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Journal of Engineering Sciences, Assiut University, Vol. 38, No. 5, pp.1181-1195, September 2010.
1181
SIMULATION OF GAS-STEAM TURBINE COMBINED CYCLE
WITH WATER INJECTION AT COMPRESSOR INLET
B. Saleh Mechanical Engineering Dept., Faculty of Engineering, Assiut
University, Assiut, Egypt, bahaa_saleh69@yahoo.com
(Received July 21, 2009 Accepted August 2, 2010)
This investigation presents a computer simulation of gas-steam turbine
combined cycle used in hot dry weather countries. Water is injected into
compressor inlet air to alleviate the inlet air temperature and in turn
improve the gas turbine output power and efficiency. The gas turbine
simulation model is used following [1] and a standard Rankine steam
cycle is employed in cascade with the gas turbine cycle. The turbine
exhaust gases, instead of disposing to the ambient, are used in three heat
exchangers arranged in series to provide all the required heat for steam
cycle. The heat exchangers are simulated by the effectiveness-NTU
method to obtain the off-design performance. The steam properties and
the psychometric air properties are evaluated during the simulation by a
linked computer code developed by the author. The results show that, the
energy recovery in steam turbine produce power as much as one half of
gas turbine output power. A combined efficiency in the range of 45% is
reached at design points.
KEYWORDS: Combined cycles, Power cycles, Gas turbine, Steam
turbine, Heat recovery steam generator
NOMENCLATURE
Latin symbols A heat transfer area
m mass flow rate
c specific heat 1T Inlet temperature difference
C heat capacity rate 2T
2T exit temperature difference
q heat flux mT
Logarithmic mean
temperature difference U overall heat transfer coefficient
Acronyms
CC combined cycle GT gas turbine
CIT compressor inlet temperature ST steam turbine
HRSG heat recovery steam generator TIT turbine inlet temperature
-NTU effectiveness - number of transfer units
B. Saleh 1182
1. INTRODUCTION
The outstanding features of combined cycle (CC) power plants become more attractive
with its increasing usage in power market. The features include its high efficiency in
utilizing energy resources, low environmental emissions, short duration of
construction, low initial investment cost, low operation and maintenance cost, and
flexibility of fuel selection, etc. Thus, CC power plants are quite competitive in the
power market. In gas turbine (GT) power plant, heat recovery schemes are one of the
most important ways of increasing the efficiency of the power generation, usually with
the aim of improving performance.
The main disadvantage of the steam turbines (ST) is the fact that they operate
at relatively low temperatures. Even at steam ultra-supercritical parameters, 280 bar
and 580 C, the overall efficiency of these turbines does not exceed 48.5% when
operating with a vacuum condensation. The gas turbines are free of this drawback. The
temperature of flue gases at turbine inlet can reach up to 1400 C. An important
disadvantage of GT power plants is the need of high excess air, over the stoichiometric
amount, required to maintain the temperature at the combustion chamber outlet at the
required level. In order to provide this excess, from 50 to 66% of the mechanical
energy produced in the turbine is consumed in the air compressor. This leads to a
decrease in the thermal efficiency of GT that reaches up to 38.5%. In the case of waste
heat utilization by a ST cycle, the combined efficiency can reach 58.5% [2, 3, 4].
The deterioration of GT cycle efficiency due to relatively high ambient
temperatures is considered as a challenging issue. In hot weather countries especially
in relatively dry environments, water injection into the compressor inlet air is a
common technique. The water injection improves the GT performance due two folds.
The water injection reduces the inlet air temperature and will improve the combustion
characteristics in the combustion chamber and hence reduces the flue gases emissions.
However, the hot exhaust gases still carry a big portion of energy added in the
combustion chamber which is considered as an energy loss in addition to its effect as
thermal emissions. The cascading of the GT cycle with ST is considered to be a good
solution whenever it is appropriate.
The influence of operating parameters of the heat recovery steam generator
(HRSG) on the overall performance of the CC was studied by [5]. They used an
optimization method to achieve high efficiency and cost reduction. Also, an
optimization scheme of the HRSG in the view of thermodynamic and thermoeconomic
aspects were carried out by [6]. The results of the optimization lead to a meaningful
increase of the thermal efficiency of the plant that approaches the 60%.
