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Energy 35 (2010) 341–350

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Energy

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Energy and exergy analyses of an externally fired gas turbine (EFGT) cycleintegrated with biomass gasifier for distributed power generation

Amitava Datta*, Ranjan Ganguly, Luna SarkarDepartment of Power Engineering, Jadavpur University, Salt Lake Campus, Kolkata 700098, India

a r t i c l e i n f o

Article history:Received 9 January 2009Received in revised form1 August 2009Accepted 25 September 2009Available online 28 October 2009

Keywords:Gas turbineExternal firingBiomassGasifier

* Corresponding author. Tel.: þ91 33 23355813; faxE-mail address: amdatta_ju@yahoo.com (A. Datta)

0360-5442/$ – see front matter � 2009 Elsevier Ltd.doi:10.1016/j.energy.2009.09.031

a b s t r a c t

Biomass based decentralized power generation using externally fired gas turbine (EFGT) can be a tech-nically feasible option. In this work, thermal performance and sizing of such plants have been analyzed atdifferent cycle pressure ratio (rp¼ 2�8), turbine inlet temperature (TIT¼ 1050–1350 K) and the heatexchanger cold end temperature difference (CETD¼ 200–300 K). It is found that the thermal efficiency ofthe EFGT plant reaches a maximum at an optimum pressure ratio depending upon the TIT and heatexchanger CETD. For a particular pressure ratio, thermal efficiency increases either with the increase inTIT or with the decrease in heat exchanger CETD. The specific air flow, associated with the size of theplant equipment, decreases with the increase in pressure ratio. This decrease is rapid at the lower end ofthe pressure ratio (rp< 4) but levels-off at higher rp values. An increase in the TIT reduces the specific airflow, while a change in the heat exchanger CETD has no influence on it. Based on this comparison, theperformance of a 100 kW EFGT plant has been analyzed for three sets of operating parameters anda trade-off in the operating condition is reached.

� 2009 Elsevier Ltd. All rights reserved.

1. Introduction

Small scale decentralized power generation is gaining impor-tance for distributing electricity in the remote areas far from thecentralized grid [1–4]. The delivery of grid power to the remoteareas, particularly in the hilly terrain, is extremely uneconomic [5].On the contrary, the installation of small capacity plants catering tothe local needs using the local resource can be an attractive alter-native for remote places. Biomass is one of the important availableprimary resources, which generally exists in abundance in thevillages and already serves as the source of energy e.g. in cooking.

Energy from the biomass can be thermochemically recoveredfor the generation of electricity either through direct combustion orthrough gasification and subsequent combustion of the producergas. In large scale, biomass gasification can be used for powergeneration in a combined cycle [6,7]. On the other hand, pistonengines or micro gas turbines are suitable for small capacitydistributed generation. Producer gas can be used in conventionaldiesel engines in the dual fuel mode or in producer gas engines forthe generation of power [8]. However, such engines having

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reciprocating components require more maintenance and abun-dance of cooling water, which make them unsuitable for remotelocations.

The use of biomass as fuel in conventional (internally fired) gasturbine engines entails various problems [9]. Firstly, the gasturbines are sensitive machines that require extremely clean gas toavoid damage to the turbine blades (such as erosion, incrustation,and corrosion) and blockage of filters and fuel injectors. Thisrequires installation of expensive gas clean up system, consisting ofscrubbers, ceramic filters, cyclones etc., at the gasifier outlet.Secondly, the low calorific value of the producer gas, obtained frombiomass gasification, necessitates a high fuel flow. It calls fora design modification in the combustor and the turbine inlet guidevanes, otherwise the change in the mass balance between thecompressor and the turbine moves the compressor operating pointtowards surge [9]. These problems are resolved, if the biomass canbe conveniently used as a fuel in an externally fired gas turbine(EFGT) engine.

In an EFGT cycle [9], the high pressure air from the compressor isheated in a heat exchanger before admitting to the turbine. Theturbine essentially handles clean air and the turbine exhaust air issubsequently used to burn the fuel in a combustion chamber. Thecombustion product is employed as the hot stream of the heatexchanger, before being released from the power cycle. The cyclecan employ dirty and low cost fuels, as the combustion products do

Nomenclature

AHE Heat exchanger surface areaAFSt Stoichiometric Air-fuel ratioCETD Cold End Temperature Difference of heat exchangerCp Specific heat at constant pressureei Specific thermomechanical flow exergy at state iech Specific chemical exergyEFGT Externally Fired Gas TurbineEN Energy released with exhaust gash Enthalpyhf Enthalpy of formationK Equilibrium constantMj Molecular weight of species jPi Pressure at state i or Partial pressure for species iDP Pressure droprp Pressure ratioR Universal gas constants EntropyTi Temperature at state iTIT Turbine Inlet TemperatureU Overall heat transfer coefficient of heat exchangerw Specific workW WorkX Number of molesZ Moisture content in the as-fired biomass (by mass)

Greek Letters4 Equivalence ratiog Ratio of specific heatshc,isen Isentropic efficiency of compressorht,isen Isentropic efficiency of turbineh Efficiency

Sub-scriptsa AirB Producer gas after gasification of biomassC CompressorCC Combustion chamberf FuelG Gasifierg Product gasHE Heat exchangerin Inputi Index for thermodynamic state pointo Reference stateT Turbinew Water

Super-scriptsc Cold side of the heat exchangerh Hot side of the heat exchanger

A. Datta et al. / Energy 35 (2010) 341–350342

not enter the turbine. Although the presence of ash in the productsmay cause erosion and fouling of the heat exchanger tubes, whilecorrosive products eats away the tube material, maintenance of theheat exchanger is much less troublesome than that for the turbine.

