ASME90 Gt 335 Bolland

7/28/2019 ASME90 Gt 335 Bolland

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A Comparative Evaluation of AdvancedCombined Cycle AlternativesOLAV BOLLANDThermal Energy DivisionDepartment of Mechanical EngineeringThe Norwegian Institute of Technology7034 Trondheim, Norway

ABSTRACTThis paper presents a comparison of measures to improve theefficiency of combined gas and steam turbine cycles. A typicalmodern dual pressure combined cycle has been chosen as areference. Severai alternative arrangements to improve the efficiencyare considered. These com prise the dual pressure reheat cycle, thetriple pressure cycle, the triple pressure reheat cycle, the dualpressure supercritical reheat cycle and the triple pressure supercriticalreheat cycle. The effect of supplementary firing is also considered forsome cases. The different alternatives are compared with respect toefficiency, required heat transfer area and stack temperature. A full.exergy analysis is given to explain the performance differences forthe cycle alternatives. The exergy balance shows a detailedbreakdown of all system losses for the HASG, steam turbine,condenser and piping.NOMENCLATUREA =heat transfer areaA = nondimensional heat transferarea, see Eq. (16)c = efficiency moisture

correction factorCC = Combined CycleC. = specific heate = specific exergyE = exergyH = specific enthalpy of exhausth = specific enthalpy of H20HP = high pressureIP = intermediate pressureLHV= Lower heating valueLP = low pressurem = mass flowp = pressureQ = heat flowA = gas constantAH = reheat pressures =specific entropyt = emperatureT = temperature

kJ/kgJKkJ/kgJskJ/s ,k WkJ/kgkJ/kg

kJ/kgkg/sbarkJ/s ,k WkJ/kg/KkJ/kg/sCK

u = total heattransfer coefficientW =workx = steam quality1'] = efficiencySUBSCRIPTS

aauxCCCexfGT

ambientauxiliarycoldcombined cycleexhaust gasfuelgas turbinehot

m2fKkWkg/kg

HHASG=ppreh

heat recovery steam generatorpinch-point temperature differencereheatSC steam cyclest steam1,2,3 cycle state points in Fig. 3INTRODUCTIONIn the past two decades, combined gas and steam turbine cycle (CC')plants have successfully been put into operation with very good fuelutilization compared to other types of thermal power plants. CC plantscan either be used for the generation of electricity only, or for thegeneration of electricity and heat. Among thermal power plants whichare commercially available, the CC is the type which generateselectricity with the highest efficiency. The future outlook for CC plantsin Europe is very good. A number of such plants are expected to bebuilt during the 1990s. Great Britain and the Netherlands arecountries where CCplants are becoming very popular.In the early seventies, CCs were built with typical electricalefficiencies about 40 %. Aecently, CCs have been built with electrical

'The terms GuD (Gas IJI1d Damp!) and STAG (SIeam And Gas) are alsocommonly used.'Presenled al Ihe Gas Turbine and Aeroengine Congress and Exposilion-June 11-14, 1990-Brussels, BelgiumThis paper has been accepled for publicalion in Ihe Transaclions of the ASMEDiscussion of il will be accepted at ASME Headquarters unlil September 30, 1990


