a working fluid in the Rankine cycle. dimethyl ether (Kustrin et al. (c) with regeneration. such as the Kalina cycle. Using advanced materials. 1. (b) with superheat and reheat.2.. a superheat temperature of 816 °C was obtained at a 4 MW demonstration plant (Rice. A mixture of more than two components was also suggested in the so-called SMR-cycle (Verschoot and Brouwer. Regarding the other possibility. 1994). Regarding the devastating impact of chlorofluorocarbons on the ozone layer.
Flow scheme of the Kalina cycle. as shown in Fig. condensed and pressurized before entering the vapour generator. 1. The vapour is mixed with the basic solution. The ammonia in the mixture begins to vapour first. This reduces the temperature mismatch between the topper's waste heat and the fluid in the boiler. The main flow passes the recuperator and is flashed in the separator.3a) versus the Kalina cycle (Fig. . Vapour turbine Heat Separator HRVG Recuperator Working solution Basic solution Enriched vapour Lean solution Condenser Absorber Fig. A diagram of heat utilization in the Rankine cycle (Fig.10 COGENERATION. 1.3b) clearly shows the advantage of the latter. and allows an efficiency rise in the bottoming cycle. The composition of the fluid is changed in the cycle at different points. Heat utilization line in: (a) the Rankine and (b) the Kalina cycle. COMBINED AND INTEGRATED CYCLES T T Q Q a) b) Fig. In the DCSS the stream from the turbine is cooled in the recuperator. 1.3. 1.4. 1. its concentration decreases. In most Kalina cycle studies a mixture of water and ammonia has been used. and then mixed with a lean solution of ammonia in order to raise the condensing temperature.4. Part of the flow is sent to dilute the ammonia-rich stream coming from the separator. fluid of variable composition is used to provide a better match between the temperatures of the hot and cold flows (Kalina. and the distillation condensation subsystem (DCSS). 1984). and the boiling temperature of the mixture increases. a steam-ammonia turbine. The resulting basic solution is condensed in the absorber and brought to the recuperator under pressure. A basic Kalina cycle consists of a heat recovery vapour generator (HRVG). and as it boils off.
such as the gas turbine (Marston and Hyre. modern gas turbines offer high efficiency (up to 42%) and a considerable unit power output (up to 270 MW). . 1994). isentropic expansion to initial pressure (process 3–4). 1. 1991).4 Joule-Brayton cycle The Joule. and. (b) with reheat. which eliminates the heat transfer area. The Joule cycle is realized in a gas turbine plant.. as well as for cogeneration (Olsson et al. 1. solar and low temperature heat recovery applications (Olsson et al. 1.BASIC CYCLES T 3 3 b 3 11 a 4 2 1 2 1 4 4 2 c 1 d s a) b) c) Fig.5a). The waste heat is rejected into the atmosphere. finally. where the fuel is burned directly in the working fluid. 1997). 1994). 1991).5.. and in geothermal (Marston and Sanyal. Although the exhaust is released at temperature of 400 to 600 °C and represents appreciable energy loss.. 1995). Basic and advanced Joule cycles: (a) basic cycle. The high air excess ratio needed for turbine cooling is responsible for a relatively high oxygen content in the exhaust gases (14–16%) and a larger mass flow than in a comparable gas engine. or Brayton cycle is a gas power cycle that consists of the following processes: isentropic compression of the working fluid to the maximum working pressure (line 1–2 in Fig. directfired power plants (Kalina. Modifications within the Kalina cycle have also been analyzed (Glasare et al. Some typical modern gas turbines are listed in Table 1. (c) with intercooling Several investigations showed a favourable performance of the Kalina bottoming with various topping cycles. 1993). These specific features of the Joule cycle can be utilized by means of supplementary firing and waste heat recovery.1 (Gas Turbine World Handbook. heat addition at constant pressure along line 2–3. which makes the cooler unnecessary. Both these factors make the gas turbine plant compact and less expensive than a steam plant with the same power output. the heat release at constant pressure (line 4–1).
2 34. the ABB GT26 is three times lighter than the Intercooler Air Fuel Fuel Stack Recuperator Fig.0 35. the gas turbine achieves around 75% of the fullload efficiency. Turbine LM1600PA OGT15000 LM2500PE GT10 FT8 RB211 LM5000PC LM6000PC Trent PG6101FA 701DA PG9311FA 701F2 V94.8 34. The expansion work can be increased by means of reheating (Fig. and at 30% load this drops to 50% of the nominal efficiency (Kehlhofer.9 51. % 35. For example.2 34.6 34. .0 253.5 34.6.2 136.8 37.2 70.2 36. COMBINED AND INTEGRATED CYCLES Table 1.0 265. intecoling and recuperation.1 38. Therefore. such as the controlled inlet guide vanes and multi-shaft designs.8 17.7 255.9 226.3A GT26 701G1 Manufacturer General Electric Mashprom/Orenda General Electric ABB United Technologies Rolls-Royce General Electric General Electric United Technologies General Electric Westinghouse General Electric Westinghouse Siemens ABB Westinghouse Power. arrangements.9 41. 1.5 38.2 41.12 COGENERATION.5 27.1 35. the plant becomes compacter: for example.5 38. this makes it possible to provide full-load efficiency within a broader load range by varying reheat fuel flow.8 24.7 One important disadvantage is that a gas turbine does not perform well in part-load operation. The Heron gas turbine. Performance specifications of some modern gas turbines. Other modifications of the cycle include reheat. Because of the increased specific work output due to reheat. 1991).2 38.1 22.1. 1. MW 13.0 Efficiency. at 50% load.6 25. moreover.5b). are employed to improve the part-load performance.7 37.0 271.5 43.
Efficiency can be raised when recuperation of the exhaust flow is employed. the maximum permissible temperature in the cycle can be as high as 2500 °C.5c). this allows higher turbine inlet temperatures. if air cooling is applied. The use of recuperation is limited. and constant volume heat rejection to the atmosphere (Fig. Gas and Diesel engines are commonly used to generate electric power at power ranges below . 1. 1995). General Electric PG9311FA with approximately the same power output. which are significantly higher compared to 25–32% for other turbines of this power range size. The recuperated turbines are expected to obtain efficiencies from 39% (WR–21) to 43% (Heron) and even as much as 50% (the ATS Solar prototype expected to be available by the year 2000). Another technique to increase the specific work output is intercooling. 1.6) described by Poolman (1993).7. Otto and Diesel cycles. however. An in-depth analysis of the regenerated gas turbine cycle can be found in Beck and Wilson (1996). The engines have a rather high efficiency and provide good performance at part-load conditions. This concept is used in several gas turbines. such as the 1. The compressor outlet air becomes colder and. the 21 MW Westinghouse/Rolls-Royce WR–21 (Shepard et al. isentropic expansion of the heated gas to produce work.4 MW Heron (Fig. constant volume heat addition.7b). or ATS series Solar turbines in the 1 to 25 MW size range (Stambler.7a).BASIC CYCLES 13 p 3 2 3 2 4 1 4 1 v a) b) Fig. 1995). so that the cylinder is exposed to high temperature for only a short period during the whole cycle. and to decrease it intercooling may be applied. 1. by the compressor outlet temperature T2.5 Otto and Diesel cycles In the Otto and Diesel cycles the work is obtained in a sequence of non-flow processes. Since both engines operate in a reciprocating manner. The Otto cycle (the spark-ignition cycle) consists of isentropic compression. The Diesel compression-ignition cycle resembles the Otto cycle.. 1. 1. 1. but heat addition occurs here instead at a constant pressure (line 2–3 in Fig. which diminishes the work required by the compressor (Fig.
as shown in Fig. COMBINED AND INTEGRATED CYCLES T 3 4 Heater Expander Regenerator Cooler 2 1 Compressor s Fig. The latter makes it possible to incorporate a heat recovery arrangement as a bottoming cycle. 70 MW with efficiency up to 50%. which makes the equipment bulky and heavy when large power output is required.8. the latter offers the highest power density. helium and hydrogen.6 Stirling cycle A cycle consisting of two constant volume processes and two isothermals is called the Stirling cycle (Fig. . Therefore. Diesel and Gas Turbine Worldwide Handbook. Of air. The main disadvantage of reciprocating engines is the slow work rate. and rejected isothermally in process 1–2. a free-piston design can be impemented. 1997). The Stirling cycle. For stationary applications the engine may use helium or less costly methane. methane. Different gases may be employed as the working fluid. which is important in mobile applications. 1. Further improvements of Otto (Diesel) engine can be achieved by increasing the compression ratio.8. 1995. 1. Along with the crankshaft arrangement. the 51 MW MAN K 98 engine weights more than 1500 tons.8). the Stirling engine generates waste heat at two levels: the lower one of the cooling circuit. allowing a compact design and low noise level.14 COGENERATION. (Gas Turbine World Handbook. In Otto (Diesel) cycle waste heat is generated at two temperature levels: by a lowtemperature flow of coolant (90–125 °C) and by a medium-temperature flow of exhaust gas (200–400 °C). while the ABB GT8C gas turbine with the same output just 95 tons. Heat exchange occurs between processes 2–3 and 4–1. 1. Heat is added from an external source at constant temperature in process 3–4. 1997). One end of the regenerator must be continuously maintained at the upper temperature which is only limited by metallurgical restrictions. whereas the gasoline engines have smaller power output (up to 500 kW) and lower efficiency (Anonymous. Just as an Otto (Diesel) engine. the cycle possesses the Carnot efficiency (TADDED – TREJECTED)/TADDED. This makes the engine suitable for heat recovery over a broad range of temperatures. 1. For example. turbocharging and intercooling.
1.4. 1997). Schematic of a fuel cell. 1.3). the Ericsson cycle has never gained acceptance as a heat engine.9. this externally-fired engine can provide an efficiency comparable to that of the Otto (Diesel) cycle and much lower emissions. Because an infinite number of the intercoolers seems unrealistic.8 Fuel cells Fuel cells convert the chemical energy of the fuel directly into electricity. An important advantage of the Stirling cycle is a flat part-load characteristic: the engine maintains nominal efficiency even at 50% load (Penswick. The chemical reaction is nearly reversible. and therefore. the reversed Stirling cycle has been successfully used for refrigerating. and are not restricted by the Carnot efficiency.7 Ericsson cycle To improve the efficiency of the Joule cycle the infinite number of intercoolers and reheat stages together with a heat exchanger between lower. modern Stirling engines have not reached such a mature phase of development as Otto or Joule cycles. and commercial units are only available at a relatively low power level of several hundred watts. Nevertheless. 1. this form of energy conversion can be considered the most effective one.and higher-pressure streams can be applied to the cycle to provide heat addition and rejection at constant temperature. . Besides. In spite of difficulties in designing an efficient and compact heat exchanger. in an air liquefaction plant. An arrangement with water injection at each compressor stage can approximate compression at a constant temperature (see Section 3. as for example. in practice.BASIC CYCLES 15 Load Anode Fuel Exhaust Cathode Oxydant Exhaust Electrolyte Fig. However. The same holds true for the reheat stages. thus approaching the Carnot cycle. An efficient coal-fired plant can be implemented using the Stirling cycle for power systems less than 10 MW. it is seldom greater than two. where the Rankine cycle's hardware becomes costly and inefficient. and the higher-temperature one of the heater's exhaust.
coal gas Methane. 1993) Type of cell Fuel Power range MW Efficiency present. Solid Polymer Fuel Cell (SPFC). The fuel (hydrogen. methanol) are processed in a reforming reaction to form hydrogen and carbon monoxide. AFC and SPFC operate at rather low temperatures of 60–100 °C and are best suited for transport or domestic applications. % 40–45 45–55 45–50 Efficiency in 2015. COMBINED AND INTEGRATED CYCLES Table 1. The former operate at a temperature under 200 °C. Phosphoric Acid Fuel Cell (PAFC). Fuel cells have the useful characteristics that their efficiency increases at part-load and that the efficiency is independent of scale because of their modular construction. due to the high degree of irreversibilities of the anodic oxidation reactions of hydrocarbons.9). Current and projected performance specifications of fuel cells (from Van Barkel et al. methanol Methane.1–5 1–100 1–100 Type of cell Lifetime present.16 COGENERATION. Other fuels (methane. thus utilizing the waste heat. % 45–50 55–60 55–65 Phosphoric Acid Molten Carbonate Solid Oxide Methane. NLG/kW 2300 1800 1800 Phosphoric Acid Molten Carbonate Solid Oxide 35 000 15 000 22 000 A fuel cell consists of electrodes put into an electrolyte. NLG/kW 4200 3900 3900 Capital cost in 2015. . In the ionizing reactions electric current is generated (Fig. Basically. methane) is brought to the anode. the low-temperature systems can be used as a bottoming cycle. and minimal water use are among other important advantages. the latter above 600 °C.. hrs 80 000 35 000 35 000 Capital cost present. coal gas 0. 1. these can be divided into two groups: low-temperature (AFC. This determines the application of a fuel cell in a combined cycle. or coal-derived fuel can be used in a fuel cell.2. Different fuels such as hydrogen. SPFC. very low pollutant emissions. hydrocarbons. The type of the electrolyte determines the type of fuel cell. Low maintenance. and Solid Oxide Fuel Cell (SOFC). and the oxidant (air.. and the high-temperature ones as a topping one. which permit a fuel cell plant to be located in urban areas. However. 1993). While the low-temperature cells are fed with hydrogen which is produced externally. hydrogen has proved to be the most suitable fuel (Srinivasan et al. the high-temperature ones can supply enough heat for internal reforming. oxygen) to the cathode. low noise and vibration. hrs Lifetime in 2015. The most developed cells types are the Alkaline Fuel Cell (AFC). PAFC) and high-temperature (MCFC and SOFC) fuel cells. Molten Carbonate Fuel Cell (MCFC).
1. In a reversed Carnot cycle a wet vapour is compressed isentropically and passed to the . The results of the study by Van Brakel et al. Barclay. The plant is designed to operate with an overall efficiency of 50% (Anonymous.11. A Carnot cycle-based heat pump. An absorption heat pump. 1996). Rankine. these figures are projected to be greatly reduced with increasing production volume (Van Brakel et al. 1. 1995. 1. Condenser Heat Generator Throttle valve Evaporator Absorber Fig. 1995).10.. Joule or Stirling) to supply heat at an elevated temperature at the expense of work done on the working fluid. Considerable improvements in power density and operational lifetime are expected in the future development of fuel cells. Capital costs around 3000 NLG/kW for the year 2000 and 1500–2000 NLG/kW in 2030 have been given. A 2 MW demonstration direct carbonate fuel cell plant has been opened in Santa Clara. Hirschenhofer and McClelland.9 Heat pumps Heat pumps use a reversed cycle (Carnot.BASIC CYCLES 17 T Condenser 3 2 3 2 Tprocess M Expander 4 Evaporator Compressor 1 4 1 Tambient s Fig. are presented in Table 2. 1993.
hydrolysis-dehydrolysis (Murata et al. 1. Then the fluid is expanded to its original pressure and evaporated at constant pressure to the initial state (Fig. expanded. where a combination of endothermic and exothermic reactions is employed (Fig. condensed. 1. The pump work required to bring a liquid to the desired pressure is much less than the work required to compress a vapour. A concept of CO2 conversion chemical heat pump using activated ceramics was described by Tamaura and Tsuji (1993).. 1. the ideal COP of the Canot heat pump in this case is about 6. evaporated. A typical temperature range for a heat pump lies between 5 and 60 °C. 1993. . the vapour can be drawn off. the better the COP. A chemical heat pump. but allows a compact construction. condenser.12).2) The amount of work is always less then the heat supplied.11). A chemical heat pump enables heat transportation over long distances without the need for insulation due to a negligible amount of sensible heat in comparison to the latent heat contents of a fluid (Hasatani. Then. The coefficient of performance (COP) of a Carnot heat pump can be defined as COP = Q ADDED Q ADDED T MAX = = W Q ADDED − Q REJECTED T MAX − T MIN (1. and COP is thus always greater than unity. etc. The work done by the compressor can be considerably reduced when the working vapour is first dissolved in a liquid before compression. and dissolved again.12. A number of reactions have been proposed: dehydrogenation-hydrogenation. Some simplification can be made in the expansion part when the expander is replaced by a throttle valve. This principle is used in an absorption heat pump (Fig. It is also obvious that the smaller is temperature difference between the upper and the lower temperatures.18 COGENERATION. The heat pumps can be applied in virtually any cycle to supply heat at a low-temperature level for industrial or domestic needs. This introduces irreversibilities in the cycle and diminishes the amount of heat extracted in the evaporator. 1993).. The ratio of output to input expresses the performance of a heat pump. 1. 1995). Yamashita and Saito. 1995). there is some heat needed in the generator to drive off the vapour from the solution.10). but on the other hand. COMBINED AND INTEGRATED CYCLES Low temperature heat Endothermic reactor Process heat Exothermic reactor Fig. therefore. namely heat supplied QADDED to work input W. The waste heat can be also brought to a higher temperature by a chemical heat pump. depolymerization-polymerization (Kawasaki et al.
It performs quite well within different temperature ranges (efficiency up to 45%) and in part load. just as the Stirling cycle. offer better COP values. It performs rather well in part load. good part-load performance. however. a high operating temperature (up to 650 °C) inevitably involves the use of very high pressure (up to 350 bar). External firing makes it possible to use any fuel. and up to 70 MW for gas and Diesel engines). but requires a bulky steam boiler. The projected power plant size may be as high as 100 MW. The Otto (Diesel) represents another mature energy conversion technology of high efficiency (up to 50%) and a broad power range (up to 0. The cycle is well-developed. a compact design. The Stirling cycle is well-known. steam turbine and condenser with the associated cooling system. high efficiency (better than 50%). They promise. but requires a high-quality fuel such as natural gas or oil distillate. and the possibility of integration with other cycles. such as the Rankine or Stirling cycles. is still in development and only one demonstration plant has been built so far. The heat pump technology is reaching its mature state. The fuel cells are. and thus a higher efficiency without the need for costly multiple-pressure and reheat arrangements. however. Regarding fuels. still in the development phase. The heat pumps that employ thermodynamic cycles with a constant temperature of heat addition and rejection. poor part-load operation of the gas turbine remains a major drawback. and further implementation of heat pumps in industries and domestic application is expected.5 and 280 MW and operates at best with an efficiency of 42%. but the power size of the engine is limited by several kilowatts. modularity. and offering a maximum efficiency at a level of 47%. has not yet been developed to the levels of the Otto (Diesel) or Joule cycle. The Kalina cycle. since heat addition occurs from an external heat source.10 Conclusions The basic thermodynamic cycles can be summarized as follows: The Rankine cycle (as a steam plant) is the most developed cycle covering the power range of 10–1000 MW. low emissions. which increases the mean temperature difference in the boiler. The Otto (Diesel) engine has a good part-load performance. Maximum cycle temperature can be as high as 2000 °C.5 MW for gasoline engines. but the power density is lower than that of the Joule cycle. The Joule-Brayton cycle operates at temperatures up to 1500 °C and moderate pressures up to 40 bar. Nevertheless. Also. again thanks to the water properties heat rejection can easily be obtained at a temperature close to ambient. to provide an easier emission control and operation at a low noise level. The Kalina cycle offers a variable-temperature profile of evaporation. .BASIC CYCLES 19 1. Due to intrinsic steam properties. It covers a power range between 0. The working parameters of this cycle are close to that of the Rankine's. Another drawback is a constant temperature of evaporation. whereas the temperature is still much lower than that of the fuel flame (2000–2400 °C). any sorts and qualities are suitable for this cycle. but nonetheless.
K. F. D. Y.20 COGENERATION. Chapman & Hall. New York. 1997. and Wilson D. Gas Turbine World 1997 Handbook. Japan. Horn. 6. 1995. March. "The Coming of Age of Fuel Cells". A. 1996. Kalina. Morgantown... Heron Exergy BV. Nagoya.. Japan. 1995. Mechanical Engineering.W. Beck. Watanabe. Modern Power Systems. M. 8. 12. pp.I. Hirschenhofer. J. December 4–6.H.. 97–102. ASME Publications. 1995. New York. 1995. Corman. "The Selection of Working Fluids Other Than Steam for Future Power Generation Cycles". M. Nagoya. "A Theoretical Comparison of the Rankine and Kalina Cycles". 1995. J. G. Aichi. CT. Anonymous. 1966. "Studies on Paraldehyde Depolymerization for Chemical Heat Pump". 1996. I. in: Analysis of Energy Systems – Design and Operation. "Direct Carbonate Fuel Cell Brings Distributed Power Nearer". 1996. November. Pergamon Press. 5. 41–47.I. G.R. in: Second Law Analysis – Industrial and Environmental Applications. Kustrin. 13. Olsson... Proceedings of the International Symposium on Advanced Energy Conversion Systems and Related Technologies RAN'95. "Kalina Cycle Tested at Canoga Park". M. 10. Combined Heat Process – An Exergy Approach. Gas-Turbine Regenerators. Modern Power Systems. Von Spakovsky. and McClelland. and Tribus. Bercic. 19. October. "The Heron Turbine".M.. and Norris. presented at the Advanced Turbine Systems Annual Program Review Meeting. 4. R. 737–742. R.D. 7. 1980. Fairfield.. vol. 298–305. G. Hasatani. Proceedings of the International Symposium on Advanced Energy Conversion Systems and Related Technologies RAN'95..G.S. 3.. 1993. December 4–6. "Properties of Dimethyl Ether As a Working . 11. E. and Svedberg.. Kalina.. ASME Publications.. 1985. Haywood. 1984. 17..H. and Kanzawa. El-Sayed. ASME Paper 93-GT-156. August. August. T.K. and Tuma. Barclay. "Kalina Cycles and System for Direct-Fired Power Plants". 18. London. A. pp.1. 1992. Pequot Publishing..C. 2. "Combined Cycle System with Novel Bottoming Cycle". October 17–19. Kawasaki. 16.. "H Gas Turbine Combined Cycle". 1991. 1997. Diesel & Gas Turbine Worldwide Catalog. Oxford.. USA. 15. "New Diesel Engines Target Asian Market". AESVol. Mechanical Engineering Publications Ltd.191.. The Chemical Engineer.. H. Wisconsin. pp. T. Glasare. Journal of Engineering for Gas Turbines and Power. New York. M. 9. Aichi. pp.. 106.. Analysis of Engineering Cycles. West Virginia. A.B. COMBINED AND INTEGRATED CYCLES References 1. 1995. "Highly Efficient Conversion Technologies for Energy Utilization". "Analysis of Kalina Cycle Designs with Modifications Including a Distillation Column". Modern Power Systems. Brookfield. 14. Anonymous. Diesel & Gas Turbine Publications. Company Information. AESVol.J.25/HTD-Vol. Anonymous.
J. Stecco. Rice.. Marston. and Kameyama. Recuperation". and Chiprich. K. and Visser. Blomen and M. 1995. New York. Florence. pp. M.. "Gas Turbine Bottoming Cycles: Triple-Pressure Steam Versus Kalina". Kennewick.R. pp. Longman Scientific & Technical. Olsson.. Toyoshima. G. and Tsuji.. Murata. Srinivasan. Dejfors. 34. the Netherlands. Y. Rogers. 1994. in: Verslag van de Tweede Energieconversiedag. Terai. vol. Marston. "Optimization of Kalina Cycles for Geothermal Application". Tamaura. Olsson. "An Integrated Gas Turbine – Kalina Cycle for Cogeneration". 117. Proceedings of the Florence World Energy Research Symposium FLOWERS'94. E.. G. 30. 28. P. K. Dave. Penswick B. Proceedings of the International Symposium on CO2 Fixation and Efficient Utilization of Energy. S. 1995a. 385–394. "Kalina Cycles for Power Generation from Industrial Waste Heat". Stirling Technology Company. R. Gas.H. and Svedberg. Mugerwa. Analysis.C. 32. 1994.K... ASME Publications. Italy. 1991. T. M.. I. Y.G.E. 1997. 1993. S. UK.B... G.K. eds. Proceedings of the Florence World Energy Research Symposium FLOWERS'94. C. Work and Heat Transfer. 1995. Desideri. New York. B. E.P. C. "Overview of Fuel Cell Technology".. Verschoot. "SMR-Cyclus Haalt Meer Rendement uit Restwarmte" (SMR-Cycle Shows Better Efficiency for Waste Heat Recovery). 117.. Murugesamoorthi. 33. Bowen. Tokyo. I. November.B. in: Fuel Cell Systems.A. 1993. Y..N. 1994. January–February. D.J. 33. WA. 27. Journal of Engineering for Gas Turbines and Power. "CO2 Conversion Chemical Heat Pump System Using the Activated Ceramics". Y. Thermodynamics and the Design.. Yamamoto. E. Stambler.B... 1993. 's-Hertogenbosch. Japan.. M. July 6–8. Essex. November–December.. H.. "Small Engine ATS Design Project Stresses Intercooling.M. Italy. "Experimental and Simulative Study on Chemical Heat Pump with Reaction Couple of Cyclohexane Dehydrogenation and Benzene Hydrogenation for Effective Use of Industrial Waste Heat of Solar Energy". Polytechnisch Tijdschrift. 21. Japan. Proceedings of the International Symposium on CO2 Fixation and Efficient Utilization of Energy. ASME Paper 91-CT-202... Engineering Thermodynamics. Gas Turbine World. K. 20. C. S. Shepard. 1993. and Hyre. and Brouwer. (in Dutch). July 6–8. 23. 1992. Florence. 26..A. October.. "Brandstofcellen: Stand van Zaken en Toekomst"(Fuel Cells: Present and Future). 1994 (in Dutch). 1997. and Mayhew... L.. 25.. Jansen. "Design and Development of the WR–21 Intercooled Recuperated (ICR) Marine Gas Turbine". 1993.J. A. 29.J. J. vol. pp. E. Journal of Engineering for Gas Turbines and Power. .J. "Kleinschalige Gasturbinetechniek" (Small-Scale Gas Turbines). Plenum Press... 27. 119. E. Thorin. 557–562. April.. Tokyo. 97–104. and Svedberg. 24. and Improvement of Energy Systems. pp..BASIC CYCLES 21 Fluid in an Organic Rankine Cycle"... 10–15. Poolman. Journal of Engineering for Gas Turbines and Power. "Split Stream Boilers for High-Pressure/High-Temperature Topping Steam Turbine Combined Cycles".M. U. and Appleby. Van Brakel. 31. vol. A. Parthasarathy. and Sanyal. S.F..L. 22.S. AES-Vol. Personal communication. (in Dutch).J.H.
Y. COMBINED AND INTEGRATED CYCLES 35.22 COGENERATION. 1993.. Proceedings of the International Symposium on CO2 Fixation and Efficient Utilization of Energy. M. "Catalyst Study on Liquid-Phase Dehydrogenation of 2-Propanol for a Newly-Proposed Chemical Heat Pump". . Japan. and Saito. Tokyo. Yamashita.
