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

In this work, thermal performance and sizing of such plants have been analyzed at different cycle pressure ratio rp¼ 28, turbine inlet temperature TIT ¼ 1050–1350 K and the heat exchange

Energy and exergy analyses of an externally fired gas turbine (EFGT) cycle

integrated with biomass gasifier for distributed power generation

Department of Power Engineering, Jadavpur University, Salt Lake Campus, Kolkata 700098, India

Article history:

Received 9 January 2009

Received in revised form

1 August 2009

Accepted 25 September 2009

Available online 28 October 2009

Keywords:

Gas turbine

External firing

Biomass

Gasifier

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 at different cycle pressure ratio (rp¼ 28), turbine inlet temperature (TIT ¼ 1050–1350 K) and the heat exchanger cold end temperature difference (CETD ¼ 200–300 K) It is found that the thermal efficiency of the EFGT plant reaches a maximum at an optimum pressure ratio depending upon the TIT and heat exchanger CETD For a particular pressure ratio, thermal efficiency increases either with the increase in TIT or with the decrease in heat exchanger CETD The specific air flow, associated with the size of the plant equipment, decreases with the increase in pressure ratio This decrease is rapid at the lower end of the pressure ratio (rp<4) but levels-off at higher rpvalues An increase in the TIT reduces the specific air flow, while a change in the heat exchanger CETD has no influence on it Based on this comparison, the performance of a 100 kW EFGT plant has been analyzed for three sets of operating parameters and

a trade-off in the operating condition is reached

Ó 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 the

areas, particularly in the hilly terrain, is extremely uneconomic[5]

On the contrary, the installation of small capacity plants catering to

the local needs using the local resource can be an attractive

alter-native for remote places Biomass is one of the important available

primary resources, which generally exists in abundance in the

villages and already serves as the source of energy e.g in cooking

Energy from the biomass can be thermochemically recovered

for the generation of electricity either through direct combustion or

through gasification and subsequent combustion of the producer

gas In large scale, biomass gasification can be used for power

engines or micro gas turbines are suitable for small capacity

distributed generation Producer gas can be used in conventional

diesel engines in the dual fuel mode or in producer gas engines for

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

The use of biomass as fuel in conventional (internally fired) gas

turbines are sensitive machines that require extremely clean gas to avoid damage to the turbine blades (such as erosion, incrustation, and corrosion) and blockage of filters and fuel injectors This requires installation of expensive gas clean up system, consisting of scrubbers, ceramic filters, cyclones etc., at the gasifier outlet Secondly, the low calorific value of the producer gas, obtained from biomass gasification, necessitates a high fuel flow It calls for

a design modification in the combustor and the turbine inlet guide vanes, otherwise the change in the mass balance between the compressor and the turbine moves the compressor operating point towards surge[9] These problems are resolved, if the biomass can

be 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 is heated in a heat exchanger before admitting to the turbine The turbine essentially handles clean air and the turbine exhaust air is subsequently used to burn the fuel in a combustion chamber The combustion product is employed as the hot stream of the heat exchanger, before being released from the power cycle The cycle can employ dirty and low cost fuels, as the combustion products do

* Corresponding author Tel.: þ91 33 23355813; fax: þ91 33 23357254.

E-mail address: amdatta_ju@yahoo.com (A Datta).

Contents lists available atScienceDirect

Energy

j o u r n a l h o m e p a g e : w w w e l s e v i e r c o m / l o c a t e / e n e r g y

0360-5442/$ – see front matter Ó 2009 Elsevier Ltd All rights reserved.

not enter the turbine Although the presence of ash in the products

may cause erosion and fouling of the heat exchanger tubes, while

corrosive products eats away the tube material, maintenance of the

heat exchanger is much less troublesome than that for the turbine

of closed and open cycle externally fired gas turbine plants with

direct combustion of biomass in a circulating fluidized bed furnace

It is found that the efficiency reaches a maximum value at an

compared the thermodynamic performance of an externally fired

gas turbine cycle with direct combustion of biomass against an

internally fired cycle firing either natural gas or producer gas from

biomass gasification The study was performed for the simple gas

turbine cycle as well as for the combined cycle operation with

a steam based Rankine cycle at the bottom The results showed

promising performance for the EFGT plant particularly considering

the renewable and environment-friendly attributes of the biomass

feasibility of the external firing of biomass in gas turbines The

authors concluded that cogeneration based on EFGT on the scale of

a 100 kW externally fired gas turbine plant fuelled with biomass

and having an integral dryer for biomass The influence of

param-eters like pressure ratio, turbine inlet temperature and temperature

difference in the heat exchanger on the thermal efficiency for

electrical generation was analyzed It was found that the dry

biomass produces efficiency in the range of 22–33% and the

inte-gration of the dryer improves flexibility in the plant operation

performance of an externally fired micro gas turbine pilot plant of

80 kW capacity fired with natural gas The paper demonstrated the

feasibility of operation and control of the gas turbine plant of small

capacity

All the literatures on EFGT universally claim that one of the biggest challenges in the design lies in developing the high temperature heat exchanger that is capable of achieving high turbine inlet temperature and at the same time withstands the stresses imposed by the working conditions and the constituent of

the cost of material are the two important considerations that decide the economy of the plant The use of nickel based super alloys in the heat exchanger allows the turbine inlet temperature to