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efficiencies above 50 %. Two plants are worth mentioning in thisrespect: Pegus 12 (220 MW) in Utrecht, the Netherlands, which wasbeing put into operation in February 1989 (Frutschi and Plancherel[5]). Second, a 1350 MW plant which is being built in Am.barli, Turkey(Ham ann and Joyce [6]). For both plants the guaranteed net eff iciencybased on the lower heating value is above 51 %, which can beregarded as 1989 "state of the art" for CCs.The increase in the electrical efficiency of CCs in the last few yearshas mainly been caused by gas turbine improvements. Increasedfiring temperatures have been introduced for gas turbines withrelatively moderate pressure ratios, which has resulted in exhaust gastemperatures above 500C. This type of gas turbine improvementhas a positive influence on the electrical efficiency of the steam cycle.This paper deals with the potential for improving the steam cycleefficiency in a large CC (> 400 MW), and thereby increasing the CCelectrical efficiency. The steam cycle can be improved by, amongother things, decreasing the temperature differences in the heatrecovery steam generator (HRSG) and by lowering the condenserpressure. This paper concentrates on another alternative: Increasingthe CC electrical efficiency for a given gas turbine, by improvementsin the steam cycle c o n f i g u r a t j ~ n .WJJy. is it necessary to improve the efficiency of the CC ? ModemCCs already have a very high efficiency and most improvements inefficiency have disadvantages with respect to investment costs,complexity and reliability. The economics of a power plant govern howthe plant should be built. In this respect there are four main factors tobe considered: Fuel costs, capital costs, time of construction andenvironmental issues. The motivation for further improvements in CC'efficiencies is an expected growth in energy prices and environmentalaspects. The latter are becoming more and more important, andimproved fuel utilization is one measure to reduce emissions fromthermal power plants.GAS TURBINES USED IN THE STUDYTwo gas turbines were used for the study: Siemens V94 2 andSiemens V94 3. The former of these is at present one of the largestgas turbines in operation anywhere in the world, and is arepresentative choice for a modem CC gas turbine. The latter is animproved model from Siemens, with a high firing temperature. Thismachine represents a new generation of gas turbines now be ngintroduced to the market. This new generation consists of machineslike the 150 MW class Fr 7F nd the 210 MW Fr 9F from GeneralElectric and Aisthom; Westinghouse and Mitsubishi with the 15 MWMF-l l l and 150 MW W501 F; Siemens with the 60 MW V64.3, 155MW V84.3 and the 190 MW V94.3. ABB is probably going to makethe 200 MW Type 15.The !wo gas turbines selected for this study are calculated at ISOconditions with inleVoutiet pressure drops of 10/40 mbar. Thethermodynamic cycle data are given in Table 1 (Rukes [11)).Table 1 Performance data for the selected gas turbines

V94.2 V94.3 2Net power output 144.3 189.5 MWNet e f f i c i ency 32.5 > 35 %Exhaust gas mass flow 504.0 565.2 kg/sExhaust gas temperature 553.1 563.4 CPre s su r e r a t io 10.7 16.0 -

'According to Siemens, performance data for the V94.3 are preliminary.

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STEAM CYCLE CONFIGURATIONSLarge modern CCs are normally built with a dual-pressure steambottoming cycle. Fig. 1 shows a typical distribution of losses ofavailable work3 for such a steam cycle. Obviously, the HRSG lossesconstitute a main share (=40%) of the steam cycle loss of availablework. Figs. 2a-1 show TQ-diagrams and the cycle flowsheet forproposed changes in configuration to decrease loss of available workand thereby improve the steam cycle efficiency. The following cycleconfigurations are shown in Figs. 2a-l:

2a-b Dual pressure cycle2c-d Dual pressure reheat cycle2e-f Dual pressure supercritical reheat cycle2g-h Triple pressure cycle2i-j Triple pressure reheat cycle2k-1 Triple pressure supercritical reheat cycle

The choice of preheating and deaeration system needs to beexplained. The steam for deaeration is generated in a flashtankoutside the HRSG. Hot pressurized water from the LP-economizer isthrottled before entering the flashtank. The steam which is flashed offgoes to the deaerator, and the water leaving the flashtank is used topreheat the feedwater com ng from the condenser. To prevent theexhaust gas moisture from condensing, the feedwater should beheated up to a temperature above the dewpoint of the exhaust gas.This is done in the feedwater preheater with a circulation loop(dashed line in flowsheet figures). The circulation ratio is such that thefeedwater temperature entering the HRSG is above the dewpoint ofthe exhaust gas. In this study the feedwater temperature entering theHRSG is set to 60C, which is weU above the dewpoint of theexhaust gas; normally about 40C.

STACK 9.3%CONDENSER 21.3%

MECH. & GENERAl. 3.6 %

PIPING &. VALVES 3.5%

PUMPS 0.5%TURBINE 21.5%

~ ____ RSG 40.0%Figure 1 Breakdown of exergy losses in a dual pressure steamcycle

COMPUTATIONAL MODELThe HRSG model is separated into!wo basic types of computationalmodeis: The subcritical pressure stage and the supercritical pressurestage. The former consists of an economizer, an evaporator withforced circulation and a superheater. The latter consists of a oncethrough heat exchanger. For the HRSG-calculations, the exhaustgas/live-steam approach temperature difference, pinch-pointtemperature difference, economizer approach temperature difference,live-steam pressure and pressure drops are given as input. For thesubcritical pressure stage the computational procedure is quitestraightforward. The procedure applied in this study is similar to themelhod presented by Chin and EImasri (3). The computational

'The terms "available work", "availability" and "exergy" are synonymous and aregoing to be used interchangeably.