The advantage resulting from combining cycles was illustrated by Kehlhofer (1991) for three single cycles (a gas turbine plant. others reject heat at nearly ambient temperature.1d) covers the largest area on the T–s diagram.or low-temperature bottoming cycle. a steam boiler plant. The steam turbine plants (Fig. 2. but rejects heat at quite a high temperature level of 800 K (Fig. 2. in order to approach higher efficiency.1b and Fig. Being divided into topping and bottoming cycles according to their temperature range. Of the combined cycles. 2.Chapter 2 Combined Cycles Abstract This chapter describes possibilities and potentials in combining basic thermodynamic cycles in order to obtain better performance. Therefore. but their waste heat is released at a temperature close to that of the environment. a combined cycle is required with a high-temperature topping cycle and a medium. a gas turbine cycle with external firing and one with air bottoming were selected for the further consideration within the NECT programme. a number of possible combinations were defined. and a steam plant with reheat) and a combined one. and this results in . but possess a moderate maximum temperature. The combination of these cycles (Fig.1a). the integrated concept of a gas turbine and a waste incinerator.1 Introduction As shown in the previous chapter. 2. some basic cycles have a maximum temperature closer to the fuel flame temperature. The resulting combined cycles were evaluated for potential performance gains and feasibility. Another combination. is addressed in Chapter 6. a gas turbine plant operates within a broad range of temperature. but a rather high temperature of heat rejection. 2. Performance of these cycles is analyzed in Chapters 5 and 7.1c) reach a maximum working temperature of only 800 K. from 300 to 1300 K. As seen on a T–s diagram.
The high exhaust temperature of a gas turbine indicates its low efficiency. 1991).2. % (a) 1000 520 48 (b) 600 300 50 (c) 680 300 56 (d) 1000 300 70 Fig. as shown in Fig. Comprehensive thermodynamic and economic analysis of combined cycles together with practical examples are given by Horlock (1992). COMBINED AND INTEGRATED CYCLES T 1300 K T 650 K 800 K 550 K 300 K 800 K 300 K s (a) (b) s T 1300 K T 800 K 650 K 650 K 800 K 750 K 500 K 300 K 300 K s (c) (d) s Plant Average temperature of heat supplied. the highest Carnot efficiency of 70%. 2.1.24 COGENERATION. The drawback of one cycle may become a benefit when combined with another cycle. To determine what cycle is better suited for topping or bottoming application. K Carnot efficiency. The highly efficient aero-derivative gas turbines provide worse performance in a combined-cycle configuration than industrial turbines with lower efficiency and higher exhaust temperatures. 2. K Average temperature of heat rejected. Comparison of single cycles with a combined cycle (after Kehlhofer. whereas the best single-cycle efficiency approaches a value of 56%. These considerations should be taken into account when searching for the optimal parameters of combined cycles. but is advantageous for the steam bottoming cycle. they can be ranked according to their operating temperature range. .
The Joule (gas turbine) cycle along with the Otto and Diesel cycles can be better applied as topping cycles.COMBINED CYCLES 25 T.1). For the reference fuel cell system a high-temperature scheme was chosen. The Otto and Diesel cycles are comparable in terms of operating parameters. which makes use of a heat exchanger located in the furnace of a Table 2. and medium. Combined cycles considered.2. 2. In some configurations. in a Joule cycle with external firing. a number of combinations can be defined (Table 2. just as the Stirling cycle. for example. High-temperature cycles are good candidates for topping. such as the Solid Oxide Fuel Cell (SOFC). so they were regarded as the same cycle. When these cycles are put into a matrix.1. °C 2000 1500 Otto (Diesel) cycle 1000 Joule cycle (gas turbine) High-temperature fuel cells Stirling cycle 500 Joule cycle (air bottoming) Rankine cycle (water/steam) Kalina cycle 0 Rankine cycle (organic) Low-temperature fuel cells Fig. This is the case. both cycles are integrated so that it is ambiguous which is the topper and which is the bottomer. Thermodynamic cycles arranged according to their temperature range. The Rankine cycle is suitable both for topping and bottoming. Topping cycles Rankine Otto/Diesel Joule w w w w w w w w w w Fuel cell w w w w w w w Bottoming cycles Rankine Kalina Joule Otto/Diesel Stirling Fuel cell Heat pump w w w w w .and low-temperature cycles for bottoming. and the Molten Carbonate Fuel Cell (MCFC).
COMBINED AND INTEGRATED CYCLES Steam topping cycle Fuel Organic bottoming cycle Air Stack Fig.3. 1966) (Fig. 2. A pilot 40 MW plant was developed and built by the General Electric Company in the late 1940s that used mercury vapour for the topping application. steam boiler. In the case of steam topping at moderate conditions.3). due to the variable chemical composition of the latter (see Section 1. The thermodynamically attractive properties of mercury (critical temperature of 1500 °C.2 Rankine/Rankine For the topping cycle a working fluid should be selected with a high critical temperature so that heat transfer takes place at the maximum allowable temperature under the saturation line. but its very low saturation pressure at top temperature (0. is undesirable.3 Rankine/Kalina The Kalina cycle can provide a better match between the temperature of the flue gases and that of the working fluid. The application of this bottoming cycle in direct-fired power plants (Fig. 1980). ammonia or an organic medium can be applied as the working fluid in the bottoming Rankine cycle (Horn and Norris. such as the potassiumsteam-organic Rankine cycle.3). The concept of steam topping at high pressure (350 bar) and temperature (above 700 °C) with flow splitting was outlined by Rice (1997). 2. saturation pressure at 550 °C of 14 bar) made it possible to combine the mercury topping cycle with the steam bottoming.1 bar at 560 °C) and other disadvantageous properties inhibit its use for this application.4) can improve the performance of the whole system by 23% compared to the . or when a fuel cell is installed before the combustion chamber of a gas turbine.26 COGENERATION. Plant efficiency of 37% was mentioned in the literature (Haywood. Potassium vapour can also be considered as a fluid for the topping cycle. since the complexity of triple-cycle combinations. A steam topping/organic bottoming (Rankine/Rankine) cycle . 2. 2. In the following description of combined cycles mostly dual cycles are considered. 2.
An arrange- . 1995).. The Rankine/Kalina combined cycle. Fuel Air Stack Stirling engine Fig. 1991. Rankine bottoming cycle (Kalina.COMBINED CYCLES 27 Kalina cycle Fuel Air Stack Fig. the Kalina cycle will obtain approximately 20% greater efficiency than the triple-pressure Rankine cycle. Kalina and Tribus. 2. The Rankine/Stirling combined cycle. 2. but limited by the small power size of the Stirling engine. which simplifies operation and maintenance. 2. 1992). 1996).4 Rankine/Stirling The use of the Stirling engine in combination with the Rankine cycle is theoretically attractive. Then. The use of the Kalina cycle has been recently licensed by ABB for direct coal-fired power plants (Stambler. at the same top temperature of 650 °C.4. The steam boiler can be completely replaced by the Kalina-cycle vapour generator. while the former requires no extremely high pressure (Davidson et al. No part of the Kalina system operates at vacuum conditions.5.
can be considered in order to obtain a higher maximum temperature in the Stirling cycle. where the gas turbine exhaust is fed into the existing steam boiler's furnace. The bottoming cycle can be implemented either as a heat recovery steam generator (HRSG) behind a gas turbine. before the superheater.6.3. and water/steam for the bottoming. the Joule/Rankine combined cycle is the most developed and wide-spread. Another position of the heater. steam turbine) have led to wide acceptance of this scheme. Possible Joule/Rankine combinations. Rankine bottoming cycle Fuel Stack HRSG Air ST Gas turbine Fig. and well-developed technologies (gas turbine.28 COGENERATION. for example. oil-. Inexpensive and readily available media. The Joule/Rankine combined cycle. waste heat boiler. Joule topping cycle direct-fired HRSG unfired with supplementary firing Steam boiler (repowering) gas-. The use of steam as a working or cooling medium within the gas turbine is realized in integrated schemes. . Possible combinations of these two cycles are presented in Table 2.5 Joule/Rankine Of the various combined cycles. such as air for the topping cycle. coal-fired waste-fired Integrated with gas turbine steam and water injection blade cooling combustor cooling w w w w w w w w w w w w indirect-fired ment where the Stirling engine's heater is mounted in a steam boiler is shown in Fig.5. 2. 2. 2.3. COMBINED AND INTEGRATED CYCLES Table 2. or in a repowering scheme.
the same efficiency is cited for the Westinghouse steam-cooled W701G gas turbine (Gas Turbine World Handbook. these high efficiency values can be achieved at large units above 300 MW.1). A dual-pressure arrangement can increase the power output of the steam cycle by up to 10%. MW 15–50 51–100 101–250 above 250 Efficiency. Prices and efficiencies versus power output of gas turbine/HRSG combined-cycle plants. General Electric expects even higher efficiencies with the new 'H' technology.and triple-pressure triple-pressure with reheat Price.5. With a single-pressure HRSG typically about 30% of the total plant output is generated in the steam turbine. more than one pressure level is used. However. 2.7. MW Power range. 1995). Modern gas turbine plants with a triple-pressure HRSG with steam reheat can reach efficiencies above 55% (see also Section 3.1 Conventional scheme (GT/HRSG) In a typical scheme. and an additional 3% can result by choosing a triple-pressure cycle (Marston and Hyre. In order to provide better heat recovery in the steam boiler. exhaust heat from the open gas turbine circuit is recovered in a steam generator (Fig. % 40–50 44–52 50–52 56–58 HRSG type double pressure double pressure double. 2.6). USD/kW 900 900 800 800 700 700 600 600 500 500 400 400 300 10 100 300 1000 Power. 1997). . Based on the data from the 1997 Gas Turbine World Handbook. USD/kW 650–850 600–800 400–750 350–450 Fig.5% efficiency of a combined cycle plant built around their GT24/26 reheat gas turbines. ABB claims 58.3.COMBINED CYCLES 29 Price. A gas turbine with steam cooling of the turbine blades and nozzles combined with an advanced HRSG is expected to operate at a efficiency level of 60% in the near future (Corman. 1995). 2.
air cannot be heated to the required gas turbine inlet temperature. 2. 2. since the heat recovery occurs only within the liquid state. Here. Such a configuration is analyzed for the case of a solid waste incinerator in Chapter 6. The Joule cycle with external firing combined with the Rankine cycle. The turbine exhaust is blown into the boiler at elevated temperature and serves as a hot windbox (Fig. 2. the cycle should provide an efficiency better than 50% and/or a price lower than 600 USD/kW.3 Externally-fired gas turbine A combination of an externally-fired gas turbine with a direct-fired boiler allows combustion of low-grade fuels to drive turbomachinery and to generate steam in the steam boiler. In the current state of the technology. 2. After compression. 1200–1300 °C. air is heated indirectly in a heat exchanger. Huang and Naumowicz. 2. Thus.5.5. A performance analysis of such a system can be found in the literature (Trumpler. (2) Exhaust heat can alternatively be recovered by preheating the boiler feed water for the existing steam plant. COMBINED AND INTEGRATED CYCLES Stack Fuel Air Natural gas (optional) Fig. 1995).8. A further increase of the air temperature up to 1300 °C can be accomplished by the use of supplementary firing of . and then expands in the turbine to the atmospheric pressure. the maximum temperature that the feed water at subcritical plants can reach is just 374 °C.. this combination is employed to repower an existing steam plant. An overview of equipment-only prices and efficiencies based on The Gas Turbine World Handbook (1997) is given in Fig. 2. The table under the figure indicates that for an alternative combined-cycle to be attractive in the under 50 MW power range. (1) The gas turbine exhaust can be utilized as preheated combustion air in the steam boiler or a process furnace.8).30 COGENERATION. 1985. so the maximum practicable temperature is considered 950 °C (Seery et al.2 Repowering In addition to the classical Joule/Rankine scheme (Fig. 1992). the degree of waste heat utilization is lower than in an HRSG.6).7.
Another advantage is the smaller size of the whole unit. 1995. in combination with a General Electric MS9001FA gas turbine. Corman et al. The Joule cycle with the Kalina bottoming. 2. 1991. coal-derived gas). The externally-fired cycle with and without supplementary firing is analyzed in Chapter 5... 1985. the Kalina combined cycle rates 412 MW and 59. This concept is now being developed under the Combustion 2000 programme of the US Department of Energy by a team of manufacturers. The working fluid composition can easily be changed in order to obtain the optimal performance in respect to alterations in load or ambient conditions. Therefore. 1995).. Because the exhaust pressure of the vapour turbine in the Kalina cycle is above atmospheric pressure. Rumminger et al. Bjorge. 1994. engineering companies and universities. For example. 1997).9. 1995).COMBINED CYCLES Fuel Air Vapour turbine Gas turbine HRVG Stack Separator 31 Fig. or by the use of advanced ceramic materials in the heat exchanger.. In these studies was found that the Kalina cycle can produce 10–30% more power than a Rankine cycle. The operation of the 3 MW pilot plant built in Canoga Park. no vacuum is needed to be maintained in the condenser during operation. The footprint of the Kalina plant is about 60% of the size of a Rankine plant design (Stambler. 1992. 1995). or stand-by periods.6% efficiency versus 396 MW and 57. Marston and Hyre. 1992. and the cost of electricity 10% lower than current coal-fired plants (Ruth.3% in the conventional triple-pressure combined cycle (Bjorge et al. Kalina et al. California.. clean fuels (natural gas. SO2 emissions less than 26 g/GJ. was demonstrated in 1992-1994 (Anonymous. 1995). More than . the Kalina cycle has been investigated in several studies (El-Sayed and Tribus. The High Performance Power System (HIPPS) is planned to be commercially available by 2005 with the following specifications: thermal efficiency (based on HHV) of greater than 50%. the start-up procedure can be performed in a much shorter time. NOX emissions less than 26 g/GJ. Olsson et al.6 Joule/Kalina As a bottomer for a gas turbine. 2.
heats the air in the secondary gas turbine cycle. Recent studies (Bolland et al.2%. or a water processing unit. COMBINED AND INTEGRATED CYCLES Fuel Air Exhaust Air Air-out Intercooler Fig. which. The exhaust of the primary gas turbine is sent to a heat exchanger. introduction of the air bottoming cycle with two intercoolers led to an increase in power from 5. Hirs et al. No traces of corrosion were noticed in samples from different parts of the unit. . condenser).8 to 43. 1995) showed the feasibility of this configuration... In comparison to the Joule/Rankine combined cycle. The scheme can also be applied for cogeneration. For example.10. this scheme does not require bulky steam equipment (boiler. for the Allison 571K topping gas turbine. 2. and allows unmanned operation. 1995. These reported an increase of power by 18 to 30% depending on the number of intercoolers. leaving the cycle at 200–250 °C. can be used for process needs that require heat of such temperatures. A General Electric demonstration plant is planned to be put into operation by 1998.9 to 7. The exhaust air. California) to design and market gas turbine powered Kalina combined cycle plants. steam turbine. 2. Inc. Hayward. Air is expanded in the turbine to generate additional power. The plant is expected to be in the 100 MW size range (Bjorge et al. 1997). 4000 hours of operation confirmed the high degree of the heat recovery and system operability under different work conditions.. A gas turbine with air bottoming (the Joule/Joule cycle). Intercooling in the air compressor reduces the required expander work. in turn. 2.7 Joule/Joule Two Joule cycles can be combined by an air-gas heat exchanger (Fig.32 COGENERATION. and an efficiency growth of up to 10% points. In 1993 General Electric signed an exclusive world-wide licensing agreement with the Kalina cycle patent owner (Exergy.10).5 MW and in efficiency from 33. Comparable results were obtained with the General Electric LM2500 topping turbine..
Joule/Otto (Diesel) combined cycles: (a) pressurization by a gas turbine. When the outlet air is used in food processing industries.8 Joule/Otto(Diesel) Preheating of the inlet air of an Otto (Diesel) engine can sufficiently improve its performance. Uncontaminated hot air is of special interest for food processing or pharmaceutical industries. when the exhaust gases. 2. (b) turbocharging. the firing should proceed indirectly. The performance analysis of this cycle is given in Chapter 7.COMBINED CYCLES Air Fuel Exhaust 33 Fuel Air Gas turbine Fig.12. Because the bottoming cycle has a low temperature ratio. The exhaust heat can be recovered either in a heat exchanger (Fig. or directly. The Otto (Diesel) engine inlet air preheating by a gas turbine. expansion work insignificantly exceeds the work consumed by the compressor. are passed into the Otto engine's combustion chamber. the cycle is sensitive to turbomachinery efficiency. Therefore. 2.11. Another possibility offers a pressurization scheme. The gas turbine exhaust can be applied in order to increase the inlet flow temperature. Fuel Air Air Exhaust Fuel Exhaust Gas turbine a) b) Fig. 2. and necessitates the use of the most efficient equipment.11). where air is extracted from the compressor and fed into the Otto (Diesel) . 2. containing 14–16% oxygen. This sensitivity can be diminished by burning additional fuel and raising the inlet temperature in the bottoming cycle to 1000–1200 °C.
1994). this heat can be recovered in a waste heat boiler (Fig. 2. the heater of the latter can be placed either in the combustor of the turbine. Simple turbocharging and utilization of engine exhaust is covered in Section 2. . or after the expander in the exhaust flow (Fig. thus. because it does not require a considerable air flow for cooling: the air excess is typically 30–40%.7%.34 COGENERATION.5 MW of electric power that can be generated in the bottoming steam or organic Rankine cycle (Vernau. performance of the Rankine cycle rapidly declines when the exhaust temperature is decreased (Woodward. Just as the Joule/Joule cycle. 1984).10 Otto(Diesel)/Rankine The exhaust gas from the Otto/Diesel engine has a temperature of 300–500 °C and. Moreover. Gas turbine waste heat recovery by the Stirling engine. Hence a 10 MW engine can provide about 1–1. the engine outlet flow expands through the low-pressure stage of the gas turbine (Fig. COMBINED AND INTEGRATED CYCLES Fuel Air Gas turbine Exhaust Stirling engine Fig. 2.9 Joule/Stirling In a combination of a gas turbine and a Stirling engine. The arrangement is determined by the optimal performance of the combined cycle and by the materials used in the Stirling heater's head. Such a combined-cycle plant will obtain an efficiency of 47.13.. engine. As much as 9 MW can be recovered by a bottoming Stirling cycle from the exhaust of a Rolls-Royce RB211 gas turbine of 27 MW (Walker et al. However.13).14a). the engine releases a smaller amount of exhaust gases than a comparable gas turbine.12a). this combination provides a compact and simple heat recovery scheme. compared to 200–350% for a gas turbine. 2. 2. 1994).13 on the Otto (Diesel)/Joule combined cycle. 2. Subsequently. 2.
3 to 36. Otto (Diesel) engine waste heat recovery: (a) by the Rankine cycle. 2. where it is recovered to drive the piston. An enhancement of a turbo-charged Diesel engine by means of inlet air humidification has led to an increase in efficiency (from 33.12 Otto(Diesel)/Stirling Yet another Otto engine heat recovery configuration is presented in Fig. this concept seems most suitable when applied in a vehicle or other small size applications.14. On the other hand. 2. The use of the ammonia-steam mixture may improve heat recovery. 2.5%). Woodward (1994) noted that high cylinder exhaust pressure and low temperature favour the Joule alternative. 2.11 Otto(Diesel)/Kalina A combination of the Otto and Kalina cycles can be considered comparable with the Otto/Rankine combination.COMBINED CYCLES Fuel Air Steam generator Exhaust Air Fuel 35 Otto engine Stirling engine a) b) Fig. Taking into consideration the small size of the Stirling engine. as well as in the reduction of NOX emissions (from 12 to 3 g/kWh). as reported by Rosen and Olsson (1996). .13 Otto(Diesel)/Joule The implementation of an expander in the Diesel cycle allows to expand the exhaust to the atmospheric pressure and thus provide additional power to the cycle (Fig. 2.12b).13b. and power output (from 185 to 205 kW). high cylinder gas temperature and low pressure make the Rankine cycle more attractive for heat recovery. 2. but because the exhaust flow of the Otto engine is not significant. the improvement may not be justified by the complexity of the Kalina bottoming unit. The gas engine exhaust provides heat to the Stirling engine's heater. Comparing the use of the Joule-Brayton cycle versus the Rankine cycle in the bottoming configuration. (b) by the Stirling cycle.
For a smaller-scale unit of 20 MW. COMBINED AND INTEGRATED CYCLES Air Fuel Fuel cell Stack Steam turbine Fig. SOFC with steam bottoming (Fuel cell/Rankine).36 COGENERATION. the figures were 46% and 47%.14 Fuel cell/Rankine An evaluation study of a fuel cell system with steam bottoming (EPRI. and the local cell impedance of 0. average stack temperature of 950 °C. 2. respectively. 1992) indicated that the plant had the potential to maintain high efficiencies over a broad range of load conditions. assuming fuel utilization of 80%. The study.15. Water Stack Steam generator Fuel cell Fuel Air A C Fig. 2. 1994) an efficiency of 55% was reported. The schematic diagram of the plant is given in Fig. showed that a large-scale plant (300 MW) would have 49% efficiency in full load and around 52% in 25%-load operation. based on Westinghouse atmospheric tubular solid oxide fuel cell technology. Fuel cell in the Joule cycle. In another study (Mozaffarian. The capital requirement for the combined cycle plant was estimated to range from 800 to 1200 USD/kW for the large unit. . 2.7W·cm2.15.16. and 1500 to 2100 USD/kW for the small one. 2.
The specific costs of a fuel cell/gas turbine combined cycle from 1600 NLG/kW to 2500 NLG/kW were cited.17. The schematic of the system is presented in Fig. Jansen and Mozaffarian. 2.. 1994. resulting from a reforming reaction of methane. a parallel concept can be considered. a turndown ratio of 4:1. Jansen et al. NOX emissions lower than 5 ppm. By the year 1999 the natural-gas fuelled plant is expected to have efficiency above 60%. steam injection. Other advanced concepts with anode/cathode gas recirculation. intercooling and recuperation (Harvey and Richter. therefore. A study by Westinghouse (Parker and Bevc. Several studies showed an advantageous performance of such a concept. 1996) were reported to obtain efficiencies around 70%. operating at elevated pressure allows to integrate a gas turbine within the system.16 Fuel cell/Joule A fuel cell system. The lowest cost of electricity was obtained at a value of fuel utilization of 55%. and installed costs of 1000 USD/kW. 2. In addition to the series scheme. the Kalina bottoming cycle offers 10–30% more power and 3–9% higher efficiency than the Rankine bottoming cycle. be seen as a choice for the future power plants. Fuel cell combined with an Otto engine and a waste heat boiler. thus improving performance.16. and can. The gas mixture. hydrogen is fed to a low temperature fuel . the findings of the reports cited in that section can be extrapolated for the fuel cell/Kalina combination with a correction for higher efficiency. 2.COMBINED CYCLES 37 Fuel Air Water Steam generator Exhaust Fuel cell Fig. Then. In the Argonne National Laboratory's report (Minkov et al. is separated by the use of a membrane into hydrogen and carbon dioxide. 1995. 1988) the optimal fuel utilization in a MCFC-gas turbine combined plant was discussed. This combination is claimed to have the highest efficiency of any options discussed in this chapter. The use of the fuel cells integrated with combustors allows efficiency to approach 70%.6.. Therefore. 2. 1996) reported on a 3 MW SOFC combined cycle plant based on a Heron turbine.15 Fuel cell/Kalina As already mentioned in Section 5.
Fuel Oxydant
Fuel cell
Stirling engine
Fig. 2.18. The Fuel cell/Stirling combined cycle.
cell, while CO is burned in the combustor of the gas turbine. The most effective polymer membranes cannot be applied here directly, since they operate at temperatures below 150 °C. The use of other materials, such as ceramic, glass, or metal, will allow separation of the reformed gas at higher temperatures. However, such membranes are costly and less effective, and the implementation of this scheme may be hindered by economic reasons.
2.17 Fuel cell/Otto(Diesel)
A number of configurations can be considered, regarding the combination of a topping fuel cell and a bottoming Otto (Diesel) engine. High-temperature waste heat from the fuel cell could preheat the inlet air of the Otto (Diesel) engine, and the engine's exhaust would be recovered to generate steam for the reforming process (Fig. 2.17). Alternatively, the fuel, partially utilized in the fuel cell, could be passed through a steam boiler directly into the Otto engine.
2.18 Fuel cell/Stirling
This combination can provide high performance on a small scale, determined by the size of the Stirling engine, for example, in automotive vehicles (Fig. 2.18). Since both represent a relatively new technology, this concept is not expected to be realized in the immediate future.
2.19 Fuel cell/Fuel cell
Partially utilized fuel and oxidant from a topping fuel cell can be directed to a bottoming cell to complete oxidation. For example, the SOFC anode flow containing unused fuel and H2O/CO2 mixture are passed to MCFC, and the cathode flow of unused air and H2O is sent to the low-temperature PAFC. Such a scheme allows operation at even higher efficiency than a combination with a heat engine, although its realization does not seem to be feasible within the next 10 years.
2.20 Conclusions
Combined cycles offer better performance than single cycles. Moreover, shortcomings of the topping cycle can often be compensated by the bottoming cycle. Combining high-temperature cycles with those of medium- and low temperature provides the most effective way in approaching Carnot efficiency, and thus better utilization of the fuel exergy. However, the possibilities for combination may be limited by various factors, such as the status of development, power output, fuel requirements, or part-load characteristics. Considering these factors the following combined cycles were found of interest for future development. The Joule/Rankine cycle, in the form of an externally-fired combined cycle, can provide high efficiency and low emission release, while burning low-grade fuels. This cycle is considered in Chapter 5. In addition, the gas turbine can be used to repower an existing steam plant, or to enhance a solid waste incinerator. The latter option with possibilities for superheating in an HRSG behind the gas turbine is discussed in Chapter 6. The Joule/Joule cycle, a gas turbine with air bottoming, offers a performance comparable to that of the steam bottoming cycle without the use of steam-water equipment. The air, leaving the bottoming cycle at a temperature above 200 °C, can be utilized in food processing industries. Analysis of this cycle is given in Chapter 7. The Joule cycle with a topping high-temperature fuel cell represents a highly effective combined cycle that, furthermore, has low emissions and a broad turn-down ratio. The development of this cycle falls under the scope of another project within the NECT programme, and therefore, will be not covered in the following work. Several other promising cycles have not been selected for further development due to a small power size (the Stirling cycle), or because of the advanced state of development elsewhere (Kalina cycle).