Anheden [10] presented thermodynamic and economic analysesof closed and open cycle externally fired gas turbine plants withdirect combustion of biomass in a circulating fluidized bed furnace.It is found that the efficiency reaches a maximum value at anoptimum pressure ratio of the cycle. Ferreira and Pilidis [9]compared the thermodynamic performance of an externally firedgas turbine cycle with direct combustion of biomass against aninternally fired cycle firing either natural gas or producer gas frombiomass gasification. The study was performed for the simple gasturbine cycle as well as for the combined cycle operation witha steam based Rankine cycle at the bottom. The results showedpromising performance for the EFGT plant particularly consideringthe renewable and environment-friendly attributes of the biomassfuel. Bram et al. [11] reviewed the technological and economicfeasibility of the external firing of biomass in gas turbines. Theauthors concluded that cogeneration based on EFGT on the scale of100–200 kWe offers good prospects from both economic andtechnical aspects. Cocoa et al. [12] evaluated the performance ofa 100 kW externally fired gas turbine plant fuelled with biomassand having an integral dryer for biomass. The influence of param-eters like pressure ratio, turbine inlet temperature and temperaturedifference in the heat exchanger on the thermal efficiency forelectrical generation was analyzed. It was found that the drybiomass produces efficiency in the range of 22–33% and the inte-gration of the dryer improves flexibility in the plant operation.Traverso et al. [13] presented the steady state and transientperformance of an externally fired micro gas turbine pilot plant of80 kW capacity fired with natural gas. The paper demonstrated thefeasibility of operation and control of the gas turbine plant of smallcapacity.

All the literatures on EFGT universally claim that one of thebiggest challenges in the design lies in developing the hightemperature heat exchanger that is capable of achieving highturbine inlet temperature and at the same time withstands thestresses imposed by the working conditions and the constituent ofthe combustion product [9–12]. The size of the heat exchanger andthe cost of material are the two important considerations thatdecide the economy of the plant. The use of nickel based superalloys in the heat exchanger allows the turbine inlet temperature toreach 800–825

�C, while more advanced oxide dispersion (ODS)

alloys withstand temperature up to 1100�C at the turbine inlet [10].

The turbine inlet temperature may be as high as 1300�C with

ceramic heat exchanger materials [14], but prolonged operationwith such exchangers is yet to be firmly tested. Increase in theturbine inlet temperature is favorable towards achieving higherplant efficiency but it complicates the equipment design. Anuncooled micro gas turbine can sustain a maximum turbine inlettemperature of 950

�C, while further increase in the temperature

requires turbine blade cooling arrangement [13]. Since all thesemodifications towards performance improvement bear consider-able cost implications, such modifications always needs a priorievaluation, based on energy and exergy based performance anal-ysis of the cycle.

In the present work, we have conducted the energy and exergybased performance analysis of an externally fired gas turbine cyclerunning on biomass as fuel. The effects of operating parameters,like pressure ratio, turbine inlet temperature, heat exchanger coldend temperature difference, on the thermal efficiency and specificair flow for the cycle have been analyzed. The main focus of thepresent study is to identify the ideal operating parameters for theuse of a EFGT plant for decentralized power generation supplyingthe local needs in the remote areas, where extending the gridpower is uneconomic. Accordingly, the performance parameters fora 100 kW gas turbine plant have been evaluated with selective sets

A. Datta et al. / Energy 35 (2010) 341–350 343

of operating conditions. An integral gasifier has been considered inthe cycle for the gasification of the biomass fuel prior to itscombustion. This is because of the fact that the operation andcontrol of a direct biomass combustor (like a CFB combustor as in[10]) at a small scale (as considered here) involves additionalcomplexities and more number of skilled personnel that isunavailable in remote areas at low cost. On the other hand, there isdeveloped technology of biomass gasifier at small scale [15] whichcan be integrated with the proposed gas turbine plant. An exergybased accounting has been performed for the cycle to find out themajor irreversibilities in the cycle. The exergetic efficiencies of theindividual plant equipment are also compared at different cycleoperating conditions.

2. Theoretical formulation

2.1. Description of the proposed cycle

Fig. 1(a) illustrates the schematic description of the externallyfired gas turbine cycle analyzed, while Fig. 1(b) shows the differentprocesses on a temperature-entropy (T-s) plane. In the power cycle,the ambient air is compressed in a centrifugal compressor over thepressure ratio (rp) of the cycle. A part of the air is extracted from anintermediate stage of the compressor for the gasification of the

T

s

1

2 3

4

56

A

CT

CC

G

HE

Biomass

B

1

3

4

5

62

T5

T1

P2

P1+hHEPΔ

CETD

P1

P2-cHEPΔ

P1+hHEPΔ +ΔPCC

T3 (=TIT)

Air

ExhaustGas

a

b

Fig. 1. (a) Schematic diagram and (b) Temperature-Entropy diagram of the EFGT cycle.C-Compressor, CC- Combustion Chamber, HE-Heat Exchanger, G-Gasifier, T-Turbine,TIT-Turbine Inlet Temperature, CETD-Cold End Temperature Difference of HeatExchanger.

biomass stock, while the remaining air undergoes the fullcompression. The compressed air is then heated in an indirect heatexchanger before entering the turbine. After expansion in theturbine, the air is fed into the combustion chamber, where theproducer gas, generated from gasification of the biomass, is burnt.The high temperature products gas of combustion is then passedthrough the heat exchanger in order to heat the air, and finallyreleased into the atmosphere.

The following assumptions have been made for the analysis ofthe cycle:

1. Air is admitted to the compressor (state 1, refer Fig. 1b) atatmospheric condition, P1¼101.325 kPa, T1¼300 K.

2. The compression process is adiabatic with an isentropicefficiency of 87% [9].

3. The gasification process is adiabatic and chemical equilibriumis reached in the producer gas at the gasifier exit. A totalpressure drop (DPG) of 16 mm Hg column (i.e. 2.13 kPa) [16] isconsidered across the gasifier.