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

with such exchangers is yet to be firmly tested Increase in the turbine inlet temperature is favorable towards achieving higher plant efficiency but it complicates the equipment design An uncooled micro gas turbine can sustain a maximum turbine inlet

modifications towards performance improvement bear consider-able cost implications, such modifications always needs a priori evaluation, based on energy and exergy based performance anal-ysis of the cycle

In the present work, we have conducted the energy and exergy based performance analysis of an externally fired gas turbine cycle running on biomass as fuel The effects of operating parameters, like pressure ratio, turbine inlet temperature, heat exchanger cold end temperature difference, on the thermal efficiency and specific air flow for the cycle have been analyzed The main focus of the present study is to identify the ideal operating parameters for the use of a EFGT plant for decentralized power generation supplying the local needs in the remote areas, where extending the grid power is uneconomic Accordingly, the performance parameters for

a 100 kW gas turbine plant have been evaluated with selective sets

Nomenclature

Greek Letters

hc,isen Isentropic efficiency of compressor

ht,isen Isentropic efficiency of turbine

Sub-scripts

Super-scripts

of operating conditions An integral gasifier has been considered in

the cycle for the gasification of the biomass fuel prior to its

combustion This is because of the fact that the operation and

control of a direct biomass combustor (like a CFB combustor as in

complexities and more number of skilled personnel that is

unavailable in remote areas at low cost On the other hand, there is

can be integrated with the proposed gas turbine plant An exergy

based accounting has been performed for the cycle to find out the

major irreversibilities in the cycle The exergetic efficiencies of the

individual plant equipment are also compared at different cycle

operating conditions

2 Theoretical formulation

2.1 Description of the proposed cycle

Fig 1(a) illustrates the schematic description of the externally

fired gas turbine cycle analyzed, whileFig 1(b) shows the different

processes on a temperature-entropy (T-s) plane In the power cycle,

the ambient air is compressed in a centrifugal compressor over the

pressure ratio (rp) of the cycle A part of the air is extracted from an

intermediate stage of the compressor for the gasification of the

biomass stock, while the remaining air undergoes the full compression The compressed air is then heated in an indirect heat exchanger before entering the turbine After expansion in the turbine, the air is fed into the combustion chamber, where the producer gas, generated from gasification of the biomass, is burnt The high temperature products gas of combustion is then passed through the heat exchanger in order to heat the air, and finally released into the atmosphere

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

1 Air is admitted to the compressor (state 1, refer Fig 1b) at atmospheric condition, P1¼101.325 kPa, T1¼ 300 K

2 The compression process is adiabatic with an isentropic efficiency of 87%[9]

3 The gasification process is adiabatic and chemical equilibrium

is reached in the producer gas at the gasifier exit A total pressure drop (DPG) of 16 mm Hg column (i.e 2.13 kPa)[16]is considered across the gasifier

4 The ultimate analysis of the dry biomass fuel (wood) shows

a gravimetric composition of C: 50%, H: 6% and O: 44%, while the 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 the inlet pressure[12]

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

8 Complete combustion takes place in the combustion chamber under adiabatic condition A pressure drop of 0.5% of the inlet pressure takes place across the combustion chamber

101.325 kPa, respectively

2.2 Energy analysis of the cycle 2.2.1 Air compressor

cycle pressure ratio (rp), which is varied in the range of 2–8 The temperature of air (T2s) at pressure P2for the isentropic compres-sion is calculated considering a third order polynomial variation of

Cp air ¼ aairþ bairT þ cairT2þ dairT3 [18] The actual work done on the compressor per kmol of admitted air (wC) is calculated using the isentropic efficiency of the compressor (hc,isen)

The specific compressor work can be expressed as:

wC ¼

ZT 2

T 1

Cp airdT

¼ aairðT2 T1Þ þ

b

air

2



T2 T2

þcair 3



T3 T3

þ

d

air

4



T4 T4

(1)

temperature of air (T2) at the compressor outlet

The air, extracted at the intermediate state point ‘A’ from the compressor, is used for the gasification of the biomass stock The

across the gasifier and feed the producer gas to the combustion chamber Therefore,

T

s

1

4

5 6

A

CC

G

HE

Biomass

B

1

3

4

5

6 2

T 5

T 1

P2

P1+ΔP HE h

CETD

P1

P2-ΔP HE c

P1+ΔP HE h +ΔPCC

T 3 (=TIT)