1. Anonymous, "Kalina Cycle Tested at Canoga Park", Modern Power Systems, March, 1992. 2. Anonymous, "San Diego Repowers with 501Fs for Power and Efficiency Boost", Gas Turbine World, July-August, 1994. 3. Bjorge, R., "Vested Interest in Kalina Combined Cycles", Gas Turbine World, November-December, 1995. 4. Bjorge, R., Boericke, R., O'Connor, M.F., and Smith, R.W., "Kalina Combined Cycle Power Plant Design and Performance Characteristics", Proceeding of the International Conference Power-Gen Europe '97, Madrid, Spain, June 17-19, 1997. 5. Bolland, O., Førde, M., and Hånde, B., "Air Bottoming Cycle. Use of Gas Turbine Waste Heat for Power Generation", ASME Paper 95-CTP-50, 1995. 6. Corman, J.C., "H Gas Turbine Combined Cycle", presented at the Advanced Turbine Systems Annual Program Review Meeting, Morgantown, West Virginia,
October 17-19, 1995. 7. Corman, J.C., Bjorge, R.W., and Kalina, A., "Kalina Cycle Looks Good for Combined Cycle Generation", Modern Power Systems, July, 1995. 8. Davidson, M.J., Peletz, L.J., and Kalina, A.I., "Kalina Cycle System Advancements for Direct Fired Power Generation", Proceedings of the International Conference Power-Gen '96, Orlando, Florida, December 4-6, 1996. 9. El-Sayed, Y.M. and Tribus, M., "A Theoretical Comparison of the Rankine and Kalina Cycles", in Analysis of Energy Systems - Design and Operation, AESVol.1, pp. 97-102, ASME Publications, 1985. 10. EPRI, "Evaluation of the Westinghouse Solid Oxide Fuel Cell Technology for Electric Utility Applications in Japan", EPRI Final Report TR-100713, 1992. 11. Gas Turbine World 1997 Handbook, Pequot Publishing, Fairfield, CT, USA, 1997. 12. Harvey, S.P. and Richter, H.J., "A Detailed Study of a Gas Turbine Cycle with an Integrated Internal Reforming Solid Oxide Fuel Cell", Proceedings of the 29th Intersociety Energy Conversion Engineering Conference IECEC'94, Montrey, California, August 7-11, 1994. 13. Haywood, R.W., Analysis of Engineering Cycles, Pergamon Press, Oxford, 1980. 14. Hirs, G.G., Wagener, M.T.P.A., and Korobitsyn, M.A., "Performance Analysis of the Dual Gas Turbine Combined Cycle", in: Thermodynamics and the Design, Analysis, and Improvement of Energy Systems (ed. by R. Krane), AES-Vol. 35, pp. 255-259, ASME Publications, New York,1995. 15. Horlock, J.H., Combined Power Plants, Pergamon Press, Oxford, 1992. 16. Horn, G. and Norris, T.D., "The Selection of Working Fluids Other Than Steam for Future Power Generation Cycles", The Chemical Engineer, November, pp. 298-305, 1966. 17. Huang, F.F. and Naumowicz, T., "Overall Performance Evaluation of an Externally-Fired Air-Turbine Combined-Cycle Power Plant Based on First-Law as well as Second-Law Analysis", Proceedings of the International Symposium on Efficiency, Costs, Optimization and Simulation of Energy Systems ECOS'92, Zaragoza, Spain, June 15-18, 1992. 18. Jansen, D., Mozaffarian, M., and Van Schie, N., "High Temperature Fuel Cells. A Competing Technology for Gas Turbines? If You Can't Beat Them, Join Them!", Proceedings of the 2nd International Fuel Cell Conference, Kobe, Japan, February 5-8, 1996. 19. Jansen, D. and Mozaffarian, M., "Advanced Fuel Cell Energy Conversion Systems", Energy Conversion and Management, Vol. 38, No. 10–13, pp. 957–967, 1997. 20. Kalina, A.I., "Kalina Cycles and System for Direct-Fired Power Plants", Second Law Analysis - Industrial and Environmental Applications, AES-Vol.25/HTDVol.191, pp. 41-47, ASME Publications, 1991. 21. Kalina, A.I. and Tribus, M., "Advances in Kalina Cycle Technology (1980-1991). Part I: Development of a Practical Cycle, and Part II: Iterative Improvements", Proceedings of the Florence World Energy Research Symposium FLOWERS'92, Florence, Italy, June 7-12, 1992. 22. Kehlhofer, R., Combined-Cycle Gas & Steam Turbine Power Plants, The Fairmont Press, Lilburn, Georgia, 1991.
29. 107. 25. E.R. D. Krumpelt. Fauvel. J. Gas Turbine World.R. Nagoya. vol. and Yoshimura. Olsson. pp. London. Combustion and Engines. 1995. 861-871. Efficient Electricity from Coal". Vernau. Rice. vol. Lucerne. Proceedings of the 2nd International Fuel Cell Conference. B. Stambler. and Bevc. 1984. Rosen. Journal of Engineering for Gas Turbines and Power. 35. 1996. and Bornstein. Using Humidification of the Compressed Air". October 3-7. I.. Desideri. S. Mozaffarian.W.J.. Milton. M. Minkov.. J. Journal of Engineering for Gas Turbines and Power.O. 1996. 1994. and Olsson. M. "An Integrated Analysis of the Kalina Cycle in Combined Cycles". July. Nagoya. Walker.G. Marston. 1994. Daniels. Thermodynamics. V... 27.D. P. Proceedings of the International VDI Seminar on ORC-HP-Technology. Stecco.. P. O. The Stirling Alternative. ASME Paper 91-CT-202. 33. 26. 1995. G.. 1995. Proceedings of the First European Solid Oxide Fuel Cell Forum. 1991.. 117. L. Lutz. Vol. G. Journal of Engineering for Gas Turbines and Power. Zurich. 30.. "SureCELL™ Integrated Solid Oxide Fuel Cell/Combustion Turbine Power Plants for Distributed Power Applications". 31. August 7-11.. June 25-27. R.. and Bingham. Reader. 385-394. Kobe. G.. Proceedings of the International Symposium on Advanced Energy Conversion Systems and Related Technologies RAN'95.. and Svedberg. "Gas Turbine Bottoming Cycles: Triple-Pressure Steam Versus Kalina". 1995.. 36. Holowczak... "Solid Oxide Fuel Cell for Combined Heat and Power Applications". A. I. pp. "Split Stream Boilers for High-Pressure/High-Temperature Topping Steam Turbine Combined Cycles".. 119. Sweden.. C.E. Parker.. J. M. M. Stockholm. Ruth. 37. September 10-12.G. Monterey. Rumminger. Aichi. "Recovery from Exhaust Gas on a Diesel Engine". 1988. 34. Seery. F.A.J.. A. Power Engineering. "Kalina Cycle Provides 25% More Power and 3% Better Efficiency". 28. Proceedings of the ECOS '96 Conference. Japan.P.. 1985. N. L. Gordon & Breach Scientific .G. February 5-8.Kalina Cycle for Cogeneration". Refrigerants and Heat Pumps. "The US Department of Energy's Combustion 2000 Program: Clean. and Hyre. E. Trumpler. Switzerland. Power Systems. Chapman & Hall. "A Cost-Effective Indirect Coal-Fired Gas Turbine Power and Water-From-Air Cycle".H. Japan.. Proceedings of the 29th Intersociety Energy Conversion Engineering Conference IECEC'94. "Engineering Development of a Coal-Fired High Performance Power Generating System". 10-15. Proceedings of the International Symposium on Advanced Energy Conversion Systems and Related Technologies RAN'95. "Topping Cycle Fuel Cells Effective Combined With Turbines"..R... "An Integrated Gas Turbine . Sangiovanni. July-August.COMBINED CYCLES 41 23. Japan. 24. December 4-6. California. U.M.E.. 1995. Dibble. 32. W. pp. "A Novel Technique for Reducing the NOx Emissions and Increasing the Efficiency of a Turbo-Charged Diesel Engine. A. E. December 4-6. Aichi... and Asbury. 1997.S..K..
Vol.. Switzerland. 1994. "Evaluation of Brayton and Rankine Alternatives for Diesel Waste Heat Exploitation". .B. 39-45. Yverdon. 116. J. 38. 1994. Journal of Engineering for Gas Turbines and Power.42 COGENERATION. pp. COMBINED AND INTEGRATED CYCLES Publishers. Woodward.
In spite of such a high efficiency value (with a TIT of 1350 °C. which is being carried out by gas turbine manufacturers.78). engineering companies.5% (Gas Turbine World Handbook. offers various way of improving cycle performance.Chapter 3 Advanced Cycles Abstract This chapter gives an overview of improvements that can be implemented within the basic and combined cycles described in the previous chapters. chemical recuperation. 3. research organizations and universities. including General Electric. 1997). . offer a combined-cycle package of similar efficiency at a power range above 250 MW. gas turbines with steam injection. and to obtain comparable efficiency at a smaller scale. Since the Rankine and Joule cycles represent the most developed technologies. the modifications covered here are primarily related to these two cycles. Main performance parameters are given along with a discussion on advantages and disadvantages of these schemes. and Westinghouse-Siemens. the Carnot factor is 0. which is caused by technical and economic constraints. four cycles were selected for the next phase of the NECT programme for technical and economic evaluation. and partial oxidation. where performance values are not so spectacular.1 Introduction The last generation combined-cycle plants based on gas and steam turbines can operate at efficiencies up to 58. These include multi-pressure steam boilers. At the scale under 30 MW combined-cycle efficiency is typically about 42%. Because the New Energy Conversion Programme is aimed at distributed power generation. air humidification. Based on conclusions of this chapter. Research and development work. All major gas turbine manufacturers. ABB. innovative cycles are being proposed in order to rise this level of efficiency even higher. the cycle innovations are interesting for an implementation within just this power range.
Another two studies resulted from the Collaborative Advanced Gas Turbine programme (Davidson et al. natural-gas-fuelled advanced gas turbines and cycles for use in utility and distributed applications in the next decade (Cohn et al. Gas turbinebased power plants will remain the most efficient energy conversion technology in the coming decade.... Advanced cycles reviewed in other studies.. which is the major constraint on the turbine efficiency.2 Background Several studies on the advanced energy conversion technologies have been performed in the recent years.. 1997 w w w w w w w w w w w w w w w w 3.1. which are jointly sponsored by DOE. (1995) and Yang (1995) described advances in gas turbine technology. 1995 Stambler. 1997 w w w w w w w w w w w w w w HAT Evaporative cooling w w w w w w w w w w w w w Kalina Chemical Partial Fuel cellCeramic cycle recuperation oxidation gas turbine gas turbine Stecco. 1992 Hodrien et al. 1994 Bannister et al. or to enhance the latter. and manufacturers.1. 1995 Chiesa et al. 1995). 1992). 1995 Yang. considerable investments in material research are aimed on increasing the firing temperature. Chiesa et al.. utilities. 1995 Yang.. 1992 Stambler. The California Energy Commission evaluated some promising new gas turbine-based technologies (Stambler. GT with GT with GT with thermal STIG intercooling reheat recuperation Stecco. Regarding the gas turbine technology. 1995 Davidson et al.44 COGENERATION. The aforementioned reports allow some conclusions to be drawn. 1995 Lior. 1997) and the Advanced Turbine System programme (Stambler. Advances in fuel cell development will allow the fuel cells to compete with gas turbine systems. 1992 Stambler. 1994). 1992 Hodrien et al. and Lior (1995) presented some novel energy conversion approaches. 1994 Bannister et al. 1995 Davidson et al. The technologies covered in these studies are summarized in Table 3.. a comparable increase in efficiency can be . Hodrien and Fairbairn (1994) gave a general overview of power generation systems from steam power plants to fuel cell combined cycle systems.. COMBINED AND INTEGRATED CYCLES Table 3.. Stecco (1992). 1995 Stambler. However. The programmes aim to accelerate the commercial availability of higher efficiency.
In order to simplify the heat recovery steam boiler the concept of water flashing was proposed by . (1997). As can be seen from Table 3. In modern power plants it is solved by the introduction of the second or third pressure level in order to reduce the mean temperature difference in the boiler. the cycles with intercooling.3 Advanced Rankine cycles 3.1. Such an arrangement results in heat utilization with a minimum temperature difference and still avoiding the use of mixed working fluid. achieved at lower cost by the use of intercooling. Operating within the subcritical region with pressures up to 180 bar and a superheat temperature of 542 °C.1. partial oxidation. and/or recuperation. The most advanced cycles. the HAT cycle. as in the Kalina cycle.2 Water flashing The multi-pressure system's complexity inhibits its use at small scale. 3. chemical and thermal recuperation have attracted most of the interest. 3. were reported to have a potential to reach efficiencies around 60% in a five-year period and are expected to enter commercial service shortly after 2000 (Hodrien and Fairbairn. reheat. To obtain a higher value. a steam bottoming cycle with five and even more pressure levels with steam turbo-chargers was proposed by Jericha et al. Triple-pressure heat recovery steam generator with reheat. The highest efficiency of a commercially-available advanced combined-cycle plant (ABB GT-26 gas turbine with a triple-pressure reheat HRSG) of 58.3. 1994). 1997).3. An example of such an arrangement is given in Fig.1.1 Multi-pressure steam boiler The drawback of the Rankine cycle in a bottoming application due to heat addition at a constant temperature has already been noted in Chapter 2. 3.5% was reported (Gas Turbine World Handbook. 3. a combined-cycle plant was reported to obtain overall efficiency of about 60%.ADVANCED CYCLES 45 Stack Fuel Air LP IP HP Gas turbine Steam turbine Fig. the chemically-recuperated and the partial oxidation gas turbines.
Instead of a separate high pressure section in the steam generator.3.2b). it is accomplished by several means. Since the same flow is fed through all tube bundles of the boiler.5% efficiency for a typical combined cycle plant based on the boiler with water flashing. In the following study Grimaldi and Desideri (1994) reported that steam recompression bottoming cycle is comparable to the conventional dual-pressure steam bottoming one. 3. 3. (a) Water flashing. This is particularly useful for small power plants where a low steam/water flow rate makes impractical splitting of the flow into different pressure levels. Normally. Dechamps reported 51. This arrangement adds up to 2 percentage points to the single-pressure combined cycle efficiency. while requiring a 20% smaller surface area of the boiler at the expense of the steam compressor. 3. or by using steam injection in a gas turbine. (b) Steam recompression Dechamps (1994). At a high-pressure level of 35 bar and a low pressure of 10 bar.3 Steam recompression Another concept to simplify the heat recovery steam generator was proposed by Cheng (1978).4 Steam flow splitting In a cogeneration plant special attention is paid to the plant's ability to operate under various heat loads.2a). and can be comparable with the dual-pressure HRSG schemes. dumping the excess steam in a dump condenser. A study by Grimaldi and Manfrida (1992) showed good thermodynamic performance of the plant. . 3. lower pressure level by means of extracting the saturated.2. the pressure in the boiler is raised by means of the steam compressor (Fig. 3. such as expanding of the excess steam through the condensing part of a steam turbine. a combined-cycle gas turbine plant will operate at 57% efficiency. COMBINED AND INTEGRATED CYCLES Steam compressor M Blowdown Heat Stack Heat Stack Steam generator HP LP Flash tank a) b) Fig. heat recovery is improved.3. This scheme creates the second.46 COGENERATION. pressurized water from the evaporator's drum and flashing it in the flash tank (Fig.
unless the turbine is specially designed to work in this mode. At combined-cycle cogeneration plants. Moreover. 3. Maintaining a constant flow through the main steam turbine allows operation at fullload efficiency at different heat load conditions.4 Advanced Joule cycles 3. The use of a condensing turbine requires an expensive low-pressure section and a condenser. 3. by steam injection (see also the following section). Steam injection is more effective than water injection. Another way to provide operating flexibility is the concept of steam flow splitting and recompression (Lievense and Hirs. 1986). heat load can be controlled by supplementary firing. The compressor is driven by a back-pressure steam turbine.ADVANCED CYCLES Heat Steam boiler Stack 47 ST Splitter Process Fig. The excess steam is recompressed to initial pressure by a steam compressor. since the steam generated in the HRSG is fed into the turbine.3. A combination of these methods with steam splitting will permit an effective and flexible operation of the cogeneration plant. Steam injection in a gas turbine is possible to a certain extent. at full heat load the condensing section presents a parasitic loss. 1989). it helps to lower NOX formation in the combustion chamber and to cool the blades more effectively than air. Such a cycle is . A simplified scheme of the concept is presented in Fig.4. and in industrial gas turbines since the 1960s (Foster-Pegg. and/or by putting the gas turbine in part-load operation. However. which gives additional power to the cycle.1 Water and steam injection The positive effect of the steam or water injection on the performance of a gas turbine is well-known. Water injection has been used for power augmentation in aircraft engines since the 1950s. each of those has its drawbacks. The injection increases the mass flow and the specific heat of the working fluid.3. The dump condenser appears even a worse solution. 3. thus improving the heat recovery. Steam flow splitting in a cogeneration plant. Along with this.
When compared with the combined cycle.4 (Cheng. 3. 1992. Larson and Williams. 3.4. 1978. while intermediate-pressure and low-pressure steam are often expanded in the first gas turbine stages. By this means the power output is increased from 34 MW to 52 MW and efficiency from 37 to 43% (Tuzson. often designated as the STIG (STeam Injection Gas turbine) cycle. There are several gas turbines speciallydesigned for steam injection. Steam-injected gas turbine configurations. and up to 6.5% of the air flow at low-pressure. COMBINED AND INTEGRATED CYCLES Water Stack Steam generator Fuel Air HP LP Gas turbine Fig. or Ruston TB5000. .. 1987). Rice. Steam injection for power augmentation and NOx reduction. High-pressure steam can be injected into the combustor.48 COGENERATION. Allison 501-K.5. Water Stack Steam generator Stack Water Steam generator LP ST Turbo charger HP ST Air Fuel Air Fuel Gas turbine Gas turbine a) b) Fig. 1995). 3. as shown in Fig. the STIG plant proves to be economically competitive in power range under 150 MW (Van Laar et al. In the LM5000 STIG turbine about 7% steam by weight of the air flow is injected at high pressure. 1988). such as the General Electric LM2500 and LM5000 STIG series.
proposed by Foster-Pegg (1989). to be recognized that expansion of steam in the gas turbine proceeds to the atmospheric pressure and in a less efficient manner than in the steam turbine. The analysis of this concept for different classes of gas turbines (industrial and aero-derivatives) showed that the dual-recuperated intercooled aftercooled steam-injected (DRIASI) cycle can provide comparable or superior efficiencies to those of combined cycles for small systems up to 30 MW. 1995) reported 52. the turbocharged steam injected gas turbine (Fig. 3. 1991). Therefore. The introduction of intercooling and reheat in a STIG turbine allows to reduce the power consumed by the compressor from 50% of the total output for modern engines down to 30%. however. Therefore..5b). recuperation. based on the data of the existing turbines from 3 to 47 MW. For larger systems. 3.2% efficiency at a TIT of 1370 °C and a pressure ratio of 34. the Turbo-STIG. An additional increase in efficiency by about 3 percentage points was expected with the introduction of a reheat stage (Macchi et al.ADVANCED CYCLES Water 49 Stack Intercooler Air Fuel Fig. efficiency of the gas turbine becomes less dependent on the compressor characteristics and the work ratio is considerably increased. and 53. and efficiency to 55% (Cohn. Whereas in the combined-cycle plant steam leaves the steam turbine at much lower pressures. and presented in Fig.6.6. His calculations. Other STIG modifications include: a STIG turbine with a topping steam turbine (Fig. 3. It needs.2% efficiency at a TIT of 1500 °C and a pressure ratio of 45. Another study on the intercooled STIG (Chiesa et al. A power plant that combines steam injection. the performance of the DRIASI cycle was found to be inferior both to combined cycles and to steam-injected turbines.5a). 1995). 1984). a gas turbine with steam injection will always have a lower efficiency than that in combined-cycle operation. 3.. where HP steam is first expanded in a back-pressure steam turbine and then is injected into the combustor (Rice. and water injection is described by Bolland and Stadaas (1995).6%. The power of a conceptual intercooled LM5000 STIG was reported to grow to 110 MW. . thus providing more power. showed that the Turbo-STIG configuration would result in an average increase in power of 95% and an average efficiency rise from 30% to 42. Dual-recuperation intercooled aftercooled steam-injected cycle (DRIASI).
27% increase in power and a 0. However. Every 10 mbar outlet loss will decrease efficiency by 0. 3. 3.7).0% decrease in efficiency and 1.7.7–0. (1996) a 3.50 COGENERATION. By means of supercharging.4. 1996). Kolp et al.9% and power output by 0.8.. Fuel Inlet air cooler Air Separatorcondenser Stack Water Fig. as an inlet air filter. COMBINED AND INTEGRATED CYCLES Fuel Air Water M Rotational Particle Separator Gas turbine Fig.68% increase . 1995). The use of the rotating particle separator can serve two purposes.1 µm and larger from gases (Brouwers. The rotating separator permits an effective separation of solid and liquid particles of diameter 0. additional fuel consumption will be required in order to bring the working fluid to the turbine inlet temperature. and a rather small increase in efficiency is expected. In Laagland et al. a 10 mbar loss will cause a 1–2 °C temperature rise in the stack. introducing an evaporative cooler. and as a supercharger (Fig. allows even more power to be produced. 3. The advantages of air cooling for power augmentation are described by several authors (De Lucia et al.6% (Gas Turbine World Handbook. 1997). An enhancement. 1993.6% reduction in power output. the inlet pressure in the system can be increased and that results in additional power generated in the gas turbine..8–1. In addition. Supercharging a gas turbine with the Rotational Inlet Air Filter and an evaporative cooler. 3. For a single-pressure combined cycle every 10 mbar inlet loss will result in a 0.2 Supercharging and evaporative cooling The pressure losses in the inlet air filter and in the heat recovery steam generator lead to a certain decrease in efficiency and power. Gas turbine with water injection in the compressor.
3. 3.8). The intercooling is accomplished by injecting water at the compressor stages.9) can provide even higher efficiency. and an isentropic efficiency of compressor and expander of 0. the use of heat exchangers for this purpose will result in a re-design of the existing engine. 3. Water in the exhaust is recovered in a separating-condensing unit. such as the one discussed below in section 3.. which results in nearly isothermal compression (Fig.4. 1983. The Humid Air Turbine cycle.4.67% and 0. in efficiency of a gas turbine are reported.85. 3. the humid air turbine (HAT) cycle provides a substantial power boost of the system and an efficiency rise of several percentage points.3 Wet compression The benefits of intercooling have already been discussed in the previous chapter. a pressure ratio of 30. and an increase in size and complexity. can be applied instead. Introduction of a reheat stage and an increase in the pressure ratio to 100 improve efficiency to 48% in a simple cycle. 3. However.ADVANCED CYCLES Economizer Stack Saturator Recuperator 51 Water Intercooler Air Aftercooler Fuel Fig. A more advanced concept with intercooling (Fig.11%. 1980).4. 1987). The . respectively.9.7 on gas turbine cycles with multi-stage combustion. known as wet compression. Anonymous. as reported by Qun et al. A much simpler approach.4 Humid Air Turbine cycle Originally proposed as the evaporative-regenerative cycle (Mori et al. the efficiency of a plant can reach 43% (Poletavkin. This concept is employed in more complex configurations. The wet compression was found feasible in experiments carried out on a Kawasaki S1A-02 gas turbine. In the case of the combined cycle the improvements were 2. (1997). Assuming a turbine inlet temperature of 1100 °C.
The exhaust heat is then recovered in the recuperator and in the economizer to preheat the water for air saturation (Rao. air is first compressed in the low-pressure compressor of the combustion turbine and then enters the intercooler. after preheating in the gas turbine exhaust. cooled in the aftercooler and then fed to the air saturator. and can also be started up in much shorter time than a conventional combined cycle plant (Nakhamkin and Gulen. Several authors reported the net electric efficiency of the HAT cycle.. The Cascaded Humidified Advanced Turbine (CHAT). whereas a study by Horlock (1997) suggests that the optimum pressure ratio is rather low. 3. By varying the water content the HAT-plant can be put in partload operation without penalizing efficiency. The water content increases the mass and the specific heat of the flow. 1991).10. The cooled compressed air is further compressed in the high-pressure compressor. The humid air leaving the saturator.. and the use of the recuperator gives higher efficiency. 1995) for a high-pressure cycle. . is fed to the combustor. The heat of compression is recovered for air saturation by circulating water and makeup water which is passed to the saturator. 1995).52 COGENERATION. It varies from 54% for a low-pressure ratio turbine (Stecco et al. COMBINED AND INTEGRATED CYCLES Economizer Stack Saturator Recuperator Water Intercooler Aftercooler IPC HPC Fuel HPT Fuel Air LPC LPT Fig. The hot gas exiting the combustor expands through the gas turbine driving the compressors and providing power. being in the range of 8 to 10. The air is contacted over packing with water heated by the various heat sources. which leads to additional power. 1993) to 57% (Chiesa et al.
. The typical steam to fuel ratio lies in the range of 2:1 to 4:1. separated from the exhaust gases. In addition to the injected water in the gas turbine exhaust. the cycle efficiency of 63–65% is projected for a CHAT plant of 500 MW (Nakhamkin et al. (1995) reported an efficiency of 55%. The flow scheme is similar to the wet compression cycle (Section 3. For a 300 MW HAT-plant specific capital costs of 400 US dollars per kW have been cited (De Biasi.5% and a capital cost of USD 800/kW (Anonymous.5 Semi-closed gas turbine In a semi-closed gas turbine cycle the gas turbine exhaust is cooled down to the condensation temperature of water vapour. A condenser with direct contact or indirect heat exchange can be considered. 1995). the exhaust must be cooled to 30–40 °C (Nguyen and Den Otter. 1996). treated. Due to the high moisture content in the working fluid a modified design of expanders is required.3). Being proposed as the Cascaded Humidified Advanced Turbine (CHAT) cycle. Facchini et al. Advantages of the semi-closed cycle include combustion at nearly stoichiometric conditions. Due to the acid content in the flue gases the condensing surfaces should be protected from corrosion. 3. there is also some water in the inlet air and some water formed in the combustion process. 3. the water injection is optional. To provide a continuous supply of water without the use of external sources. 1997). 1992). low emissions. Turbo Power and Marine together with Flour Daniel are evaluating FT4000 gas turbine options in the HAT cycle at 200 MW and 55. That means a water consumption of 3–5 ton/h for a gas turbine of 25 MW. good partload efficiency. and smaller size relative to a conventional combined-cycle plant with an open-cycle gas turbine. Nemec and Lear. Efficiencies reported were 2–3% points lower than those of the reference combined-cycle plants. avoiding the complications that high turbine inlet temperatures imply on hardware (Fig.3% efficiency (Day and Rao. Nakhamkin et al. 1996. and part of the exhaust gas is recirculated through the compressor (Gasparovich. high specific power. 3. for example by a Teflon coating. this water can be reclaimed in a recovery unit where some of the moisture content is condensed. with the addition of a heat recovery steam generator before the condenser-separator. 1997). 1968.. and reused (as shown in Fig.4. Nevertheless. Further development of the HAT-cycle is being carried out by Westinghouse for W501D5 and W501F gas turbines. If necessary.4.8).6 Water recovery The previously reviewed cycles with water/steam injections consume large quantities of water. The performance of a direct contact counter-current con- .10). the plant employs proven industrial components. When an advanced turbine with a TIT of the LP expander of 1500 °C and an overall pressure ratio of 80 is employed. 1993). A small-scale CHAT plant of 12 MW is expected to be ready for commercial installation in 1998 with an efficiency of 44.4.ADVANCED CYCLES 53 The cycle does not require expensive steam/water equipment that simplifies the scheme and lowers operating and maintenance costs. 3.