4. The ultimate analysis of the dry biomass fuel (wood) showsa gravimetric composition of C: 50%, H: 6% and O: 44%, whilethe calorific value of the biomass (on dry basis) is 449568 kJ/kmol (i.e. 18732 kJ/kg) [17].

5. The moisture content in the biomass is 20% on mass basis.6. The pressure drop in heat exchanger cold side is 3% of the inlet

pressure, while on the hot side the pressure drop is 1.5% of theinlet pressure [12].

7. The expansion process in the turbine is adiabatic with anisentropic efficiency of 89% [9].

8. Complete combustion takes place in the combustion chamberunder adiabatic condition. A pressure drop of 0.5% of the inletpressure takes place across the combustion chamber.

9. The reference temperature To and pressure Po are 25�C and

101.325 kPa, respectively.

2.2. Energy analysis of the cycle

2.2.1. Air compressorThe compressor delivery pressure (P2) is evaluated using the

cycle pressure ratio (rp), which is varied in the range of 2–8. Thetemperature of air (T2s) at pressure P2 for the isentropic compres-sion is calculated considering a third order polynomial variation ofthe molar specific heat of air with temperature as,Cpair ¼ aair þ bairT þ cairT2 þ dairT3 [18]. The actual work done onthe compressor per kmol of admitted air (wC) is calculated using theisentropic efficiency of the compressor (hc,isen).

The specific compressor work can be expressed as:

wC ¼ZT2

T1

Cpair dT

¼ aairðT2 � T1Þ þ�

bair

2

��T2

2 � T21

�þ�cair

3

��T3

2 � T31

�

þ�

dair

4

��T4

2 � T41

�(1)

Eq. (1) is solved by Newton–Raphson method for the actualtemperature of air (T2) at the compressor outlet.

The air, extracted at the intermediate state point ‘A’ from thecompressor, is used for the gasification of the biomass stock. Thepressure PA should be sufficient to overcome the pressure dropacross the gasifier and feed the producer gas to the combustionchamber. Therefore,

Table 1Product gas composition from gasification of rubber wood using the present modeland from the works of Jayah et al. [20] and Sharma [21].

Dry Gas Composition Jayah et al. [20]Experiment

Sharma [21]EquilibriumModel

PresentModel

Moisture content¼ 16%, A-F Ratio¼ 2.2H2 18.3 19.35 18.97CO 20.2 19.34 24.75CO2 9.7 11.18 8.01CH4 1.1 0.25 0.39N2 50.7 50.19 47.88

Moisture content¼ 18.5%, A-F Ratio¼ 2.03H2 17.2 19.85 20.91CO 19.6 19.64 23.79CO2 9.9 11.01 9.25CH4 1.4 0.26 0.99N2 51.9 49.26 45.06

A. Datta et al. / Energy 35 (2010) 341–350344

PA ¼ P1 þ�

DPhHE þ DPCC

�þ DPG (2)

where, DPhHE , DPCC DPG are the pressure drops across the heat

exchanger hot side, combustion chamber and gasifier respectively.The temperature of extracted air (TA) is obtained using a similarapproach for T2 described above.

2.2.2. GasifierThe biomass feedstock (wood) is fed to the gasifier in a sub-

stoichiometric environment. The gasifier environment is describedby the equivalence ratio (f), which is defined as the ratio of stoi-chiometric air-fuel ratio to actual air-fuel ratio. We have consideredf¼ 3.33 for our calculation.

A representative chemical formula is considered for the drybiomass fuel as CHQOR, using the mass percentage of carbon,hydrogen and oxygen, respectively, from the ultimate analysis ofthe fuel [19].

The number of moles of oxygen for the gasification of 1 kmol ofdry biomass (CHQOR) is calculated as

XO ¼AFSt,Mf

4:76,Maf(3)

where, AFSt is the stoichiometric air-fuel ratio for the fuel used andMf and Ma are the molecular weights of fuel and air, respectively.

The amount of moisture (in kmol) fed with every kmol of dryfeedstock is

B ¼ ðZ=ð100� ZÞÞ,�

Mf =Mw

�(4)

where, Z is the moisture content (mass percentage) in thebiomass (as-fired) and Mw is the molecular weight of the watervapour.

The global gasification reaction can be expressed as follows:

CHQ OR þ BH2Oþ XOO2 þ 3:76XO,N2

¼ X1H2 þ X2COþ X3CO2 þ X4H2Oþ X5CH4 þ 3:76XO,N2

(5)

where X1, X2, X3, X4 and X5 are the number of moles of H2, CO, CO2,H2O, CH4, respectively, produced on gasification of one kmol ofwood.

The values of X1 through X5 are solved considering the carbon,hydrogen and oxygen balances from the chemical reaction (Eq. (5))and the chemical equilibrium in the product gas following themethanation reaction and water gas shift reaction [17,19] as below:

C þ 2H24CH4 (6)

COþ H2O4CO2 þ H2 (7)

The equilibrium constants for methanation reaction (K1) andwater gas shift reaction (K2) are expressed as follows:

K1 ¼PCH4

=Po�PH2

=Po�2 ¼

X5Po

X5

j¼1

Xj

X1P4(8)

K2 ¼�PCO2

=Po��

PH2=Po�

ðPCO=PoÞ�PH2O=Po

� ¼ X1X3

X2X4(9)

