1. Applied Thermal Engineering 60 (2013) 285e294
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Applied Thermal Engineering
journal homepage: www.elsevier.com/locate/apthermeng
Using coolant modulation and pre-cooling to avoid turbine blade
overheating in a gas turbine combined cycle power plant fired
with low calorific value gas
Ik Hwan Kwon a, Do Won Kang a, Tong Seop Kim b,*
a Graduate School, Inha University, Inha-ro 100, Nam-gu, Incheon 402-751, Republic of Korea
b Department of Mechanical Engineering, Inha University, Inha-ro 100, Nam-gu, Incheon 402-751, Republic of Korea
h i g h l i g h t s
The blade overheating problem in firing low calorific value gas in gas turbine was examined.
Several measures to suppress blade overheating were compared.
Coolant modulation was shown to result in a much lower power penalty than under-firing.
Pre-cooling of the coolant reduces the power penalty further by reducing the coolant supply.
a r t i c l e i n f o
Article history:
Received 2 April 2013
Accepted 8 July 2013
Available online 16 July 2013
Keywords:
Gas turbine
Combined cycle
Low calorific gas
Turbine blade overheating
Under-firing
Coolant modulation
Pre-cooling
a b s t r a c t
Overheating of turbine blades is one of the major concerns in using low calorific value fuels in gas
turbines. In this work, we examined the deviation of operating conditions of a gas turbine fired with a
low calorific value gas fuel, with a focus on the turbine blade temperatures. Several measures to suppress
blade overheating were compared in terms of the power output and efficiency of the gas turbine
combined cycle plant. Blade overheating can be prevented by decreasing the firing temperature without
the need for hardware modifications, but the accompanying power reduction is considerable. As a
remedy to this large reduction in power, modulation of the coolant supply to each blade row was
simulated, and a much lower power penalty was observed. Moreover, pre-cooling of the coolant en-hances
the power output further by reducing the coolant supply. Pre-cooling recovers 80% of the
available maximum augmentation of the combined cycle by simply switching the fuel from natural gas to
low calorific value gas. Pre-cooling also provides higher overall combined cycle efficiency compared to
under-firing.
2013 Elsevier Ltd. All rights reserved.
1. Introduction
Fuel diversity is the major advantage of gas turbines over other
types of power generators. In addition to natural gases, a wide
range of low calorific gaseous fuels (such as synthetic gases from
coal and biomass gasification, and various kinds of biogas) can be
used in gas turbines. The integrated gasification combined cycle
(IGCC) is considered to be the most environmentally friendly
method of using coal. Several full-size plants are under operation
and a number of projects are ongoing worldwide. Various perfor-mance
analyses and comparisons have been undertaken during the
past decade including performance summaries and modeling of
existing plants [1,2], and examinations of the effects of major
design parameters such as the integration between a gas turbine
and auxiliary components [3,4]. Design limitations regarding the
operating conditions of major components, especially turbine and
compressor, have been studied [5,6]. In addition, the influence of
syngas composition on the performance and operability of gas
turbines has been examined [7]. Attempts to use biomass as a fuel
in gas turbines and combined cycle power plants have also been
initiated recently. Various basic studies on the use of biomass in gas
turbine-based power plants have been published, such as per-spectives
on the use of biomass in combined cycle plants [8],
different strategies for using biomass [9], and the influence of firing
biomass on gas turbine components [10]. Also, the possibility of co-firing
biomass with natural gas has been investigated [11,12]. The
* Corresponding author. Tel.: þ82 32 860 7307; fax: þ82 32 868 1716.
E-mail address: kts@inha.ac.kr (T.S. Kim).
1359-4311/$ e see front matter 2013 Elsevier Ltd. All rights reserved.
http://dx.doi.org/10.1016/j.applthermaleng.2013.07.008

2. 286 I.H. Kwon et al. / Applied Thermal Engineering 60 (2013) 285e294
use of biogas (digester and landfill gas) in relatively small gas
turbine-based combined heat and power systems has also been
studied [13,14].
Even though low calorific value gas fuels are good resources for
use in gas turbine-based power plants, there are concerns
regarding the effect of these fuels on the operability and lifetimes of
gas turbine components. Overheating of hot sections such as tur-bine
blades is one of the major concerns, as pointed out in recent
publications [5,6,10,15]. The common understanding is as follows.
The calorific values of syngases are much lower than that of natural
gas, which gas turbines are designed for. Therefore, when the fuel is
switched from natural gas to a low calorific value gas, more fuel
must be supplied to the combustor, resulting in greater mass flow
in the turbine. The larger turbine mass flow results in a rise in the
compressor pressure ratio if there is no hardware modification in
the turbine. This causes an increase in the cooling air temperature,
which increases the blade metal temperature over the design
temperature in coal and biomass integrated gasification power
plants [6,10]. It has been reported that firing biogas might cause a
similar overheating problem of hot sections [14]. Thus, hot section
overheating is a common phenomenon when firing low calorific
value gas in a gas turbine. Therefore, fuel switching should be
accompanied by proper measures to suppress the overheating.