However. and sometimes CO2 into a hydrogen and carbon monoxide fuel mixture that can be burned in the combustor.54 COGENERATION. COMBINED AND INTEGRATED CYCLES Water Fuel Flue gas recycle M Exhaust Boiler Reformer Exhaust Air Air Reformer Fuel Gas turbine Gas turbine a) b) Fig. the gas turbine exhaust temperature is not high enough for a complete reforming reaction. this option was proposed for exhaust heat recovery in remote compressor station applications (Botros et al. At 550 °C only 20% of the fuel is reformed.4.7 Chemically-recuperated gas turbine The chemically recuperated gas turbine (CRGT) uses a reforming process to convert methane. Nevertheless.7% was reported. 1997).5% in simple cycle. (b) with flue gas recycling. In order to increase the temperature some additional firing can be applied. and it makes the turbine operate in a steam injection mode.. The chemical recuperation in gas turbines: (a) with steam reforming. This is a little less than that of the conventional combined cycle (47.11a). 1994) showed that a power plant based on a LM5000 gas turbine could reach an efficiency of 47% in comparison to 39. For the same gas turbine with steam injection efficiency of 46. the hydrogenrich fuel has greater flammability than methane and supports combustion at a lower flame temperature. a pinch-point difference of the condenser) can be found in the study by Bettagli and Facchini (1994). steam generated in an HRSG is mixed with natural gas in a reformer (Fig. In the scheme with steam reforming. water.11. 3. Janes (1990) estimated NOX production as low as 1 ppm. which potentially reduces NOX formation. 3. water content in the gas turbine exhaust flow. Recuperation that proceeds thermally and chemically results in a higher degree of heat recovery than in standard recuperation schemes. Different reforming schemes have been proposed. An analysis of a simple CRGT without intercooling or reheat (Kesser et al.. 3. This endothermic reaction absorbs heat at a temperature lower than the combustion temperature and in this manner increases the fuel's heating value. . denser in relation to various factors (water input temperature. This indicates that 10–20% of the conversion rate in the reformer is not sufficient to utilize the advantages of chemical recuperation. Moreover.9%).
Fuel Air Partial oxidation reactor
Conventional combustor
Fig. 3.12. Schematic of the gas turbine with partial oxidation.
Another scheme uses exhaust gas recuperation (Briesch et al., 1995), a portion of the flue gas is compressed, mixed with natural gas, heated with exhaust heat from the combustion turbine, mixed with the air from the compressor, and sent to the combustor (Fig. 3.11b). When the mixture is heated in the presence of a nickel-based catalyst, hydrogen and carbon monoxide are produced. The reaction is accelerated at low excess oxygen, low pressure, and high mass ratio of recycled exhaust gas to methane. Therefore, the best results are achieved when the fuel is burned at the stoichiometric ratio, and the exhaust gas is used to reduce the turbine inlet temperature. Briesch indicates the typical value of recirculation as over 50% of turbine flow. This means that both the air compressor flow and the exhaust flow are less than half that of conventional cycles with the same turbine size. In a recent paper (Newby et al, 1997) a comparison was given between simple, advanced and combined cycle plants based on the Westinghouse 501F turbine. Efficiency figures were: 35.7% in the conventional simple cycle, 38.7% in the fluegas reforming simple cycle, 45.6% in the STIG cycle, 48.7% in the steam reforming cycle, 56.8% in the conventional combined cycle, and 57.1% in the combined cycle with flue-gas reforming.
3.4.8 Gas turbines with multi-stage combustion
Approaching isothermal heat addition in a gas turbine can be accomplished by increasing the number of expansion stages, as done in the reheat cycle. Usually, air cooling is used in to maintain the working temperature within the metallurgical limits. This can be alternatively obtained by means of substoichiometric combustion, i.e. in a reheat cycle with fuel excess. The fuel-rich mixture is first partially burned in one or more combustors, and the final oxidation takes place in the last stage combustor (Fig. 3.12). Then, the exhaust heat is recovered in a bottoming cycle. Cancelling the excess air results in a decrease in work required by the compressor, reduced dimensions of the turbomachinery, and low NOX emissions. The concept of partial oxidation is implemented in the 'Chemical' gas turbine (Arai et al., 1995; Yamamoto et al., 1995). After substoichiometric combustion in the first burner, the hot gas containing unburned fuel expands through the high temperature turbine to the atmospheric pressure. The reduced atmosphere in the first stage allows the
Stack HRSG
Fuel-rich gas turbine
Fuel-lean gas turbine
Fig. 3.13. The Chemical gas turbine.
use of carbon-reinforced composites, which can withstand temperatures up to 1700 °C. The gas is then recompressed and passed on to the second, lean-mixture burner. The gas is expanded through the second turbine, and subsequently used to generate steam in the boiler of the Rankine cycle (Fig. 3.13). Lior (1995) found that such a system could reach efficiency level of 66%, assuming a turbine inlet temperature of 1500 °C and an effectiveness of the air-gas heat exchanger of 90%. Harvey et al., (1995) investigated a three-stage turbine with partial oxidation in the first two stages and chemical recuperation with recycled exhaust gas and water injected between compressor stages. Realistic assumptions made for the equipment characteristics (turbomachinery efficiency of 0.9, a turbine inlet temperature of 1260 °C, a pressure ratio of 20, and minimum ∆T in heat exchangers of 25 K) result in a 65% efficiency for an optimized scheme. Other multi-stage combustion turbine cycles were proposed in an attempt to approach isothermal heat addition. El-Masri and Magnusson (1984) described a multireheat gas turbine that performs the combustion/work extraction at a nearly constant temperature. However, the cycle requires high pressure ratio in the range of 40–100. Calculations made by El-Masri and Magnusson for a plant consisting of a gas turbine with a TIT of 1300 °C, a pressure ratio of 100, and a steam cycle at 650 °C, showed 65% efficiency. The combustion of hydrogen and oxygen produces steam of high temperature. When this reaction proceeds in a large number of stages nearly isothermal superheating of steam can be achieved. A plant with 15 stages, the maximum temperature of 525 °C, the maximum pressure of 180 bar, and the condensing temperature of 33 °C obtained 49.2% efficiency in calculations by Cicconardi et al. (1998). The gas turbine cycle with partial oxidation is covered in detail in Chapter 8.
CO2 condenser
Water Liquid CO2
Fig. 3.14. Zero-emission power cycle with CO2 recirculation.
3.4.9 CO2 gas turbine
A power plant where CO2 acts as a working fluid in a gas turbine was proposed by several authors (Yantovski et al., 1992; De Ruyck, 1992; Mathieu et al., 1995). In the example by Yantovski, oxygen from the air separation unit is delivered to the combustors at a pressure of 240 bar, along with the fuel and recycled CO2. The working fluid expands in the double reheat turbine (240/60/15/4 bar), the exhaust is cooled down in a heat exchanger, and water separation occurs (Fig. 3.14). In the multi-stage intercooled compressor the fluid is brought to a pressure of 60 bar, and at 20 °C liquefaction takes place. The liquid CO2 is pumped to a storage area and partially returned to the cycle. Assuming a turbine inlet temperature of 1300 °C, isentropic efficiencies of the gas turbine stages of 80%, 85% and 90%, and an isentropic efficiency of the oxygen compressor of 80%, Yantovski (1994) reported a total plant efficiency of 55%. Since a gas turbine operating at 240 bar and 1300 °C is not feasible, a simplified option with a working pressure of 60 bar was proposed, which can achieve 52% efficiency. The scheme by De Ruyck employs an evaporative water injection and that by Mathieu a heat recovery steam generator. They operate at pressures of 50–70 bar and have efficiencies of 45–47%. CO2 storage, such as depleted gas reservoir, must be considered for these cycles.
3.5 Other cycles 3.5.1 Ejector topping power cycle
Instead of expanding hot gases in a turbine, this cycle uses the gases to compress another gas in an ejector. In spite of relatively low efficiency, the ejector can tolerate higher temperatures than a turbine, and it can use working fluids that have better ther-
3. and the steam flow splitting as a means of a flexible control at cogeneration plants. can also be applied in a power cycle. 3. three technologies seem of interest in future considerations: the steam recompression and the water flashing as a means of creating a second pressure level in a boiler. System efficiency varies with topping boiler conditions. reforming with CO2-recycle appreciably diminishes the discharge . and helium as the secondary one (Fig.58 COGENERATION. The use of the steam flow splitting allows to extend the heat load range without a significant loss in efficiency. High-temperature ejector topping cycle. where a conventional double.15. The first two are especially attractive at small combined-cycle plants. The system has no moving parts and operates at intermediate pressure.15). The advantages of such an arrangement need to be evaluated in order to make the final conclusions about this concept. The Heron turbine can be considered as a good candidate for chemical recuperation. The steam-methane reforming process.and triple-pressure boiler is not economically justified. The components can be constructed from ceramic or graphite composites. modynamic properties than steam or air. As for the Joule cycle. COMBINED AND INTEGRATED CYCLES Na Ejector Na+He He Fuel Air H2O Fig. The cycle proposed by Freedman and Lior (1994) uses sodium as the primary fluid. 3. some enhancements of the gas turbine technology were considered. which is well-known in the process industry. Assuming a steam bottoming efficiency of 40%. at 1700 °C and 50 bar the total efficiency approaches 51%. since it has relatively high temperature of the power turbine exhaust (above 600 °C) which is favorable for the reforming process.6 Discussion and conclusions Among the concepts of the advanced Rankine cycles described in this chapter. A performance analysis of such systems showed their competitiveness and feasibility. Moreover. This is achieved by recompressing the excess process steam back to initial boiler pressure.
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1997). additional options with supplementary firing to double steam production are also examined. Criteria such as the fuel utilization efficiency. and electric and heat-activated heat pumps. Several typical industrial cogeneration schemes. The plants under consideration include conventional boiler. The last decade has witnessed a rapidly growing number of cogeneration units. In The Netherlands the amount of electricity produced at the CHP installations has grown from 9% of the total generated power in 1980 to 30% in 1997.Chapter 4 Industrial Cogeneration Abstract Combined generation of heat and power employs various schemes to satisfy district heating or process needs. 4. providing 6500 MWe (Holwerda. Current CHP penetration in the European market averages 9%. gas turbine schemes with a heat recovery steam generator and a back-pressure steam turbine.1 Introduction Combined production of heat and power (CHP) is generally regarded as an environmentally and economically attractive energy technology and is widely used for process and space heating needs. The gas turbine-based schemes were found the most efficient in fuel-saving and exergy terms. total fuel consumption. are assessed to quantify the differences between the options using different methodologies. and the European Commission would like this to reach 18% by 2010 (Anonymous. Further increase of the cogeneration . which provide process steam of 210 °C and 10 bar. and exergetic efficiency are discussed. fuel saving. The fuel savings of the case with the lowest heat-to-power ratio appeared the most susceptible to an improvement in utility efficiency. A number of performance assessment parameters are used to compare the schemes with each other and with separate generation plants. In the case of the gas turbine-based schemes. 1997). A comparison between the fuel energy utilization efficiency and exergetic efficiency for the cases considered is given.
The price ratio reflects the present day economy. the fuel savings. in a study on global CHP integration in an area where various industries and dwellings for 600 000 inhabitants were located (Haskoning. as well as to compare steam of different parameters in evaluation of internal steam prices within a chemical concern (Donders. 1994. Sarabchi and Polley. 1997) and optimization of cogeneration plants (Pak and Suzuki. in which heat is mainly produced by direct combustion boilers.2 Performance criteria Combined generation of heat and power implies production of two different kinds of energy. Of those parameters. such as the PURPA efficiency (derived from the Public Utility Regulatory Policy Act) that gives heat a factor of 0. in a free market would result in a different heat-to-power price ratio. Thermodynamics suggests the use of the exergetic factor which exactly indicates the quality of heat in terms of its work potential. which deals only with the quantitative side of energy. or energy efficiency.5 (Huang. the most straightforward criterion is based on the First Law. deals with the fuel consumption of the plants. (2) 'factored' efficiency. (3) exergetic efficiency. A common denominator should first be established to analyze the efficiency of a cogeneration plant. or fuel energy savings. 1994).66 COGENERATION. A number of criteria can be used to assess the plants's performance. The increasing number of cogeneration plants. or by politics. The first. The factor is determined by economics as the price ratio of heat to power. exergetic efficiency is now becoming more often used by engineering companies and industries. this considerable amount of power should be generated efficiently. After first being used as an academic criterion. 1985. 1995). Horlock. 1997). it cannot provide an adequate performance figure. which consume less fuel per unit of heat. 1998). This was shown. by Huang (1990) and Sarabchi (1992) for a basic gas turbine cogeneration plant. The demand for heat and power at an industrial site can be satisfied either by . An even more correct performance value is obtained if the exergy content of fuel is also taken into account: ηEX Q PCG + ECG = F ECG (4. Habib. It is defined as: η FU = QCG + PCG FCG (4. for example. 1996). the most frequently used are: (1) fuel utilization. as summarized in Huang (1996) and Horlock (1997). Obviously.2) Introducing the Second Law analysis has proved to be a valuable methodology for evaluation (Kotas. where a weight factor is given to heat. This drawback has led to the introduction of 'factored' efficiency.1) Since the fuel utilization efficiency gives the same value to work and heat. COMBINED AND INTEGRATED CYCLES share in the total power production is expected. 1978. which is usually taken as 1/3 (Timmermans. and (4) avoided fuel costs. Exergy was used. To evaluate the performance of a CHP plant an objective criterion is required. for example. The fourth criterion. 4.
and one for the demand side λD. Eqn. if the cogeneration plant cannot meet heat or power demands (Fig. thermal efficiency ηQ is the ratio of heat produced to the fuel input. The fuel consumed by the cogeneration plant can be expressed either as the ratio of the heat produced to the thermal efficiency. and two heat-to-power ratios. Here. or both. if all plants compared generate heat of the same quality. or by a separate production scheme. if the cogeneration plant is unable to supply all heat and power required. and electric efficiency ηP is the ratio of power to the fuel input of a plant. 4. 4. it is still suitable for comparison studies. The general expression for the total fuel consumption is P Q FTOTAL = ( FSEP + FSEP ) + FCG (4.1.3) The first two terms indicate the amount of fuel consumed by the separated scheme in order to fulfill the demand.4) If the fraction of heat generated by a CHP plant with respect to the heat required is designated as y.INDUSTRIAL COGENERATION Cogeneration plant Process demand 67 FCG ηQCG ηPCG HEAT FUEL FQSEP Separate production Boiler QD ηQSEP Utility POWER FPSEP η P SEP PD Fig. By introducing thermal and electric efficiencies for cogeneration and separate plants. Although the thermal efficiency does not reflect the quality of heat produced. Cogeneration versus separate production a cogeneration plant.1). one for the cogeneration plant λCG. then the expression for FTOTAL can be written as . or as the ratio of the power produced to the electric efficiency.3 becomes FTOTAL = = PD − PCG QD − QCG QCG + + Q P ηSEP ηQ ηCG SEP 1 P ηSEP Q −Q 1 QCG 1 ⋅ QD − QCG + D Q CG + Q λ CG ηSEP ηCG λD (4. 4.
5) If the cogeneration plant supplies all heat required. Thus. where all required heat and power are generated separately. when λCG ≠ λD). the term 1/λCG is omitted. it is also of interest to compare a cogeneration scheme with a reference separate production scheme in the case when both supply the same amount of heat. The total fuel consumption can be compared with the reference case.e. y = 1. and the fuel that would be consumed by a utility to generate power equal to that of the CHP plant. FPSEP. Eqn.8) The first term is a sum of the amount of fuel consumed by the reference boiler plant SEP.5 is reduced to 1 1 1 1 FTOTAL = QD ⋅ P ⋅ − + Q η SEP λ D λ CG ηCG (4. COMBINED AND INTEGRATED CYCLES 1 1 y 1− y y FTOTAL = QD ⋅ P ⋅ − + Q + Q η SEP λ D λ CG η SEP ηCG (4. The second term is the fuel consumed by the cogeneration plant. An expression of the relative fuel consumption factor can be defined as: 1 Φ= η FTOTAL = FREF P SEP 1 1 1 ⋅ − + Q λ D λ CG ηCG 1 1 + Q P η REF λ D η REF (4.6) For the specific case of a heat pump. Introducing efficiencies.68 COGENERATION. the fuel energy saving ratio (FESR) is defined as the ratio of savings to the separate scheme's fuel consumption (Horlock.7) Whereas the relative fuel consumption factor can assess any combination of combined and separate production of heat and power (i. the fuel savings becomes: FQ ∆F = F Q SEP P P 1 ηCG 1 ηCG + P FCG − FCG = QD ⋅ Q + Q ⋅ P − 1 η SEP η SEP ηCG η SEP (4. 4.e. FQCG.9) Since ∆F is an absolute value. the difference in fuel consumption is the fuel savings of the CHP plant: Q P ∆F = ( FSEP + FSEP ) − FCG (4. 1997): FESR = ∆F P F + FSEP Q SEP . because the heat pump does not produce any power. i.
when heat and power demands are "matched". P 1 1 η CG FESR = Q + Q ⋅ P − 1 η SEP η CG η SEP P 1 η CG Q + Q P η SEP η CG ⋅ η SEP = 1− ηQ CG P η Q ⋅ η SEP SEP P P ⋅ η SEP + η Q ⋅ η CG SEP (4. the critical cogeneration efficiency can be derived (Korobitsyn and Hirs. 4. The last decade has been marked by an increasing number of highly efficient combined-cycle power plants being put into operation. High-temperature fuel cell Heat flow HEAT Power flow Back-pressure ST Tprocess Stirling engine Condensing ST To Heat pump Fig. and a zero-saving line can be plotted in the coordinates of the thermal and electric efficiencies of the cogeneration plant (Wiemer. Possible cogeneration combinations. FTOTAL = FCG.e. the relation between FESR and Φ is simply: FESR = FREF − FCG FCG = 1− = 1− Φ FREF FREF (4. 4.INDUSTRIAL COGENERATION 69 FUEL Utility POWER Direct combustion IC engine.10: P ηCG ηQ CG + Q =1 P η SEP η SEP (4. Gas turbine.10) In the case. hence it seems necessary to evaluate the influence of an improvement in separate electric production (utility) efficiency on the fuel savings of a CHP plant. i.2. 1995). when fuel savings is zero. can be obtained from Eqn. .12) Considering efficiencies of the separate production as constant.11) A critical condition. 1997).
4 MWe 125. and that the return flow constituted 100% of the process flow. the reference case. All systems comprise a deaerator [1. A heat recovery steam generator features the approach and pinch temperature of 10 K.2).1 kg/s 462°C 37. Heat-activated heat pump system.29% 2.08% 6. by direct combustion or by multi-level cascading (Fig.2 bar]. COMBINED AND INTEGRATED CYCLES Table 4. The steam turbine has an isentropic efficiency of 80% and expands steam from 80 bar/430 °C to 10 bar/203 °C. As a process simulator. with supplementary firing Calculations were made using Dutch natural gas as a fuel with a lower heating value of 38 MJ/kg.1).70 COGENERATION. Also. Performance specifications of the LM6000 gas turbine. with supplementary firing the same as case GT-B2-ST. Gas turbine with dual pressure HRSG [10/80 bar] and back-pressure steam turbine. Conventional boiler [80 bar] with back-pressure steam turbine.66% 4.88% Exhaust gas composition: O2 CO2 N2 H2O 14. 203°C]. three modifications with a duct burner were also considered: GT-BF GT-BF-ST GT-B2F-ST the same as case GT-B. with supplementary firing the same as case GT-B-ST. Gas turbine with single pressure HRSG [80 bar] and back-pressure steam turbine. Thermal efficiency of the reference conventional boiler was set to 95%.3 Configurations Modern power technology offers many possibilities to realize the cogeneration concept via different routes: from fuel as a source to heat as a sink. Of these options. Each of these delivers 40 t/h of superheated process steam [10 bar. Gas turbine with single pressure HRSG [10 bar]. a number of typical schemes were chosen for an analysis as listed below: CB CB-ST GT-B GT-B-ST GT-B2-ST Conventional boiler [10 bar]. Parameters: Electric power Exhaust flow Exhaust temperature Simple-cycle efficiency 38. two heat pump systems were considered: EHP HAHP Electrically-driven heat pump system.97% 76.3 and 4. The flow diagrams of the cases are presented in Figures 4. 4. the heat-balance package Gate Cycle was used (Enter Software.4. nor deaerator vent flows. . It was assumed neither boiler blow-down. 1995). The gas turbine simulation was based on the performance data of the General Electric LM6000 engine (Table 4. In addition to these.1.
Flow diagrams of Cases CB. CB-ST.INDUSTRIAL COGENERATION 71 Steam boiler 10 bar 203°C 90°C Fuel Deaerator Process Air Stack 105°C Case CB 80 bar 430°C Steam boiler ST 10 bar 203°C Fuel 90°C Air Stack Process Deaerator 105°C Case CB-ST Stack Fuel Air HRSG Gas turbine Deaerator 10 bar 203°C 105°C 90°C Process Case GT-B Fig. 4. . GT-B.3.
and the process demand and return temperatures: .5). Both heat pump systems were based on the same vapor compression cycle (Fig. 4.72 COGENERATION.4. Flow diagrams of Cases GT-B-ST and GT-B2-ST. COP of the cycle is a function of the Carnot factor. temperature differences in the heat acceptor and rejector. COMBINED AND INTEGRATED CYCLES Stack Fuel Air HRSG Gas turbine 80 bar 430°C Deaerator 105°C ST 10 bar 203°C 90°C Process Case GT-B-ST Stack Fuel Air HRSG Gas turbine 80 bar 430°C Deaerator 105°C ST 10 bar 203°C 90°C Process Case GT-B2-ST Fig. 4.
25 is obtained. heat-to-power ratio. 1. assuming an electric driver efficiency of 0. For the heat-activated heat pump: efficiency of the driver or the absorption unit. .4. 4. Then. TRETURN − ∆TACC COPHP = 1 + f CARNOT ⋅ (TD + ∆TREJ ) − (TRETURN − ∆TACC ) (4.95. and utilization of its waste heat. and includes the utilization of waste heat. For the heat-activated heat pump. the overall COP of the electric heat pump system at these levels becomes 0. This analysis was also applied to a case of doubling steam production by means of supplementary firing (HRSG options). 4.4. fuel consumption and generated power were linearly scaled to fulfill the steam demand.4 Performance analysis The analysis was based on a constant steam demand of 40 t/h.5 and 0. Schematic and T-s diagram of the vapor compression heat pump. thermal and electric efficiencies based on exergy values.5 and 0. the following factors were taken into account. and fuel consumption (in thermal and exergetic megawatts). respectively. A summary of results is given in Table 4. When a simulation model had a different steam production.5 is obtained for the heat pump cycle operating within the process temperature range (90 °C and 203 °C).13) By taking 0. The latter two parameters were calculated for three levels of utility efficiency (0. Therefore. a system COP of 1. a COP of 2. the stack temperature. and 1. In this case.95. 1994) and ∆T of 10 K. A similar COP value is representable for the absorption heat pump system. the flow values were also scaled down to the constant steam supply of 40 t/h.2. Regarding system COPs.6 as the Carnot factor. 0. 0. which adds to system COP.6). For the electric heat pump: electric efficiency of the separate power production (utility efficiency) and efficiency of the electric driver.INDUSTRIAL COGENERATION 73 3 Condenser T 2 TDEMAND 3 2 ∆TREJ Throttle valve TRETURN Compressor 4 ∆TACC 4 1 Evaporator 1 s Fig. all options were compared with the conventional boiler as a reference. The table lists the following parameters: heat and power production. which is typical for such a cycle (Walker et al.6. and utility efficiency levels of 0. which is driven by a prime mover (such as a gas engine) at an efficiency of 0.3..19. thermal and electric efficiencies.42. the fuel energy saving ratio and the relative fuel consumption factor.5.
026 12.2.134 0.026 42.110 .5.6).421 0. Main results of the simulations.002 0.510 0.119 0.360 9.105 0.026 21.188 54.546 0.171 144 0.602 0.104 0.127 6.327 0.808 116 0.450 47.795 0.500 64.360 9.229 0.976 GT-B2-ST 27.188* 1.867 0. CB 27.026 28.026 23.943 74 COGENERATION.127 0.191 0.858 0.905* 1.070 0.026 23.874 0.026 4.143 66.524 0.073 0.ex °C 0.959 0.820 0.269 34.285 0.190 0.111 75.975 67.959 0.303 0.043 0.164 0.909 0.946 0.420 0.6 Notes: *) Based on a value of utility efficiency ηPSEP of 0.812 0.303 0.001 0.250 1.250 0.444 208 0.905 0.580 148 0.030* 23. .4.684 1.681 0.234 0.828 0.819 0.360 9.857 0.4 0.469 0.026 24.045 0.720 1.026 17.427 0.972 0.888 22.379 1.934 0.804 0.360 9.943 GT-B2F-ST 27.026 30.397 0. 0.876 0.476 0.839 0.283 0. COMBINED AND INTEGRATED CYCLES Heat Power Fuel Stack ηQ ηP λCG FUE ηQEX ηPEX ηEX FESR at ηPSEP: Φ (λD=5) at ηPSEP: MW. Values of the fuel energy saving ratio FESR and the relative fuel consumption factor Φ are given for three levels of utility efficiency (0.909 1.5 CB-ST 27.070 0.953 0.405 0.421 0.360 9.918 0.4 at ηPSEP: 0.841 0.257 0.699 0.th MW.958 GT-BF 27.280 64.100 45.783 0.360 9.5 and 0.4 0.717 29.th MW.0.188* 0.141 0.847 0.889 155 0.348 0.863 0. 4.795 0.729 0.014 GT-BF-ST 27.167 0.5) 0.500 0.822 0.482 0.700 52.690 0.397 0.890 0.700 101.el MW.491 0.405 0.669 0.186 0.086 0.122 0.142 0.971 GT-B-ST 27.448 235 0.964 0.5 0.269 0.914 0.923 0.136 0.ex MW.524 0.308 0.225 33.641 0.256 0.953 0.384 0.448 0.868 0.951 0.533 121 0.017 0.378* 1.360 9.360 9.922 0.921 HAHP 27.266 2.378* 0.360 9.086 0.588 105.153 0.362 1.902 0.374 0.201 0.261 0.360 9.864 0.250 0.750 73.979 GT-B 27.493 0.404 0.6 0.261 0.339 1.569 0.177 141 0.5 0.Table.6 Φ (λD=0.949 0.365 0.170 0.805 0.950 EHP 27.748 1.
the case should compete with utility in electric efficiency. 4. In Fig.2 (Fig. The options with the highest FESR are the gas turbine-based plants with a double-pressure HRSG and a steam turbine. the fired option shows a higher ratio of fuel savings. It can be seen from the table that case HAHP uses the least. but a value of 0.5.6). At a utility efficiency of 0.42 is not sufficient for this condition. and case GT-B-ST the most fuel to fulfill the heat demand. 4.2). At the same time. case GT-B-ST.INDUSTRIAL COGENERATION 75 Fig.6 (Table 4.4. The options based on the conventional boiler and heat pumps show the highest fuel utilization efficiency. caused by exclusively generating high-pressure steam in HRSG for the steam turbine.1 and 0. its FESR value becomes negative at ηPSEP = 0. and affects these cases most. When the cases are examined under different levels of utility efficiency ηPSEP. and sup- . Regarding FESR. 4.64. Having the lowest heat-to-pressure ratio. case GT-B-ST generates the most power with λCG = 0. For case CB-ST with the highest λCG. Fuel energy saving ratio of the cogeneration options. In the configuration of a gas turbine with a single-pressure HRSG and a steam turbine. no great differences were observed between unfired and firing cases. appeared to have the lowest values. GT-B2-ST and its modification with supplementary firing GT-B2F-ST. but when evaluated in exergetic terms. some sensitivity of the options with a high heat-to-power ratio is revealed. It is especially apparent in case GT-B-ST: while being quite beneficial at ηPSEP = 0.6. the improvement in utility efficiency is much less noticeable.6 the line of zero FESR value becomes steeper. all cogeneration options obtain FESR values between 0.
