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

Energy 35 (2010) 341–350

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Energy
journal homepage: www.elsevier.com/locate/energy

Energy and exergy analyses of an externally fired gas turbine (EFGT) cycle
integrated with biomass gasifier for distributed power generation
Amitava Datta*, Ranjan Ganguly, Luna Sarkar
Department of Power Engineering, Jadavpur University, Salt Lake Campus, Kolkata 700098, India

a r t i c l e i n f o

a b s t r a c t

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

Biomass based decentralized power generation using externally fired gas turbine (EFGT) can be a technically feasible option. In this work, thermal performance and sizing of such plants have been analyzed at
different cycle pressure ratio (rp ¼ 2À8), 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 rp values. 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.

Keywords:
Gas turbine
External firing
Biomass
Gasifier

1. Introduction
Small scale decentralized power generation is gaining importance for distributing electricity in the remote areas far from the
centralized grid [1–4]. The delivery of grid power to the remote
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 alternative 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
generation in a combined cycle [6,7]. On the other hand, piston
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
the generation of power [8]. However, such engines having

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

E-mail address: (A. Datta).
0360-5442/$ – see front matter Ó 2009 Elsevier Ltd. All rights reserved.
doi:10.1016/j.energy.2009.09.031

reciprocating components require more maintenance and abundance of cooling water, which make them unsuitable for remote
locations.
The use of biomass as fuel in conventional (internally fired) gas
turbine engines entails various problems [9]. Firstly, the 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

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

Nomenclature
AHE
AFSt
CETD
Cp
ei
ech
EFGT
EN
h
hf
K
Mj
Pi
DP
rp
R
s
Ti
TIT
U
w
W
X
Z

Heat exchanger surface area
Stoichiometric Air-fuel ratio
Cold End Temperature Difference of heat exchanger
Specific heat at constant pressure
Specific thermomechanical flow exergy at state i
Specific chemical exergy
Externally Fired Gas Turbine
Energy released with exhaust gas
Enthalpy
Enthalpy of formation
Equilibrium constant
Molecular weight of species j
Pressure at state i or Partial pressure for species i
Pressure drop
Pressure ratio
Universal gas constant
Entropy
Temperature at state i
Turbine Inlet Temperature
Overall heat transfer coefficient of heat exchanger
Specific work
Work
Number of moles
Moisture content in the as-fired biomass (by mass)

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.
Anheden [10] presented thermodynamic and economic analyses

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
optimum pressure ratio of the cycle. Ferreira and Pilidis [9]
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
fuel. Bram et al. [11] reviewed the technological and economic
feasibility of the external firing of biomass in gas turbines. The
authors concluded that cogeneration based on EFGT on the scale of
100–200 kWe offers good prospects from both economic and
technical aspects. Cocoa et al. [12] evaluated the performance of
a 100 kW externally fired gas turbine plant fuelled with biomass
and having an integral dryer for biomass. The influence of parameters 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 integration of the dryer improves flexibility in the plant operation.
Traverso et al. [13] presented the steady state and transient
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.

Greek Letters
Equivalence ratio

Ratio of specific heats
Isentropic efficiency of compressor
Isentropic efficiency of turbine
Efficiency

4
g
hc,isen
ht,isen
h

Sub-scripts
a
Air
B
Producer gas after gasification of biomass
C
Compressor
CC
Combustion chamber
f
Fuel
G
Gasifier
g
Product gas
HE
Heat exchanger
in
Input

i
Index for thermodynamic state point
o
Reference state
T
Turbine
w
Water
Super-scripts
c
Cold side of the heat exchanger
h
Hot side of the heat exchanger

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 combustion product [9–12]. The size of the heat exchanger and
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

reach 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 operation

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

temperature of 950 C, while further increase in the temperature
requires turbine blade cooling arrangement [13]. Since all these
modifications towards performance improvement bear considerable cost implications, such modifications always needs a priori
evaluation, based on energy and exergy based performance analysis 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


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

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
[10]) at a small scale (as considered here) involves additional
complexities and more number of skilled personnel that is

unavailable in remote areas at low cost. On the other hand, there is
developed technology of biomass gasifier at small scale [15] which
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, while Fig. 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

a

G
Exhaust
Gas

5

6

HE

B
CC

3

T

C

Air

c
P2- ΔPHE

T5

5

T3 (=TIT)

3

P2

1. Air is admitted to the compressor (state 1, refer Fig. 1b) at
atmospheric condition, P1 ¼101.325 kPa, T1 ¼ 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.