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Figure 2a Flowsheet diagram for the dual pressure cyele

Figure 2c Flowsheet diagram for the dual pressure ref/eat cyele

2e Flows!leet diagram for the dual pressure supercriticalcyele

3

600 r----------------------------------,

500

U 400

wer 350::)f--


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Figure 2g Flowsheet diagram for the triple pressure cyele

Figure 21 Flowsheel diagram for the triple pressure reheat cyele

FiguI'o 2k Flowsheet diagram for the triple pressure supercriticalreheat cyele

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U 4CD

CLcr 350CJf-ITcLlCl:2:LC

Figure 2h U ~ , ( j I ~ 1 G r i ' l m for the trip/e pressure eyele

IFAT TRI\NSH,H MW

Figure 2j TO-diagram for the triple pressure reheat eyele

U 400

er:)1-Illi JCl:2Ulr-

'00

HEAT T R A N S F ~ R MWFigure 21 TO-diagram for the triple pressure supereritical reheatcyele


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". EXHAUST GAS

TRANSFERRED HEATFigure 3 TO-diagram for super-critical pressure stageprocedure for the supercritical p n : l ~ ~ u r e stage is more complex sincethe steam' is not undergoing any sudden change of phase and thereis no distinct "pinch-point". The steam temperature at the pinch-pointis therefore not as easily found as for the subcritical case. Fig. 3shows a Ta-diagram for a supercritical pressure stage. Obviously,the slopes of the two curves have to be equal at the pinch-point.Further, the distance between the two curves has to be the specifiedpinch-point temperature difference, and there must be a heat transferbalance between exhaust gas and steam. With these threerequirements formulated in Eqs.(1-3), the steam temperature (t2s,) atthe pinch-point can be found.

m.,(h 1,-h2 ) '" mex '(H 1-H 2).. '" ~ . e x - ~

(1 )(2)(3)

If constant specific-heat is assumed for the exhaust gas, the steamtemperature at the pinch-point can be found more easily.(4)

However, for this study the formulation for variable specific heat isapplied. It should be noted that when applying Eq. (1) caution mustbe made when evaluating the specific heat (Cp2 .s') for steam near thecritical temperature. The specific heat changes very rapidly as thefunction of temperature above and near the critical point. For asupercritical stage with reheating, the same procedure can be appliedexcept for Eq. (2) which has to be rewritten and an extra heat balanceequation needs to be added (Eq. 6).(h1 -h 2 HH 1-H3) '" (H 1-H 2)'(h 1 -h3..) (5)mex'(H 1-H 3) '" ms,(h 1s,-h3 ) + mreh(hl.reh-h3.reh) (6)

The HRSG heat transfer area calculation is carried out by anintegration method for each type of heat exchanger (superheater,evaporator, economizer and preheater). Acounterflow heat exchangermodel is applied. The integration model ensures that the effect ofvariable specific heat capacities is taken into account. This isimportant for high pressure superheated steam and water nearsaturation. A counterflow heat exchanger model may not be accuratein all cases, but when comparing heat transfer areas as in this study,-Waier ai supercritical pressure can hardly be defined as either water or steam,but is here referred to as steam.

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such a model should be sufficient.The steam turbine expansion is broken into a number of sectionswhich correspond to the number of HRSG pressure stages. Eachsection is computed by an individual dry isentropic step efficiency.The efficiency for the LP section is corrected for moisture if theexpansion crosses into the wet region. The LP expansion is brokeninto steps, and the efficiency for each step is corrected for moisturewhen the exit quality is below that for the onset of condensation. This"Wilson line" quality is normally between 0.95-0.98. The efficiencydegradation is assumed to be an exponential function (7) of meanstep steam quality.

(7)The exponent c is typically in the range 1 0-1 .3, and is chosen to be1.15 for this study. That means the penalty for moisture is a decreasein isentropic efficiency with a factor typically in the range 0.63-0.75for every extra percent of moisture. The steam turbine model alsotakes into account a throttle valve loss, steam leakages through thesteam turbine seais, LP section leaving loss and steam turbine andgenerator auxiliary power requirements.The condenser model is a water-cooled counterflow heat exchanger.The cooling water pressure drop and the required pump work arecalculated.Heat and exergy balances are carried out for all components in themodel. To ensure that there are no errors in the model, an overallsystem heat balance is carried out as well as an overall exergybalance.The net efficiency of the CC is here defined by Eq. (8).