We investigated the degree of turbine blade overheating in a
modern state-of-the-art gas turbine for combined cycle power
plants, and comparatively analyzed several measures to suppress
overheating. Decreasing the firing temperature would be the
simplest way to avoid turbine overheating [6,7,15], but this reduces
the performance of the gas turbine noticeably, especially with
respect to power output [6,7]. Alternative methods are associated
with the turbine coolant supply. The flow rate of the coolant can be
modulated to supply an appropriate amount of coolant to keep the
blade temperature below a target value. The feasibility of coolant
flowrate control was examined in a previous study [6], wherein the
authors used a simplified approach that focused on the first stage
nozzle blade. The present study adopted a more detailed analysis
based on a calculation for each cooled blade row. Another distinct
feature of this study is the adoption of coolant pre-cooling. Turbine
coolant can be pre-cooled by water or steam from the bottoming
cycle before it is supplied to the turbine. Pre-cooling would mini-mize
the coolant supply because with a lower the coolant tem-perature,
less coolant is needed. We demonstrated the relative
advantage of modulating coolant flow rates and pre-cooling the
coolant compared to simple under-firing. A full gas turbine com-bined
cycle plant was modeled. Changes in plant performance and
the operating condition (especially the turbine blade temperatures)
when using a low calorific value gas were analyzed. In addition, the
effects of different methods to restore the blade temperatures to
the reference values were simulated and compared.
In the detailed turbine design and analysis stage, blade tem-perature
distribution and lifetime analysis using numerical
methods, especially the conjugate heat transfer analysis [16e19],
can be used to optimize the cooling system accounting for the
operating condition change. However, performing a numerical
analysis is beyond the scope of this study, and thus we have focused
only on thermodynamic system level analysis in this paper. The
result of this study may provide useful basic data for the detailed
numerical analysis.
2. System modeling
2.1. Gas turbine
Fig. 1 shows the gas turbine combined cycle system considered
in this study. The performance of the entire system was simulated
using GateCycle [20]. A state-of-the-art F-class gas turbine that is
widely used for combined cycle plants was adopted. Design spec-ifications
were taken from a manufacturer’s report [21e23] and the
open literature [24]. The engine consists of an eighteen-stage
compressor with a pressure ratio of 16, and a three-stage turbine.
The design fuel is a natural gas consisting of 90.1% CH4 by mole and
other miscellaneous hydrocarbons, and its lower heating value
(LHV) is 49,244 kJ/kg. The turbine blade cooling was modeled as
close to the actual design as possible using the reference data, as
depicted by the coolant lines shown in Fig. 1. Of course, the coolant
pre-cooling lines were not adopted in the reference engine, but
Nomenclature
A area (m2)
C cooling constant, absolute velocity (m/s)
CC combined cycle
CDP compressor discharge pressure
CDT compressor discharge temperature
cp specific heat (kJ/kg K)
ECO economizer
EVA evaporator
GT gas turbine
HP high pressure
IGCC integrated gasification combined cycle
IP intermediate pressure
LCG low calorific value gas
LHV lower heating value (kJ/kg)
LP low pressure
m_ mass flow rate (kg/s)
NG natural gas
P pressure (kPa)
PR pressure ratio
R gas constant (kJ/kg K)
SH superheater
ST steam turbine
T temperature (K)
U blade speed (m/s)
V relative velocity (m/s)
W_ power (MW)
a absolute flow angle
b relative flow angle
f cooling effectiveness
g specific heat ratio
h efficiency
k constant
Subscripts
1 nozzle inlet
2 nozzle outlet, rotor inlet
3 rotor outlet
a axial component
b turbine blade
c coolant
d design point
g gas
in inlet
rel relative total property
N asymptotic

3. I.H. Kwon et al. / Applied Thermal Engineering 60 (2013) 285e294 287
Fig. 1. Schematic of system configuration (dotted lines are only for the coolant pre-cooled case).
were used only in the pre-cooled case simulated in this study. For
each turbine stage, the cooling of nozzle and rotor blades was
separately modeled to predict the temperature variation of each
blade row. Five rows (2.5 stages) were cooled by air from different
sources, as shown in Fig. 1. The second stage nozzle/rotor and the
third stage nozzle were cooled by air bled from the compressor
middle stages. Since the exact locations were not described in the
manufacturer’s references, we selected appropriate stages that
have sufficient pressure to be injected into the corresponding tur-bine
sections. The simulated design performance of the gas turbine
is shown in Table 1. All of the three major performance parameters
(power output, thermal efficiency, and exhaust temperature) were
in good agreement with the reference data, which demonstrates
the feasibility of the reference gas turbine modeling.
When a low calorific value gas is supplied to the combustor as
fuel in an existing gas turbine designed for natural gas, the oper-ating
conditions of both the compressor and the turbine deviate
from their design conditions. Therefore, a full off-design analysis is
required to perform a realistic simulation. The compressor was
modeled using the performance map shown in Fig. 2. We used a
multi-stage axial compressor map with a similar design pressure
ratio embedded in GateCycle [20], with proper scaling, taking into
account the design point (pressure ratio and mass flow) of the gas
turbine used in this study. The off-design operation of the turbine
was modeled by the following constant swallowing capacity
(choking condition), which is very reasonable for heavy-duty in-dustrial
gas turbines [20]:
p
kAinPin
m_ in
ffiffiffiffiffiffi
Tin
¼ constant; where k ¼
ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
g
R
2
g þ 1
gþ1
g1
s
(1)
The coolant flow rate for each off-design operation was calculated
as follows [20]:
Table 1
Gas turbine specifications.