Air

Exhaust

Gas

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 Heat

PA ¼ P1þ

DPh

HEþDPCC

where, DPh

exchanger hot side, combustion chamber and gasifier respectively

The temperature of extracted air (TA) is obtained using a similar

2.2.2 Gasifier

The biomass feedstock (wood) is fed to the gasifier in a

sub-stoichiometric environment The gasifier environment is described

by the equivalence ratio (f), which is defined as the ratio of

stoi-chiometric air-fuel ratio to actual air-fuel ratio We have considered

f¼ 3.33 for our calculation

A representative chemical formula is considered for the dry

hydrogen and oxygen, respectively, from the ultimate analysis of

the fuel[19]

The number of moles of oxygen for the gasification of 1 kmol of

dry biomass (CHQOR) is calculated as

XO ¼ AFSt,Mf

where, AFStis the stoichiometric air-fuel ratio for the fuel used and

Mfand Maare the molecular weights of fuel and air, respectively

The amount of moisture (in kmol) fed with every kmol of dry

feedstock is

B ¼ ðZ=ð100  ZÞÞ,

Mf=Mw



(4)

where, Z is the moisture content (mass percentage) in the

vapour

The global gasification reaction can be expressed as follows:

CHQORþ BH2O þ XOO2þ 3:76XO,N2

¼ X1H2þ X2CO þ X3CO2þ X4H2O þ X5CH4þ 3:76XO,N2

(5)

where X1, X2, X3, X4and X5are the number of moles of H2, CO, CO2,

wood

The values of X1through X5are solved considering the carbon,

and the chemical equilibrium in the product gas following the

methanation reaction and water gas shift reaction[17,19]as below:

water gas shift reaction (K2) are expressed as follows:

K1 ¼ PCH4=Po



PH2=Po2 ¼

X5Po

X5

j ¼ 1

Xj

K2 ¼



PCO2=Po

PH2=Po

ðPCO=PoÞ

PH2O=Po ¼XX1X3

In Eqs.(8)and(9), Pirepresents the partial pressure of species i,

while Pois the reference pressure P4is the pressure at the gasifier

exit (which is equal to the combustor pressure) The equilibrium

energy balance equation is drawn to evaluate the gasification

follows:

hfwoodþ BhfH2Oþ XO

0

BZ

T A

T o

CpO2dT

1 C

A þ 3:76XO

0

BZ

T A

T o

CpN2dT

1 C

¼ X1

0 B

@hfH2þ

ZT g

T o

CpH2dT

1 C

A þ X2

0 B

@hf COþ

ZT g

T o

CpCOdT

1 C

þ X3

0 B

@hfCO2þ

ZT g

T o

CpCO2dT

1 C

A þ X4

0 B

@hfH2Oþ

ZT g

T o

CpH2OdT

1 C

þ X5

0 B

@hfCH4þ

ZT g

T o

CpCH4dT

1 C

A þ 3:76XO

0

BZ

T g

T o

CpN2dT

1

where hfwood, hf

CO2 and hf

CH4 represent enthalpies of formation of wood, moisture, hydrogen, carbon monoxide, carbon dioxide and methane, respectively The enthalpy of formation of wood has been derived from the heating value of the fuel The composition and temperature of the producer gas are

gasification of rubber wood using the present gasifier model under two different moisture content and air-fuel ratio The composition of the biomass is taken from the earlier work of Jayah

biomass The corresponding results from the experiments of Jayah

same conditions are also given for comparison The results show that the present gasifier model predicts the gas composition fairly well

2.2.3 Turbine

the heat exchanger The temperature of air at the inlet to the

considering the variable specific heat and the isentropic turbine efficiency (ht,isen) The actual temperature of air at the turbine

Newton-Raphson method

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

Dry Gas Composition Jayah et al [20]

Experiment

Sharma [21]

Equilibrium Model

Present Model Moisture content ¼ 16%, A-F Ratio ¼ 2.2

Moisture content ¼ 18.5%, A-F Ratio ¼ 2.03

ZT 3

T 4

CpairdT ¼ aairðT3T4Þþ

b

air

2



T2T2

þcair 3



T3T3

þ

d

air

4



T4T4

(11)

2.2.4 Combustion chamber

The combustion chamber is fed with the air from the turbine

exhaust and the producer gas from the gasifier Complete

combustion has been assumed in the combustion chamber

following the chemical equation

ðX1H2þ X2CO þ X3CO2þ X4H2O þ X5CH4þ 3:76XON2Þ

þ XO0ðO2þ 3:76N2Þ/X6CO2þ X7H2O þ X8O2

þ 3:76

XOþ XO0

X0

turbine for each kmol of wood fired in the cycle Assuming an

adiabatic condition in the combustor, the energy balance is given

as:-X

j

Xjðhfjþ

ZT g

T o

CpjdTÞ







Fuel

j

Xjðhfjþ

ZT 4

T o

CpjdTÞ







Air

j

Xjðhfjþ

ZT 5

T o

Cp jdTÞ







Products

(13)

fuel (the producer gas), air or product gas mixture and hfj and Cp j

are the heat of formation and temperature dependent specific heat

of that component Putting the number of moles of different

components from Eq.(12), it is found that Eq.(13)reduces to one

involving T5and X0

O 2.2.5 Heat exchanger

The hot combustion gases leaving the combustor enters the heat

exchanger at state 5 and leaves at state 6, heating the compressed

air from state 2 to state 3 without any heat loss to the surrounding

(SeeFig 1)