9. The reference temperature To and pressure Po are 25 C and
101.325 kPa, respectively.

2.2. Energy analysis of the cycle

1

b

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:

4

A
2

2.2.1. Air compressor
The compressor delivery pressure (P2) is evaluated using the
cycle pressure ratio (rp), which is varied in the range of 2–8. The
temperature of air (T2s) at pressure P2 for the isentropic compression is calculated considering a third order polynomial variation of
the molar specific heat of air with temperature as,
Cpair ¼ aair þ bair T þ cair T 2 þ dair T 3 [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:

4

wC ¼

CETD

T

6

2
h
P1+ ΔPHE
+ΔPCC
h
P1+ ΔPHE

P1

T1

343

1

s
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
Exchanger.

ZT2

Cpair dT

T1




 c 

bair  2
T2 À T12 þ air T23 À T13
¼ aair ðT2 À T1 Þ þ
2
3



dair  4
T2 À T14
þ
4

(1)

Eq. (1) is solved by Newton–Raphson method for the actual
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
pressure PA should be sufficient to overcome the pressure drop
across the gasifier and feed the producer gas to the combustion
chamber. Therefore,


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





h
PA ¼ P1 þ DPHE
þ DPCC þ DPG

(2)

h , DP
where, DPHE
CC 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 similar
approach for T2 described above.

2.2.2. Gasifier
The biomass feedstock (wood) is fed to the gasifier in a substoichiometric environment. The gasifier environment is described
by the equivalence ratio (f), which is defined as the ratio of stoichiometric air-fuel ratio to actual air-fuel ratio. We have considered
f ¼ 3.33 for our calculation.
A representative chemical formula is considered for the dry
biomass fuel as CHQOR, using the mass percentage of carbon,
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

AFSt ,Mf
¼
4:76,Ma f

(3)

where, AFSt is the stoichiometric air-fuel ratio for the fuel used and
Mf and Ma are 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 À ZÞÞ, Mf =Mw

(4)

where, Z is the moisture content (mass percentage) in the
biomass (as-fired) and Mw is the molecular weight of the water
vapour.
The global gasification reaction can be expressed as follows:

CHQ OR þ BH2 O þ XO O2 þ 3:76XO ,N2
¼ X1 H2 þ X2 CO þ X3 CO2 þ X4 H2 O þ X5 CH4 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 of
wood.
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 the
methanation reaction and water gas shift reaction [17,19] as below:

C þ 2H2 4CH4

(6)

CO þ H2 O4CO2 þ H2

(7)

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

K1 ¼ À

PCH4 =Po
Á2 ¼
PH2 =Po

X5 Po

5
X

Xj

j¼1

X1 P4

À

K2 ¼

ÁÀ
Á
PCO2 =Po PH2 =Po
X X
À
Á ¼ 1 3
X2 X4
ðPCO =Po Þ PH2 O =Po

(8)

(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 gasifier
exit (which is equal to the combustor pressure). The equilibrium
constants K1 and K2 depend on the gasification temperature. An

energy balance equation is drawn to evaluate the gasification
temperature (Tg) (assuming no heat loss from the gasifier) as
follows:

0
hfwood þ BhfH

2O

B
þ XO @

ZTA
To

0
B
¼ X1 @hfH þ
2

ZTg
To

0
B
þ X3 @hfCO þ

ZTg

2

To

0
B
þ X5 @hfCH þ

ZTg

4

1

0

C
B
CpO2 dT A þ 3:76XO @
1

2O

0

C
B
CpCH4 dT A þ 3:76XO @

To

C
CpN2 dT A

ZTg

1
C

CpCO dT A

To

0

C
B
CpCO2 dT A þ X4 @hfH
1

1

To

0

C
B
CpH2 dT A þ X2 @hfCO þ
1

ZTA

þ

ZTg

1
C

CpH2 O dT A

To

ZTg

1

C
CpN2 dT A

ð10Þ

To

where hfwood, hfH O , hfH , hfCO , hfCO and hfCH represent enthalpies of
2
2
2
4
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
obtained by solving for the values of X1 through X5 and Tg
simultaneously. Table 1 shows the product gas concentration on
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
et al. [20] and Sharma [21], who worked with gasification of