Parameter Reference Modeling
Ambient condition 273.2 K, 1013 kPa, 60% RH
Pressure ratio 16 16
Compressor isentropic efficiency (%) NA 83.7
Total coolant flow relative to inlet air (%) NA 17.4
Turbine inlet temperature (K) NA 1670.2
Turbine rotor inlet temperaturea (K) 1600.2 1600.2
Number of turbine stages 3 3
Turbine stage efficiency (%) NA 88.5
Exhaust gas flow (kg/s) 445.0 444.9
Net power (MW) 171.7 171.5
LHV efficiency (%) 36.5 36.7
a Temperature at the first stage rotor inlet.
2.5
2.0
1.5
1.0
0.5
0.0
1.2
1.0
0.8
0.6
0.4
0.2
0.6 0.7 0.8 0.9 1.0 1.1 1.2
PR/PR
d
d
Relative corrected mass flow
90
95
100
105
110
relative speed
efficiency
pressure ratio
Fig. 2. Compressor performance map.

4. 288 I.H. Kwon et al. / Applied Thermal Engineering 60 (2013) 285e294
m_ c ¼ m_ c;d
Pc
Pc;d
!
Tc;d
Tc
0:5
(2)
Once the turbine inlet temperature is given, thermodynamic
matching between the compressor map and the turbine charac-teristic
equation determines the operating condition of the gas
turbine. Fig. 3 exemplifies the feasibility of the off-design calcula-tion.
It shows the variation of the full load (fully fired) performance
of the gas turbine versus the ambient temperature. The simulated
variations of the power output and the efficiency are in very good
agreement with the manufacturer’s reference data [23], which
proves the validity of the off-design calculation based on the soft-ware
used in this study.
2.2. Fuel
We selected coal syngas as a typical low calorific value gas fuel.
Table 2 shows the compositions and calorific value of the fuel [1]
used in this study. The table lists the properties of the low calo-rific
value gas, and those of natural gas. Hydrogen and carbon
monoxide are the major components, and the calorific value
(shown as the lower heating value in the table) is about one-sixth
that of natural gas.
2.3. Bottoming cycle
A triple pressure bottoming steam turbine cycle was used, as
shown in Fig.1. The major design parameters of the bottoming cycle
and the predicted the combined cycle design performance using
natural gas are listed in Table 3. Due to the gas-side pressure drop at
the heat recovery steam generator, the gas turbine power in the
combined cycle plant is slightly less than the reference power
shown in Table 1. The simulated combined cycle efficiencywas very
close to that reported in the literature (56.5%) [24].
In the simulated case in which the turbine coolant is pre-cooled,
some of the intermediate pressure and low pressure water streams
from the bottoming cycle were used for pre-cooling, as indicated by
the dotted lines in Fig. 1. Thus, no heat loss outside the entire
combined cycle system was allowed. The recovery of the thermal
energy released from the cooling air by the water/steam of the
bottoming cycle was beneficial in terms of the overall plant per-formance
in the conventional natural gas-fired gas turbine and
combined cycle plants [25].We adopted this observation to the low
calorific gas-fired system as a way to minimize the performance
penalty while maintaining the target blade temperature. We
assumed that the bottoming steam cycle was optimally designed
with respect to the cycle parameters (steam pressures and tem-peratures,
condenser pressure, temperature difference, pressure
drop, etc.) given in Table 3 for each gas turbine condition.
2.4. Turbine blade cooling
The variation in the temperature of each turbine blade row was
investigated using a cooling model [26]. The model describes a
relationship between the cooling effectiveness and the ratio of
thermal capacities (the mass flow multiplied by the specific heat)
between the coolant and the mainstream gas. The cooling effec-tiveness
is defined by
f ¼ Tg Tb
Tg Tc
(3)
Once the temperatures of the mainstream gas, cooling air, and
blade metal are given at the design point, the cooling effectiveness
can be specified. Table 4 shows the coolant properties for each
blade row. The gas and coolant temperatures are total tempera-tures.
In case of the rotor blades, the total temperature relative to
the rotating frame (i.e., the relative total temperature) should be
used. The blade temperature in Eq. (3) represents an average
temperature. In real engines, the blade surface temperature must
have a distribution affected by the non-uniformity of the gas and
120
115
110
105
100
95
90
85
80
Reference
Simulation
-10 0 10 20 30 40
Ambient Temperature(
o
C)
Relative variation (%)
Power output
Efficiency
Fig. 3. Example of off-design calculation: performance variation versus ambient
temperature.
Table 2
Fuel compositions and heating values.
Component Mole fractions (%)
NG LCG
CO 35.1
CO2 13.1
H2 31.4
H2O 16.4
N2 þ Ar 0.19 3.9
CH4 90.09 0.07
C2H6 6.04
C3H8 2.54
C4H10 1.12
Others 0.02
LHV (kJ/kg) 49244.2 8624.7
Table 3
Bottoming cycle specifications and combined cycle plant performance.