A heat balance across the heat exchanger gives

4:76XO0

ZT 3

T 2

CpairdT ¼ Xg

ZT 5

T 6

leaving the combustor Following the reaction Eq.(12),

Xg ¼ X6þ X7þ X8þ 3:76

XOþ XO0

(15)

entering the heat exchanger

Equations(13)and(14), representing the energy balance of the

combustor and the heat exchanger, are simultaneously solved using

an iterative technique to obtain the values of T5and X0

O

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 defined with respect to a restricted dead state, which is characterized by the reference pressure and temperature of the dead state The specific thermomechanical flow exergy at any state is calculated from the generalized equation given as follows:

where ‘i’ represents the state point (e.g 1 through 6 and A, as given

inFig 1) at which the exergy is evaluated and ‘o’ is the state point at the exergy reference environment

hi ho ¼

ZT i

T o

si so ¼

ZT i

T o

CpdT

T  Rln

Pi

The chemical exergy is defined with respect to the true dead state, which considers the chemical composition of the reference environment in addition to the reference pressure and temperature

possesses chemical exergy at restricted dead state when the partial pressure of the components in the system differs from the partial pressure of the same components in the reference environment However, in the combustion literature chemical availability is associated when useful work could be extracted through chemical

We have followed the latter concept in evaluating chemical exergy

in this work

The chemical exergy of the wood (ech

wood) is obtained from its lower heating value using a multiplication factorb[25], which is given by

b ¼

1:044 þ 0:0160H

C 0:34493

O C



1 þ 0:0531H

C



1  0:4124O

C

(19)

The producer gas from the gasifier possesses chemical exergy (ech

B) in addition to the thermomechanical exergy (eB, which is due

to the elevated pressure and temperature of the gas and the mixing

of the constituents) The specific chemical exergy of the producer gas is given by

X1þ X2þX3þ X4þ X5þ 3:76XOe

ch

H 2

X1þ X2þ X3þ X4þ X5þ3:76XO

ech CO

X1þ X2þ X3þ X4þ X5þ 3:76XOe

ch

where, ech

H 2ech

COand ech

CH 4represent the specific chemical exergy of H2,

CO and CH4, respectively[22]

2.4 Performance parameters Finally, the cycle performance parameters have been evaluated based on one kmole of dry biomass fed to the plant The actual work done on the compressor is expressed as,

WC ¼ 4:76XO0

ZT 2

T 1

Cp airdT þ 4:76XO

ZT A

T 1

While the actual work done by the turbine is

WT ¼ 4:76X0

O

ZT 3

T 4

The thermal efficiency (hth) of the EFGT cycle is obtained using

the turbine and compressor work and the calorific value of the fuel

The energy delivered with the exhaust gas from the cycle, which

can be subsequently recovered as waste heat in a downstream

process is

j

Xj

ZT 6

T o

where, Xjis the number of moles of the jth species on the product

side of Eq.(12)and Cp jis the respective specific heat

The exergy input into the plant (ein) for every mole of biomass

fed to the cycle is given as ðech

woodþ 4:76fXOþ X0

Oge1Þ A part of the input exergy is actually converted into useful work, while the other

parts are lost with the exhaust gas and are destroyed due to the

irreversibilities in different components of the plant The useful

exergy and the exergy lost as fractions of the input exergy are

ðW T W C Þ

e in andEN

e in respectively The former also represents the

exer-getic efficiency of the cycle In addition to these, exergy has been

destroyed in each of the components of the cycle due to process

irreversibilities The expressions of exergy destruction in the

In addition to the exergy destruction, the expressions of the

exergetic efficiency of the individual components are also

evalu-ated as indicators of their deviation from ideality, while operating

between the corresponding thermodynamic states The

expres-sions of exergetic efficiency of the individual plant components are

also shown inTable 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 A

performance comparison is eventually made with reference to

a 100 kW unit for distributed power generation Simple operation

and low cost are the two key factors in choosing the plant operating parameters for distributed generation in remote areas In this effort, both the thermal performance and sizing of the plant are taken into account The former is represented by the thermal efficiency of the plant and is an indicator of the operating cost (fuel cost) for

a particular plant capacity The plant size is compared on the basis

of specific air flow (i.e air flow per unit energy output) through the turbine Lower value of the specific air flow indicates smaller size of

operating parameters based on which the performance analysis has been performed here The estimated producer gas temperature from the gasifier at the corresponding conditions is 1006 K The influence of three salient operating parameters, viz., the pressure ratio of the cycle (rp), turbine inlet temperature (TIT) and the heat exchanger cold end temperature difference (CETD) are the three critical operating parameters, whose influence on the cycle performance are investigated

Fig 2shows the variation in the cycle thermal efficiency with the pressure ratio at three different turbine inlet temperatures, viz