In Eqs. (8) and (9), Pi represents the partial pressure of species i,while Po is the reference pressure. P4 is the pressure at the gasifierexit (which is equal to the combustor pressure). The equilibriumconstants K1 and K2 depend on the gasification temperature. An

energy balance equation is drawn to evaluate the gasificationtemperature (Tg) (assuming no heat loss from the gasifier) asfollows:

hfwood þ BhfH2Oþ XO

0B@ZTA

To

CpO2dT

1CAþ 3:76XO

0B@ZTA

To

CpN2dT

1CA

¼ X1

0B@hfH2

þZTg

To

CpH2dT

1CAþ X2

0B@hfCO

þZTg

To

CpCO dT

1CA

þ X3

0B@hfCO2

þZTg

To

CpCO2dT

1CAþ X4

0B@hfH2O

þZTg

To

CpH2O dT

1CA

þ X5

0B@hfCH4

þZTg

To

CpCH4dT

1CAþ 3:76XO

0B@ZTg

To

CpN2dT

1CA ð10Þ

where hfwood, hfH2 O, hfH2

, hfCO, hfCO2

and hfCH4represent enthalpies of

formation of wood, moisture, hydrogen, carbon monoxide, carbondioxide and methane, respectively. The enthalpy of formation ofwood has been derived from the heating value of the fuel.

The composition and temperature of the producer gas areobtained by solving for the values of X1 through X5 and Tg

simultaneously. Table 1 shows the product gas concentration ongasification of rubber wood using the present gasifier modelunder two different moisture content and air-fuel ratio. Thecomposition of the biomass is taken from the earlier work of Jayahet al. [20] and Sharma [21], who worked with gasification ofbiomass. The corresponding results from the experiments of Jayahet al. [20] and the equilibrium model of Sharma [21] under thesame conditions are also given for comparison. The results showthat the present gasifier model predicts the gas composition fairlywell.

2.2.3. TurbineThe turbine inlet pressure P3 is calculated considering a 3%

pressure drop from the compressor exit (P2) in the cold side ofthe heat exchanger. The temperature of air at the inlet to theturbine (T3) is an input parameter for the analysis. The actualwork done by the turbine per kmol of air (wT) is evaluatedconsidering the variable specific heat and the isentropic turbineefficiency (ht,isen). The actual temperature of air at the turbineoutlet (T4) is found by solving the following equation usingNewton-Raphson method.

A. Datta et al. / Energy 35 (2010) 341–350 345

wT ¼ZT3

T4

Cpair dT¼aairðT3�T4Þþ�

bair

2

��T2

3�T24

�þ�cair

3

��T3

3�T34

�

þ�

dair

4

��T4

3�T44

�(11)

2.2.4. Combustion chamberThe combustion chamber is fed with the air from the turbine

exhaust and the producer gas from the gasifier. Completecombustion has been assumed in the combustion chamberfollowing the chemical equation

ðX1H2 þ X2COþ X3CO2 þ X4H2Oþ X5CH4 þ 3:76XON2Þþ X0OðO2 þ 3:76N2Þ/X6CO2 þ X7H2Oþ X8O2

þ 3:76�XO þ X0O

�N2 (12)

X0O denotes the kmoles of O2 entering the combustor from theturbine for each kmol of wood fired in the cycle. Assuming anadiabatic condition in the combustor, the energy balance is givenas:-

Xj

XjðhfjþZTg

To

Cpj dT�����Fuel

þX

j

XjðhfjþZT4

To

Cpj dT�����Air

¼X

j

XjðhfjþZT5

To

Cpj dT�����Products

(13)

where, Xj represents the number of mole of the jth component infuel (the producer gas), air or product gas mixture and hfj

and Cpj

are the heat of formation and temperature dependent specific heatof that component. Putting the number of moles of differentcomponents from Eq. (12), it is found that Eq. (13) reduces to oneinvolving T5 and X0O.

2.2.5. Heat exchangerThe hot combustion gases leaving the combustor enters the heat

exchanger at state 5 and leaves at state 6, heating the compressedair from state 2 to state 3 without any heat loss to the surrounding(See Fig. 1).

A heat balance across the heat exchanger gives

4:76X0O

ZT3

T2

CpairdT ¼ Xg

ZT5

T6

Cpg dT (14)

where, Xg represents the number of moles of hot exhaust gasesleaving the combustor. Following the reaction Eq. (12),

Xg ¼ X6 þ X7 þ X8 þ 3:76�XO þ X0O

�(15)

Cpg represents the molar specific heat of the exhaust gas mixtureentering the heat exchanger.

Equations (13) and (14), representing the energy balance of thecombustor and the heat exchanger, are simultaneously solved usingan iterative technique to obtain the values of T5 and X0O.

2.3. Exergy analysis of the cycle

Since the power cycle involves the gasifier and the combustor,both the thermomechanical exergy and chemical exergy are

considered in the analysis. The thermomechanical exergy is definedwith respect to a restricted dead state, which is characterized by thereference pressure and temperature of the dead state. The specificthermomechanical flow exergy at any state is calculated from thegeneralized equation given as follows:

ei ¼ hi � ho � Toðsi � soÞ (16)

where ‘i’ represents the state point (e.g. 1 through 6 and A, as givenin Fig. 1) at which the exergy is evaluated and ‘o’ is the state point atthe exergy reference environment.

hi � ho ¼ZTi

To

CpdT (17)

si � so ¼ZTi

To

CpdTT� Rln

Pi

Po(18)

The chemical exergy is defined with respect to the true deadstate, which considers the chemical composition of the referenceenvironment in addition to the reference pressure and temperature[22]. In true thermodynamic sense a multicomponent systempossesses chemical exergy at restricted dead state when the partialpressure of the components in the system differs from the partialpressure of the same components in the reference environment.However, in the combustion literature chemical availability isassociated when useful work could be extracted through chemicalreaction [23,24] at reference temperature and pressure conditions.We have followed the latter concept in evaluating chemical exergyin this work.