HP pressure (bar) 180
IP pressure (bar) 40
LP pressure (bar) 30
Condenser pressure (bar) 0.07
Steam temperaturea (K) 838.9
Pinch temperature difference (K) 11.1
Gas-side pressure drop (bar) 0.042
GT power (MW) 168.2
ST power (MW) 94.3
Total CC plant power (MW) 262.5
CC Plant LHV efficiency (%) 56.1
a Both at HP and IP inlets.

5. I.H. Kwon et al. / Applied Thermal Engineering 60 (2013) 285e294 289
coolant temperatures, and the temperature of local hot spot must
be higher than the average temperature. However, we did not
consider a detailed temperature distribution because it can only be
predicted through a computational analysis, which is beyond the
scope of the present study. Examples of conjugate heat transfer
analyses can be found in literature [16e19]. Accordingly, we
adopted the average temperature approach and compared the
average temperatures among different operating strategies.
To estimate the rotor-relative total temperature, a velocity tri-angle
between the rotor inlet and outlet was considered at every
turbine stage, as shown in Fig. 4. The absolute total temperature at
the rotor inlet (T02) is the mixed-out temperature after nozzle blade
cooling. Based on reasonable design values for the blade speed (U),
flow coefficient (Ca/U), loading coefficient (stage work/U2), and
reaction, all of the absolute and relative velocities (C, V) and angles
(a, b) were determined at the design point. The relative total
temperature was obtained using the relation
T02:rel ¼ T2 þ
V2
2
2cp
(4)
At each off-design condition, the absolute nozzle exit flow angle
(a2) was assumed to remain constant at the design value (constant
flow deviation). Then, the absolute flow speed (C2) was calculated
using the mass flow equation. U is fixed because the rotational
speed of the shaft should be kept constant in gas turbines used for
electric power generation (3600 rpm for a 60 Hz machine that
directly drives the generator without a gear box). As a result, the
relative speed V2 at the off-design conditionwas calculated. Finally,
the relative total temperature of the rotor blade (T02,rel) was pre-dicted
and used for Tg in Eq. (3).
The cooling performance was predicted using the following
equation [26], which describes a relationship between the cooling
effectiveness and the ratio of thermal capacities between the
coolant and the mainstream gas:
m_ c$cp;c
¼ f
C
m_ g$cp;g
fN
f
(5)
where fN represents the asymptotic cooling effectiveness corre-sponding
to a very high thermal capacity ratio, and C represents the
technology level of the cooling scheme. The cooling performance is
usually presented as a curve showing the functional relations be-tween
the cooling effectiveness and the thermal capacity ratio, as
exemplified in Fig. 5. The trend of the cooling curve is similar to
those of real engines [27]. Thus, Eq. (5) can be used to simulate the
behavior sufficiently well as illustrated in Refs. [6,7]. The value of
fNwas set to 0.92 for the first and second stages, and to 0.83 for the
last stage, with a reference to the literature [28].
At the design point, C of each blade row was determined using
Eq. (5), using all of the other parameters given by the cycle calcu-lation.
The calculated C values are 0.062 and 0.061 for the first stage
nozzle and rotor; 0.079 and 0.077 for the second stage nozzle and
rotor; and 0.058 for the third stage nozzle, respectively. The design
temperatures of the first and second stage blade rows were set to
870 C (1143.2 K), and that of the third stage nozzle bladewas set to
750 C (1023.2 K). Then, the complete turbine blade cooling model
for each blade row was established. For an off-design condition, all
of the parameters in Eqs. (3) and (5) (except for the blade metal
temperature (Tb)) were known from the cycle calculation. Then, the
metal temperature was predicted using Eq. (3).
The estimations of the variations in turbine blade temperatures
described in this section were performed by using the macro-function
in GateCycle [20]. For low calorific value gas-fired opera-tion,
four cases were simulated. Except for the under-firing case, the
turbine inlet temperature remains at the design value. The first is
the baseline case in which the fuelwas switched from natural gas to
syngas. In the second case, we reduced the turbine inlet tempera-ture
to keep all of the blade temperatures below the design values.
In the third case, all of the cooled blade rows were kept at the
design values by modulating (actually increasing) the cooling air
flows. Finally, the effect of coolant pre-cooling in the third case was
investigated. Table 5 summarizes the five cases.
Table 4
Coolant properties for each cooled blade row.
Stage Blade Coolant to gas
mass flow ratio
Coolant
temperature (K)
Gas
temperature (K)
1st Nozzle 0.088 691.4 1670.2
Rotor 0.050 691.4 1422.1
2nd Nozzle 0.032 600.4 1312.0
Rotor 0.020 600.4 1158.7
3rd Nozzle 0.007 520.1 1065.1
Fig. 4. Velocity triangle of a turbine stage.
φ
Fig. 5. Example of cooling effectiveness curve.

6. 290 I.H. Kwon et al. / Applied Thermal Engineering 60 (2013) 285e294
The net plant power output and efficiency are defined as
follows:
W_ net ¼ W_ GT þW_ ST W_ Aux (6)
hnet ¼
W_ net
ðm_ $LHVÞ
fuel
(7)
The auxiliary power is the sum of all of the power consumptions
of the additional components (mostly pumps).