1050 K, 1200 K and 1350 K Still higher turbine inlet temperature is

Table 2

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

O Þe 1  X O e A  X 0

O e 2  þW C 4:76½X O e A þ X 0

O e 2  ðX O þ X 0

O Þe 1 

W C

Gasifier jwood þ 4:76X O e A  X B ðe ch

B þ e B Þ;

where, X B ¼ X 1 þ X 2 þ X 3 þ X 4 þ X 5 þ 3.76X O

X B ðe ch

B þ e B Þ

e ch wood þ 4:76X O e A

Turbine 4.76X0O (e 3  e 4 )W T

W T

4:76X 0

O ðe 3  e 4 Þ

O e 4 þ X B ðe ch

4:76X 0

O e 4 þ X B ðe ch

B þ e B Þ Heat Exchanger X g (e 5  e 6 ) þ 4.76X0O (e 2  e 3 ) Xge6þ 4:76X

0

O e 3

X e þ 4:76X 0 e

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

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

Calorific Value 449568 kJ/kmol

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

Properties of Ambient Air

Composition (by vol.)

Equipment performance Isentropic efficiency of compressor (hc,isen ) 87%

Isentropic efficiency of turbine (ht,isen ) 89%

Pressure drop across the gasifier (DP G ) 16 mm Hg column Pressure drop at heat exchanger

cold side ðDP c

HE Þ as % of inlet pressure

3 Pressure drop at heat exchanger

hot side ðDP h

HE Þ as % of inlet pressure

1.5 Pressure drop across combustor (DP CC )

as % of inlet pressure

0.5

Operating parameters with range Equivalence ratio at the gasifier (4) 3.33 Compressor Pressure ratio (r p ) 2–8 Turbine inlet temperature (TIT) 1050–1350 K Heat exchanger cold end temperature

difference (CETD)

200–300 K

possible in today’s gas turbine technology[26] However, it requires

expensive turbine materials and extensive cooling arrangement for

the turbine blades, thereby increasing the capital cost as well as the

complexity of operation It is observed that for a particular turbine

inlet temperature, the efficiency first increases with the increase in

pressure ratio to attain a maximum value and then decreases with

further increase in the pressure ratio On the other hand, higher

turbine inlet temperature ensures higher thermal efficiency at all

pressure ratios As a result, the efficiency peaks of 24.3%, 29.7% and

34.4% are obtained for TITs of 1050 K, 1200 K and 1350 K,

respec-tively The maximum efficiency is reached in the pressure ratio

range of 3–4 for the three TITs considered here

The variation in the pressure ratio influences the specific air

consumption in the cycle and therefore the size of the plant

components The pressure at the inlet to the turbine is different at

different pressure ratios, while at the exit of the turbine the

pres-sure remains the same for all the cases At the high prespres-sure end of

the turbine the size at different conditions are compared using the

specific air consumption by volume, while for the low pressure end

the specific consumption by mass determines the size Also, an

increase in the pressure increases the metal thickness of the

shows the variation of the specific air flow by mass (kg/kWh)

entering the turbine against the pressure ratio (rp) at different

turbine inlet temperature It is observed that at a particular turbine

inlet temperature the specific air mass flow first decreases with the

increase in the pressure ratio The decrease in mass flow is found to

be rapid at the lower end of pressure ratio and gradually decreases

as the pressure ratio is increased Beyond a pressure ratio value the

mass flow begins to increase with further increase in pressure ratio

The pressure ratio at which the reversal in the trend of mass flow

variation occurs is lower at lower value of turbine inlet temperature

TIT ¼ 1200 K and 1350 K this reversal is not observed till rp¼ 8.0) It

turbine inlet temperature decreases the specific air consumption at

the turbine inlet

kWh) at the inlet to the turbine with changing pressure ratio, at

decrease in the specific air consumption by volume is monotonic

rapid at the lower end of the pressure ratio range, the incremental

change becomes less at higher rp The specific volume flow of air at the turbine inlet is guided by the pressure and the specific mass

low and the specific mass flow is high, both contributing to the

sharply increases the specific volume flow of air At higher values

curves are nearly flat, and the pressure is high For TIT ¼ 1050 K and 1200 K, although the mass flow of air observes a gradual

effect of increasing pressure, and the volume flow continues to decrease (though only at a slow rate)

Therefore, as observed fromFigs 3a and b, the specific mass and volume flow rates of air are high at low values of rp(e.g at rp¼ 2.0)

leads to a marginal decrease in the specific volume flow, but the increased pressure warrants thicker walls for the high pressure components of the cycle Therefore, though there may be

a marginal advantage in the reduction in volume flow rate (and hence the plant size) beyond a particular pressure ratio, the higher wall thickness will offset the cost benefit

0

0.1

0.2

0.3

0.4

0.5

TIT=1050 K TIT=1200 K TIT=1350 K

r p

Fig 2 Variation of thermal efficiency (h) of the EFGT cycle with the pressure ratio (r p )

at different turbine inlet temperatures (TIT).