A CC simulation model for four GT configurations was investigated in [7].
Their results showed that the reheated GT is the most desirable overall, mainly because
of its high turbine exhaust gas temperature and resulting high thermal efficiency of the
bottoming steam cycle. The optimal GT cycle will lead to a more efficient CC power
plant.
References [8, 9, 10] have emphasized that the operation of a CC thermal-
power plant is influenced by the conditions that are present at the place where it is
installed, mainly ambient temperature, atmospheric pressure and the air relative-
humidity. These parameters affect the generated electric-power and the heat-rate
during operation. Among these variables, the ambient temperature causes the greatest
SIMULATION OF GAS-STEAM TURBINE COMBINED CYCLE … 1183
performance variation during operation. Since the ambient air conditions play an
important role on the performance of CC power generation plants, their effects still
need further investigations.
In the current investigation, a combined gas-steam turbine model is used to
simulate the performance (output power and efficiency) of power station at design and
off-design operations. The simulation model consists of GT cycle model and a
bottoming ST cycle model coupled by the HRSG. The GT model is studied before in
details [1] and the ST model is implemented in this study along with the operational
characteristics of the HRSG. Also the steam condenser with a cooling tower is
simulated in the view of ambient air psychometric parameters. The investigated steam
cycle is the typical Rankine cycle where the heat addition occurs in three sequential
heat exchangers. The heat rejection from the ST cycle is removed by a typical water-
cooled condenser. The cooling water is circulated through a cooling tower, in which
the ambient psychometric properties have been considered. The impact of GT
operation on its combined ST results on only two parameters. These parameters are the
flue gasses temperature and the mass flow rate. The design point parameters for ST are
implemented to match that of the GT. The HRSG is thermally designed for the design
point parameters.
2. COMBINED CYCLE SIMULATION MODEL
The simulation model consists of two parts. The first part for the GT model is taken
from [1]. Their model has been used to simulate the GT cycle under different ambient
conditions. They primarily, investigated the effects of ambient conditions on the
compressor performance following its manufacturer map. They studied different
control operations for both compressor and combustion chamber operation. The
compressor was operated under constant speed and constant pressure ratio whilst the
combustion chamber was simulated for either constant fuel or fixed exit temperature
operations. The second part is the cascading of the ST cycle in the bottom of the
forementioned GT cycle. The present simulation investigates the overall performance
of the CC. The simulation accounts for the coupling of the operational parameters for
both top and bottoming cycles. Primarily, the coupling mechanism is the HRSG. The
degree of coupling depends basically on the HRSG performance at design and off-
design operations.
The HRSG consists of three counter flow heat exchangers, as shown in Fig. 1.
Following the GT, the first heat exchanger is used as super heater for steam cycle, the
second as evaporator and the third as economizer. The three heat exchangers are
designed in accordance with the parameters at the design point of the GT and hence,
the ST design point is obtained. For selected fixed inlet pressure and temperature to the
ST, the steam mass flow at design point is obtained following the HRSG design
parameters.
The present simulation has been conducted for the following design parameters
for GT cycle: compressor inlet air at 40 °C, 20 % RH without water injection,
compressor isentropic efficiency 0.80%, turbine inlet temperature and pressure 1300
°C, 9.4 MPa, turbine mass flow rate 18.11 kg/s, and turbine isentropic efficiency
0.85%. The outlet parameters for GT cycle at design condition: flue gases temperature
B. Saleh 1184
767 °C, turbine power 12.53 MW, compressor power 6.28 MW, net power 6.25 MW,
and cycle thermal efficiency 26.1%. The design parameters for ST cycle are: turbine
inlet temperature and pressure 500 °C, 12.5 MPa, steam mass flow rate 3.83 kg/s,
turbine isentropic efficiency 0.85%, and condenser pressure 9.14 kPa. The outlet
parameters for ST cycle at design condition: HRSG chimney temperature 100°C,
turbine power 4.24 MW, and cycle thermal efficiency 35.1%.
Fig. 1. Schematic diagram of combined cycle
At design point, the following relations estimate each heat exchanger heat
transfer area:
),1(mT
qUA
(1)
)2(
ln2
1
21
T
T
TTTm
(2)
where:
q: heat flux,
1T : inlet temperature difference,
UA: product of overall heat transfer coefficient by heat transfer area
2T : exit temperature difference,
mT : logarithmic mean temperature difference.