45 0.4 0.3 0.40 CB-ST GT-B GT-B-ST GT-B2-ST EHP HAHP 0.4 0.6 0. (b) and (c) λD = 0.35 0. ηPSEP GT-B2-ST EHP HAHP CB-ST GT-B a) 1.65 Utility efficiency.2 -0.50 0.2 0. Φ 1.4 0.2 Relative fuel consumption factor.80 0. Φ 1.45 0. Φ 1. 4.7.45 0.8 0.8 0.5 0. ηPSEP c) Fig.0 0.2 Relative fuel consumption factor.55 0.5.35 0.0 0.4 0.55 0.00 0. . Influence of utility efficiency on the relative fuel consumption for the demand heat-to-power ratio: (a) λD = 5.65 Utility efficiency.90 0.05 Relative fuel consumption factor.6 0.60 0.5 P SEP CB-ST GT-B GT-B-ST GT-B2-ST EHP HAHP 0.3 0.95 0.65 Utility efficiency.2 0.6 0.76 COGENERATION. η b) 1.4 0.85 0.55 0.0 -0.75 0.0 0. COMBINED AND INTEGRATED CYCLES 1.6 GT-B-ST 0.
and at ηPSEP = 0.1% Return 1. Usually. the heat-to-power ratio λD is about 4 to 5 (Wiemer.0% Steam boiler 58.4% Fig.8. At a typical value of 5 (Fig.7% Fuel 100% Steam turbine 3. 4. Their performance is so good that they consume only 20 to 40% fuel of the reference scheme. the .2% 42. a clear advantage of gas turbine-based plants are noted.9% Fuel 100% Heat 32.INDUSTRIAL COGENERATION Steam boiler 77 68.7a).7% Heat 27.7 for two conditions. plementary firing has some positive effect in reducing the stack temperature from 235 to 208 °C (case GT-BF-ST). The influence of utility efficiency on the relative fuel consumption is presented in Figure 4. 4.5% Power 12.35. especially at lower utility efficiency values. Grassmann diagrams of cases CB and CB-ST.0% Return 1. 1997). The first one is the situation when only local process needs are satisfied.
consumption factor even becomes negative. and it remains steady at 82-84% of the fuel consumption of the reference.8% 22. Grassmann diagrams of cases GT-B and GT-B-ST.0% Heat 14.4% Steam turbine 14. and above 0.4% Gas turbine 41.9.0% Heat 8.0% Return 0.5% Power 36. The heat-activated heat pump has a value of factor Φ that is independent from utility efficiency.5% Fuel 100% HRSG 8.0% Power 4.78 COGENERATION.4. .0% 22. because excessive power production in these cases is performed at electric efficiencies higher than that of the utility. The conventional boiler with a steam turbine stands out above the reference scheme. COMBINED AND INTEGRATED CYCLES Gas turbine 41.6 EHP turns out to be the most efficient of the other options. but its advantage is rather minor: it consumes just 86% to 93% of the reference.4% Return 0.5% Fuel 100% HRSG 8. But this is not the case for the electric heat pump: it becomes advantageous only when ηPSEP exceeds a value of 0.5% Power 36.4% Fig. 4.
intersecting other options' lines at ηPSEP of about 0. i. 0.3% 21.INDUSTRIAL COGENERATION Gas turbine 32.9% Power 6.4% 8. 4. Similar trends are observed in Figures 4. 4.7c). Grassmann diagrams of cases GT-BF-ST and GT-B2-ST. for example. . 4.2% Heat 13. case GT-B-ST loses its advantage at a utility efficiency of 0.8% 1.55-0.553.2% 30. As has already been noted before. Still.7c.0% Power 4.0% Return 0.10.5% Fuel 100% HRSG Steam turbine 4.7a and 4.e. as an enlarged view shows (Fig. the CHP plants' advantage is no longer so profound.9% 14. the gas turbine plants offer better performance (80% to 90% of the reference's fuel consumption).8% Return 0.2% 8.6.7b shows a small difference among them and a comparable value of the relative fuel consumption.5. When more power than heat is required.8% Steam turbine Duct burner HRSG Gas turbine 41.7% 17. A look at Fig.9% Fuel 21.5% Power 36.8% Heat 3.4% Fig. λD is.5% Fuel 78.5% 79 Power 28.0% Heat 8.5% 22.
188 1. 4.80 COGENERATION. It confirms the evident advantage of the gas turbine plant in general. whereas option GT-B-ST generates steam of a higher exergy value. the higher exergetic quality of this steam makes it comparable with the low-pressure steam flow in GT-B configuration. COMBINED AND INTEGRATED CYCLES 1. HRSG.493 0. Exergy flows through the power plants are presented in Grassmann diagrams (Figures 4.4 0. which has an efficiency of 45%.482 0. That is explained by the lower steam production of case GT-B-ST.6 0.863 1.384 0.2 1 0. a duct burner. The final chart (Fig. and a steam turbine.795 0.524 0.4 Fuel utilization efficiency 1.5 Exergy analysis Exergy analysis. The following assumptions were made: (1) only chemical exergy was used for fuel.11) presents a comparison of overall exergy efficiency versus fuel utilization efficiency (the First Law efficiency). 4. the losses vary from 57% to 67%. Although the mass flow of high-pressure steam in case GT-B-ST accounts for 63% of the flow in case GT-B. Comparison of fuel utilization efficiency and exergy efficiency for the options considered 4. (2) only physical exergy was used for flue gas and water/steam flows. In a gas turbine the losses amount to 41.10). applied to a cogeneration plant as a whole.500 Exergetic efficiency 0. of the unfired double-pressure HRSG scheme.805 0.783 0. in particular. The efficiencies of the gas turbine alternatives lie in the range of 50% with two extremes: option GT-BF (single-pressure HRSG with supplementary firing and without steam turbine). was also used to analyze performance of separate components of a power plant.949 0.397 0. an equal exergy ratio for HRSG is noted in options GT-B and GT-B-ST.250 Fig. The plant was divided into black boxes. Regarding the performance of HRSG.953 0.303 0.11.448 0.8 to 4.491 0. and. Substantial exergy losses can be seen to occur in the conventional boiler. a gas turbine.4%. Introducing supplementary firing deteriorates exergy efficiency in all gas turbine plants.868 0.690 0.5%.8 0. and option GT-B2ST (double-pressure unfired HRSG and steam turbine) with an efficiency of 52.2 0 GT-BF CB-ST GT-B-ST GT-BF-ST GT-B2-ST GT-B2F-ST HAHP CB GT-B EHP 0. . such as: a boiler.378 0. Depending on the steam pressure.
The electric heat pump system is directly dependent on utility efficiency and can outperform the other options only when the utility reaches an efficiency value of 0.. consuming about 83% fuel of the reference. Personal communication. 4.1. This calls for the same programme of repowering low-efficiency cogeneration plants. . and only the exergy criterion can demonstrate the imperfections of these configurations. Donders. Anonymous.INDUSTRIAL COGENERATION 81 The heat pump systems proved to have a performance comparable to a simple cogeneration scheme using a conventional boiler with a steam turbine. 1995. 1997. California. Supplementary firing at gas turbine-based plants was found not attractive in exergy terms: the unfired schemes prove to be better in all cases. while at λD = 5 the relative consumption increases to 95%. November. 2. Their exergetic efficiencies lie between 38% and 40%. Akzo Nobel Chemicals BV. February. "European Commission Promotes CHP".6 Conclusions Different methodologies demonstrated the apparent benefits of the gas turbine plant featuring the dual-pressure unfired HRSG concept in comparison with the other options considered. References 1. The high FUE value of the conventional boiler or heat pump system reflects only the quantitative. A conventional boiler-based CHP installation with a high heat-to-power ratio is much less sensitive to an improvement of utility efficiency. as was done in the 1970–1980s for direct-fired utility plants. These figures would be higher at a smaller temperature difference between the process and return flows. 1998. A low heat-to-power ratio of a cogeneration plant necessitates its operation at high electric efficiency in order to provide fuel savings as utility efficiency grows. Modern Power Systems. Enter Software. This scheme shows the best fuel saving ratio. However. energy efficiency decreases. It was shown that the fuel utilization efficiency based on the First Law failed as a means for comparing different CHP configurations. and has one of the lowest relative fuel consumption factor at different levels of utility efficiency together with the best exergetic performance value. Version 4. 3. 5. Menlo Park. and not the qualitative side of the process. A remarkable difference can be seen between energy and exergy efficiencies. in terms of technology development only cogeneration plants with exergetic efficiency close to that of utility should be used. The heat-activated heat pump demonstrates a better performance than the reference scheme. Y. This suggests the general use of the exergy criterion for assessing cogeneration plants. and even an opposite trend for most of the options should be noted: while exergy efficiency rises.6.5. when the demand heat-to-power ratio is 0. Gate Cycle User's Guide.
4. Habib, M.A., "First- and Second-Law Analysis of Steam-Turbine Cogeneration Systems", ASME Journal of Engineering for Gas Turbines and Power, Vol. 116, pp.15-19, 1994. 5. Holwerda, B., "Overcapaciteit. Who Cares?", Energie- en Milieuspectrum, September, 1997, (in Dutch). 6. Horlock, J.H., Cogeneration - Combined Heat and Power (CHP), Krieger Publishing Company, Malabar, Florida, 1997. 7. Huang, F.F., "Performance Evaluation of Selected Combustion Gas Turbine Cogeneration Systems Based on First- and Second-Law Analysis", ASME Journal of Engineering for Gas Turbines and Power, Vol. 112, pp. 117-121, 1990. 8. Huang, F.F., "Performance Assessment Parameters of a Cogeneration System", Proceedings of the ECOS '96 Symposium, Stockholm, Sweden, June 25-27, 1996. 9. Korobitsyn, M.A. and Hirs, G.G., "Analysis of Cogeneration Alternatives", ASME Paper 95-CTP-11, 1995. 10. Kotas, T.J., The Exergy Method of Thermal Plant Analysis, Butterworths, London, 1985. 11. Pak, P.S. and Suzuki, Y., "Exergetic Evaluation of Methods for Improving Power Generation Efficiency of a Gas Turbine Cogeneration System", International Journal of Energy Research, Vol. 21, pp. 737-747, 1997. 12. Sarabchi, K., "Parametric Analysis of Gas Turbine Cogeneration Plant from Firstand Second-Law Viewpoints", IGTI-Vol.7, pp. 485-491, ASME Cogen-Turbo, 1992. 13. Sarabchi, K. and Polley, G.T., "Characteristics Charts for Preliminary Design and Selection of a Gas Turbine Cogeneration Plant", ASME Paper 95-GT-156, 1995. 14. Timmermans, A.R.J.,"Combined Cycles and their Possibilities", Lecture Series on Power Generation, von Karman Institute for Fluid Dynamics, Rhode-Saint-Genèse, Belgium, 1978. 15. Walker, G., Reader, G., Fauvel, O.R., and Bingham, E.R., The Stirling Alternative. Power Systems, Refrigerants and Heat Pumps, Gordon and Breach Science Publishers, Yverdon, Switzerland, 1994. 16. Wiemer, W., "Liever Warmte Pompen dan Koppelen aan Kracht (Heat Pumps versus Cogeneration)" Energie- en Milieuspectrum, September, 1997 (in Dutch).
Chapter 5
Externally-Fired Combined Cycle
The most advanced systems for the generation of electricity from coal are the Integrated coal Gasification Combined Cycle (IGCC) and the Pressurized Fluidized Bed Combustion (PFBC) system. They both represent state-of-the-art technology and provide electric power at 41-46% LHV efficiency. However, the high level of complexity leads to lower reliability and higher equipment and maintenance costs. A simpler system, based on a coal-fired steam boiler integrated with a gas turbine, has been proposed. In such a plant, known as the externally fired combined cycle (EFCC), compressed air is heated indirectly in a heat exchanger located in the coal furnace. The hot air is then expanded in a turbine, and the exhaust is passed on to the coal furnace as preheated combustion air. Steam produced in the boiler generates additional power in a steam turbine. The exclusion of coal combustion products from the gas turbine avoids the expense of hot gas clean-up and corrosion of turbine blades by coal ash. In order to obtain a high level of efficiency the air entering the gas turbine should have a temperature close to the turbine inlet temperature, which can be as high as 1300° C for modern turbines. Such temperatures can be accomplished by the use of ceramic heat exchangers, or alternatively, by supplementary firing of a premium fuel, such as natural gas, in the air flow. To analyze the potential of an externally-fired gas turbine system with supplementary firing, a computer model was set up. Using realistic assumptions, the optimal parameters were found, and the influence of supplementary firing was calculated. It was shown that for an EFCC, supplementary firing provides a transition solution. To compete with other power plants in the future, a high-temperature ceramic heat exchanger would be required. Finally, a basic exergy analysis was carried out to assess the EFCC concept and to determine exergy losses within the systems under consideration. The overall exergy efficiency was found to be comparable with those of the IGCC or PFBC systems.
Stack Stack Coal HRSG ST Furnace ST Existing boiler
Natural gas
Natural gas
Fig. 5.1. Schematic diagram of an externally-fired combined cycle plant: (a) with a specially-designed air heater; (b) in repowering configuration.
5.1 Introduction
Coal represents about 70% of the world's proven fossil fuel resources and provides 31% of the world's electricity (Ramezan et al., 1996). Electric power from coal is predominantly generated in direct coal-fired power plants, where hot products of coal combustion exchange heat with steam in a boiler, and the steam produces work in a steam turbine. Due to thermodynamic and metallurgic constraints, the efficiency of such plants is rather low. Modern coal-fired power plants obtain an efficiency of 3840% (LHV) operating at 250-300 bar and 530-560 °C. Higher efficiencies can be achieved by further increasing steam parameters, but even for supercritical boilers operating at pressures around 350 bar and temperatures of 700 °C efficiency is expected to be not higher than 47% (LHV). These figures contrast with the efficiency of natural gas fired combined cycle plants that at relatively moderate parameters convert 4552% of the fuel's lower heating value into electric power, while today's more advanced plants can achieve even 57% efficiency (Farmer, 1996). An intrinsic handicap of the direct-fired boiler is a large temperature difference between the combustion gases and the working medium, which is more than 1200 K for a typical plant. This work potential is better utilized in a gas turbine cycle that covers a temperature range from 1400 °C to 400 °C. Since coal as a solid fuel is not quite suitable for a gas turbine, its chemical energy can be delivered to the turbine in an indirect way. There are several technologies known today that provide an effective conversion of coal into electricity by employing the gas turbine cycle. These are the Integrated Coal Gasification Combined Cycle (IGCC), the Pressurized Fluidized Bed Combustion (PFBC), and the Externally-Fired Combined Cycle (EFCC) plants.
Just as the IGCC technology. and/or passed to a heat recovery steam generator. 1996). After expansion. 1993. This concept is also feasible for repowering of existing steam plants (Klara et al.1a).2. Externally-fired combined cycle makes use of a gas turbine as well. and it is has a rather high level of complexity and capital costs. Coal gasification proceeds in a fixed bed. . which is situated in the coal furnace (Fig. pressurized bed combustion is also characterized by its complexity and high capital costs. Today's PFBC plants operate with an efficiency of 37-38% (LHV).2. Stambler. 1996). The gas turbine is driven by compressed air which is heated in the fluidized bed. 5. (b) an externally-fired cycle In an IGCC plant coal is gasified.1b). In this case. an implementation of the repowering concept may require major reconstruction of the steam boiler. air is heated in the existing furnace (Fig. However. fluidized bed. and the next generation plants are expected to reach a level of 40-45% (Almhem and Lofe.. Combustion of coal in a fluidized bed under pressure integrated with a gas turbine is another advanced coal conversion technology. Then. cleaned-up and fed to the gas turbine. Nevertheless. as shown in Fig. most of the power is generated by the steam turbine. 5.EXTERNALLY-FIRED COMBINED CYCLE 85 T T Air heater Q a) b) Q Fig. Heat addition in the gas turbine cycle occurs indirectly in a high temperature heat exchanger. exhaust heat from the gas turbine is recovered in a steam boiler. 5. but avoids coal combustion gases entering the turbomachinery. an IGCC plant contains a number of critical components for synthesis gas cooling and clean-up. air can be partly redirected to the furnace to supply preheated air for combustion. Heat exchange between air and the flue gases occurs in the gaseous state. it does not require high pressures at high temperatures. Modern coal gasification power plants show an LHV efficiency of 42 to 45% (Maude. However. Although air has a lower specific heat than steam. which accounts for up to 80% of the plant's total output. A comparison with a Rankine cycle shows an advantage of the externally-fired cycle in heat utilization. or in an entrained flow system. 1996). instead of a specially-designed high temperature air furnace. 5. so the average temperature difference in the heat exchangers is smaller than in the case of steam. Flue gas heat utilization in: (a) a steam boiler.
3 16 90 90 100 30 482 70 85 53 30 100 60 471 34 39.5 89. Vandervort et al.7 1371 40 40 47 1285 12 40 48. 15% steam injection 2.7 COGENERATION. HAT-cycle 4.4 Huang et al. 35% natural gas 8. Wiechers9 Year 1989 1990 1991 1991 1992 1993 1994 1995 1995 1995 1996 1996 1997 Notes: 1.5 1288 15 180 579 50 84 47. atmospheric fluidized bed.1. Recycle: fraction of gas turbine outlet flow sent to the furnace.5 1350 1300 90 5 980 ε % 80 SF Gas turbine Recycle Steam turbine Plant efficiency T air p ratio η comp η turb p in T in p cond η turb η HHV η LHV °C % % % bar °C mbar % % % 10 88 88 100 46 9. in HAT-cycle with 16% water injection efficiency is 52. HAT-cycle with 20% water injection 6. SF: a hyphen means that no supplementary firing is provided.5 15 89.1 91. FT4000 gas turbine.5 1150 111 788 90 2 1260 100 800 1260 85 3 1260 90 1260 927 980 4. LM6000 gas turbine. 1% steam injection 5.8 12 88 90 100 48.8 100 239 558 50 49 14 100 140 540 50 85 52 1165 15 85. . Foster-Pegg1 Parson et al. Mathieu et al.5 90.3 860 7.6 88 49 130 520 40 46.2% 3. 24% natural gas Heat exchanger: hot side ∆T and average pressure drop are given.3 100 41 482 48. Overview of EFCC plants in literature Heat exchanger ∆T ∆p/p T air °C % °C 4.8 Consonni et al.6 Ruth7 Klara et al. 35% natural gas 7.7 80 85 100 36.2 De Ruyck et al. 35% natural gas 9.1 86/80 88 0 45 482 75 33 13.5 1000 100 4.86 Table 5.3 Huang et al.5 Seery et al. COMBINED AND INTEGRATED CYCLES Authors LaHaye et al.
. 1993). With the exception for the heat exchanger.EXTERNALLY-FIRED COMBINED CYCLE 87 Air heating occurs in the most critical component of the system. with funding from the European Union (Consonni et al. The authors. 1995). 1989. 5. 1996). 1995. in particular. all other equipment is of proven technology. cycles with turbine inlet temperatures of 1200-1300 °C obtained an LHV efficiency of 48-52%. 1991. the developer of the high-temperature ceramic heat exchanger. The group is headed by the Morgantown Energy Technology Center (LaHaye and Zabolotny. Another group. 1993).. The air heater is being designed by the Foster Wheeler and United Technologies teams (Ruth. As for the turbine inlet temperature. 1994. In the United States this concept has been developed by a group of companies. A pressure ratio of 12-15 was chosen in the most studies. using biomass fuel (Eidensten et al. under the management of the Pittsburgh Energy Technology Center.. 1995). De Ruyck et al. a summary of which is given in Table 5. . A ceramic heat exchanger made from silicon carbide composites is capable to operate at even higher levels.. work on EFCC is being carried out in Italy by a consortium of Ansaldo. Vandervort et al. while the cycle pressure ratio does not considerably influence the efficiency. To improve the corrosion resistance of silicon carbide ceramic oxide coatings can be used (Van Roode. Huang and Naumowicz (1994). including Hague International. Yan et al. More advanced oxide dispersion alloys withstand temperatures around 950 °C (Klara. an EFCC plant can obtain an efficiency of 40-49% (LHV). Depending on the turbine inlet temperature and other parameters. 1995). The Swedish National Board for Industrial and Technical Development NUTEK is promoting an externally-fired evaporative gas turbine cycle.. 1996a). In literature. development of the ceramic heat exchanger still requires some engineering effort: such problems as relatively large dimensions of the exchanger and the ducting. In Europe. (1996). thermal expansion of the ceramic components at start-up and stop conditions. using comparable assumptions.1. came out with similar performance figures. Nickel-based superalloys allow operation at 800-825 °C. For example. However. ENEL. etc. Seery et al. it is limited by the heat exchanger materials. The advantages of air humidification were outlined in the works by Parsons and Bechtel (1991). Temperatures up to 1075 °C were achieved in the experiments reported by Solomon et al. the high temperature heat exchanger. The Netherlands Agency for Energy and Environment Novem has included EFCC in its New Energy Conversion Technologies program for technical and economic evaluation (Korobitsyn. 1996). Parsons and Bechtel. is working with industries to develop a similar power cycle called the High Performance Power System (HIPPS). Huang and Naumowicz (1992) concluded that a high efficiency of the cycle is determined by a high turbine inlet temperature.. Seery et al. (1991). but much interest has been regained in the last decade.2 Background An externally-fired gas turbine cycle was proposed long ago as a possible realization of the Joule-Brayton cycle. and the level of 1370 °C and higher is anticipated (Vandervort. and research institutions. the cycle is described in a number of publications. 1993.
Klara et al. a part of which can be made of metal.6% 13. may be prohibitive. They were compared with other power cycles. no firing GTCC natural gas fired combined cycle plant IGCC coal gasification combined cycle plant The options with a gas turbine were based on the specifications derived from those of a Siemens V94. A number of configurations were defined to estimate the performance of EFCC for different types of the air heater: (1) a metallic one made from nickel-based alloys operating at 800°C. (1996). 1996). This possibility is discussed by Ruth (1995). the amount of fuel consumed in the supplementary burner is determined by the temperature level in the air heater. and its condensing pressure was set at 0.5 or several times higher than that of a metallic heat exchanger (Consonni and Macchi.2 engine: Compressor isentropic efficiency Turbine isentropic efficiency Pressure ratio Turbine inlet temperature Combustion chamber pressure drop Combustor efficiency 88. The heat exchanger featured 90% effectiveness and a pressure drop of 5%. air heater temperature of 800°C. The configurations were denoted as: PC pulverized coal fired power plant EFCC-800 EFCC. (2) a metallic one made from oxide dispersion alloys operating at 980°C. and others.0% 85. which is 2. slag behaviour on ceramic surface and others. supplementary firing to 1165°C EFCC-980 air heater temperature of 980°C.3 Plant configurations At a given turbine inlet temperature. COMBINED AND INTEGRATED CYCLES connections and seals.88 COGENERATION. and elaborated in this chapter. The current price per square meter of ceramic surface. The use of supplementary firing in order to raise the air temperature to the gas turbine inlet conditions can be considered as an interim solution.05 bar.7 1165 °C 0. should be taken into account.27 bar 99% The steam section of all configurations consisted of a single-pressure boiler that provided superheated steam of 550 °C and 140 bar to a condensing steam turbine. and (3) a ceramic (silicium carbide) heater at 1165°C. supplementary firing to 1165°C EFCC-1165 air heater temperature of 1165°C. The turbine had an isentropic efficiency of 88%. Supplementary firing was provided by burning methane (the lower heating value of 50 . 5.
2. a weighted 2. % Methane fraction.75 315 270 45 39.0 34.1 100.EXTERNALLY-FIRED COMBINED CYCLE 89 Table 5. meaning about 30% lower fuel consumption (Table 5. Depending on the coal fraction in the total fuel input.58 425 281 143 GTCC 33.2 100.air heater 100.0 50.2 61. Summary of results PC EF-800 EF-980 EF-1165 ST fraction.8 0.0 0.9 60.0 0.70 357 276 80 39.82 MJ/kg).2). the flue gas Efficiency.15 252 252 0 IGCC 49. external firing options cannot reach the level of a natural gas fired combined cycle due to a higher internal consumption rate for coal preparation.boiler .0 0.9 48. .5% penalty was subtracted from the gross plant efficiency due to the flue gas desulphurization (FGD).3 60. the Shell technology based on an entrained flow gasifier was chosen.0 0. Power plants efficiency versus coal fraction in the total fuel input.8 51.3.6 46. % Surface. % Fig. For the coal gasification plant.7 75.8 50. The gasifier operated at 1500 °C and 30 bar. % GT fraction. 5.0 45.4 Discussion Simulation results show a considerable performance gain of EFCC plants over the pulverized coal plant. % LHV efficiency. % Coal fraction. and the pressure losses across the gasifier of 5 bar were assumed.2 66. Part of the gas turbine outlet flow that is passed to the coal furnace was controlled so that the flue gas temperature at the steam boiler inlet remained at 630 °C.0 100.1 47.0 100.0 45. % LHV 60 60 Advanced power plants GTCC 50 EFCC-800 50 IGCC EFCC-1165 40 40 PC EFCC-980 30 0 20 40 60 80 30 100 Coal fraction. 5. m2/kW .84 223 223 0 38.4 24. At the same time.