0 10 20 30 40 50

TIT=1050 K TIT=1200 K TIT=1350 K

r p

r p

0 10 20 30 40 50 60

70

TIT=1050 K TIT=1200 K TIT=1350 K

3/kWh)

a

b

Fig 3 (a) Variation of specific air flow by mass (kg/kWh) with pressure ratio (r p ) for the EFGT cycle at different turbine inlet temperatures (TIT) (b) Variation of specific air flow by volume (m 3 /kWh) with pressure ratio (r p ) for the EFGT cycle at different turbine inlet temperatures (TIT).

The heat exchanger is one of the most critical equipment in the

EFGT cycle Considering the cost of the material for the high

temperature heat exchanger its size requires to be optimized

However, the design of the heat exchanger also influences the

thermal performance of the power cycle, by influencing the exhaust

gas loss from the cycle.Fig 4shows the variation in the cycle thermal

efficiency with pressure ratio at different cold end temperature

difference (CETD) of the heat exchanger for a particular turbine inlet

temperature The results show the same trend in the variation of

efficiency with pressure ratio at all the CETD values, with the

maximum efficiency reached at an optimum pressure ratio The

optimum pressure ratio is found to be 4.0 for the three different

CETD values of 200 K, 250 K and 300 K considered However, with

the increase in the CETD at a particular pressure ratio, the efficiency

is found to decrease When the CETD is high more amount of the

energy is wasted through the exhaust gas stream, reducing the net

work produced in the cycle In fact for particular rpand TIT, the state

points 2, 3 and 4 shown inFig 1do not change with the variation of

CETD However, the variation in the temperatures across the heat

exchanger changes the quantity of air flow governed by the energy

balance across the heat exchanger It is observed that the number of

moles of air flowing through the turbine per unit mole of dry

biomass feed (X0

variation in CETD does not change the specific air flow rate through

the turbine as the corresponding state points remain identical

Based on the above discussion, it can be proclaimed that the

depending on the TIT and CETD At the low pressure ratio of 2–3, the

size of the equipment will be large because of the high value of the

specific air flow Conversely, a high pressure ratio increases the wall

thickness of the equipment, thereby increasing the cost and weight

Considering all these facts, we have chosen rp¼ 4.0 as the optimum

value of the pressure ratio for the EFGT cycle Two different turbine

inlet temperatures (1200 K and 1350 K) and two different CETD

values (200 K and 300 K) are chosen to compare the performance

Accordingly, three sets of cycle operating conditions with different

turbine inlet temperatures (TIT) and heat exchanger cold end

temperature differences (CETD) have been identified (as given in

Table 4) to compare the performance of a 100 kW EFGT plant

3.2 Energy and exergy based analysis of a 100 kW biomass fired

EFGT plant

Table 4lists the important performance parameters for the three

cases for a 100 kW EFGT based micro gas turbine plant running at full

load It is seen fromTable 4that when the TIT is 1350 K and CETD is

200 K (Case 2), the thermal efficiency of the plant attains the highest value of 34.33% Accordingly, the fuel flow rate and the exhaust heat loss are the lowest The air flow rate is also the lowest for this case, indicating smaller size of the components, like compressor and turbine On the other hand, the logarithmic mean temperature difference (LMTD) of the heat exchanger based on the temperature differences at the hot and cold ends is low, giving a high overall (UAHE) value for the heat exchanger, where U and AHEare the overall heat transfer coefficient and the heat transfer surface area of the heat exchanger, respectively If we consider a nearly constant value of the overall heat transfer coefficient (U) for all the cases, then case 2 performance data calls for the largest size of the heat exchanger The higher turbine inlet temperature and the increased size of the heat exchanger required for this case is indicative of a high cost of the plant

In case when the TIT is 1200 K and the CETD 300 K (Case3) the thermal efficiency of the plant is the lowest (24.18%) The fuel flow rate and the exhaust heat loss are the maximum in this case The corresponding air flow rate is also the highest among the three sets compared 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 A thermal efficiency of 29.68% has been achieved in this case The heat exchanger LMTD is 131 K giving overall UA as 3.44 Therefore, considering the capital and operating cost of the plant, case 1 is the better choice of plant operating condition

A second law based performance analysis for the three cases

balance has been made as fractions of the exergy input to the cycle The fraction of the input exergy converted into useful work determines the exergetic efficiency of the cycle The remaining part of the input exergy is either lost in the exhaust heat or destroyed through irreversibilities in various compo-nents It is observed from the results of the three cases that the maximum exergetic efficiency is attained in case 2, where the turbine inlet temperature is the highest The exergy loss in the exhaust is the highest in case 3, where the exhaust gas leaves the cycle at the maximum temperature (because of the highest

three cases

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

0

0.1

0.2

0.3

0.4

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 (r p )

at different heat exchanger cold end temperature difference (CETD).