Tlcc '"WGr + Wsc - WAlJX

m,.LHVWGPW SC '" power output at the generator terminalsWNJX auxiliary power demand and pump work

EXERGY ANALYSIS

(8)

Traditional first-Iaw cycle analysis based upon componentperformance characteristics coupled with energy balances invariablylead to a correct final answer. However, such analysis cannot locateand quantify the sources of loss which lead to that result. This isbecause the first lawembodies no distinction between work and heat,no provision for quantifying the Quality of energy. These limitationsare not a serious drawback when dealing with familiar systems, sincean intuitive understanding of the dif ferent parametric influences onsystem performance and a second-Iaw qualitative appreciation of"grade-of-heat" and effect of pressure loss can be developed. Whenanalyzing novel and complex thermal systems, however, such anunderstand ng must be supplemented by more rigorous quantitativemethods. Second-Iaw analysis, or exergy analysis, provides thesetools. Second-Iaw analysis is no substitute for first-Iaw analysis,rather, a supplement.The quantity energy can be split into exergy and anergy.

Energy '" Exergy + AnergyAnergy is energy in equilibrium with the ambient conditions andcannot be converted to work, while exergy is the proportion of energywhich theoretically can be converted to work. The exergy of a flowstream for a given pressure Pl and temperature t1can be computedby the following expression:


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(9)where,

ha = h(p.,t.)Sa = s(P.,t.)

For an ideal gas, the term S,-S. can be wrilten:

S,-Sa = ~ p ( f ) f T ' d T - R J d ~ / P (10)a aTo quantify the loss of exergy for a component, an exergy balanceis applied. Fig. 4 shows a counterflow heat exchanger and thequantities related to an exergy balance. The loss of exergy for theheat exchanger is

(11 )

EH 2 PH,2 l2 SH,2 HEAT EH,I PH1 TH,I SH,1

Ec,1 PC.I l i SC,I EXCHANGER EC2 PC,2 TC2 SC,2/Figure 4 Exergy balanee of a counterflow heat exchangerThe steam cycle is a closed loop. Exergy in a stream leaving acomponent is transferred to the next component, and is therefore notlost from the system. The exception is at points where exergy isdeliberately rejected from the system sueh as the HRSG stack andthe condenser cooling water. The loss of exergy for all othercomponents in the steam eycle can be calculated in a similar way asshown in the above example. When adding up all losses of exergyplus the generated work, this should equal the exergy of the gasturbine exhaust gas entering the HRSG. At this point the exergy.efficiency can be defined

Wnet,out EHASG,ln - L El a .'I1exergy = -E- - = E

HASG,in HRSG,ln(12)

and further, the maximum thermal effieieney referred to the exhaustgas is given by

where,llthermal,max = EHASG,ln

HASG,in

0HRSG,in = mex ' (H HRSG,in -Ha)

(13)

(14)and where, HHRSG,in is referred to ambient conditions which means

Ha = OThe maximum thermal efficiency deseribes the thermodynamiclimitation for the conversion of heat to work. The exergy efficiencyaccounts for internal steam cycle "imperfeetions", such as heat

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exchanger temperature differences, "mismatched" heat exchangertemperature profiles, heat losses, pressure losses, mixing losses,mechanical losses, generation of entropy in compression andexpansion processes and rejection of exergy to the surroundings. Byeombining Eqs. (12) and (13) a relation belween work and heat isobtained