Such judgement can be done by using a mean efficiency relative to a coal fraction in the total fuel input. a lower flow across the expander. An economic analysis should evalu- Specific heat exchange area. This means that the effect of supplementary firing in EFCC plants is rather limited in a long term. When compared with advanced plants.5 Steam boiler area 300 250 1 200 150 0.90 COGENERATION. COMBINED AND INTEGRATED CYCLES 450 2 Specific gas consumption.3. a specific natural gas consumption in kW(thermal) per kW(electric) is plotted. For example. Regarding the surface of heat exchangers.4. desulphurization. Another reason is a higher FGD penalty levied on efficiency by the increased coal consumption. schemes can be evaluated whether the use of natural gas in the EFCC is justified in comparison with two separate plants. externally-fired options have an advantage above the mean efficiency line between PC and GTCC. Bearing in mind that a EFCC plant's ultimate goal is to use 100% coal. it constitutes 34% of the total surface versus 14% in the case with maximum supplementary firing (EFCC-800). while the area of the air heater varies considerably (Fig.4). and in case EFCC-800 even goes marginal. This is caused by the reduction of the natural gas consumption. It ranges from 5. Nevertheless.5 100 50 0 GTCC EFCC-800 EFCC-980 EFCC-1165 PC 0 Fig. the advantage becomes much smaller.6% points in case EFCC-800 to 10. a natural gas-fired combined cycle and a direct-fired plant. In the 'coal-only' EFCC. losses in the air heater. As shown in Fig. As more coal is used in the externally-fired cycle. a small drop in efficiency can be noted. kW/kW Gas consumption Air heater area 350 1. etc. m2/kW 400 .7% points in case EFCC-1165. 5.5 kW of natural gas consumed by EFCC-980 is equivalent to 63 square meters of ceramic surface in case EFCC-1165. 0. 5. such as IGCC. the boiler surface area was found of comparable size for all plants. Specific methane consumption and specific surface area of the steam boiler and the air heater per 1 kW of the net power output. depicted by a dashed line on the chart. On the same chart. It allows a comparison between the gas consumption and the surface area. and only the development of a high-temperature heat exchanger for a 'coal-only' EFCC can provide its competitiveness with respect to other power cycles. EFCC efficiencies were found comparable with that of the IGCC plant. and therefore. 5.
and 0.5% 4. the fraction of the turbine outlet flow sent to the furnace was set to 0. case EFCC-980 was compared with an IGCC plant.3% and 41. 35.5 Exergy analysis To assess systems exergetically. The difference is due to a different air heater's duty.335. As expected.8% Power 25.5%.6% 25. They are the greatest losses within the systems. and EFCC-1165. the largest losses were found in the combustion process.3% Power 19.1% 2. This figure agrees with those reported by Chen (1997) and Eidensten and Yan (1997) for similar systems. EFCC-960.66 for EFCC-800.5 and 5. 5.5.9%. but a different distribution of exergy losses.EXTERNALLY-FIRED COMBINED CYCLE 91 13.7% 14. 0.1% 1. The analysis showed an equivalent efficiency of the plants.5% 19.1% Natural gas 21. The results of the analysis are presented in the Grassmann diagrams (Figures 5. The total losses in the coal furnace and the gas turbine for case EFCC-980 amount to 38. To maintain the boiler inlet temperature at 630 °C in all EFCC configurations.7% 8. 5.6% 4.2% Gas clean-up Stack 3. and can be diminished by . Grassmann diagram of an EFCC power plant. as soon as the price of alloys and ceramic surfaces will be available. respectively.6). This value is also comparable with the losses in the gasifier and the gas turbine of the IGCC plant that make up 36.4% 44. assuming that they both are based on commercially available hardware. respectively.6% 13. ate an economic optimum.6%. An air heater with a larger duty requires more preheated air coming from the gas turbine outlet.494.5% Coal 78.5% Fig.
8% 2.92 COGENERATION.7% Power 22.6 Conclusions Externally-fired combined cycle plant.3% 3.3% 2.6% Fig. Possible problems caused by the use of ceramics include limitation on the maximum pressure. which surface area can be half of that of the steam boiler. Other measures. since it can be better utilized in high efficient gas-fired combined-cycle plants. which is rather limited in a 'coal-only' option. COMBINED AND INTEGRATED CYCLES 2. slag deposition on the ceramic surface. In a long term. such as recuperation or exhaust gas recycling.8% Coal 100% 2. At the same time.8% Power 22. based on currently available hardware and can operate at efficiencies similar to those of advanced coal plants such as IGCC.6 Grassmann diagram of an IGCC power plant. the use of natural gas for firing would not be justified. A high-temperature clean-up system will reduce losses related to removal of aggressive compounds and particles. supplementary firing allows a flexible control over the power output. manufacturing of ceramic connections and seals. 5.0% 3. however.6% 7. The 'coal-only' plant necessitates the use of a costly ceramic-based heat exchanger. Supplementary firing in the externally-fired combined cycle power plants provides a short-term solution. may also be considered. 5. The efficiency of a plant with supplementary firing is remarkably higher than that of a pulverized coal plant. choosing the most efficient equipment for these key components.0% 10. as materials for the high temperature heat exchanger are being developed.7% Stack 3. except for the air heater. Exergetic efficiency of heat recovery can be improved by choosing more pressure levels in the HRSG. .4% 1.0% 25.4% 2.3% ASU 0.
Proceedings of the ECOS'92 Symposium. 443-452. F. R. Power-Gen International '96 Conference. Eidenstein. Klara. Consonni. "Second Law Analysis of Externally Fired Combined Cycle for Cogeneration".. S. and Yan. and Naumowicz. Farmer.D. "Biomass Externally Fired Gas Turbine Cogeneration". Using the results of exergy analysis. Yan. 2. 7. pp.. 12.G. 1996.. Klara. and Lofe. 1996.J. Part A: Thermodynamics and Technological Issues" . some improvements can be considered in order to reduce irreversibilities within a system. 3. 1990. VanBibber. Zaragoza. IGTI-Vol.S.. 5.. In exergetic terms. J. and Farina. 1991. E. 1996. Part B: Alternative Configurations and Cost Projections". 10. Proceedings of the International Symposium on Thermodynamic Analysis and Improvement of Energy Systems TAIES '97. P. 112. 6. 1996. J. Foster-Pegg. Almhem.. and Svedberg. Final Report ECN-CX-96-060.. Florida. Spain. J. M. G. 8. "A Small Air Turbine Power Plant Fired With Coal in an Atmospheric Fluid Bed".A. Maniatis. EFCCs were proved to be comparable with IGCC plants. . Korobitsyn.EXTERNALLY-FIRED COMBINED CYCLE 93 An economic analysis is needed to find an optimum between an expensive heat exchanger surface and the consumption of a premium fuel for additional firing.W.R. J. ASME Paper 94-GT-345.. Netherlands Energy Research Foundation ECN.. 13. "Overall Performance Evaluation of an Externally-Fired Air-Turbine Combined-Cycle Power Plant Based on First-Law as well as Second-Law Analysis". 6. 1996..F. E. S..R.. G. ASME Paper 96-GT-93. LaHaye. Baron. 1994.. "Advanced Power Generation: The Potential of Indirectly-Fired Combined Cycles". L. 11. ASME Paper 95-GT-261. "Externally-Fired Combined Cycles (EFCC). China. T. ASME Cogen-Turbo. "Portugal IPP 1000-MW 'BOT' for PowerGen Operational in 1998". 9. E... K. Izsak. March-April. P. 1997. and Pottie. Power-Gen International '96 Conference.. "Externally-Fired Combined Cycle (EFCC)".H. "Externally-Fired Combined Cycles (EFCC). J. December 4-6. J. References 1. 1992....M. and Figueroa J. J. Huang. Wherley. 4. June 10-13. K. 1995.M. F. December 4-6. De Ruyck. M. "Review on New and Improved Combinations of Existing Energy Conversion Technologies". Journal of Engineering for Gas Turbines and Power.. Gas Turbine World. "PFBC Power Plants: A Competitive Alternative". Macchi. Petten. L. 21-27... R. "A Biomass-Fueled Cogeneration Plant Based on an Evaporative Gas Turbine Cycle at the University of Brussels". M. "HIPPS Repowering: A Strategy for Life Extension With Increased Efficiency and Reduced Emissions". Orlando. and Zabolotny. 1996. and Macchi. pp. L. and Wherley. Consonni.. Eidensten.. June 15-18. Orlando. Ward. M.. Vol. Beijing. ASME Paper 96GT-92. Florida. The Netherlands.
1996. 1989.. "Engineering Development of a Coal-Fired High Performance Power Generating System". Journal of Engineering for Gas Turbines and Power. 24. "Low Emission Boiler System: Coal-Fired Power for the 21st Century". J. M. Eidensten. .. 25. IGTI Vol. 23. Price. "The U.J. M. December 4-6. 1996. I. Gas Turbine World. G. International Symposium on Advanced Energy Conversion Systems and Related Technologies RAN'95. S. Yan. The Netherlands. J. Hungary. Wiechers. Van Roode.. London. Budapest. and Bechtel... D. "Exergy Analysis of Coal Conversion Processes". Cosgrove.R.S. pp.. 14. J. Nagoya. "An Investigation of the Heat Recovery System in Externally Fired Evaporative Gas Turbines". Indirect-Fired. Stambler. G. Japan... 15.... Van Bibber. M. 1997. 20. and Stala. Orlando. International Symposium on Advanced Energy Conversion Systems and Related Technologies RAN'95. S. IEA Coal Research.. Parsons.. April 25-28. S. PowerGen Europe'96 Conference. 1996.A. 22. Japan. Pines.. Ramezan...R. Advanced Power Generation . and Kim. C. Zhao. Coal-Fueled Gas Turbines". "Progress in IGCC and Advanced Cycles Outlined at EPRI Meeting". Maude. 139-147.A Comparative Study of Design Options for Coal..J. 115... ASME Paper 93-GT359. 1993.. 16.J. "Externally-Fired Combined Cycle: Repowering of Existing Steam Plants". and Higgins. and Bornstein.L. J. "Ceramic Oxide Coatings for the Corrosion Protection of Silicon Carbide". Florida.F. 1996. Power-Gen International '96 Conference..A. 18. Stoddard L. 17. Buggeln. 2231. P.J. C.. Ruth. N. and Desoleil. Shamroth. Solomon.. 1993. January-February.. December 4-6. Vol.T. Efficient Electricity from Coal". 1993.. L. 1995.R. Department of Energy's Combustion 2000 Program: Clean.L. L. Seery. T. 118.. L.. pp. J. 4. Sangiovanni. Delft. 1995. ASME Paper 91-GT288. and Svedberg. 1991. J. Holowczak. Bary.. Vol.. MSc Thesis. D.E.. Journal of Engineering for Gas Turbines and Power.. Nagoya.E. White. 21. E. Delft University of Technology. M. COMBINED AND INTEGRATED CYCLES Proceedings of the ASME Cogen Turbo Symposium. December 4-6. "A Coal-Fired Heat Exchanger for an Externally Fired Gas Turbine". 19. ASME Paper 95-GT-72.C. Serio.S. R. P.94 COGENERATION. Y. Mathieu. "Off-Design Performance of EFCC Plants". "Performance Gains Derived from Water Injection in Regenerative. G. Vandervort.
The hot windbox configuration was found the most effective solution. It also can manage disposal of various wastes. The results showed that the integrated options can effect a substantial increase in efficiency. An exergy analysis of the cases was carried out. Luxembourg and Denmark recover energy from over 70% of their waste . so that steam of a higher energy content becomes available for electricity production. 6. semi-solids. Performance of these configurations together with two modified schemes was analyzed in terms of efficiency. This can be accomplished in a number of ways. the exhaust from a gas turbine plant is sent through a superheater section to the waste incinerator's boiler providing preheated combustion air. or for process needs. steam generated at a waste-to-energy plant is superheated in a combined-cycle plant that operates in parallel. In one case. but also produce an appreciable amount of heat which can be recovered in a steam boiler. incineration remains the most effective volume reducing technology. natural gas consumption and boiler surface area. sludge. and gases. or recycling. In the other case. An increase in steam temperature can be achieved by external superheating in a heat recovery steam generator positioned behind a gas turbine. The postcombustion systems of an incineration plant control undesirable airborne emissions. composting. liquid wastes. such as combustible solids. In this way Switzerland. The aggressive nature of the flue gases does not allow the temperature of steam in the boiler to rise above 400 °C. offering a smaller boiler surface area along with a moderate rate of natural gas consumption. In addition.Chapter 6 Integration of a Gas Turbine and a Municipal Solid Waste Incinerator Abstract Modern municipal waste incinerators not only considerably reduce waste volume. or process needs. the heat released from combustion can be recovered in a boiler to supply steam to a steam turbine.1 Introduction Of waste processing technologies such as landfill.
Simplified solid waste incinerator schematic (Reference case). The surface temperature of the tubes can be adjusted with the use of a de-superheater. metals 3-10%. (Wolpert. glass 3-8%. Pirson and Braker. several arrangements in the boiler design should be provided. These limits constrain the high efficiency of the MSW incineration boiler. The evaporator's tubing is to be located in the second pass. COMBINED AND INTEGRATED CYCLES Steam Secondary air Waste Water Stack Primary air Slag Fig. . while for waste of high and medium heating value. etc. Anonymous. Pirson and Braker. yard and food wastes 20-35%. plastic 6-10%. For waste with a low heating value. such as location. 1994). The composition of the municipal solid waste (MSW) varies considerably and depends on many factors. 1994.. 1994.1). RamsgaardNielsen. In order to avoid slag deposits.1. 1989). On the other hand. 6. and to maintain a sufficient residence time for after-burning. temperature of the flue gas in the stack should not drop below 200 °C due to the risk of condensation of aggressive compounds. the primary air constitutes 80% of the total air flow. origin of the waste. 6. local policy. The magnitude of the heating value affects the ratio between primary and secondary air in the incinerator. A general overview of the new thermal treatment technologies can be found in Pfeiffer and Van Egmond (1995). the superheater's tube bundles in the second or the third pass. passing through the boiler. is cooled from 1000 °C in the radiation section to 600°-800 °C at the entrance of the convection area. 1993. the percentage lies between 50% and 70%. Typical waste consists of: paper and paperboard 20-40% by mass. 1996). An incineration plant can process as much as 2500 tons of refuse per day (Tillman et al. The flue gas. The heating value of the waste varies accordingly from 6 to 14 MJ/kg. Heat recovery and power generation have become a common practice at incineration plants. 1994). and many waste-to-energy (WTE) plants have been built around the world (Isles. and inert matter up to 10% (Tillman et al. After passing the evaporator's bundles the gas flows through the superheater and economizer (Fig. wood 2-6%. erosion and corrosion in the boiler. where flue gas temperature drops to an acceptable level. 1989. textile 2-6%.96 COGENERATION.. Some improvement of the cycle can be achieved by raising the temperature of the steam externally or by preheating the combustion air. The aforementioned precautions result in a relatively low temperature of the superheated steam of around 400 °C. or both.
are described in the literature: the lower temperature steam from the incinerator can be passed either to the intermediate pressure section of a steam turbine (Ramsgaard-Nielsen. air fans can be kept in a stand-by position. in order to provide the same amount of oxygen to the boiler. a 30% larger flow from the gas turbine is required. In the latter configuration.24 4. where further superheating of steam coming from the incinerator takes place. An improvement in the total plant efficiency from 41. Mass% Yard and food waste Paper and paperboard Plastic Glass Metal Textile Combustibles Non-combustibles 30. therefore.10 0. where the heat recovery steam generator (HRSG) operates in parallel with the incineration boiler. .9% together with a 28% power increase was reported by Pijpker and Keppel (1986). consuming up to 80% of the fuel input.2 Gas turbine integration The high temperature of a gas turbine's exhaust makes it well-suited to integration with an MSW incinerator. 6.0 12.0 12. Another approach is to use of the gas turbine exhaust as the inlet flow for the incinerator's furnace. in the both integrated configurations.INTEGRATION OF GAS TURBINE AND WASTE INCINERATOR 97 Table. or be further superheated in the HRSG (Andersson. exhaust heat reduces heat demand in the boiler. 1996. This concept is often used for steam plant repowering.2% higher efficiency than the mean efficiency of two separate plants. A gas turbine trip can affect the steam boiler reliability.28 17.0 5. Considering the combustion of waste. This can be accomplished either by the use of air blowers or by expansion to a pressure above the atmospheric level. or operated in parallel with the gas turbine. When the gas turbine flow is too large for a given boiler. Plant configurations.0 3. Davidse and Roukema (1984) estimated that the gas turbine flow should be about 106% of the normal air flow for the same heat load in the boiler's furnace.83 0.3% to 45. Composition of the municipal solid waste. and less oxygen is needed. Utilization of gas turbine exhaust heat as preheated combustion air is another possible integration scheme. some precautions should be made.17 0.1. Exhaust heat can be recovered in the waste heat boiler. 1994). The gas turbine exhaust contains 14-16 vol% oxygen. the integrated plant was reported to have a 2.0 3. the gas turbine plays a dominant role.95 26.0 29.57 6. For example.96 0. Thus. 1994). compared to 21 vol% in the air.0 Molar% Carbon Hydrogen Oxygen Nitrogen Sulfur Phosphorus Chlorine Water Ash 31. Van Wijk. Such a hot windbox configuration is dependent on the gas turbine performance. the exhaust flow should have a pressure high enough to pass air through the waste layer on the grates.90 17.0 6. At the same time. However.
6. if external superheating is applied. integrated plants were compared with a conventional MSW incinerator. where it is not exposed to the corrosive gases.96 0. .3 Configurations To estimate the effect of gas turbine integration in different configurations. 6. This modification of the previous case is referred to as Case 2. The performance of the hot windbox scheme can be further improved. the gas turbine exhaust heat is partially recovered in a superheater. In Case 3. following which it is passed to the MSW incinerator (Fig. where it is superheated to 520 °C. The specifications of the reference incineration plant. this heat can be utilized in the incineration boiler as secondary air. steam temperatures higher than 400 °C can safely be achieved. while increasing the steam temperature to the same level as that in the HRSG. saturated 0. The simulation models based on the gas turbine were scaled to have the same fuel (waste) input for comparison analysis.3). the gas turbine plant based on a General Electric Frame 6 turbine.7 bar.5 bar 200 °C 0. Waste incineration boiler: HP steam LP steam Condenser pressure Deaerator pressure Minimum stack temperature Combustion efficiency ST isentropic efficiency Gas turbine: Net power output Efficiency (LHV) Exhaust temperature Exhaust flow 38 MW 0. 6. COMBINED AND INTEGRATED CYCLES Table.4 kg/s 40 bar.2.319 545 °C 139. and the heat recovery steam generator are given in Table 6. or to the convection section of the boiler. 6. or 33 tons per hour.85 HRSG: Approach temperature Pinch temperature Heat transfer coefficients: · economizer · evaporator · superheater Superheater temperature Minimum stack temperature 65 W/m2K 55 W/m2K 35 W/m2K 520 °C 80 °C 15 °C 25 °C part of the flow can be bypassed to a stack.2. The amount of waste being processed by all plants was assumed to be 230 000 tons of solid waste per year. The superheater has a simpler design and a much smaller surface area than a heat recovery boiler.98 COGENERATION. and then sent to the steam turbine (Fig. Plant specifications.1 bar 3. 400 °C 3. If the steam from the incineration boiler is passed to a superheater located in the gas turbine exhaust duct.1. The steam coming from the incineration plant is passed to the HRSG. if the lower heating value of the waste averages 10 MJ/kg. The composition of the waste is presented in Table 6. This amount of refuse is equal to a fuel input of 92 MWth. Case 1 represents a combination of a gas turbine plant with a heat recovery steam generator and a waste incinerator. Since the exhaust from the HRSG has a temperature of 80-100 °C.2).
INTEGRATION OF GAS TURBINE AND WASTE INCINERATOR 99 Waste ST Fuel Air HRSG Gas turbine Stack Air Stack Fig. ST Fuel Air Waste Stack Gas turbine Fig. 6. HRSG-Incineration boiler parallel configuration with an exhaust bypass (Case 4). The hot windbox configuration with superheating (Case 3).3.4. HRSG-Incineration boiler parallel configuration (Case 1). 6. . Waste ST Fuel Stack Air HRSG Gas turbine Fig. 6.2.
89 15.49 92.24 Case 1 Case 2 Case 3 Case 4 In this configuration a large steam flow from the incineration boiler does not allow the steam temperature in the superheater to rise above 486 °C.12 92. m2/MWe 294 710 686 340 518 24. which constitutes 12-15% of the total power generated.64 41. Pressure losses in heat exchangers and through the waste combustion grates.00 104.51 34.92 46.44 65. which consumes to 53. Here.81 28.00 100.28 29. Summary of results (scaled).4.4 Performance analysis The results of the analysis showed that integration of an MSW incinerator with a gas turbine plant leads to a significant increase in efficiency.73 196. % Power output. Internal consumption. MWth · incineration boiler · gas turbine Total MSW share. sea level).29 29.36 47. Reference Fuel input.23 92. The use of the HRSG exhaust as the secondary combustion air in Case 2 had a little effect on the total efficiency due to its rather low temperature: the difference between Case 1 and Case 2 is . The cases were simulated using GateCycle heat balance software package (Enter Software.39 47.00 92.91 24.00 28.54 41.43 92.00 101. COMBINED AND INTEGRATED CYCLES Table 6.76 53.73 46. As shown in Table 6. relative humidity 60%. % Natural gas share.29 36.100 COGENERATION. Thus.00 92.37 79.14 47. The simulations were made under ISO conditions (15 °C.91 40.00 22. The schematic of this case is given in Fig. % · based on total input · based on MSW Specific surface area. another scheme with an HRSG and an exhaust bypass is considered.88 52.96 78.57 52.9% in the reference case up to 37. 1995). was omitted as well. In all integrated configurations the steam pressure in the incineration boiler was raised from 40 bar to 80-100 bar in order to take full advantage of superheating.92 32. 6. The gross plant efficiency (based on LHV) improved from 24.00 0. a superheat of 520 °C can be achieved by controlling the amount of the flue gas that is directed to the furnace via the bypass.14 192. MWe · steam turbine · gas turbine Total Efficiency.39 193.44 140.79 33.42 46. blow-down.3 this figure is due to the gas turbine contribution.44 81.88 37. Thus.00 100.92 0.3% in Case 3 and 41.00 0.3.46 31. 6. and deaerator vent flows were not taken into account.00 48.2% of the total fuel input. only gross efficiency is given in the analysis.10 22.3% in Case 4.48 52.
The latter represents a line between two extremes: a combined cycle plant that operates on 100% natural gas and has an efficiency of 52%. The MSW fraction in the total fuel input versus gross efficiency.6% points in comparison to the reference.INTEGRATION OF GAS TURBINE AND WASTE INCINERATOR Gross efficiency. Case 4. To determine the first term the gas turbine share should be subtracted from the total power production. 6. % Fig.5). nonetheless. if the gross efficiencies of the cases are plotted against the MSW fraction in the total fuel input together with the mean efficiency of separate plants (Fig.3%. This share can be found by multiplying the gas turbine fuel input by a combined-cycle efficiency.2% points. so that the expression for the MSW-based efficiency becomes: η MSW = PTOTAL − FNG ⋅ ηCC FMSW (6.5% for Case 3.3%. however at the expense of the highest natural gas consumption rate. The waste share in this case is 65%.5.1) When 52% is taken as the combined-cycle efficiency. while maintaining a high efficiency. % Combined cycle 50 50 101 40 Case 4 Case 1 Case 3 40 30 30 0 20 40 60 80 100 MSW fraction. Excluding the gas turbine part. The effect of integration can be illustrated. has the best overall efficiency of 41.6% for Case 1 and 29. On this basis Case 3 represents a good match between the MSW fraction in the total fuel input and the plant efficiency. this means an increase of 4. waste-based efficiency can be defined as the ratio between the power generated by the MSW combustion and the MSW fuel input. and a waste incineration plant (the reference case) that processes 100% waste at . the plant efficiency is as high as 37. Case 3. the calculated MSW-based efficiency varies between 28. where an exhaust bypass is employed in order to rise the superheated steam temperature. In the best case. Since the prime objective of a waste incineration plant is to process waste. only 0. 6. the natural gas share should be limited.
Comparison of the MSW fraction in the total fuel input. and the total plant efficiency is illustrated in Fig.6% advantage.9% efficiency. the specific surface area is just 15% larger than that of the reference. pressure. Case 3 shows a 3. Case 3 appears to be the most advantageous among the options considered. This results in a lower MSW fraction in the fuel input and a larger boiler surface area.102 COGENERATION. 6. different. 6.6.3 the surface area is expressed in specific units. % Surface.2 and 2. m2/MW. of course. where the incineration plant operates in parallel with an HRSG. chemical composition) from the GateCycle models were post-processed to obtain exergy values of the flows. Considering the total boiler surface area. temperature. A comparison of the MSW fraction in the fuel input. require twice as much area as the reference plant. Case 2 is not shown due to a minor distinction from Case 1. The flow parameters (mass flow. the schemes. m2/MW 900 800 700 600 500 400 300 200 100 0 Case 4 Fig. the specific boiler area. exergy analysis was performed in order to assess the systems thermodynamically. enthalpy. exceeds the reference by 76%. Case 4. COMBINED AND INTEGRATED CYCLES Gross efficiency 100 90 80 70 60 50 40 30 20 10 0 Reference Case 1 Case 2 Case 3 Surface area 1000 MSW fraction. the gross plant efficiency (LHV).3 percentage points respectively. which has only a superheater. % LHV efficiency. all integrated schemes require a larger surface than the conventional incineration plant. at other levels of the combined-cycle plant efficiency. 6. In Case 3.6. The benefits of integration are especially noticeable in Case 3: while Case 1 and Case 4 stand out above the line by 2. In Table 6. The hybrid scheme.5 Exergy analysis In addition to the conventional enthalpy-based analysis. As is seen from the table. . 24. The chart indicates that the high efficiency of Cases 1 through 4 are obtained at the expense of the gas turbine plant. and the specific boiler surface area. The values of the integration advantage are.
6. The exergy values were calculated as a percentage of the total fuel input. Grassmann diagram of the parallel configuration (Case 1). 6.2% MSW burner MSW 47.7-6.5% HRSG Steam turbine Natural gas 52.2%.9 as Grassmann diagrams.INTEGRATION OF GAS TURBINE AND WASTE INCINERATOR MSW burner 103 Boiler MSW 100% 54.4% Stack 1.1% 19.4% Fig.7% Power 23.5% BFW 1. The analysis showed that the largest loss of 45% occurs in the burner because of the highly irreversible nature of the combustion process (Fig.16% 14. The results of exergy analysis are presented in Fig.5% Boiler 26. The loss in the burner can be diminished to some extent by preheating the combustion air and by recycling the flue gas (Harvey et al.5% 33.8. 1995). The use of the gas turbine exhaust as inlet air provides both measures. 6. As seen in Fig. . Gas turbine Power 16.8% BFW 2. this leads to an improvement of burner exergetic efficiency from a value of 54.8% BFW 3..8% to 64.1% Stack 0.9.8% Stack 3% Steam turbine 38.7).7% Power 23.7.3% Fig. 6. Grassmann diagram of the conventional waste incinerator (Reference case). 6.
owing to a large temperature difference between the combustion gases and the working medium. 6.8). 6.6 percentage points. 6. In the reference case about 13% of the total fuel exergy is destroyed in the boiler.6% Superheater 9.2 % MSW 65.3% Power 10.9.8% to 39. In addition. However. integration added 12-15% to gross plant efficiency. such as in Case 1. Among the options studied.2% Stack 2% Boiler 37. If steam is superheated in an HRSG. however.2 and 3.7% 46% BFW 3% 33. COMBINED AND INTEGRATED CYCLES Gas turbine Natural gas 34. this scheme has half the boiler surface area than that of other integrated options. this option requires a ducting. integrated schemes showed an advantage in efficiency ranging between 2. the exergy efficiency of the plant grows from 23. Another major loss was found in the boiler.104 COGENERATION.8% Fig. .6 Conclusions Integration of a gas turbine and an MSW incinerator results in a considerable rise in efficiency. and an excessive gas turbine outlet pressure should be provided to pass the exhaust gas through the waste layer on the grates. the hot windbox configuration with external superheating offers high efficiency along with lower natural gas consumption.3% (Fig. When compared with the mean efficiency of separate plants with respect to the MSW fraction in the total fuel input. so external superheating should be applied in order to increase the exergy of the steam. For the configurations studied. Grassmann diagram of the hot windbox configuration with external superheating (Case 3).3% Power 25.5% Exhaust 6% Steam turbine MSW burner η=64. a significant amount of natural gas is needed. This loss cannot be avoided within the incineration boiler. This improvement is caused by a more effective gas turbine plant. which feature a full-scale HRSG together with an incineration boiler. This is especially noticeable in the schemes where the incineration boiler operates in parallel with the HRSG.