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

Case 1: r p ¼ 4, TIT ¼ 1200 K CETD ¼ 200 K

Case 2: r p ¼ 4, TIT ¼ 1350 K CETD ¼ 200 K

Case 3: r p ¼ 4, TIT ¼ 1200 K CETD ¼ 300 K Fuel (biomass)

flow rate, kg/s

0.0216 0.0186 0.0265 Air flow rate, kg/s 0.5867 0.4712 0.5942 Thermal Efficiency, % 29.68 34.33 24.18 Exhaust heat, kW 238.92 193.09 315.65 Rate of heat exchange

across heat exchanger, kW

450.5 440.03 450.72

Heat Exchanger hot end temperature difference, K

(UA HE ) overall for heat exchanger, W/K

of the compressor and turbine are only a little The fraction of

exergy destructed in the gasifier is the same in the three cases,

since the operating parameters of the gasifier has been considered

identical A sizeable amount (15.39%) of the input exergy is

des-tructed in the gasifier owing to the gasification reactions that take

place there The exergy destruction in the combustor is the

highest in all the three cases, amounting to 19.6%, 17.57% and

18.98% of the input exergy, respectively The destruction of exergy

in the combustion chamber is due to the heat exchange between

the streams and chemical reactions that take place Operating the

combustor at higher temperature and higher temperature of the

air fed to the combustion chamber decrease the exergy

destruc-tion in the combustor Exergy destrucdestruc-tion in the heat exchanger

increases when the temperature difference between the two

streams exchanging heat increases Accordingly, the maximum

fraction of the exergy destruction in heat exchanger occurs in case

3, where the LMTD is also the highest More than 10% of the input

exergy is destroyed in the heat exchanger for this case For the

other two cases (i.e case 1 and case 2) the exergy destroyed in the

heat exchanger are 8.94% and 8.45% of the input exergy

respectively

Fig 6 shows the exergetic efficiencies for the individual

components for the three cases The individual exergetic

effi-ciency value of the equipment indicates the deviation from

ideality 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 the combustion chamber is attributed to the irreversibility pertaining

to the chemical reactions occurring there The exergetic efficiency

of the compressor is identical for all the three cases (91.5%), since

it operates at same pressure ratio and isentropic efficiency Similarly, the exergy efficiencies of the gasifier are the same for the three cases as the operating pressure, gasifier equivalence ratio and the properties of the biomass are considered to be the same The exergetic efficiency for the turbine is the highest (96.4%) in case 2 where the turbine operates with the highest inlet temperature For this condition, the air temperature at the turbine outlet also remains higher than the other conditions As

a result, the combustion chamber operates with the maximum air preheat in case 2 The flame temperature in the combustor also becomes the maximum in this case As the chemical reaction occurs at high temperature the associated irreversibility becomes less and the combustion chamber exergetic efficiency attains the maximum value for case 2 The exergetic efficiency of the heat exchanger largely depends on the mean temperature difference between the streams across the exchanger Lower mean temperature difference is indicative of lower irreversibilities This

is evident in the result as the heat exchanger in case 2 (the case with lowest LMTD among the three) shows the highest exergetic efficiency

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

Exergy out 23.54%

Compressor 2.24%

Gasifier 15.39%

Turbine 2.28%

Combustor

19.60%

Useful 28.01%

Heat Exchanger 8.94%

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

Useful 32.40%

Exergy out 22.01%

Compressor 2.07%

Gasifier 15.39%

Turbine 2.11%

Combustor 17.57%

Heat Exchanger 8.45%

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

Useful 22.82%

Exergy out 29.05%

Compressor 1.83%

Gasifier 15.39%

Turbine 1.85%

Combustor 18.98%

Heat Exchanger 10.08%

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

4 Conclusions

A thermodynamic analysis has been performed for an externally

fired gas turbine (EFGT) cycle with an integrated biomass gasifier

The effects of operating parameters like pressure ratio (rp), turbine

inlet temperature (TIT) and cold end temperature difference (CETD)

of the heat exchanger on the thermal efficiency and specific air flow

have been studied The thermal efficiency of the cycle is found to be

within 16–34% for the range of operating parameters under

investigation The cycle thermal efficiency is the maximum at an

optimum pressure ratio of the cycle (in the range of 3–4) for

a particular turbine inlet temperature and cold end temperature

difference across the heat exchanger At a particular pressure ratio

of the cycle the thermal efficiency increases either with the

increase in the turbine inlet temperature or with the decrease in

the cold end temperature difference of the heat exchanger

The specific air flow at the turbine inlet is evaluated to compare

the size of the plant equipment It is found that the specific air flow

by volume decreases with the increase in pressure ratio sharply at

decrease with the increase in rpat the lower end of rp, while the

curves become flatter and even rises gradually beyond a particular

pressure ratio The increase in the turbine inlet temperature

decreases the specific air flow at the entry to the turbine However,

the cold end temperature difference across the heat exchanger does

not 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 thermal performance and sizing have been compared based on the thermal efficiency, air flow rate and heat transfer area of the heat exchanger Moreover, an exergy balance has been carried out for each of the cases to account the useful exergy, exergy loss and exergy destruction Major exergy destruction is found to occur in the gasifier, combustor and the heat exchanger