W net,out = 0HASG,ln' 'I1therrrel,max' 'fJexel1lY (15) .RESULTS AND DISCUSSIONParametric studies were carried out for combined cycles us ng theSiemens V94.2 and V94.3 gas turbines (representing eurrent "stateof the art" and advanced technology, respeetively). Thermodynamiceycle data for the gas turbines were presented earlier, and theAppendix contains additional partieulars for the various cyelecomponents.Net exergy effieieneies (Eq. 12) and net combined eycle effieieneies(Eq. 8) were calculated for the various combined cycles. Theseefficiencies are presented in Figs. 5 and 6 as tunetions of the designHP-pressure for the HRSG, using the V94.3. At eaeh HP-pressurethe other live-steam pressures (RH,IP,LP) are optimized with respectto cycle efficieney. Figs. 7 through 9 com pare the effeets of the twogas turbines on net cyele efficiency, staek temperature, and therequired HRSG heat transfer area in eaeh of the cycles.Introducing reheat improves the efficiency by 0.2-0.4 percentagepoints compared to the non-reheat cyeles, both for the dual and triplepressure cycles. The difference in efficiency belween dual and triplepressure cycles is about 0.5-0.6 percentage points except for smal/erHP-pressures where this difference tends to decrease. Supercriticalreheat cycles give a higher efficiency than the subcritieal eycles.There is also a significant difference in effieiency between the dualand triple pressure supercritical reheat cycles, which is about 0.5percentage point.The variation in exergy efficiency is from 65 % for the dual pressurecycle at 60 bar HP-pressure and up to nearly 71 % for the triplepressure supereritical reheat cycle. The differenees in exergyeffieiency eorrespond, of eourse, to what has been explained for Fig.5. It is interesting here to make a comparison between the steameycle and the Kalina-cyele, where in the latter eyele a mixture ofammonia and water is used as working fluid. For the Kalina-cycle,Steeco and Desideri [12) have calculated the exergy effieiency to be63.2 % when utilizing exhaust gas heat at 577C. It should beemphasized, however, that the assumptions used in the present workdiffer from those of the reference [12).In Figs. 5 and 6 the graphs are marked in order to represent a"reasonable" choiee of HP-pressure for eaeh type of eyele. Theseehoices are mainly based on the steam turbine exit quality for thenon-reheat eyeles. For the subcritical reheat eycles the HP-pressuresare chosen to be 140 bar. Increased live-steam pressures normallyimplies higher HRSG eost, but for the subcritieal reheat cycles theHRSG HP-stage has a smaller mass flow and also a smallervolumetric flow compared to the non-reheat cycles. The increasedtube thickness (and weight) due to higher pressure will be opposedby smaller tube diameters. Table 2 summarizes the results from thecalculations with the chosen HP-pressures which are marked in Figs.5 and 6.Figs. 7- 9 are graphical presentations of the data in Table 2. Whencomparing the CC performance with respeet to gas turbine technology(V94.2 versus V94.3), the differences in effieieney (Fig. 7) are about2.0-2.1 percentage points. These differences are larger than anyofthe differenees between the cycles with a given gas turbine. Thepotential for CC effieiency improvement therefore mainly relies on gas


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'*-UZW[ )CLCCUJiruZ

HP -PF1ESSlJRE balFigure 5 Combined Cyele net errieieney as funetion of HP-pressure for the V94.3 gas turbine

)uzwClCLLIIJJ)C)rr:CCIxWi-UJ:7

DUAL Pf-1ESSUf1E Flfii[AT (YCLE

HP -PFiESSURt'. balFigure 6 Steam cyele net exergy affieieney as function of HP -pressure for tlJe V94.3 gas turbine

55.5

>- 550UZW 5 ' ~ . 5uIL 6-1.0UJ[-UJ tJ3,5ZlLi_JU()CjUl:2LlJLoU

r - - - - - - - - - - - - - - - - - - - - - - - - - - - - - - - - - - - ..-------------- - - --'- .-.----,1m SiEM[N:; V94.2llII SIfM[US V94.J

DUAl PRESSURF. DUAl PRESSUHE DUAL PRESSVkE fRI ?l E Pf1ESSlJRE mIP!...L H\hSI,"lF T;1IPI c PF1f:SSlJRfCYCLE REH[AT CyeLf SUPfRCRITICAt CYCI_E RcrlfA.T CYCLE S0P ::I;C!-1lfICAlC'leLE n[flf:A,r (yeU'

Figure 7 Combined Cyele net efficieney for spee/fjely ehosen HP -pressure for t/le V94.2 and V94.3 gas turbin es

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aUUJII:::Jf--CCWCL2:Wi--v 'Uf--U )

CYCLE DUAL PRESSUliE lfllf.'LE PRE:SSURE TI1iiU: PRESSURE -f"HIiLE PFiES$UR[Rc:HEAT CYCI t: S U ~ ' E n C R I T I C A L CYC! E AEHEAT CYCI i SIJPEHCR:TICALCYCLE REHEA)