1.N. Pasini. Ghezzi. "Application of Gas Turbine Generators For Preheating Combustion Air With Conventional Boilers".. 8. In the parallel configurations the area can be twice as large as that of the reference. 117. 13. Journal of Engineering for Gas Turbines and Power. L. H. 10. 1986.INTEGRATION OF GAS TURBINE AND WASTE INCINERATOR 105 The use of the HRSG outlet gases as the secondary combustion air in the incinerator had a negligible effect on efficiency. 4&5. Modern Power Systems. 12.. and Keppel. F. Exergy analysis showed that the substantial exergy loss in the burner can be reduced by 9. "Towards a New Technology for the Thermal Treatment of Municipal Solid Waste in the Netherlands". Vol. 5. May. though its waste fraction in the total fuel input was the lowest of all the cases. "Amsterdam to Have First 140 MWe Gas Turbine". May 16-18. Menlo Park. The 28th Intersociety Energy Conversion Engineering Conference IECEC'93. Atlanta. "Sønderborg CHP Plant Burns Gas and Refuse". R. 1996. 1993.A. S. References 1. C. G. Rossi. L.. Modern Power Systems. San Diego.. 1989. and Richter. Knoche. K...F.. S.M.. Modern Power Systems. 1994. Harvey. August 8-13. 24-30.. J..-P. 1996. Oxford. Gate Cycle User's Guide. A. D. "Profitable Waste Recovery". S. and Vick. "Power for ENW from Alkmaar Waste Incinerator". pp. Enter Software. Vol. Incineration of Municipal and Hazardous Solid Wastes. 1993.. The Netherlands. The scheme with partial exhaust diverting was found quite effective. "Waste Incineration Turns to Cogeneration".. 1995. May. November. Modern Power Systems. 1994. Pirson. U. Pijpker. Academic Press. J. Tillman.E. California. The second largest exergy loss that occurs in the incinerator boiler due to the large temperature difference is regarded as unavoidable. 1984. 4. pp. Modern Power Systems.E. "The GAVI Wijster 735 000 t/a Waste to Energy Plant". Version 4. 3.. B.. and Degli Antoni Ferri. February. Andersson.B.J. Ramsgaard-Nielsen. 891-911. Anonymous. "Reduction of Combustion Irreversibility in a Gas Turbine Power Plant Through Off-Gas Recycling". ASME Paper 84-GT-161. Davidse. J.J. A. For the windbox option the increase is only 15%. P. Waste Incineration Handbook. Georgia. 7. and Van Egmond. Pfeiffer..P. if the gas turbine exhaust is applied as pre-heated combustion air. The integrated options made use of heat recovery steam generators that lead to an increased boiler surface area.. 1995. "Analysis of Gas Cycles Utilization for Heat Recovery in Incineration Plants". The Power-Gen Europe'95 Conference.. . September. and Roukema. International Power Generation. 9. W. Cheremisinoff.. 2. 6.4% points. Amsterdam. and Bracker. 11. Isles. Butterworth-Heinemann. 1995. 1992. April.
Energietechniek. Part II. . 1994.5.. Van Wijk. pp. Vol. 782785.Last Developments". V. 15. "Koppeling Warmte/Kracht met Afval" (Cogeneration Integrated with Waste Incineration). Wolpert. "Incineration of Municipal Solid Waste Combined with Energy Production . .M. G.. Renewable Energy. (in Dutch). 1994. COMBINED AND INTEGRATED CYCLES 14. February.106 COGENERATION.
Chapter 7
Air Bottoming Cycle
The heat recovery steam generator and the steam turbine of the conventional combined cycle plant can be replaced by the air bottoming cycle (ABC). In the ABC, the exhaust of an existing, topping gas turbine is sent to a gas-air heat exchanger, which heats the air in the secondary gas turbine cycle. These two turbines make up the dual gas turbine combined cycle (DGTCC) plant, which is characterized by the absence of water and steam equipment, low working pressure and a short start-up time. The plant allows unmanned operation, and can be implemented in regions where water resources are limited. In addition, air which leaves the secondary cycle at a temperature of 200 to 270 °C can be used for process needs. That makes it possible to implement the DGTCC as a cogeneration plant. To analyze the performance of the dual gas turbine cycle, a number of simulation models, based on existing gas turbines, were made. Energy and exergy analyses of the DGTCC with various topping gas turbines were performed. The analyses showed that performance of the DGTCC is comparable with that of the conventional combined cycle plant with steam bottoming.
7.1 Introduction
The concept of an air turbine with an external heat source is not new: it dates back to the invention of the gas turbine. Nevertheless, interest in the idea has grown in the last decade: the air turbine is a key component in the externally-fired combined cycle (Chapter 7, this thesis) and in the air bottoming cycle. The latter was discussed by Farrell (1988) and by Wicks (1991). One of the implementations is the use of air bottoming with the topping gas turbine cycle (Anonymous, 1991). The exhaust gases from the gas turbine are recuperated in a heat exchanger by heating air in the secondary, air turbine cycle. The heated air expands in the turbine, supplying additional
Air 1 3 2 3
4 Exhaust Air 21 Heat exchanger 22 23 4 Air-out 24 1 Intercooling (optional) 2 22 21
Fig. 7.1. Gas turbine with the air bottoming cycle.
power (Fig. 7.1). Such a configuration is known as the dual gas turbine combined cycle (DGTCC) (Wicks and Wagner, 1993; Weston, 1993). In a conventional combined cycle plant the exhaust heat of the gas turbine is utilized by the heat recovery steam generator and steam turbine. In addition to the boiler and the steam turbine, the steam bottoming cycle (SBC) incorporates a condenser, a water treatment unit, pumps, piping, etc., resulting in considerable costs. At the same time, an advanced unfired SBC provides only a third of the total plant power. It also has some drawbacks in operation and maintenance: the start-up period of the steam boiler can take some hours, the high pressure equipment requires strict ASME code compliance and licensed operators. Substitution of the SBC by a bottoming cycle which requires no water, nor steam, seems an attractive alternative for power generation in the under 50 MW range. The use of air bottoming with small topping gas turbines offers an increase in power and efficiency without the complexity of the SBC. The present chapter discusses the configuration and performance of the combined-cycle plants based on small and medium gas turbines.
7.2 Background
Wicks (1991) derived the concept of the air bottoming cycle from the theory of the ideal fuel-burning engine by comparing the engine with the Carnot cycle. While the Carnot cycle is ideal for heat sources and heat sinks of constant temperature, this is not the case, when considering fuel burning engines. In an ideal fuel-burning engine, the combustion products are created, and present a finite size hot reservoir, that releases heat over the entire temperature range from their maximum to ambient temperature. When this is put into a T-s diagram, a triangular area can be drawn, instead of the rectangular area in the case of the Carnot cycle. Thus, the cycle consists of isothermal compression, heat addition, and isentropic expansion. This three-process cycle may not be practical for a single cycle engine because very high pressure ratio will be required. This concept is more realistic for a heat recover-
Recuperated GT
Dual GT
a 2 b 1
4 2
4 2
Fig. 7.2. T-s diagram of the recuperated gas turbine cycle, the dual gas turbine combined cycle, and the gas turbine with steam bottoming cycle.
ing bottoming cycle. Since the maximum cycle temperature will be the topping cycle exhaust temperature, rather than the topping cycle firing temperature, the corresponding pressure ratio will not be that high. The gas turbine cycle consisting of an adiabatic compressor and an expander, and a heat exchanger can be applied as an air bottoming cycle. The isothermal compression would require continuous cooling during the compression process. This can be approached with an increasing number of intercoolers, or by the introduction of wet compression (Poletavkin, 1980). The air bottoming cycle can be compared with both the recuperated gas turbine and the steam bottoming Rankine cycle. As seen on the T-s diagram for the same topping gas turbine (Fig. 7.2), heat recovery in the case of the recuperated cycle is limited by the compressor outlet temperature and cannot proceed below that temperature level (line 4-b). The dual gas turbine combined cycle has no such limitation, and the use of intercooling can improve the heat recovery even further. In comparison with the steam bottoming cycle, it is clear that the steam boiler has much smaller pinch temperatures than the ABC, although the average ∆T of the former is still comparable to that of the latter, because the heat exchange in the air cycle occurs in a single gas phase. Along with the utilization of a gas turbine exhaust, other sources of heat can be considered: such as waste heat from a chemical reactor, or an incinerator's furnace. For better ABC performance, the heat temperature, however, should exceed 400 °C. The air bottoming cycle was proposed to increase efficiency of the simple-cycle gas turbine units operating on Norwegian oil platforms in view of the CO2 tax (Bolland, 1995). Converting these units into conventional combined-cycle plants was found not feasible due to the considerable weight of the steam bottoming equipment and the boiler feed water requirements. The same considerations inhibit the use of the steaminjected gas turbine (STIG) cycle in off-shore applications (Spector and Patt, 1997), whereas the air cycle offers less weight and an efficiency close to that of the SBC. The dual gas turbine combined cycle can be implemented as a combined heat and power plant by supplying the outlet air flow, which has a temperature above 200 °C, for heating needs. The hot air may be used either directly as a product, or be seen as a heat transfer medium.
35 0.0 ηC.20 0. k as (γ-1)/γ.9.05 0. optimization to the maximum power results in the maximum efficiency.40 Dimensionless specific work.1) If a temperature of 480 °C is taken as a typical value of the top cycle temperature (the temperature ratio equals 2.30 0. twice as much work can be obtained from the cycle.00 1 2 3 4 5 ηC. Introducing γ as the ratio of specific heats cP/cV.25 0. 7. ηT = 0.10 0. COMBINED AND INTEGRATED CYCLES 0. The chart shows that by improving the efficiency of the compressor and expander from 0.15 0. w/cPT1 0. ηT = 0. Also.2) . 7. it can be noted that the curve around the optimum point becomes flatter as efficiencies grow.8 to 0. Since this is the bottoming cycle. t as the temperature ratio T23/T21 (using the designations in Fig.8 6 7 8 9 10 Pressure ratio Fig. The operational parameters of the ABC were set to typical values of the simulated hardware.1). 7. (7.3).9 ηC. Specific work of the ABC at different levels of turbomachinery efficiency.8). 7.110 COGENERATION. the specific work of the ABC per unit air flow can be described by W 1 = t ⋅ ηT ⋅ (1 − r − k ) − ⋅ ( r k − 1) c p T21 ηC ( t ⋅ ηT ⋅ ηC − r k )( r k − 1) = r k ⋅ ηC (7. then a number of specific-work curves can be plotted for different turbomachinery efficiencies (Fig. The top cycle temperature T23 is determined by the heat exchanger properties and can be set either by a fixed difference between the hot inlet and cold outlet temperatures ∆T: T23 = T3 − ∆ T .3. ηT = 1.3 Thermodynamic considerations A cogeneration plant based on an existing gas turbine engine with the air bottoming was modeled and compared with a conventional gas turbine cogeneration plant.
4 Configurations Because the topping gas turbine outlet temperature determines the performance of the bottoming cycle most. and with two intercoolers.61 21.4 33.00 Efficiency Pressure ratio Exhaust flow Exhaust temp.1 12.3 35.8 10. The intercooler cooled the air down to 30 °C.7 33. as well as with one.9.5 9.6 68. Performance specifications of the selected gas turbines.3) so that the top cycle temperature can be expressed as: T21 ⋅ r k ⋅ (1 − ε ) T23 = ε ⋅ T3 + ηC (7.93 5. at a pressure ratio of 4. 1995) at ISO conditions.4 28.7 54.0 18. ∆T equals 50 K. Heat exchanger effectiveness was set to 85%.9 30.0 kg/s 18. turbines with an exhaust temperature above 500 °C were chosen from the the small and medium power range. In the bottoming cycle both compressor and expander had an isentropic efficiency of 90%.4 15. Regarding the compressor. For example. and an effectiveness of 0. The combined heat and power concept with the ABC was compared with a cogeneration plant that incorporates a steam bottoming cycle.AIR BOTTOMING CYCLE 111 Table 7. Their performance data are presented in Table 7.1.91 12. % LHV 26.5 78. 7.1. The comparison was made on the . water was used as an intercooling medium.78 3. Engine Power MW Solar Centaur H Allison 501-KB5 Allison 571-K Mitsubishi MF-111A GE LM2500 ABB GT-10 3. When intercoolers were applied. The pressure losses at the air side of the heat exchanger were assumed to be 120 mbar. a compressor isentropic efficiency of 0.8 °C 514 549 533 547 529 541 or by the heat exchanger effectiveness ε ε= T23 − T22 T21 ⋅ r k − T23 ⋅ ηC = T3 − T22 T21 ⋅ r k − T3 ⋅ ηC (7.7 15.56 24. schemes were analyzed without intercooling. Simulations were made with the use of the Gate Cycle code (Enter Software.85. and at a pressure ratio of 6 the difference is 40 K. and 20 mbar on the gas side.6 18.4) The temperature difference in this case is not fixed.9 14.
and the approach temperature of 5 K. basis of an equal amount of exergy delivered to the process at the same temperature.75 2 intercooler 1. optimum values were found for each topping gas turbine in three bottoming configura- .4).25 1 no intercooling 0. 7. COMBINED AND INTEGRATED CYCLES Process Stack Fuel Air Deaerator Return Fig.5.112 COGENERATION. In the evaluations. A single-pressure HRSG was chosen as typical for smaller gas turbines (Fig. 7. a continuous blow-down was set to 1% of the total flow through the evaporator. The HRSG is characterized by the pinch temperature of 8 K. and the deaerator pressure was maintained at 1. MW 1 intercooler 1.75 1 2 3 4 5 6 7 8 9 10 11 12 Pressure ratio Fig. Superheating of 10 K was applied to the process steam to prevent condensation in the distribution grid.4.5 ABC power. Excessive steam was sent to the condensing steam turbine (an isentropic efficiency of 80%) to generate additional power.05 bar. 5. 7.2 bar. The reference cogeneration plant with the steam bottoming cycle. for both ABC and SBC the same topping gas turbines were used.5 Performance analysis Regarding performance of the air bottoming cycle at different pressure ratios. The pressure in the condenser was set at 0. In addition. 1. Optimization of the overall pressure ratio in the ABC (Allison 571-K).
7. 2 IC Fig. 7.2.2% points (2 intercoolers).6. no IC DGTCC. LHV efficiency of the simple-cycle and DGTCC plants. 6. no IC DGTCC. % 40 35 30 25 20 15 10 5 0 MF 111A 501-KB5 LM2500 Centaur 571-K GT10 DGTCC.5. Those also correspond with the figures given by Bolland (1995) for the GE LM2500. 1 IC DGTCC.3-9.AIR BOTTOMING CYCLE 113 50 45 40 35 30 25 20 15 10 5 0 MF 111A 501-KB5 LM2500 Centaur 571-K GT10 Simple cycle DGTCC. % Power increase. and 6. respectively.7% points (no intercooling) to 8. 1 IC DGTCC. since they all operate within the similar temperature range and with the same turbomachinery efficiencies.6 and 7. 7.2. LHV efficiency.9.4% points (1 intercooler) and to 9. with one and with two intercoolers). 2 IC Fig. 7. Comparable figures were obtained for other gas turbines. The chart indicates a significant effect of intercooling on the ABC performance: an addition of an intercooler increases the power output of the non-intercooled cycle by 22%. The optimum pressure of the configurations was 4. Power increase with respect to the simple-cycle power output.8-7.7. tions (without intercooling. . An example for the Allison 571-K is given in Fig.3-10. and the second intercooler adds 8% more. The application of the air bottoming cycle gives an LHV efficiency gain from 6. The results of simulation are summarized in Figures 7.
the overall efficiency of these cases remains lower than that of the better engines. a small difference can be seen in the cases without intercooling (Table 7. The ABC increases a gas turbine plant's power by 20% to 35%.5 Process °C N. Total plant Gas turbine System power MW 1. such as the Solar Centaur H and the Allison 501-KB5 (Fig.6 Cogeneration Two dual gas turbine combined cycle plants based on the Allison 571-K and the GE LM2500 were compared with conventional cogeneration plants built around the same topping turbines and providing heat of the same exergetic value and temperature.2.6 26.3 47.1 44. TPROC as in case 4 7. SBC. The increase is especially pronounced in the cases of less-efficient turbines. no intercooling GE LM2500 9.8 40.2% to 44. TPROC as in case 2 6.3 48. These turbines produce more waste heat. Comparison of ABC and SBC (Allison 571-K and GE LM 2500).1 44. and although the bottoming cycle helps to recover it. 7. Simple cycle 8. Simple cycle 2.8 35.3 21.4 26.4% electric efficiency.114 COGENERATION. SBC. For example. ABC. TPROC as in case 10 5.2). The best performance was obtained by the Allison 571-K and the GE LM2500.A.4 7. case 2 is characterized by 7 MW of power and 40. depending on the number of the intercoolers and the gas turbine's simple-cycle efficiency.6 25. which is comparable to the performance data of case 5. TPROC as in case 8 12.1 45. Whereas one intercooler improves the efficiency of a DGTCC by approximately 1.0 7. one intercooler 4. ABC.6% points compared to the non-intercooling case.5 7.5 47.A. SBC.0 46.4 49.6 43. ABC. ABC.9 41.7).0 49. two intercoolers 5. 278 220 210 273 201 N.9 27.4 42. one intercooler 10.7% points. In addition to efficiency of power generation.4 42. no intercooling Allison 571-K 3. It is evident that the use of more than two intercoolers is not worthwhile.5 Electric % 33.0 46.7%. ABC. SBC.1 7.1 MW .6 34. In comparing the performance data of ABC and SBC gas turbine plants.0 42.2 40. two intercoolers 11. 278 220 207 267 199 efficiency efficiency temperature These improvements result in overall efficiencies ranging from 36. COMBINED AND INTEGRATED CYCLES Table 7.9 7.2 46. the second one adds only 0.5 Exergetic % 32. where power and efficiency amount to 7.2 48.0 26.3 43. exergetic efficiency was also calculated as the ratio between the sum of power generated and heat supplied to the fuel input. 7. ABC.
7% Expander 1.5% 1.9% 2.8. 7. but still lies within 4%. 7.4% 193 °C Power 8.4% 229 °C Fuel 100% Exhaust 7.2% Gas turbine exhaust 26. respectively.7% 8.6% 130 °C 6.9.3% Air 8.1% Gas turbine 173 °C 4.4% 1. Grassmann diagram of the DGTCC without intercooling based on the Allison 571-K gas turbine.3% Air 0.1% 16.7% Power 6.7% 10.1% Fig.1% Expander 1. Similar results were obtained for the LM2500 turbine.4% HX Air 0. and 40.9%.8% 220 °C HX 474 °C 36. When considering ABC with two intercoolers.0% Fig.3% 9.2% Air 0.1% Intercooler 1. . Exhaust 5.2% Compressor 17. only the bottoming cycle is shown.2 bar 0.1% Compressor 1.7% Air 5.9% 0.3% 10. Grassmann diagram of the DGTCC with 1 intercooler based on the Allison 571-K gas turbine.2 bar 481 °C 278 °C 33.2% Power 32.3% 26. the difference between ABC and SBC becomes more noticeable.4% Compressor 1.8% 18.AIR BOTTOMING CYCLE 115 41.
7. These processes do not require clean air. or a nuclear reactor can be considered as a heat source. for example. The ABC provides an attractive performance enhancement of simple-cycle gas turbines without the need for a water-steam cycle.4% vs 36. intercooling is minimized or completely shut off. while maintaining a high electric efficiency: when heat production has a priority.5%. A chemical plant. the total com- . hot air from the ABC is sterile and therefore of special interest for some applications such as food or pharmaceutics industries. in an atmospheric biomass gasifier or a waste incinerator. although this represents some heat waste. are listed in Table 7. which need a large flow of clean hot air.9%.5% to 5. It should be noted that intercooling deteriorates the overall exergetic efficiency by approximately 1. because of a lower heat production rate and some heat loss occurring in the intercoolers (about 0. the combustion products from the gas turbine should not be cooled below the dew point of the acids. The simple operation of DGTCC allows unmanned operation of the plant at remote sites such as pipelines and off-shore platforms. It should be noted that the clean.7% in the intercooled case). the potato starch production. The temperature of the heated air can be reduced as low as desired without the occurrence of harmful compounds. or in paper and cardboard production. and a rise in the power output of 20-35%.3. Usually 100°C is considered a safe level. As a result. This can be useful at a site with scarce water resources or with weight and space constraints. where heat of the process is recovered to drive turbomachinery indirectly. a fluidized bed. on the contrary. This may be necessary.7 Applications Since the DGTCC produces hot air in large quantities. At a CHP plant intercooling permits a load control.9). Among options for air bottoming is its application in a thermal process. COMBINED AND INTEGRATED CYCLES Regarding exergy. 7. the air can be used to improve the efficiency of a combustion process. An exhaust flow of heated air. in the case of corrosive combustion products. The process of drying milk powder.8 Conclusions Implementation of the air bottoming cycle at a gas turbine adds 7 to 10% points to the simple-cycle efficiency. a specific property of air with respect to the gas turbine exhaust gas should be noted. and other industries. by decreasing the temperature of the compressor outlet air. To prevent condensation of corrosive acids in the stack. Thus.g. e. Intercooling.8 and 7. Finally. and more heat becomes available for the expander (33. This can also be implemented for drying purposes such as silt drying. allows better utilization of the topping gas turbine exhaust: stack losses drop from 7. thus the exhaust from the gas turbine could also be utilized.116 COGENERATION. Grassmann diagrams for cases 2 and 3 are given (Figures 7.1% of the fuel input exergy). requires no such temperature constraint. power production grows.
634.3.751. Use of Gas Turbine Waste Heat for Power Generation". ASME Paper 95-CTP-50.Y. However.785. and an economic analysis will be needed to compare the air bottoming cycle with other alternative schemes.. and Farrell.814A. Alderson. 1995. where the ABC concept can implemented. the ABC showed performance values close to and exceeding those of the SBC... such as the Allison 571-K and the General Electric LM2500. while featuring a simpler and robuster design. Bolland. "Air Cycle Thermodynamic Conversion System".Y. References 1. Anonymous. The use of intercooling considerably improves the performance of the ABC increasing bottoming cycle power output by 30% in relation to the non-intercooled case. Gas Turbine World. 4. Industry. Process heat temperatures and year load duration of the industries. 2. smaller dimensions. US patent 4.M.1. California. Gate Cycle User's Guide. N. Schenectady. 3.D.AIR BOTTOMING CYCLE 117 Table 7. 1988. N. Førde. "Low Cost 'Air Bottoming Cycle' for Gas Turbines". due to the low heat capacity of the media a considerable surface area will be required in the air-gas heat exchanger. "Air Bottoming Cycle. and the absence of a water treatment process. O. Farrell. Nevertheless. and Hånde. 5. Schenectady. . M. B. Enter Software. 1995. Menlo Park. the introduction of more than two intercoolers does not seem to be justified. The sensitivity of the ABC performance to the turbomachinery efficiency necessitates the use of the most efficient equipment.M.. 1991. General Electric Company. product Temperature °C Year load duration hours 3000-5000 4000-6000 5000-7000 5000 2000 4000-5000 5000 5000-7000 5000-7000 3000-5000 Dairy products Bakery Starch Coffee-extract Tobacco Animal food Paper and cardboard Synthetic resins Bulk polymers Sludge and silt drying 50-220 100-200 170 200-250 120 400-1000 140 140-200 50-200 350-600 bined-cycle efficiency approaches a value of 45% in the case of gas turbines with high simple-cycle efficiency. Version 4. May-June.. E. General Electric Company. W. When compared with the steam bottoming cycle. US patent 4. 1988. W. "Air Turbine Cycle".
2.. Vol. F. "Performance Analysis of Gas Turbine AirBottoming Combined System". 105-110. 9. 1980. C. Vol. 1991. R. 399-403. 37. "The Thermodynamic Theory and Design of an Ideal Fuel Burning Engine".118 COGENERATION. Spector. . Y. Vol. K. 1993.C. Boston. Parogazoturbinniye Ustanovki (Steam-Gas Turbine Power Plants).G. Najjar. Weston. "Dual Gas Turbine Combined Cycles". (in Russian). Vol.H. Wicks. F.. Wicks.. P. 1993. Nauka Publishers.. Energy Conversion and Management. pp. Russia. "A Projection of Advancements in Aeroderivative Gas Turbine Technology for the Next Two Decades (with specific emphasis on off-shore applications)". and Patt. 955-958. 10.. 1997.S. R.B. 2.. 11. Proceedings of the 28th Intersociety Energy Conversion Engineering Conference IECEC'93. Moscow. pp. 7. 474-481. Proceedings of the 25th Intersociety Energy Conversion Engineering Conference IECEC'90. Poletavkin. M. Boston. Proceedings of the 28th Intersociety Energy Conversion Engineering Conference IECEC'93. pp. 8. 1.F. pp. 1996.S. "Synthesis and Evaluation of a Combined Cycle with No Steam Nor Cooling Water Requirements". Boston. COMBINED AND INTEGRATED CYCLES 6. ASME Paper 97-GT-512. and Wagner. and Zaamout.
oxidation of the fuel is completed in the final stage. 8. ranging from 1. where a synthesis gas is formed. The absence of excess air in the first stages makes it possible to reduce the work required by the compressor and to decrease NOX formation. such a reaction proceeds at a temperature of 1300 °C and pressure up to 60 bar. Such a cycle is presented schematically in Fig. with the PO reactor as a combustion chamber in the first stage. In a partial oxidation reaction fuel is oxidized in a sub-stoichiometric atmosphere at a pressure and a temperature similar to those in the combustion chamber of a conventional turbine. Following the sub-stoichiometric stage. The temperature of the reaction is determined by the equivalence ratio. Typically. and secondary air is introduced before the final expansion stage to complete oxidation of the fuel (process b-4).1. and a gas mixture of carbon monoxide and hydrogen is produced. the ratio H2/CO is rather low. To assess the partial oxidation concept.8 (Tindall and Crews. 1995). air is fed to the partial oxidation reactor. 8. The gas then is expanded in the highpressure turbine to an intermediate pressure (process 3-a). If the temperature is lower. The high temperature of the reaction makes it possible to incorporate a PO reactor in a gas turbine cycle. The results of the comparison study and exergy analysis of the cycles considered are presented. the product may contain some unconverted methane. .Chapter 8 Partial Oxidation Gas Turbine Abstract This chapter addresses the use of the partial oxidation process within a gas turbine cycle.1 Introduction Partial oxidation is a reaction in which fuel does not burn completely due to the lack of oxygen.6 to 1. several gas turbine systems were simulated and optimized. After compression (process 1-2). Since no steam is added in this fuel reforming reaction. This can be implemented in a two-stage gas turbine cycle.