The chemical exergy of the wood (echwood) is obtained from its lower

heating value using a multiplication factor b [25], which is given by

b ¼1:044þ 0:0160

HC� 0:34493

OC

�1þ 0:0531

HC

�

1� 0:4124OC

(19)

The producer gas from the gasifier possesses chemical exergy(ech

B ) in addition to the thermomechanical exergy (eB, which is dueto the elevated pressure and temperature of the gas and the mixingof the constituents). The specific chemical exergy of the producergas is given by

echB ¼

X1

X1 þ X2þX3 þ X4 þ X5 þ 3:76XOech

H2

þ X2

X1 þ X2 þ X3 þ X4 þ X5þ3:76XOech

CO

þ X5

X1 þ X2 þ X3 þ X4 þ X5 þ 3:76XOech

CH4(20)

where, echH2

echCO and ech

CH4represent the specific chemical exergy of H2,

CO and CH4, respectively [22].

2.4. Performance parameters

Finally, the cycle performance parameters have been evaluatedbased on one kmole of dry biomass fed to the plant. The actual workdone on the compressor is expressed as,

Table 3Parameters for the analysis of EFGT cycle in the present work.

Biomass Analysis (by mass) on dry basis [17]

Carbon 50%Hydrogen 6%Oxygen 44%Calorific Value 449568 kJ/kmol

(18732 kJ/kg)Moisture content in the biomass by mass 20%

Properties of Ambient AirPressure 101.325 kPaTemperature 300 KComposition (by vol.)Nitrogen 79%Oxygen 21%

Equipment performanceIsentropic efficiency of compressor (hc,isen) 87%Isentropic efficiency of turbine (ht,isen) 89%Pressure drop across the gasifier (DPG) 16 mm Hg columnPressure drop at heat exchanger

cold side ðDPcHEÞ as % of inlet pressure

3

Pressure drop at heat exchangerhot side ðDPh

HEÞ as % of inlet pressure1.5

Pressure drop across combustor (DPCC)as % of inlet pressure

0.5

Operating parameters with rangeEquivalence ratio at the gasifier (4) 3.33Compressor Pressure ratio (rp) 2–8Turbine inlet temperature (TIT) 1050–1350 KHeat exchanger cold end temperature

difference (CETD)200–300 K

A. Datta et al. / Energy 35 (2010) 341–350346

WC ¼ 4:76X0O

ZT2

T1

Cpair dT þ 4:76XO

ZTA

T1

Cpair dT ; (21)

While the actual work done by the turbine is

WT ¼ 4:76X0O

ZT3

T4

Cpair dT (22)

The thermal efficiency (hth) of the EFGT cycle is obtained usingthe turbine and compressor work and the calorific value of the fuel.

The energy delivered with the exhaust gas from the cycle, whichcan be subsequently recovered as waste heat in a downstreamprocess is

EN ¼X

j

Xj

ZT6

To

Cpj dT (23)

where, Xj is the number of moles of the jth species on the productside of Eq. (12) and Cpj is the respective specific heat.

The exergy input into the plant (ein) for every mole of biomassfed to the cycle is given as ðech

wood þ 4:76fXO þ X0O e1Þg . A part of theinput exergy is actually converted into useful work, while the otherparts are lost with the exhaust gas and are destroyed due to theirreversibilities in different components of the plant. The usefulexergy and the exergy lost as fractions of the input exergy areðWT�WCÞ

einand EN

einrespectively. The former also represents the exer-

getic efficiency of the cycle. In addition to these, exergy has beendestroyed in each of the components of the cycle due to processirreversibilities. The expressions of exergy destruction in the indi-vidual components of the cycle are presented in Table 2.

In addition to the exergy destruction, the expressions of theexergetic efficiency of the individual components are also evalu-ated as indicators of their deviation from ideality, while operatingbetween the corresponding thermodynamic states. The expres-sions of exergetic efficiency of the individual plant components arealso shown in Table 2.

3. Results and discussion

3.1. Influence of the key operating parameters on cycle performance

The integrated model has been used to evaluate the perfor-mance of an EFGT cycle at different operating conditions. Aperformance comparison is eventually made with reference toa 100 kW unit for distributed power generation. Simple operation

Table 2Exergy destruction and second law efficiency of individual component of the EFGT plant

Component Exergy Destruction

Compressor 4:76½ðXO þ X0OÞe1 � XOeA � X0Oe2�

Gasifier jwood þ 4:76XOeA � XBðechB þ eBÞ

where, XB¼ X1þ X2þ X3þ X4þ X

Turbine 4.76X0O (e3� e4)�WT

Combustor 4:76X0Oe4 þ XBðechB þ eBÞ � Xge5

Heat Exchanger Xg (e5� e6)þ 4.76X0O (e2� e3)

and low cost are the two key factors in choosing the plant operatingparameters for distributed generation in remote areas. In this effort,both the thermal performance and sizing of the plant are taken intoaccount. The former is represented by the thermal efficiency of theplant and is an indicator of the operating cost (fuel cost) fora particular plant capacity. The plant size is compared on the basisof specific air flow (i.e. air flow per unit energy output) through theturbine. Lower value of the specific air flow indicates smaller size ofthe plant equipment and lower capital cost. Table 3 summarizes theoperating parameters based on which the performance analysis hasbeen performed here. The estimated producer gas temperaturefrom the gasifier at the corresponding conditions is 1006 K. Theinfluence of three salient operating parameters, viz., the pressureratio of the cycle (rp), turbine inlet temperature (TIT) and the heatexchanger cold end temperature difference (CETD) are the threecritical operating parameters, whose influence on the cycleperformance are investigated.

Fig. 2 shows the variation in the cycle thermal efficiency withthe pressure ratio at three different turbine inlet temperatures, viz.1050 K, 1200 K and 1350 K. Still higher turbine inlet temperature is

.