Figure 8 Stack !PflYln,1r.',,,rlln,,, (tJr specificiy et/osen NP-pressuresgas turbinesor the V94.2 and

turbine doveloprnents. \Nith the new generation oT gas lurbines, in lhisstudy represented the Siemens V94.3, it is likely to have CC netefficiencies reaching 54-55 %.Fig. 8 S!lOWS the slack temperatures for the different cydes. As canbe seen thore are only minor differences in stack temporatures whencomparing the Iwo of gas turbines. Thore is no obviousconnection belwoen slack temperature and efficiency. This meansthat the CC efficiency is no! soiely dependent on ho.'tLfllUCh of theexhaust gas energy Wflictl is utilized, bu! it is also a question of ho.wItle exhaust gas energy is utiiized.In Fig. 9 and Table 2 the HRSG heal transfer areas are given in anondimensional form, which is

(16)where is Ille steam cycla ne! power outputfor all cases. (16) defines a parameter relatingto the net power output of the steam cycla. Fig. 9 SllOWS Ihere isvery littlo difference in Ihe HR SG heat transfer area berNeen the twagas turbines. As can be seen when camparing Figs. and 9,increased efficieney comes al the expense of a larger heat transferarea. The supererilieai eyeles, especially, require a large heat transferarea for a given gain in It is interesting lo note iha! the!riple pressure subcritical reheat cyde requires less heat transfer areathan the triple pressure non-reheat cyclo. The reason for this is thal


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the relle8! results in less mass ilow going the Hp eeonomizer and whieh eonslilule most of the heattransfer area for a nonreheat eycle. Besieles, Ihe condensate flowrate is less for a rehea! cycie compareel lo a non-mhea! cycle, wlliehn1eans less heal lransfer and heat lransfer area at thH cold end of U,eHRSG.The effeet of in front or the HRSG for threedifferent types eyeles is shown Fig. 10. The seleded cyclas arethe dual pressum cyele, the lripie pressure reheat eycle and the dualpressure supercritical rehea! eyde. All three eyclas are ealculaled WiUlthe V94.3 gas lurbine. The !WO subcriticai cyeles are caleulated withdifferent along with an optimization of the oUler livesteam pressures. As can be seen from Fig. 10does no! improve lhe of the subcritieal cyc!es irrespectiveof the HPpressure wilhin the range stated in Fig. 10. On Ule alherIland, the is sJighlly improved by Uleof the dual pressure supercriiicalThe differenee be!ween the wilh respecl to efficieney isdescribed by rneans of firs l-Iaw method of analysis. A second lawor exergy analysis provides information aboul wby lrlere aredifferences in cycle performance. Fig. 11 shows a breakdown ofexergy losses for different components in the steam cycle wllh theV94.2 gas turbine, as well as the steam cycle ne! exergy efficiencies.The main for differences in efficiency between Ihe cycles isthe decrease in HRSG exergy losses, The steam turbine exergyIOSS8S do no! differ significan!ly be!ween the cycles. The varialionsean be explained the fae! that different mass fiows are expandedin !h lurbines Ula! the efficiencies vary because of differoncesin exit quality, especially between reheal and non-rehea! cycles. Thecandenser has also rather small variations in exergy losses; bu! Ihe

CYCI[

DUAl_ Pil!" SSUl1F SUPEflCfUI iCAI5 ~ . 8

64.6

54.4

580 690 eoo 610 620 610SIJPflLHM::NTAlW fiRING T E : M P c r ~ A T I J R [ t

Figure 10 Combined Cyele ne l e n ' I C I / ~ n r : v as funetion ofsupplementary firing temperature

f,40

reheat eycles have sli()hlly less condenser losses. Even if reheatcyc!es ~ l a v e a higher steam turbine exit Ihe reduc!ioneondenser steam mass flow implies a loss of exergy, Theslack exergy loss is a funclion of slack tm perature and as can beseen the triple pressure cyeles have mueh smeller losses Itlan Uledual prcssure eyeles. The exergy losses from pipes and valves 8110wnin Fig. 11 are obviously funelion of cycle eomplexity and steammass flow.