3. Results and discussion
The predicted performance of the gas turbine fired with the low
calorific value gas is compared with its design performance using
natural gas in Table 6. To achieve the same turbine inlet tempera-ture,
a much larger fuel (6.9 times) should be supplied to the
combustor in the low calorific value gas case compared to the
natural gas case, which would cause a considerable increase in the
turbine gas flow (more than a 10% increase in terms of the exhaust
gas flow). This results in a considerable improvement in the net gas
turbine power output. The predicted power output was 204.5 MW,
which is 19.2% greater than the design power output of 171.5 MW.
This power augmentation is realizable in the view point of me-chanical
design of gas turbines. Gas turbines are usually designed to
accommodate the thermodynamically available power generation
capacity in cold ambient conditions (see Fig. 3) which is much
larger than the power output at standard ambient temperature
(15 C). The gas turbine dealt with in this study is known to have a
maximum power capacity around 200 MW [29]. Some re-design
and reinforcement of the shaft mechanical system may be
required depending on the actual shaft design features. The gas
turbine efficiency was predicted to decrease slightly. This is due to
the fact that the thermal energy input required to achieve the same
combustor outlet temperature of the product gas in the low calo-rific
value gas-fired case is slightly larger because low calorific value
gas includes non-reacting (inert) components such as water vapor,
carbon dioxide, and nitrogen.
This result is positive in terms of gas turbine performance
because the power output increases substantially, but efficiency
decreases only slightly. However, the result causes a critical prob-lem
with respect to the lifetime of the hot section. The increased
turbine inlet gas flow causes the turbine inlet pressure to rise ac-cording
to the mass flow increase (see the turbine characteristic
described by Eq. (1)). Thus, the compressor pressure ratio increases
proportionally, which causes the turbine coolant temperature to
rise. Fig. 6 shows the shift of the operating point on the compressor
map. The compressor discharge pressure increases from 1620.8 to
1782.9 kPa, and the discharge temperature increases from 691 to
726 K. This is negative in terms of turbine blade temperature.
Another factor affecting the blade temperature variation is the
thermal capacity ratio between the coolant and gas flow. Fig. 7
shows this variation in the first stage nozzle. Although the main-stream
gas flow increased significantly by firing low calorific value
gas, the thermal capacity ratio did not change appreciably. This is
because the coolant flow also increased, which was affected by the
increased source pressure (see Eq. (2) e the total coolant fraction
increased from 17.3% to 19.0%). A slight decrease in the cooling
effectiveness was predicted, as shown in Fig. 7. The coolant tem-perature
was predicted to rise from 691 to 726 K as previously
mentioned, and the cooling effectivenesswas predicted to decrease
from 53.8% to 52.6%. Both of these factors (the rise in the coolant
temperature and the decrease in effectiveness) contributed to the
30 K increase in the first stage nozzle temperature, but the former
factor is dominant. Similar patterns occur in all of the other turbine
blade rows, resulting in 24e30 K increases in blade temperature.
The increase in the compressor pressure ratio could be an issue
from the viewpoint of safe engine operation. Active hardware
modifications such as an increase in the turbine annulus area [9,30]
could be an ultimate solution to the surge issue. In this study, we
focused only on the issue of turbine blade overheating because we
still have a 10% surge margin in the baseline operation and want to
present a remedy to the overheating problem. Other three opera-tions
have larger surge margins than the baseline operation.
Under-firing (reduction of the turbine inlet temperature) is the
simplest way to suppress blade overheating without engine hard-ware
modification. Fig. 8 shows the effect of reducing the turbine
inlet temperature on the temperatures of the cooled blade rows. To
reduce the temperatures of all of the blade rows below the design
temperatures, the turbine inlet temperature must be reduced to
1610 K, which is 60 K lower than the design value. Fig. 9 shows the
Table 5
Descriptions of different operating cases.
Case Fuel Description
Design NG Design operation
Baseline LCG NG is simply switched to LCG
Under-firing LCG Modulating firing temperature to keep
blade temperatures below the design values
Coolant modulation LCG Modulating coolant flow rates to keep blade
temperatures below the design values
Coolant modulation
with pre-cooling
LCG Simultaneous use of pre-cooling and
modulation of flow rate of coolant to keep blade
temperatures below the design values
Table 6
Parameter comparison between the natural gas fired case and the baseline low
calorific value gas fired case.
Parameter NG LCG
GT Power (MW) 171.5 204.5
GT Efficiency (%) 36.7 36.3
Turbine inlet temperature (K) 1670.2 1670.2
Pressure ratio 16 17.6
Exhaust mass flow (kg/s) 444.9 491.1
Fuel mass flow (kg/s) 9.5 65.3
1st stage nozzle blade temp. (K) 1143.2 1173.6
1st stage rotor blade temp. (K) 1143.2 1172.4 Fig. 6. Deviation of compressor operating condition caused by firing low calorific value
gas.