Exergetic efficiency

þWC 4:76½XOeA þ X0Oe2 � ðXO þ X0OÞe1�WC

;

5þ 3.76XO

XBðechB þ eBÞ

echwood þ 4:76XOeA

WT

4:76X0Oðe3 � e4Þ

Xge5

4:76X0Oe4 þ XBðechB þ eBÞ

Xge6 þ 4:76X0Oe3

Xge5 þ 4:76X0Oe2

0

0.1

0.2

0.3

0.4

0.5

0 2 4 6 8 10

TIT=1050 KTIT=1200 KTIT=1350 K

rp

Fig. 2. Variation of thermal efficiency (h) of the EFGT cycle with the pressure ratio (rp)at different turbine inlet temperatures (TIT).

0

10

20

30

40

50

0 2 4 6 8 10

TIT=1050 K

TIT=1200 K

TIT=1350 K

Spec

ific

air

flo

w b

y m

ass

(kg/

kWh)

rp

0 2 4 6 8 10

rp

0

10

20

30

40

50

60

70

TIT=1050 K

TIT=1200 K

TIT=1350 K

Spec

ific

air

flo

w b

y vo

lum

e (m

3 /kW

h)

a

b

Fig. 3. (a) Variation of specific air flow by mass (kg/kWh) with pressure ratio (rp) forthe EFGT cycle at different turbine inlet temperatures (TIT). (b) Variation of specific airflow by volume (m3/kWh) with pressure ratio (rp) for the EFGT cycle at differentturbine inlet temperatures (TIT).

A. Datta et al. / Energy 35 (2010) 341–350 347

possible in today’s gas turbine technology [26]. However, it requiresexpensive turbine materials and extensive cooling arrangement forthe turbine blades, thereby increasing the capital cost as well as thecomplexity of operation. It is observed that for a particular turbineinlet temperature, the efficiency first increases with the increase inpressure ratio to attain a maximum value and then decreases withfurther increase in the pressure ratio. On the other hand, higherturbine inlet temperature ensures higher thermal efficiency at allpressure ratios. As a result, the efficiency peaks of 24.3%, 29.7% and34.4% are obtained for TITs of 1050 K, 1200 K and 1350 K, respec-tively. The maximum efficiency is reached in the pressure ratiorange of 3–4 for the three TITs considered here.

The variation in the pressure ratio influences the specific airconsumption in the cycle and therefore the size of the plantcomponents. The pressure at the inlet to the turbine is different atdifferent pressure ratios, while at the exit of the turbine the pres-sure remains the same for all the cases. At the high pressure end ofthe turbine the size at different conditions are compared using thespecific air consumption by volume, while for the low pressure endthe specific consumption by mass determines the size. Also, anincrease in the pressure increases the metal thickness of theequipment casing walls, increasing their weight and cost. Fig. 3ashows the variation of the specific air flow by mass (kg/kWh)entering the turbine against the pressure ratio (rp) at differentturbine inlet temperature. It is observed that at a particular turbineinlet temperature the specific air mass flow first decreases with theincrease in the pressure ratio. The decrease in mass flow is found tobe rapid at the lower end of pressure ratio and gradually decreasesas the pressure ratio is increased. Beyond a pressure ratio value themass flow begins to increase with further increase in pressure ratio.The pressure ratio at which the reversal in the trend of mass flowvariation occurs is lower at lower value of turbine inlet temperature(for TIT¼ 1050 K the reversal occurs at rp¼ 7.0, while forTIT¼ 1200 K and 1350 K this reversal is not observed till rp¼ 8.0). Itis also important to note from Fig. 3a that the increase in theturbine inlet temperature decreases the specific air consumption atthe turbine inlet.

The variation in the specific air consumption by volume (m3/kWh) at the inlet to the turbine with changing pressure ratio, atconstant turbine inlet temperature, is shown in Fig. 3b. Thedecrease in the specific air consumption by volume is monotonicin this case with the increase in rp. While the decrease is veryrapid at the lower end of the pressure ratio range, the incremental

change becomes less at higher rp. The specific volume flow of air atthe turbine inlet is guided by the pressure and the specific massflow rate at a given TIT. At low values of rp the pressure remainslow and the specific mass flow is high, both contributing to thefact that in the small rp regime a reduction in pressure ratiosharply increases the specific volume flow of air. At higher valuesof rp, the changes are much flatter since the specific mass flowcurves are nearly flat, and the pressure is high. For TIT¼ 1050 Kand 1200 K, although the mass flow of air observes a gradualincrease with the increase in rp, such an increase is masked theeffect of increasing pressure, and the volume flow continues todecrease (though only at a slow rate).

Therefore, as observed from Figs. 3a and b, the specific mass andvolume flow rates of air are high at low values of rp (e.g. at rp¼ 2.0).Both the values decrease rapidly till rp increases to about 5.0.Further increase in rp levels-off the specific mass flow of air andleads to a marginal decrease in the specific volume flow, but theincreased pressure warrants thicker walls for the high pressurecomponents of the cycle. Therefore, though there may bea marginal advantage in the reduction in volume flow rate (andhence the plant size) beyond a particular pressure ratio, the higherwall thickness will offset the cost benefit.

0

0.1

0.2

0.3

0.4

0 2 4 6 8 10

CETD=200 K

CETD=250 K

CETD=300 K

rp

Fig. 4. Variation of thermal efficiency (h) of the EFGT cycle with the pressure ratio (rp)at different heat exchanger cold end temperature difference (CETD).

Table 4Performance parameters of 100 kW Biomass fired EFGT plant at different operatingconditions.