Though the parameters in Case 2 (TIT ¼ 1350 K, CETD ¼ 200 K) offer a higher thermal efficiency and exergetic efficiency and

a lower air flow rate, the heat exchanger size for this case is found to

be large On the other hand, the heat exchanger size for the Case 3 (TIT ¼ 1200 K, CETD ¼ 300 K) is small, but it gives the lowest thermal and exergetic efficiencies A trade-off in performance is observed for Case 1 (TIT ¼ 1200 K, CETD ¼ 200 K)

References

[1] Banerjee R Comparison of options for distributed generation in India Energy Policy 2008;34:101–11.

[2] Karki S, Madan MD, Salehfar H Environmental implications of renewable distributed generation technologies in rural electrification Energy Sources, Part B 2008;3:186–95.

[3] Nouni MR, Mullick SC, Kandpal TC Providing electricity access to remote areas

in India: niche areas for decentralized electricity supply Renewable Energy 2009;34:430–4.

[4] Sharma DC Transforming rural lives through decentralized green power Futures 2007;39:583–96.

[5] Nouni MR, Mullick SC, Kandpal TC Providing electricity access to remote areas in India: an approach towards identifying potential areas for decentralized elec-tricity supply Renewable and Sustainable Energy Reviews 2008;12:1187–220 [6] Klimantos P, Koukouzas N, Katsiadakis A, Kakaras E Air-blown biomass gasi-fication combined cycles (BGCC): system analysis and economic assessment Energy 2009;34:708–14.

[7] Rodrigues M, Faaij A, Walter A Techno-economic analysis of co-fired biomass integrated gasification/combined cycle systems with inclusion of economies of scale Energy 2003;28:1229–58.

[8] Dasappa S, Paul PJ, Mukunda HS, Rajan NKS, Sridhar G, Sridhar HV Biomass gasification technology – a route to meet energy needs Current Science 2004;87(7):908–16.

[9] Ferreira SB, Pilidis P Comparrison of externally fired and internal combustion gas turbines using biomas fuel J Energy Resources Tech Trans ASME 2001;123:291–6.

[10] Anheden M Analysis of gas turbine systems for sustainable energy conversion, Ph.D thesis, Royal Institute of Technology, Stockholm, Sweden, 2000 [11] Bram S, De Ruyck J, Novak-Zdravkovic A Status of external firing of biomass in gas turbines Proc IMechE Part A J Power and Energy 2005;219:137–45 [12] Cocco D, Deiana P, Cau G Performance evaluation of small size externally fired gas turbine (EFGT) power plants integrated with direct biomass dryers Energy 2006;31:1459–71.

[13] Traverso A, Massardo AF, Scarpellini R Externally fired micro-gas turbine: modelling and experimental performance Applied Thermal Engineering 2006;26:1935–41.

[14] LaHaye, PG, Zabolotny, E Externally-Fired Combined Cycle (EFCC) ASME-Cogen Turbo Meeting, Nice, France, August 30–September 2, 1989, pp 263–74 [15] McKendry P Energy production from biomass (part 3): gasification technol-ogies Bioresource Technology 2002;83:55–63.

[16] Dogru M, Howarth CR Gasification of hazelnut shells in downdraft gasifiers Energy 2002;27:415–27.

[17] Zainal ZA, Ali R, Lean CH Prediction of performance of downdraft gasifier using equilibrium modeling for different biomass materials Energy Conver-sion and Management 2001;42(12):1499–515.

[18] Turns SR Fundamentals of combustion McGraw-Hill; 2000.

[19] Jarungthammachote S, Dutta A Thermodynamic equilibrium model and second law analysis of a downdraft waste gasifier Energy 2007;32(9):1660–9 [20] Jayah TH, Aye L, Fuller RJ, Stewart DF Computer simulation of a downdraft wood gasifier for tea drying Biomass and Bioenergy 2003;25:459–69 [21] Sharma AK Equilibrium and kinetic modeling of char reduction reactions in

a downdraft biomass gasifier: a comparison Solar Energy 2008;82:918–28 [22] Moran MJ, Shapiro HN Fundamentals of engineering thermodynamics 4 th ed Singapore: Wiley; 2000.

[23] Rakopoulos CD, Giakoumis EG Second-law analyses applied to internal combustion engines operation Progress in Energy and Combustion Science 2006;32:2–47.

[24] Som SK, Datta A Thermodynamic irreversibilities and exergy balance in combustion processes Progress in Energy and Combustion Science 2008;34:351–76.

[25] Szargut J, Styrylska T Approximate evaluation of exergy of fuels Brennstoff Warme Kraft 1964;16(12):589–96.

[26] Williams RH, Larson ED Biomass gasifier gas turbine power generating

0.0

0.2

0.4

0.6

0.8

1.0

Heat Exchanger

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 three

different cases described in Table 4