biomass. The corresponding results from the experiments of Jayah
et al. [20] and the equilibrium model of Sharma [21] under the
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 turbine inlet pressure P3 is calculated considering a 3%
pressure drop from the compressor exit (P2) in the cold side of
the heat exchanger. The temperature of air at the inlet to the
turbine (T3) is an input parameter for the analysis. The actual
work done by the turbine per kmol of air (wT) is evaluated
considering the variable specific heat and the isentropic turbine
efficiency (ht,isen). The actual temperature of air at the turbine
outlet (T4) is found by solving the following equation using
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
18.3
H2
CO
20.2
9.7
CO2
CH4
1.1
50.7
N2

19.35
19.34
11.18
0.25
50.19

18.97
24.75
8.01
0.39
47.88

Moisture content ¼ 18.5%, A-F Ratio ¼ 2.03
17.2
H2
CO
19.6

9.9
CO2
1.4
CH4
51.9
N2

19.85
19.64
11.01
0.26
49.26

20.91
23.79
9.25
0.99
45.06


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

wT ¼

ZT3

Cpair dT ¼ aair ðT3 ÀT4 Þþ

T4



d
þ air T34 ÀT44
4



 c 

bair  2
T3 ÀT42 þ air T33 ÀT43
2
3
(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

þ XO0 ðO2 þ 3:76N2 Þ/X6 CO2 þ X7 H2 O þ X8 O2
À
Á
þ 3:76 XO þ XO0 N2

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:

ei ¼ hi À ho À To ðsi À so Þ

(12)

XO0

denotes the kmoles of O2 entering the combustor from the
turbine for each kmol of wood fired in the cycle. Assuming an
adiabatic condition in the combustor, the energy balance is given
as:-

Xjðhfj þ

j

ZTg
To

¼

X




Cpj dTÞ


Xj ðhfj þ

j

ZT5
To

þ

Xj ðhfj þ

j

Fuel




Cpj dTÞ


X

ZT4
To




Cpj dTÞ



Air

(13)
Products

where, Xj represents the number of mole of the jth component in
fuel (the producer gas), air or product gas mixture and hfj and Cpj
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 T5 and XO0 .
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
(See Fig. 1).
A heat balance across the heat exchanger gives

4:76XO0

ZT3
T2

Cpair dT ¼ Xg

ZT5

Cpg dT

(14)

T6

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

À
Á
Xg ¼ X6 þ X7 þ X8 þ 3:76 XO þ XO0

ZTi

Cp dT

(17)

To

s i À so ¼

X

(16)

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

hi À ho ¼

ðX1 H2 þ X2 CO þ X3 CO2 þ X4 H2 O þ X5 CH4 þ 3:76XO N2 Þ

345

(15)

Cpg represents the molar specific heat of the exhaust gas mixture
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 T5 and XO0 .

ZTi
To

Cp

dT
P
À Rln i
Po
T

(18)

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

[22]. In true thermodynamic sense a multicomponent system
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
reaction [23,24] at reference temperature and pressure conditions.
We have followed the latter concept in evaluating chemical exergy
in this work.
) is obtained from its lower
The chemical exergy of the wood (ech
wood
heating value using a multiplication factor b [25], which is given by

1:044 þ 0:0160

b¼



H
O
H
1 þ 0:0531
À 0:34493
C
C
C
O
1 À 0:4124

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

ech
B ¼

X1
ech
X1 þ X 2 þX3 þ X4 þ X5 þ 3:76XO H2
þ

X2
ech
X1 þ X2 þ X3 þ X4 þ X 5 þ3:76XO CO

þ

X5
ech
X1 þ X2 þ X3 þ X4 þ X5 þ 3:76XO CH4

(20)

ch
ch
where, ech
H2 eCO and eCH4 represent the specific chemical exergy of H2,
CO and CH4, respectively [22].