7. 0.65
0.60
0.55
0.50
0.45
I.H. Kwon et al. / Applied Thermal Engineering 60 (2013) 285e294 291
design
baseline
under-firing
coolant modulation
coolant modulation
with pre-cooling
NG
LCG
variations in the gas turbine power output and efficiency. With a
1610 K turbine inlet temperature, the power output was reduced by
7.4% from 204.5 MW, reaching 189.3 MW. The efficiency also
decreased slightly (from 36.3 to 36.0%). The decrease in the turbine
inlet temperature caused the pressure ratio, and thus the coolant
temperature, to decrease slightly, which provides a small advantage
in terms of the blade temperature.
Increasing the coolant flows also decreases the blade tempera-tures
by increasing the cooling effectiveness. For each blade row,
we calculated an appropriate amount of coolant flow to maintain
the blade metal temperature at the design value, and the gas
205.0
200.0
195.0
190.0
36.4
36.3
36.2
36.1
Power
Efficiency
turbine performance corresponding to such an operation was
evaluated. The cooling effectiveness of the first stage nozzle in this
coolant-modulated operation is shown in Fig. 7 together with other
cases. The required total coolant fractionwas estimated to be 23.3%.
Modulation of the coolant flow rate requires some modifications of
the coolant passage between the compressor and the turbine (e.g.,
control of the valve in the coolant flow loop). The relative advantage
of the coolant flow rate modulation compared to under-firing is
that it causes a lower power penalty from the baseline case (a
simple fuel switch from natural gas to low calorific value gas). The
gas turbine power output of the coolant-modulated case was pre-dicted
to be 196.7MW, which is larger than that of the under-firing
case by 7 MW. The efficiency was predicted to decrease slightly to
35.9%.
Based on the coolant-modulated operation, pre-cooling of the
coolant was simulated. If the coolant temperature is reduced, the
1200
1150
1100
1050
1000
1st Nozzle
1st Rotor
2nd Nozzle
2nd Rotor
3rd Nozzle
1st, 2nd stage nozzle/rotor
design temperature
3rd stage nozzle design temperature
1610 1620 1630 1640 1650 1660 1670
Blade temperature [K]
Turbine inlet temperature [K]
Fig. 8. Variations in turbine blade temperatures versus a reduction in turbine inlet
temperature.
185.0
36.0
1610 1620 1630 1640 1650 1660 1670
Power [MW]
Efficiency [%]
Turbine inlet temperature [K]
Fig. 9. Variations in gas turbine power output and efficiency versus a reduction in
turbine inlet temperature.
0.40
0.06 0.065 0.07 0.075 0.08
φ
Fig. 7. Cooling effectiveness of the first stage nozzle.
200.0
199.5
199.0
198.5
198.0
197.5
197.0
196.5
36.0
35.9
35.8
35.7
35.6
35.5
Power
Efficiency
0 20 40 60 80 100
Power [MW]
Efficiency [%]
Degree of pre-cooling [K]
Fig. 10. Effect of coolant pre-cooling on gas turbine performance.

8. 292 I.H. Kwon et al. / Applied Thermal Engineering 60 (2013) 285e294
210.0
200.0
190.0
180.0
170.0
37.5
37.0
36.5
36.0
35.5
coolant supply could be decreased while maintaining the same
blade temperatures. The reduction in the coolant flows, especially
those of the blade rows after the first stage nozzle, results in
increased turbine inlet gas flow. This enhances the gas turbine
power output. The impact of the degree of pre-cooling on gas tur-bine
performance is shown in Fig. 10. A reduction of the coolant
temperatures of all blade rows by the same amount was simulated.
A 100 K reduction would enhance the power output by 2.3 MW
(1.2%), while reducing the efficiency only slightly (0.34% points).
Fig. 11 summarizes the performance comparison among four
different low calorific value gas-fired operations and the reference
natural gas-fired operation. Fig. 12 compares the compressor
discharge pressures and the corresponding air temperatures (i.e.,
the first turbine stage coolant temperatures except for the case with
pre-cooling). Fig. 13 compares the total turbine coolant flow rates
120
110
100
90
80
70
30
25
20
15
10
5
NG LCG
and the total coolant fraction relative to compressor inlet air flow in
the five cases. The advantage of coolant modulation versus under-firing
is clear. Despite a marginal gas turbine efficiency penalty,
coolant modulation ensures a sensible advantage in terms of power
generation. For example, with coolant modulation and 50 K pre-cooling,
we obtained a 19% power augmentation and a 1% point
efficiency penalty compared to the gas turbine design performance.
Fig. 14 shows the flow rate and temperature at the gas turbine exit.
Increases of exhaust gas flow in cases of firing the low calorific
value gas is due to the increased fuel flow as we have already dis-cussed.
The baseline case has a higher exhaust temperature than
the design case even with a higher turbine expansion pressure ratio
because it exhibits a slightly higher specific heat of the turbine gas
compared to the design case due to an increase in thewater content
160.0
35.0
Power
Efficiency
Gas turbine power [MW]
Gas turbine efficiency [%]
Design Base-line
Under-firing
Coolant
modulation
Coolant
modulation
with 50K
pre-ccoling
NG LCG
Fig. 11. Gas turbine performance comparison for different operations.
2000
1800
1600
1400
1200
1000
500
480
460
440
420
400
CDP
CDT
NG LCG
Compressor discharge pressure [kPa]
Compressor discharge temperature [
o
C]
Design Base-line
Under-firing
Coolant
modulation
Coolant
modulation
with 50K
pre-cooling
Fig. 12. Compressor discharge air pressure and temperature.