Case 1: rp¼ 4,TIT¼ 1200 KCETD¼ 200 K

Case 2: rp¼ 4,TIT¼ 1350 KCETD¼ 200 K

Case 3: rp¼ 4,TIT¼ 1200 KCETD¼ 300 K

Fuel (biomass)flow rate, kg/s

0.0216 0.0186 0.0265

Air flow rate, kg/s 0.5867 0.4712 0.5942Thermal Efficiency, % 29.68 34.33 24.18Exhaust heat, kW 238.92 193.09 315.65Rate of heat exchange

across heatexchanger, kW

450.5 440.03 450.72

Heat Exchanger hot endtemperaturedifference, K

80.0 44.2 147.0

LMTD, K 131.0 103.2 214.5(UAHE)overall for heat

exchanger, W/K3.44 4.26 2.10

A. Datta et al. / Energy 35 (2010) 341–350348

The heat exchanger is one of the most critical equipment in theEFGT cycle. Considering the cost of the material for the hightemperature heat exchanger its size requires to be optimized.However, the design of the heat exchanger also influences thethermal performance of the power cycle, by influencing the exhaustgas loss from the cycle. Fig. 4 shows the variation in the cycle thermalefficiency with pressure ratio at different cold end temperaturedifference (CETD) of the heat exchanger for a particular turbine inlettemperature. The results show the same trend in the variation ofefficiency with pressure ratio at all the CETD values, with themaximum efficiency reached at an optimum pressure ratio. Theoptimum pressure ratio is found to be 4.0 for the three differentCETD values of 200 K, 250 K and 300 K considered. However, withthe increase in the CETD at a particular pressure ratio, the efficiencyis found to decrease. When the CETD is high more amount of theenergy is wasted through the exhaust gas stream, reducing the network produced in the cycle. In fact for particular rp and TIT, the statepoints 2, 3 and 4 shown in Fig. 1 do not change with the variation ofCETD. However, the variation in the temperatures across the heatexchanger changes the quantity of air flow governed by the energybalance across the heat exchanger. It is observed that the number ofmoles of air flowing through the turbine per unit mole of drybiomass feed (X0O) decreases with the increase in the CETD. Thevariation in CETD does not change the specific air flow rate throughthe turbine as the corresponding state points remain identical.

Based on the above discussion, it can be proclaimed that thecycle thermal efficiency is maximized in the rp range of 3–4,depending on the TIT and CETD. At the low pressure ratio of 2–3, thesize of the equipment will be large because of the high value of thespecific air flow. Conversely, a high pressure ratio increases the wallthickness of the equipment, thereby increasing the cost and weight.Considering all these facts, we have chosen rp¼ 4.0 as the optimumvalue of the pressure ratio for the EFGT cycle. Two different turbineinlet temperatures (1200 K and 1350 K) and two different CETDvalues (200 K and 300 K) are chosen to compare the performance.Accordingly, three sets of cycle operating conditions with differentturbine inlet temperatures (TIT) and heat exchanger cold endtemperature differences (CETD) have been identified (as given inTable 4) to compare the performance of a 100 kW EFGT plant.

3.2. Energy and exergy based analysis of a 100 kW biomass firedEFGT plant

Table 4 lists the important performance parameters for the threecases for a 100 kW EFGT based micro gas turbine plant running at full

load. It is seen from Table 4 that when the TIT is 1350 K and CETD is200 K (Case 2), the thermal efficiency of the plant attains the highestvalue of 34.33%. Accordingly, the fuel flow rate and the exhaust heatloss are the lowest. The air flow rate is also the lowest for this case,indicating smaller size of the components, like compressor andturbine. On the other hand, the logarithmic mean temperaturedifference (LMTD) of the heat exchanger based on the temperaturedifferences at the hot and cold ends is low, giving a high overall(UAHE) value for the heat exchanger, where U and AHE are the overallheat transfer coefficient and the heat transfer surface area of the heatexchanger, respectively. If we consider a nearly constant value of theoverall heat transfer coefficient (U) for all the cases, then case 2performance data calls for the largest size of the heat exchanger. Thehigher turbine inlet temperature and the increased size of the heatexchanger required for this case is indicative of a high cost of theplant.

In case when the TIT is 1200 K and the CETD 300 K (Case3) thethermal efficiency of the plant is the lowest (24.18%). The fuel flowrate and the exhaust heat loss are the maximum in this case. Thecorresponding air flow rate is also the highest among the three setscompared indicating a larger size of the turbine and compressor.While the heat exchanger LMTD for this case is high (214.5 K)indicating a smaller sized heat exchanger.

The operating parameters in case 1 offer a performance trade-off in terms of thermal efficiency and the heat exchanger size. Athermal efficiency of 29.68% has been achieved in this case. Theheat exchanger LMTD is 131 K giving overall UA as 3.44. Therefore,considering the capital and operating cost of the plant, case 1 is thebetter choice of plant operating condition.

A second law based performance analysis for the three caseshas been presented in Fig. 5, where the complete exergybalance has been made as fractions of the exergy input to thecycle. The fraction of the input exergy converted into usefulwork determines the exergetic efficiency of the cycle. Theremaining part of the input exergy is either lost in the exhaustheat or destroyed through irreversibilities in various compo-nents. It is observed from the results of the three cases that themaximum exergetic efficiency is attained in case 2, where theturbine inlet temperature is the highest. The exergy loss in theexhaust is the highest in case 3, where the exhaust gas leavesthe cycle at the maximum temperature (because of the highestCETD). Table 4 shows a comparison of the exhaust heat for thethree cases.

The major exergy destruction takes place in the gasifier,combustor and the heat exchanger, while the exergy destruction

=4, TIT=1200 K, CETD= 200 K

Exergy out23.54%

Compressor2.24%

Gasifier15.39%

Turbine2.28%

Combustor19.60%

Useful28.01%

Heat Exchanger8.94%

=4, TIT=1350 K, CETD=200 K

Useful32.40%

Exergy out22.01%

Compressor2.07%

Gasifier15.39%

Turbine2.11%

Combustor17.57%

Heat Exchanger8.45%

=4, TIT=1200 K, CETD= 300 K

Useful22.82%

Exergy out29.05%

Compressor1.83%

Gasifier15.39%

Turbine1.85%

Combustor18.98%

Heat Exchanger10.08%

Case 1: rp Case 2: rp

Case 3: rp

Fig. 5. Exergy balance of the EFGT cycle for the three different cases described in Table 4.