2.4. Performance parameters
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

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,


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

WC ¼ 4:76XO0

ZT2

Cpair dT þ 4:76XO

T1

ZTA

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

Cpair dT;

(21)

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

T1

Carbon
Hydrogen
Oxygen
Calorific Value

While the actual work done by the turbine is

WT ¼ 4:76XO0

ZT3

Cpair dT

Moisture content in the biomass by mass

(22)

Properties of Ambient Air

Pressure
Temperature
Composition (by vol.)
Nitrogen
Oxygen

T4

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

EN ¼

X

Xj

j

ZT6

Cpj dT

Equipment performance
Isentropic efficiency of compressor (hc,isen)
Isentropic efficiency of turbine (ht,isen)
Pressure drop across the gasifier (DPG)

Pressure drop at heat exchanger
c Þ as % of inlet pressure
cold side ðDPHE
Pressure drop at heat exchanger
h Þ as % of inlet pressure
hot side ðDPHE
Pressure drop across combustor (DPCC)
as % of inlet pressure

(23)

To

where, Xj is the number of moles of the jth species on the product
side of Eq. (12) and Cpj is the respective specific heat.
The exergy input into the plant (ein) for every mole of biomass
þ 4:76fXO þ XO0 ge1 Þ. A part of the
fed to the cycle is given as ðech
wood
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
ðWT ÀWC Þ
ein

and

EN
ein

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 individual components of the cycle are presented in Table 2.
In addition to the exergy destruction, the expressions of the
exergetic efficiency of the individual components are also evaluated as indicators of their deviation from ideality, while operating
between the corresponding thermodynamic states. The expressions of exergetic efficiency of the individual plant components are
also 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 performance 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

Operating parameters with range
Equivalence ratio at the gasifier (4)
Compressor Pressure ratio (rp)
Turbine inlet temperature (TIT)
Heat exchanger cold end temperature
difference (CETD)

50%
6%
44%
449568 kJ/kmol
(18732 kJ/kg)
20%
101.325 kPa

300 K
79%
21%
87%
89%
16 mm Hg column
3
1.5
0.5

3.33
2–8
1050–1350 K
200–300 K

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
the plant equipment and lower capital cost. Table 3 summarizes the
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. 2 shows 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.
Component

Exergy Destruction

Exergetic efficiency

Compressor

4:76½ðXO þ XO0 Þe1 À XO eA À XO0 e2 ŠþWC

4:76½XO eA þ XO0 e2 À ðXO þ XO0 Þe1 Š
WC

Gasifier

jwood þ 4:76XO eA À XB ðech
B þ eB Þ;
where, XB ¼ X1 þ X2 þ X3 þ X4 þ X5 þ 3.76XO
0

Turbine

4.76X O (e3 À e4)ÀWT

Combustor

4:76XO0 e4 þ XB ðech
B þ eB Þ À Xg e5

Heat Exchanger

Xg (e5 À e6) þ 4.76X O (e2 À e3)

0

XB ðech
B þ eB Þ
ech
þ 4:76XO eA
wood
WT
4:76XO0 ðe3 À e4 Þ
Xg e5
4:76XO0 e4 þ XB ðech
B þ eB Þ
Xg e6 þ 4:76XO0 e3
Xg e5 þ 4:76XO0 e2


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

a

0.5
TIT=1050 K

50
TIT=1050 K

Specific air flow by mass (kg/kWh)

TIT=1200 K
0.4

347

TIT=1350 K

0.3

0.2

0.1

TIT=1200 K

40

TIT=1350 K
30

20

10

0

0
0

2

4

6

8

0

10

2

4

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, respectively. 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 pressure remains the same for all the cases. At the high pressure 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
equipment casing walls, increasing their weight and cost. Fig. 3a
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
(for TIT ¼ 1050 K the reversal occurs at rp ¼ 7.0, while for
TIT ¼ 1200 K and 1350 K this reversal is not observed till rp ¼ 8.0). It
is also important to note from Fig. 3a that the increase in the
turbine inlet temperature decreases the specific air consumption at
the turbine inlet.
The variation in the specific air consumption by volume (m3/

kWh) at the inlet to the turbine with changing pressure ratio, at
constant turbine inlet temperature, is shown in Fig. 3b. The
decrease in the specific air consumption by volume is monotonic
in this case with the increase in rp. While the decrease is very
rapid at the lower end of the pressure ratio range, the incremental

b

70

Specific air flow by volume (m3/kWh)