60
0
Coolant flow rate
Coolant fraction
Total coolant flow rate[ kg/s]
Total coolant fraction [%]
Design Base-line
Under-firing
Coolant
modulation
Coolant
modulation
with 50K
pre-cooling
Fig. 13. Total coolant flow rate and fraction relative to compressor inlet air flow rate.
640
630
620
610
600
590
580
570
560
540
520
500
480
460
440
420
Temperature
Flow
NG LCG
Design Base-line
Under-firing
Coolant
modulation
Coolant
modulation
with 50K
pre-cooling
o
C]
Turbine exhaust temperature [
Turbine exhaust flow [kg/s]
Fig. 14. Turbine exhaust temperature and flow rate.

9. 320
300
280
260
I.H. Kwon et al. / Applied Thermal Engineering 60 (2013) 285e294 293
58.0
57.0
56.0
55.0
54.0
53.0
of the combustion gas. Under-firing, of course, causes a drop in the
exhaust gas temperature. The exhaust temperatures of the two
coolant modulated cases are lower than that of the baseline case
because of the increased coolant supply to the turbine side but they
are sufficiently higher than that of the under-firing case.
Fig. 15 shows the combined cycle performance. The trend in the
power output comparison is the same as in the gas turbine per-formance.
However, the efficiency trend is different. In particular,
coolant-modulated operation exhibits a slightly higher combined
cycle efficiency than the under-firing operation. This is because
coolant-modulated operation has a relatively higher gas turbine
exhaust temperature as shown in Fig. 14, which is definitely posi-tive
in terms of the bottoming cycle performance (i.e., it generates
more steam turbine power). The efficiency gap between normal
coolant modulation and coolant modulation with pre-cooling is
nearly insensible because the heat rejected from the coolant is
transported into the bottoming cycle. Thus, more power is gener-ated,
which minimizes the efficiency penalty due to pre-cooling.
The maximum available power augmentation by firing low calo-rific
value gas (i.e., by simply switching fuel from natural gas to low
calorific value gas) is 47MW(309MWof baseline operation versus
262MWof design operation). However, this operation significantly
affects hot section life. Coolant modulation using 50 K pre-cooling
produces 300 MW of net power output, which guarantees 80% of
the maximum available power augmentation while maintaining
blade metal temperatures at the design levels. Therefore, coolant
modulation, especially when assisted by coolant pre-cooling, is a
better solution compared to under-firing.
4. Conclusion
The results of this study are summarized as follows:
(1) Firing low calorific value gas in the gas turbine increases power
output due to the increase in fuel flow. The increased turbine
gas flow causes the pressure ratio to increase, which is the
major cause of turbine blade overheating.
(2) Under-firing could be an easiest solution to keep the blade
temperature below the design values. However, it causes a
considerable reduction in power output.
(3) Modulation of the coolant flow rates of all cooled turbine blade
rows provides a greater gas turbine power output compared to
under-firing. Moreover, coolant pre-cooling enhances the po-wer
output further by reducing the coolant supply.
(4) Coolant modulation is slightly disadvantageous in terms of gas
turbine efficiency. However, the relatively higher gas turbine
exhaust temperature provides much higher combined cycle
efficiency compared to the under-firing, and provides 80% of
the maximum available power augmentation. The recovery of
the rejected heat from the coolant by the steam cycle in the
case of coolant pre-cooling is advantageous from the viewpoint
of combined cycle performance. Therefore, coolant modula-tion,
especially when assisted by coolant pre-cooling, is a better
solution than under-firing.
Acknowledgements
This work was supported by the New Renewable Energy
Center/Korea Energy Management Corporation through the
“Design and Construction of 300 MW IGCC Demonstration Plant in
Korea” project funded by the Korean Ministry of Knowledge
Economy.
References
[1] J. Hoffmann, J. Tennant, G.J. Stiegel, Comparison of Pratt and Whitney Rock-etdyne
IGCC and Commercial IGCC Performance, 2006. DOE/NETL-401/
062006.
[2] R.A. Dennis, W.W. Shelton, P. Le, Development of Baseline Performance Values
for Turbines in Existing IGCC Applications, 2007. ASME paper GT2007-28096.
[3] J.J. Lee, Y.S. Kim, K.S. Cha, T.S. Kim, J.L. Sohn, Y.J. Joo, Influence of system
integration options on the performance of an integrated gasification com-bined
cycle power plant, Applied Energy 86 (2009) 1788e1796.
[4] M. Rieger, R. Pardemann, H. Rauchfub, B. Meyer, Effects of ASU integration on
IGCC performance and gas turbine operation, VGB Power Tech5 (2008) 102e107.
[5] E.O. Oluyede, J.N. Phillips, Fundamental Impact of Firing Syngas in Gas Tur-bines,
2007. ASME paper 2007-27385.
[6] Y.S. Kim, J.J. Lee, T.S. Kim, J.L. Sohn, Y.J. Joo, Performance analysis of a syngas-fed
gas turbine considering the operating limitations of its components,
Applied Energy 87 (2010) 1902e1911.