A. Datta et al. / Energy 35 (2010) 341–350 349

of the compressor and turbine are only a little. The fraction ofexergy destructed in the gasifier is the same in the three cases,since the operating parameters of the gasifier has been consideredidentical. A sizeable amount (15.39%) of the input exergy is des-tructed in the gasifier owing to the gasification reactions that takeplace there. The exergy destruction in the combustor is thehighest in all the three cases, amounting to 19.6%, 17.57% and18.98% of the input exergy, respectively. The destruction of exergyin the combustion chamber is due to the heat exchange betweenthe streams and chemical reactions that take place. Operating thecombustor at higher temperature and higher temperature of theair fed to the combustion chamber decrease the exergy destruc-tion in the combustor. Exergy destruction in the heat exchangerincreases when the temperature difference between the twostreams exchanging heat increases. Accordingly, the maximumfraction of the exergy destruction in heat exchanger occurs in case3, where the LMTD is also the highest. More than 10% of the inputexergy is destroyed in the heat exchanger for this case. For theother two cases (i.e. case 1 and case 2) the exergy destroyed in theheat exchanger are 8.94% and 8.45% of the input exergyrespectively.

Fig. 6 shows the exergetic efficiencies for the individualcomponents for the three cases. The individual exergetic effi-ciency value of the equipment indicates the deviation fromideality for the equipment operating across its respective ther-modynamic states. It is observed that the exergetic efficiency of

the compressor, turbine and heat exchanger remain above 90%,while those of the gasifier and the combustion chamber are less.The relatively lower exergetic efficiency in the gasifier and thecombustion chamber is attributed to the irreversibility pertainingto the chemical reactions occurring there. The exergetic efficiencyof the compressor is identical for all the three cases (91.5%), sinceit operates at same pressure ratio and isentropic efficiency.Similarly, the exergy efficiencies of the gasifier are the same forthe three cases as the operating pressure, gasifier equivalenceratio and the properties of the biomass are considered to be thesame. The exergetic efficiency for the turbine is the highest(96.4%) in case 2 where the turbine operates with the highestinlet temperature. For this condition, the air temperature at theturbine outlet also remains higher than the other conditions. Asa result, the combustion chamber operates with the maximum airpreheat in case 2. The flame temperature in the combustor alsobecomes the maximum in this case. As the chemical reactionoccurs at high temperature the associated irreversibility becomesless and the combustion chamber exergetic efficiency attains themaximum value for case 2. The exergetic efficiency of the heatexchanger largely depends on the mean temperature differencebetween the streams across the exchanger. Lower meantemperature difference is indicative of lower irreversibilities. Thisis evident in the result as the heat exchanger in case 2 (the casewith lowest LMTD among the three) shows the highest exergeticefficiency.

0.0

0.2

0.4

0.6

0.8

1.0

Compressor

Gasifier

Turbine

Combustor

Heat Exchanger

ε

rp=4, TIT=1200 K,CETD=200 K

rp=4, TIT=1350 K,CETD=200 K

rp=4, TIT=1200 K,CETD=300 K

Fig. 6. Exergetic efficiency of individual components in the EFGT cycle for the threedifferent cases described in Table 4.

A. Datta et al. / Energy 35 (2010) 341–350350

4. Conclusions

A thermodynamic analysis has been performed for an externallyfired gas turbine (EFGT) cycle with an integrated biomass gasifier.The effects of operating parameters like pressure ratio (rp), turbineinlet temperature (TIT) and cold end temperature difference (CETD)of the heat exchanger on the thermal efficiency and specific air flowhave been studied. The thermal efficiency of the cycle is found to bewithin 16–34% for the range of operating parameters underinvestigation. The cycle thermal efficiency is the maximum at anoptimum pressure ratio of the cycle (in the range of 3–4) fora particular turbine inlet temperature and cold end temperaturedifference across the heat exchanger. At a particular pressure ratioof the cycle the thermal efficiency increases either with theincrease in the turbine inlet temperature or with the decrease inthe cold end temperature difference of the heat exchanger.

The specific air flow at the turbine inlet is evaluated to comparethe size of the plant equipment. It is found that the specific air flowby volume decreases with the increase in pressure ratio sharply atthe lower end of rp, while the incremental change is marginal athigh values of rp. The specific air flow by mass exhibits a rapiddecrease with the increase in rp at the lower end of rp, while thecurves become flatter and even rises gradually beyond a particularpressure ratio. The increase in the turbine inlet temperaturedecreases the specific air flow at the entry to the turbine. However,the cold end temperature difference across the heat exchanger doesnot affect the specific air flow.

Three different sets of operating parameters, each having rp¼ 4,have finally been considered for a detailed investigation of the

performance of a 100 kW plant running on EFGT cycle. The thermalperformance and sizing have been compared based on the thermalefficiency, air flow rate and heat transfer area of the heat exchanger.Moreover, an exergy balance has been carried out for each of thecases to account the useful exergy, exergy loss and exergydestruction. Major exergy destruction is found to occur in thegasifier, combustor and the heat exchanger.

Though the parameters in Case 2 (TIT¼ 1350 K, CETD¼ 200 K)offer a higher thermal efficiency and exergetic efficiency anda lower air flow rate, the heat exchanger size for this case is found tobe large. On the other hand, the heat exchanger size for the Case 3(TIT¼ 1200 K, CETD¼ 300 K) is small, but it gives the lowestthermal and exergetic efficiencies. A trade-off in performance isobserved for Case 1 (TIT¼ 1200 K, CETD¼ 200 K)

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