2i-\1 DU!'\L P R ~ S S U R E RH--1EAT: DUAL PRESSURE SUPt=SCRITICAL RtHE,t\,!:lP mlc'l.f: PRESSURE CYCLE

3PR "TRIPL!: PRESSUF,E IlEHEAT CYCI.EJPHS c TRiPLE PHESSURE SUPERCRITIC!\I. 11F:fjC:IIl CYCI

Figure '/1 Exergy effiency and exergy loss breakdown for the considered cyeles8


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CONCLUSIONSThe combination of the first-Iaw and second-Iaw approach providesa good tool for the analysis of power cycles. The exergy balaneemethod of analysis enables all loss sources to be J.QQaied andQlliIDillled. When different steam cyele configurations are compared,the exergy analysis gives a very useful understand ng of wlnlthermodynami cal per formanee difters from ane type of eyele toanolher.The lriple pressure supercritical reheat steam eyele gives the largestincrease in efticiency compared to a "state of Ihe art" dual pressuresubcritical steam eyele. On the olher hand, the inerease .in requiredheat transfer area is large compared to Ihe gain in efficiency. Thetriple pressure subcritical reheat steam cycle seems to be veryinteresting. Compared to the dual pressure non-reheat cyele, theincrease in efficiency is approximately one percentage pointSupplementary firing is not very interesting as a measure to increaseCC efficiency with the types of gas !urbine used in this study, exceptfor the supercritical cycles. However, supplementary firing providesflexibility with respect to load control, and from a design point of view,flexibility is added to power output without having to make significantlyless effiden! cycles.When com paring the CC effieiency of the !WO gas turbines used inthis study, il is obvious U'lat the new generation of gas turbines nowbeing introduced to the market will increase the CC efficiency byaboul 2 percentage points. This new generation of gas lurbinestogether with some of the proposed steam cycle configurations willmake it possible to reaeh neI efficiencies in the range 54-55 % fo rlarge combined eycles.ACKNOWlEDGEMENTThe discussions he d with Dr. Maher EImasri wme invaluable duringthe course of this work.REFERENCES1 Bolland,O., Gl.!llEowetPJai:ltA11ernalillesJru:.Nm:wegian B o u o d r u : y ~ n s ,Siemens(KWU)-report Erlangen, FederaJ Republic of Germany, 19872 Bolland,O.,and Lken,P.A., Poteotials lo Improvem Steam C y c l e ~in.-'l1LUnlirru:LCrnIlbiruIcLGycill, SINTEF-report STF15 A87053, Trondheim,Norway, 19873 ChinJW.vV.,and Elmasii,M.A., ~ ~ J ) f Combined cycles Part 2Analysis of T l I I ' O = P J J i s s u r a l l l i l a m , . B u ~ . c : l . e , Journal of Engineering forGas Turbines and Power, Vol. 109, pp. 237-243, April 19874 EngelkeW., Dampfturbinen Jijr GuD Krafuv1Mke. BWK , Sd. 41, Nr. '7/8 -July/August, pp. 335-342, 19895 Frutschi,H.U.,and Plancherel,A.. Comparison otCQlnblnel:LC)'Cles...witlLSleam

l n j e c t l O ! l . a r u i . E . ~ a t i o o C y c ~ , IGTI-vol. 3, ASME Cogen-Turbo, pp. 137-145, 1988

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Il Harnann,B.,and Joyce,J.S" W o I l I D L J a r g a s L J 3 a l L 3 ~ _ I D L A m b a r l iC o i n b i n e i l ~ , Modem Power Systems, pp. 61-74, May 19891 Horlock,J.H., C O Q e n e r a t i o n ~ e a t and Power iCliE').

1 h e J : t n o d ) m ~ a D d Economks., Pergamon Press, New York, 1987li Johnson,D.G .. Mbglichkeiten der Kombi Kraf!werke mil Hochtemperatllr

Gasturbinen, VGS Kraftwerkstechnik 62 , pp. i 67 -173, March 19829 Kehlhofer,R., et al., G.as1urilinlmkraftwerlilLKombikraftwerke, Heizkraftwerkeu r u : L . l r n i u s t r ~ k e , Handbuchreihe Energie Technischer Verlag Resch,Verlag TUV Rheinland, Band 7, chapter 2, 198410 Kotas, T.J., I l l i . L E l r g ~ M e t h o d of T h e I T n a L J : ) J a n L A r ~ , Butterworths,London, 198511 Rukes,B., Head of Department for Calculation of Heat Cycles and HeatRemoval Systems, Siemens AG - KWU Group, The Federal Republie ofGermany, ~ C Q J D l I l U / l i c a t i H12 Steceo, SOS. ,and Desideri, U., l L I h ~ y s i s of the Kalina~ _ C o ~ and Perspectils, Presented at the ASME GasTurbine and Aeroeogine Congress and Exposition, Toronto, Canada, June i989APPENDIXInput not mentioned in the body of the paper but used in the calculations ciledare:

~ l n e s

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ASME90 Gt 335 Bolland