[7] Y.S. Kim, J.J. Lee, T.S. Kim, J.L. Sohn, Effects of syngas type on the operation and
performance of a gas turbine in integrated gasification combined cycle, Energy
Conversion and Management 52 (2011) 2262e2271.
[8] A. Franco, N. Giannini, Perspectives for the use of biomass as fuel in combined
cycle power plants, International Journal of Thermal Sciences 44 (2005) 193e197.
[9] M. Rodrigues, A. Walter, A. Faaij, Performance evaluation of atmospheric
biomass integrated gasifier combined cycle systems under different strategies
for the use of low calorific gases, Energy Conversion Management 48 (2007)
1289e1301.
[10] S. Colantoni, S.D. Gatta, R. De Prosperis, A. Russo, F. Fantozzi, U. Desideri, Gas
turbines fired with biomass pyrolysis syngas: analysis of the overheating of
hot gas path components, Journal of Engineering for Gas Turbines and Power
132 (2010) 061401-1e061401-8.
[11] A. Walter, J. Llagostera, Feasibility analysis of co-fired combined-cycles using
biomass-derived gas and natural gas, Energy Conversion Management 48
(2007) 2888e2896.
[12] A. Bhattacharya, D. Manna, B. Paul, A. Datta, Biomass integrated gasification
combined cycle power generation with supplementary biomass firing: energy
and exergy based performance analysis, Energy 36 (2011) 2599e2610.
[13] K.K. Gupta, A. Rehman, R.M. Sarviya, Bio-fuels for the gas turbine: a review,
Renewable and Sustainable Energy Reviews 14 (2010) 2946e2955.
[14] D.W. Kang, T.S. Kim, K.B. Hur, J.K. Park, The effect of firing biogas on the
performance and operating characteristics of simple and recuperative cycle
gas turbine combined heat and power systems, Applied Energy 93 (2012)
215e228.
[15] F. He, A. Li, P. Liu, L. Ma, E.N. Pistikopoulos, Operation window and part-load
performance study of a syngas fired gas turbine, Applied Energy 89 (2012)
133e141.
[16] S. Amaral, T. Verstraete, R.V. Braembussche, T. Arts, Design and optimization
of the internal cooling channels of a high pressure turbine blade e part I:
methodology, ASME Journal of Turbomachinery 132 (2010) 021013.
[17] T. Verstraete, S. Amaral, R.V. Braembussche, T. Arts, Design and optimization
of the internal cooling channels of a high pressure turbine blade e part II:
optimization, ASME Journal of Turbomachinery 132 (2010) 021014.
[18] G. Nowak, W. Wróblewski, Optimization of blade cooling system with use of
conjugate heat transfer approach, International Journal of Thermal Sciences
50 (2011) 1770e1781.
240
52.0
Power
Efficiency
NG LCG
Power [MW]
Efficiency [%]
Design Base-line
Under-firing
Coolant
modulation
Coolant
modulation
with 50K
pre-ccoling
Fig. 15. Combined cycle performance comparison among different operations.

10. 294 I.H. Kwon et al. / Applied Thermal Engineering 60 (2013) 285e294
[19] G. Nowak, I. Nowak, Shape design of internal cooling passages within a tur-bine
blade, Engineering Optimization 44 (2012) 449e466.
[20] GE Power-Enter Software, GateCycle Version 6.0, 2006.
[21] T. Ginter, T. Bouvay, Uprate Options for the MS7001 Heavy Duty Gas Turbine,
GE Power Systems, 2006. GER3808C.
[22] E. Gebhardt, The F Technology Experience Story, GE Power Systems, 2000.
GER3950C.
[23] F.J. Brooks, Ge Gas Turbine Performance Characteristics, GE Power Systems,
2000. GER3567H.
[24] R. Farmer, Gas Turbine World, Handbook, vol. 27, Pequot Publishing Inc.,
2009, pp. 75e91.
[25] I.H. Kwon, D.W. Kang, S.Y. Kang, T.S. Kim, Influence of precooling cooling air
on the performance of a gas turbine combined cycle, Transactions of the
Korean Society of Mechanical Engineering B 36 (2012) 171e179 (in Korean).
[26] T.S. Kim, S.T. Ro, Comparative Evaluation of the Effect of Turbine Configu-ration
on the Performance of Heavy-duty Gas Turbines, 1995. ASME paper
95-GT-334.
[27] K. Kawaike, N. Kobayashi, T. Ikeguchi, Effect of new blade cooling system with
minimized gas temperature dilution on gas turbine performance, Journal of
Engineering for Gas Turbines and Power 106 (1984) 756e764.
[28] A. Cohn, M. Waters, The Effect of Alternative Turbine Cooling Schemes on
the Performance of Utility Gas Turbine Power Plants, 1982. ASME 82-JPGC-GT-
19.
[29] R. Martin, J. Donohue, The 7FA Gas Turbine: A Class Reimagined, in: GE Gas
Turbine Technology Symposium, Greenville, SC, USA, 2009.
[30] R. Chacartegui, D. Sanchez, J.M. Munoz de Escalona, A. Munoz, T. Sanchez, Gas
and steam combined cycles for low calorific syngas fuels utilization, Applied
Energy 101 (2013) 81e92.