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

6

8

10

rp

rp

60

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

50
40
30
20
10
0
0

2

4

6

8

10

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

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
flow rate at a given TIT. At low values of rp the pressure remains
low and the specific mass flow is high, both contributing to the
fact that in the small rp regime a reduction in pressure ratio
sharply increases the specific volume flow of air. At higher values

of rp, the changes are much flatter since the specific mass flow
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
increase with the increase in rp, such an increase is masked the
effect of increasing pressure, and the volume flow continues to
decrease (though only at a slow rate).
Therefore, as observed from Figs. 3a and b, the specific mass and
volume flow rates of air are high at low values of rp (e.g. at rp ¼ 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 and
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.


348

A. Datta et al. / Energy 35 (2010) 341–350
Table 4
Performance parameters of 100 kW Biomass fired EFGT plant at different operating
conditions.

0.4

0.3

0.2

CETD=200 K

0.1

CETD=250 K
CETD=300 K
0
0

2

4

6

8

10

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).

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. 4 shows 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 rp and TIT, the state
points 2, 3 and 4 shown in Fig. 1 do 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 (XO0 ) decreases with the increase in the CETD. The
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
cycle 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, 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 rp ¼ 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 4 lists the important performance parameters for the three
cases for a 100 kW EFGT based micro gas turbine plant running at full

Fuel (biomass)
flow rate, kg/s
Air flow rate, kg/s
Thermal Efficiency, %
Exhaust heat, kW
Rate of heat exchange
across heat
exchanger, kW
Heat Exchanger hot end
temperature
difference, K
LMTD, K
(UAHE)overall for heat
exchanger, W/K

Case 1: rp ¼ 4,
TIT ¼ 1200 K
CETD ¼ 200 K

Case 2: rp ¼ 4,
TIT ¼ 1350 K
CETD ¼ 200 K

Case 3: rp ¼ 4,
TIT ¼ 1200 K
CETD ¼ 300 K

0.0216

0.0186

0.0265

0.5867
29.68
238.92
450.5

0.4712
34.33
193.09
440.03

0.5942
24.18
315.65
450.72

80.0

44.2

147.0

131.0
3.44

103.2
4.26

214.5
2.10

load. It is seen from Table 4 that 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 AHE are 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 tradeoff 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
has been presented in Fig. 5, where the complete exergy
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 components. 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
CETD). Table 4 shows a comparison of the exhaust heat for the
three cases.
The major exergy destruction takes place in the gasifier,
combustor and the heat exchanger, while the exergy destruction


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

Case 1: rp=4, TIT=1200 K, CETD= 200 K

349

Case 2: rp =4, TIT=1350 K, CETD=200 K
Heat Exchanger
8.45%

Heat Exchanger
8.94%
Useful
28.01%

Useful
32.40%

Combustor
17.57%

Combustor
19.60%

Turbine
2.11%

Turbine
2.28%

Gasifier
15.39%

Gasifier

15.39%

Exergy out
23.54%

Compressor
2.07%

Compressor
2.24%

Exergy out
22.01%

Case 3: rp =4, TIT=1200 K, CETD= 300 K
Heat Exchanger
10.08%

Useful
22.82%

Combustor
18.98%

Turbine
1.85%

Exergy out
29.05%

Gasifier
15.39%
Compressor
1.83%

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

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 destructed 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 destruction in the combustor. Exergy destruction 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 efficiency value of the equipment indicates the deviation from
ideality for the equipment operating across its respective thermodynamic 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.


350

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

1.0

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)

0.8

ε

0.6

0.4

References

0.2

0.0
r

so

s
pre

m

Co

r

ine

ifie

s
Ga

rb
Tu

Co

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

r

sto

u
mb

at

He

Ex

er

ng

a
ch

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.

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
the lower end of rp, while the incremental change is marginal at
high values of rp. The specific air flow by mass exhibits a rapid
decrease with the increase in rp at 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 rp ¼ 4,
have finally been considered for a detailed investigation of the

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