15 and 0. COMBINED AND INTEGRATED CYCLES Fuel 1 Air 2 Partial oxidation reactor 3 4 LPC HPC x kg/s (1-x) kg/s HPT a LPT c b Conventional combustor Fig. which in addition to soot suppression. T SC 3 RH 3 b PO 3 b a 4 a 4 4 2 2 2 c 1 1 1 s Fig. Ribesse (1971) and by V.2) illustrates the difference between the PO and reheat gas turbine cycles: in the former cycle only a part of the working fluid is brought to the maximum pressure.120 COGENERATION.6 ppm (Yamamoto et al. Fraction x is determined in the PO reaction by the required turbine inlet temperature T3. reheat.2. NOX forming under fuel-rich conditions were reported to be as low as 2. increases hydrogen content in the synthesis gas and the mass flow across the expander. T-s diagrams of the simple. 8. . many investigators indicate that soot-forming can take place in the PO process. Basic scheme of the POGT cycle. less air excess) in the second stage. and PO gas turbine cycles. 8. Experiments showed that up to 15% of the methane's carbon can be converted into soot in a reaction with an equivalence ratio of 3 (Rabou. 8. 1996). especially at high equivalence ratio values. 1975). Maslennikov (Christianovich et al. 8. The value of x lies between 0.. This problem calls for the use of steam injection.1. Ribesse's scheme represents a compressor. The T-s diagram (Fig. However. The flow split in point c is mixed again in point a. 1997).2 Background The POGT concept was described in the early 1970s by J.M. The partial oxidation gas turbine (POGT) features low NOX forming due to the substoichiometric atmosphere in the high-pressure stage and better combustion characteristics of the synthesis gas (shorter flame.22 for the temperature range of 1200 to 1400 °C.
1997). from which sulphur components in form of H2S can be removed. In his calculations. The lower heating value of the product ranges from 6 to 9 MJ/kg.3.. A simplified scheme of a repowering arrangement is shown in Fig. Maslennikov et al.4. the fuel gas can be burned in a steam power plant or in a gas turbine. Fuel Air Mixer Reformer Recycle compressor Condenserseparator Exhaust Water Fig. 8.8. thermochemical recuperation and wet compression.3. 1992. Ribesse assumed a turbomachinery polytropic efficiency of 85% and a turbine inlet temperature of 807 °C. Christianovich and Maslennikov (1976) proposed use of the partial oxidation process to reform oil residue into a fuel gas. 8. They found that the so-called incremental efficiency of the POGT unit in the retrofit modifications could be as high as 80% (calculated as a ratio of the incremental power to the incremental fuel consumption). After sulphur removal. Repowering of an existing steam boiler by the use of the PO gas turbine.PARTIAL OXIDATION GAS TURBINE 121 ST Fuel Air Fuel Air Conventional GT POGT Stack Fig. The system generates a synthesis gas which can subsequently be fed into a specific process. a PO reactor and an expander. Adding this value to the generated power gives an LHV efficiency of 36 to 38%. Gas turbine cycle with flue gas recycling. . The latter option in the form of a natural gas-fired power plant was analyzed by the same group of researchers (Maslennikov and Shterenberg.
naphtha. The partial oxidation process was outlined in a patent by Nurse (1991). The Chemical gas turbine. fuel-rich turbine of 1500 °C. (1995) studied a gas turbine cycle with thermochemical recuperation.4). The patent also describes means for removing sulphur from the synthesis gas prior to the complete combustion. COMBINED AND INTEGRATED CYCLES ST Stack HRSG Economizer Air Fuel Fuel-rich gas turbine Fuel-lean gas turbine Fig. 8. 8. The composites can withstand higher temperatures than metal materials.S. and multi-stage expansion allowed the authors to claim LHV efficiencies up to 68% in a combined-cycle configuration. In the fuel-rich section. 1997). The proposed system obtained an LHV efficiency of 65. Hodrien and Fairbairn (1993) in their survey on advanced cycles for British Gas mentioned the POGT as a highly promising cycle with a potential efficiency above 60%. recycling of the exhaust gases. (1995) proposed to employ the partial oxidation reaction in the Chemical Gas Turbine.5. 8.5. With a TIT of the first. The POGT concept was investigated by Westinghouse Electric Corporation and the Institute of Gas Technology in a study within the U. it was concluded that the gain in cycle efficiency due to the improved fuel oxidation process is rather small (1. A basic scheme of this concept is given in Fig. the syngas is expanded to the atmospheric pressure and after cooling in the recuperator and the steam generator. partial oxidation and intercooling (Fig. Yamamoto et al. Arai et al.. natural gas. but are subject to rapid degradation if exposed to a small amount of oxygen (Arai and Kobayashi. Harvey et al. However. Recent tests confirmed the material's ability to maintain its integrity even at 1800 °C (Anonymous. compressed again to the maximum pressure. coal.) for subsequent use in a gas turbine. 1997).122 COGENERATION. Department of Energy's Advanced Turbine Systems Program (Rabovitser et al. low air compressor work. Such factors as the low air excess.4%. 1996). (1997) found the thermal efficiency of the Chemical gas turbine to be as high as 44% in the simple cycle and 62% . etc.4% points). In this configuration fuel-rich combustion in the first stage allows the use of carbon fiber reinforced carbon composites (C-C composites) as a material for the turbine blades. where it serves to gasify carbonaceous fuels (oil.
and at x = 0. The results of that study showed that the basic POGT in a combined-cycle configuration has a minor advantage above the conventional gas turbine system. respectively. Regarding the fraction x of the total mass flow. and recuperator pressure loss of 5%. required compressor work. such a cycle might be seen as a sort of intermediate cycle without any specific advantages above either the simple or reheat cycle. for the cycles based on real efficiencies. Moreover. 8. 1996). Twelve configurations were simulated and optimized with the use of the Aspen Plus process simulation package (AspenTech. m= γ T1 cv (8. One of the arguments in favour of the POGT cycle was that the sub-stoichiometric combustion reduces air excess.2%. and hence. it transforms into a simple cycle. and no efficiency rise can be expected. This was shown first for the cycles based on ideal conditions (isentropic compression and expansion. recuperator ∆T of 40K. the PO cycles becomes a reheat cycle. as evident from Fig.1) . γ = . no pressure losses). and exhaust gas recycling with thermochemical recuperation resulted in LHV efficiencies up to 63. Therefore. The use of the exhaust gas recycling in a POGT configuration improved the efficiency by 3. a TIT of 1230 °C. and r as the pressure ratio: t= cp T3 γ −1 . The POGT concept in different configurations was analyzed by Deen (1996) at the Netherlands Energy Research Foundation ECN. intercooling. it also diminishes the amount of work obtained in the expander. at x = 1. Initially it was limited to one partial oxidation stage and one conventional stage without a bottoming cycle.2. m as the pressure ratio exponent. the plant was calculated to have 66% LHV efficiency. air as the working fluid with constant specific heats.1% points. Intercooling had a slight positive effect on overall efficiency.3 Configurations Another study was carried out at ECN in order to make a definitive judgement on the POGT cycle.5%.PARTIAL OXIDATION GAS TURBINE 123 in a combined-cycle configuration. A gas turbine system with partial oxidation was selected for an evaluation study within the framework of the New Energy Conversion Programme (NECT) by the Netherlands Agency for Energy and Environment Novem (Korobitsyn. real gas properties and pressure losses. reheat. the PO stage delivers a rather small fraction of work compared to the total cycle power. The equations of efficiency and specific work for the ideal simple. and then. γ as the specific heats ratio. and therefore. However. which was only 1.1% points compared to a similar conventional gas turbine cycle. and PO cycles use the following notations: t as the cycle temperature ratio. combustion chamber pressure loss of 2. The use of advanced techniques such as recuperation. TIT of the first and the second turbines of 1500 °C and 1300 °C. and a pressure ratio of 20. 1996). its influence on overall performance is quite limited. 8. All models were based on a gas turbine system with turbomachinery polytropic efficiency of 88%. The combined-cycle scheme was also modeled by Lior (1995): with a single-pressure HRSG.
2 1.4 1.4 2. COMBINED AND INTEGRATED CYCLES Temperature.124 70 COGENERATION.2) W t = t − rm +1− m c p T1 r (8.8 Dimensionless specific work output.6. reheat. and PO cycles. The specific work output is given in dimensionless form as the ratio of the net work output to the product of the air heat capacity cp and the air inlet temperature T1.0 2.4) rm (8. 8. Performance map of the simple. °C 1200 1300 1400 40 32 60 Simple cycle 1400 32 40 1200 1300 1400 40 36 32 POGT 50 12 16 24 28 Reheat cycle 16 8 12 20 40 8 4 30 4 1.0 1.6 1. Efficiency and specific work of the simple cycle can be described as: η= 1− 1 rm (8.5) W 2t = 2t − r m + 1 − c p T1 rm And for the partial oxidation cycle: .3) For the reheat cycle: 2t − r m − η = 2t − r − m 2t rm t +1 (8. W/cpT1 Fig.8 2.2 2.6 2. calculated as ideal cycles.
and PO cycles. however. 8. and only at pressures above 12 bar it produces more work. slightly higher than that of the reheat turbine. 8. The cycles were modeled with the following assumptions: .2 2. the parameter x was set at 0.2 1. Nevertheless. reheat. In the latter cycle.4 1.0 2.4 Dimensionless specific work output.PARTIAL OXIDATION GAS TURBINE 125 45 1300 1400 40 36 32 28 1200 24 20 40 Simple cycle 16 1400 1300 1200 24 20 32 24 20 40 1300 1400 40 32 35 POGT 12 16 Reheat cycle 12 16 8 30 8 1.7. η= m m P m P m P2 Pc 1 c t ⋅ 1− + xt ⋅ 1− − x ⋅ + (x − 1) ⋅ + 1 P P P P c 2 1 1 1 P m P m P k P m 2 c c c t ⋅ 1− + (1− x) ⋅ t ⋅ − + xt ⋅ 1− ⋅ t P P P P 1 2 2 1 (8. The chart indicates that at lower pressures POGT cycle resembles the simple cycle.0 1.2.8 2.6. Performance map of the simple.6 1. Efficiency of the PO cycle is.7) The performance map based on these equations is displayed in Fig. reheat and PO cycles with and without recuperation was carried out with the use of the Aspen Plus simulation package. A more realistic approach to assess performance of the simple.6) m m P m P m P2 Pc W 1 c = t ⋅ 1 − + xt ⋅ 1 − − x ⋅ + ( x − 1) ⋅ + 1 P P c p T1 P1 P1 2 c (8. The intermediate pressure in the reheat and PO cycles was optimized for the maximum work output. the PO cycle cannot reach the values of specific work obtained in the reheat cycle. calculated as real cycles. W/cpT1 Fig.
Since a polytropic expander is not included in the standard Aspen Plus unit operation models set. COMBINED AND INTEGRATED CYCLES Table 8.90 59. The chart supports the assumption that the .7/40 469 626 529 48.57 POX 40 983 14.49 32.9 (579 kJ/kg). The chart shows that the performance of the POGT cycle improves almost linearly with a rise in the pressure ratio and the turbine inlet temperature. while for both the SC and RH cycles their maximum values of specific work can be seen within the given range of pressures and temperatures.013 bar.98 24. 1300.4 Performance analysis The curves of efficiency and specific work of the cycles without recuperation are presented in Fig.88 37.5/40 9.5/40 Specific work kJ/kg 501 678 579 LHV efficiency % 39.42 56. 12.53 72.46 65. turbine cooling was not modelled.10 26. compressor 90% Polytropic efficiency. generator 98% Pressure loss in the combustion chamber 2% Minimal temperature difference in the recuperator 40K Pressure loss in the recuperator 5% For the sake of simplicity.45 46.04 483 13.3 (or 678 kJ/kg in absolute units) in the reheat cycle. 8.98 438 694 489 46. but quite noticeable in specific work.22 19. The difference between the reheat (RH) and PO cycles is not such pronounced regarding efficiency. 8.69 42.54 Pressure bar 18 6.46 24.68 26.66 47. The simple cycle (SC) shows a considerable advantage in efficiency over the other two cycles. The simulations were made using the Aspen Plus package with the Lee-KeslerPlöcker properties set. but produces much less in terms of specific work. Main results of performance and exergy analysis for the cases with a TIT of 1400 °C.41 59. 60% relative humidity.50 65. It was assumed that the methane would be delivered to the plant at required pressure. Methane was considered as a fuel.91 28.98 35.68 65. whereas for the PO cycle it is about 1.40 37. At maximum.63 N. the value of specific work equals 2.33 57.61 67.93 31.71 59.22 41.126 COGENERATION.91 18. The ambient conditions were at 1.1. 1997).A.59 69.08 58. since the pressure in the gas pipelines at an industrial consumer's site can be up to 60 bar.96 27.40 67.7.30 Exhaust temperature °C 655 887 798 Exergy: Power Exhaust Total Efficiency in CC Combustion loss % % % % % 37. 44% against 38%. 25 °C.60 27.37 36. Case without recuperation SC RH PO with recuperation SCR RHR POR 13 9.86 Turbine inlet temperatures 1200.80 25. expander 88% Mechanical efficiency. a user-defined model was set up (Kers. 1400 °C Maximum pressure 40 bar Polytropic efficiency.8/40 13.10 85.
19% Fig. In this cycle the compressor outlet temperature at 40 bar is about 640 °C. Table 8. 8. and at higher pressures it approaches the reheat cycle. and hence within a wider temperature range. 1997). In the case of the PO turbine. The high exhaust temperature enables the use of a recuperator.8. heat exchange occurs at lower pressure. Grassmann diagram of the simple cycle. 1995).57% Compressor 36.5-2. RHR. which correspond to the power maximization principle (Bejan. At these conditions.1 gives some performance values for the cycles with a TIT of 1400 °C at maximum specific work conditions.1% points advantage.09% 34.2). The schemes with recuperation (SCR. 1988) and are the most beneficial for combined cycle operation (Horlock. the simple cycle still has the highest LHV efficiency with a 1.1% 32.8% points (in POR) to the LHV efficiency.72% 2. while the level of exhaust temperature is still suitable for utilization in a steam bottoming cycle.PARTIAL OXIDATION GAS TURBINE Combustor 127 28.5 Exergy analysis An exergy analysis was performed using a FORTRAN subroutine for the Aspen Plus simulator (Folke. character of the PO cycle lies in between the SC and RH cycles: at lower pressures it is closer to the simple cycle. chemical and .4% (in SCR) to 10.46% Methane 100% Power 37. this is caused by high pressure in the compressor. In the comparable POGT cycle. The improvement in LHV efficiency of the reheat cycle is lower than in other cases. 8.96% 103. As seen from the table.66% Air 0. the recuperation was applied only to the low pressure flow (line c-a in Fig. recuperation adds from 9. and POR) were also optimized for the maximum work output.64% Exhaust 27. and thus a limited temperature range for recuperation. 8. This subroutine calculates values of physical.37% Expander 3.
and includes them in the Aspen Plus output.6%.0%. the efficiency is about 59.9% Compressor Air 0. mixing exergy. for the reheat gas turbine 59. a minor improvement was observed: the combustion losses amount to 27.2% 2. exergy efficiency figures for different types of steam bottoming cycles calculated by Bolland (1990) were used. This effect is due to the high temperature of the gases before entering the LP combustion chamber. Regarding the performance of the cycles in a combined-cycle modification.77% LP Combustor 2.10). 8. and for the POGT 57. .5% 35. The original subroutine was modified.128 COGENERATION.02% 0.34% 71.42% 33. The cases with recuperation show little difference between them. The efficiencies varied from 65% for a dual-pressure HRSG to 71% for triple-pressure supercritical reheat cycles.8-8. is responsible for the greatest exergy losses within a gas turbine plant.8% 39.73% 9. the exergy value of the exhaust was multiplied by the bottoming cycle efficiency. To estimate the performance of the cycles in the combined cycle.6%. The combustion process.4% in the simple cycle to 25% in the reheat configuration. Grassmann diagram of the reheat cycle.84% Power 36. and the reference chemical exergy values were substituted by those from Szargut (1988). which is typical for an unfired triple-pressure reheat cycle.13% Fig.72% LP Expander Exhaust 32. which proceeds in a highly irreversible manner. Since in the PO cycle most of the combustion air is not preheated. A value of 69%. Results of the exergy analysis are presented in the form of Grassmann diagrams (Figures 8.57% HP Expander 1.9.9%.27% 76.9%. COMBINED AND INTEGRATED CYCLES Methane 41.49% Methane 58. The use of staged combustion reduced the combustion losses from 28. gives the following combined-cycle efficiencies: for the simple-cycle gas turbine 56.59% HP Combustor 15.
The synthesis gas mixture can be either utilized in a direct-fired boiler. for example. by Parker and Bevc (1996).51% Power 35. the efficiency is 43.74% Air 0.03% 1. 8.93% 0.32% LP Expander 2. The exergetic efficiency of such a unit (case POX in Table 8. 8.6 %.54% HP Expander Exhaust 31. or supplied to a syngas consumer.5%).3%. Another possible implementation of the PO concept in a gas turbine is a replacement of the low-pressure conventional combustor by a high-temperature fuel cell.44% Methane 100% 96. which will reduce the exergy losses.24% Air 0.52% LP Combustor 15.12% 1.1) equals 85. The high value of the specific work is due to the sub-stoichiometric combustion (smaller air flow).71% Fig. If the efficiency of the POX unit is defined as the ratio between the power production to the difference in chemical exergy between the natural and synthesis gases: PHYS P + E EXHAUST ⋅ f CC η = CHEM CHEM E FUEL − E EXHAUST (8. As seen in Fig.PARTIAL OXIDATION GAS TURBINE 129 HP Combustor PO Reactor 12.62% 9.10.35% LP Compressor 14. so the stage can be implemented as a stand-alone apparatus by allowing expansion to the atmospheric pressure. the PO stage itself has an exergetic efficiency of about 85%.65% HP Compressor 8. an increase in overall efficiency by several percentage points can be projected. Grassmann diagram of the POGT cycle.10.8%).22% 13.8) then the POX unit will obtain an efficiency of 52.4% 0.08% 84.61% 83. Here. the same factor of combined-cycle efficiency is applied to the physical exergy of the outgoing gas (if the physical exergy of the synthesis gas is disregarded. which is made up of synthesis gas exergy (72. The positive effect of integration of a pressurized solid oxide fuel cell in a conventional gas turbine has been indicated. .1%) and power (13.
2. S.. pp. and Future". ASPEN PLUS User Guide (Release 9. pp. 8. 11. 7.J. Vol. 1990. AspenTech. but can be expected to diminish NOX forming. J. 608-616. N. E-mail Communication.. Horlock.... and therefore. 1997. Staged Combustion Through Partial Oxidation".130 COGENERATION. 1976. The low level of air excess does not improve the gas turbine efficiency. the implementation of the POGT as a subsystem in a fuel cell-gas turbine combined cycle plant is possible. Present. MSc Thesis. Also. 24-30. 9. Anonymous. 117. "Challenges for Development of Highly Efficient Gas Turbine Systems: The Chemical Gas Turbine System".. March. 1995. Harvey. "Study on Advanced Gas Turbine Systems. 1996. "Ultra High Temperature Material Is Developed by Noritake and Nagoya University". higher efficiencies of the gas turbine cycle. 12.W. Arai. reheat and POGT cycles. New York.. Journal of Engineering for Gas Turbines and Power. 1994. Massachusetts. 7. COMBINED AND INTEGRATED CYCLES 8. G.F. 117.Ya. PhD Thesis. 175-187. C. Enschede. Denver. Hodrien. 1995. Of simple. "Reduction of Combustion Irreversibility in a Gas Turbine Power Plant Through Off-Gas Recycling". and Richter. 2. "Analysis of Combined and Integrated Gas Turbine Cycles". Vol.3).H. Cambridge.A. Maslennikov. 1988. The Netherlands. University of Twente.P. University of Essen. July. "Power Into the 21st Century".. 1997 (in Japanese). The synthesis gas can be utilized in steamplant repowering schemes. Department of Mechanical Engineering. N. Knoche. F. Gas Engineering & Management. Journal of Engineering for Gas Turbines and Power. Applied Energy. "Partial Oxidation in Gas Turbine Cycles". References 1. 1997. Bejan. John Wiley & Sons.Past. and Fairbairn. R. November 3-5. Nikan Kogyo Shimbun.L. (in Dutch). the reheat gas turbine has the highest value of the specific work. or in chemical plants.. University of Twente. The Netherlands. 4.. A. P.. 10.... pp. K. and Shterenberg. ASME International Joint Power Generation Conference. 1996. Deen. Advanced Engineering Thermodynamics. Folke. 5. 6. Kers. V.C. . Bolland. Vol. "Combined Power Plants .6 Conclusions A gas turbine system with partial oxidation show similar performance as conventional gas turbine systems within a feasible range of pressures and temperatures. 1997.W. Christianovich. 3. Trondheim University. O. The high exergetic efficiency of the PO stage indicates the possibility of its use for co-production of synthesis gas and power. and Kobayashi. H. MSc Thesis. S. "Steam-Gas Power Stations with Multi-Stage Residual-Oil Combustion". V. The reduced atmosphere in the PO stage permits the use of C-C composites with higher turbine inlet temperatures. January. Germany.M. 29.
High-Efficient Steam-Gas Turbine Units for Repowering Existing Steam Plants".M. Kobayashi.L. and Miao. . 14. Chemical. "Low-Emission. November. 1991.. R. V. Petten. "Evaluation of Thermochemical Recuperation and Partial Oxidation Concepts for Natural GasFired Advanced Turbine Systems".A. Patent 4. 38. "Alternative Technologies to Steam-Methane Reforming". pp. Khinkis. Nurse.G. The Netherlands. 1992. Vyskubenko. "Gas Turbine with Catalytic Reactor for the Partial Oxidation of Natural Gas and Its Application in Power Stations".. M.. "Review on New and Improved Combinations of Existing Energy Conversion Technologies". 1996. "Advanced Energy Conversion to Power". and Tsalko. Batenin. E. Netherlands Energy Research Foundation ECN.M. 16. V..K. "Combustion with Low Heat Yield".A.R.S. and Steward. 1995. Yamamoto.. Netherlands Energy Research Foundation ECN. M. J.. 5-8.. J.. New York.. T... Kobe.Ya. and Bevc. 1988. F. V. Hemisphere Publishing. F. Yu. 15. Bannister. Japan. No. V. and Crews. and Metallurgical Processes. 1996. 1971.992. July-August. J. F. 1997. M. The Netherlands. "SureCell™ Integrated Solid Oxide Fuel Cell/Combustion Turbine Power Plants for Distributed Power Applications". 1996. Exergy Analysis of Thermal. August.R.M. 10-13..M.. Shterenberg. Japan.W. N. and Tanaka.A.. "Effects of Pressure on Fuel-Rich Combustion of Methane-Air Under High Pressure". "Advanced Gas Turbine System Utilizing Partial Oxidation Technology for Power Generation". 1997. Szargut. Report ECN-CX-060. (in Russian). Ribesse.P.L. 21.P. International Symposium on Advanced Energy Conversion Systems and Related Technologies RAN '95. Rabovitser. December 4-6. February. T. 1093-1100. R.. 17.Q. ASME Paper 96-GT-290. D. Nagoya. The Second International Fuel Cell Conference. Energy Conversion and Management. U. Parker. "Clean Electric Power Generation Process". Tindall. Petten.. 18. V. Maslennikov.999. 19. 24..A.. T. Lior.. W. Korobitsyn. 1995.M.. 1996.. Rabou. (in Dutch). 22. N. Morris. and Shterenberg. Maslennikov. N. 20. Arai. Hydrocarbon Processing. 23.PARTIAL OXIDATION GAS TURBINE 131 13. Gas Wärme International.. Vol.. ASME Paper 97-GT-378.Ya.K. Energeticheskoye Stroitel'stvo. Internal Report. L.
cp/cv difference effectiveness of a heat exchanger heat-to-power ratio efficiency Superscripts CHEM P PHYS Q chemical exergy related to power physical exergy related to heat .Appendix 1 Nomenclature Symbols c E F f H m p P Q S t T x γ ∆ ε λ η heat capacity exergy fuel consumption factor enthalpy pressure ratio exponent. (γ-1)/γ pressure power production heat entropy temperature ratio temperature fraction of flow ratio of specific heat capacities.
134 COGENERATION. COMBINED AND INTEGRATED CYCLES Subscripts ACC C CC CG D el EX FU HP REF REJ SEP T th heat pump acceptor compressor combined cycle cogeneration process demand electric exergetic fuel utilization heat pump reference heat pump rejector separate production of heat and power turbine thermal Abbreviations AFC ASME CAGT CC CFC CHP COP CRGT CRISTIG DCSS DOE DRIASI ECN EFCC EHP EPRI FGD FESR FUE GT HAHP HAT HHV HP HRSG Alkaline Fuel Cell The American Society of Mechanical Engineers Collaborative Advanced Gas Turbine programme combined cycle chlorofluorocarbons combined production of heat and power coefficient of performance chemically-recuperated gas turbine chemically-recuperated intercooled steam-injected gas turbine Distillation Condensation Subsystem (in the Kalina cycle) US Department of Energy dual-recuperated intercooled aftercooled steam-injected cycle Netherlands Energy Research Foundation externally-fired combined cycle electric heat pump Electric Power Research Institute flue gas desulphurization fuel energy savings ratio fuel utilization efficiency gas turbine heat-activated heat pump humid air turbine higher heating value high pressure heat recovery steam generator .
NOMENCLATURE 135 IGCC IP LHV LP MCFC MSW NECT Novem PAFC PC PFBC POGT RH SBC SC SF SH SOFC SPFC ST STIG TIT WTE Integrated Coal Gasification Combined Cycle intermediate pressure lower heating value low pressure Molten Carbonate Fuel Cell municipal solid waste New Energy Conversion Technologies program Netherlands Agency for Energy and Environment Phosphoric Acid Fuel Cell pulverized coal fired power plant Pressurized Fluidized Bed Combustion partial oxidation gas turbine reheat gas turbine cycle steam bottoming cycle simple gas turbine cycle supplementary firing superheater Solid Oxide Fuel Cell Solid Polymer Fuel Cell steam turbine steam-injected gas turbine gas turbine inlet temperature waste-to-energy plant .
. Mikhail Korobitsyn joined Lenenergo. University of Twente. He started his career as a maintenance engineer at a 300 MW cogeneration plant in Leningrad (St. In 1994 he started a PhD project at the Chair of Energy Technology of the Department of Mechanical Engineering.About the author After graduating from the Moscow Institute of Engineering and Management with specialization in power engineering. Since 1997 he is a research scientist at the business unit Fuels. where he was involved in several cogeneration projects. Conversion and Environment of the Netherlands Energy Research Foundation ECN. Between 1992 and 1994 he worked at Comprimo Engineers and Contractors BV in Amsterdam. Petersburg). the utility of the North-Western Russia.