WGT
WST
Fogging
chamber
Combustion
chamber
Tg4 (Tchimney) Tg1
Condenser
Tg2 Tg3
Pump
Injected water
Intake air
Gas turbine
Steam turbine
Compressor
Cooling water from
cooling tower
Steam generator Economizer Superheater
SIMULATION OF GAS-STEAM TURBINE COMBINED CYCLE … 1185
The effect of ambient conditions, used in GT, is accounted for in the operation
of the steam condenser cooling tower. The ambient conditions, dry bulb temperature
and relative humidity, are used to estimate the cooling water temperature and hence the
operating condenser pressure. The present simulation code incorporates the
psychometric relations of water vapor air mixture. Whilst the wet bulb temperature is
estimated according to the ambient RH, the cooling tower is assumed to provide the
condenser with cooling water with temperature greater than wet bulb temperature by 7
°C. The temperature difference of the cooling water across the condenser is assumed as
10 °C. The difference between the saturation temperature of the condensed steam and
the exit cooling water temperature is assumed as 5°C. This operating characteristic of
steam condenser and cooling tower describes the coupling between ambient RH and
condenser operating temperatures. This is the only effect of ambient conditions on the
steam cycle, while the great effect of the ambient conditions is fully considered in GT
simulation.
It is presumed that a power plant is basically operated on the design condition;
however, most power plants operate on the off-design condition, which is caused by
the variation of working conditions (ambient conditions and water injection at
compressor inlet). The impact of GT operation at off-design condition on steam cycle
is due to the shift of GT flue gases temperature and mass flow away from the design
point. The variation of climate parameters such as ambient temperature and relative
humidity are the main causes of GT operation at off-design. Also these variations
affect the operation of steam condenser cooling tower.
The variation of flue gases temperature and their mass flow rate influence
greatly the operation of the three heat exchangers (super heater, evaporator, and
economizer). The Effectiveness-Number of Transfer Units (-NTU) method has been
used to get the off-design parameters for each heat exchanger as in [11]. Figure 2
displays the temperature distribution for both gas and steam at design and off-design
operations. The solid lines represent the design point in which the GT operates, at
constant fuel consumption without water injection, at the forementioned design
parameters. The dashed lines represent the off-design operations which have the same
operating conditions as the design points, in addition water injection system is
employed at compressor inlet. Two samples of water injection of 0.04 and 0.14 kg/s
are shown in the figure. In all operations the ST operates at the same pressure and
temperature (12.5 MPa, 500 °C). That is the ST operations have a fixed temperature
and pressure at design and off-design operations except marginal change of feed water
temperature in lieu of variation of condenser pressure due to the effect of RH on
cooling tower performance. The off-design operations of ST have been accounted for
by varying the steam mass flow rate. It is assumed that the control of steam turbine is
adjusted for mass flow control with fixed inlet steam parameters. It is seen that the
minimum pinch point is obtained at the design point operation. The off-design steam
mass flow rate is then estimated in accordance with each heat exchanger performance
and with the limit of pinch point temperature. The pinch point is defined as the
difference between the temperature of gas leaving the evaporator (economizer side)
and the steam saturation temperature. The pinch point is the main parameter that
influences the dimension of the HRSG, and consequently its thermal performances and
its cost. The higher the value of the pinch point, the worse the efficiency of the HRSG,
but, on the other hand, the less its costs are. Usually the values of pinch point are
B. Saleh 1186
chosen according to the experience of the manufacturer, and range typically between 5
and 20 oC [4, 6, 12]. A pinch point of 10 °C has been used at the design point in the
present investigation.
Distance along heat exchanger
0
100
200
300
400
500
600
700
800
Tem
pera
ture
(o C
)
Steam design & off operation
Gas design operation
Gas off-design, 0.04 kg/s water injection
Gas off-design 0.14 kg/s water injection
Fig. 2. Heat exchangers design and off-design temperature distributions
The following relations simulate the off-design performance of the heat
exchangers:
min
min
cmC , max
max
cmC (3)
r
r
CNTU
r
CNTU
eC
e
1
1
1
1 (4)
where:
m : mass flow rate of either steam or flue gases,
c : specific heat of either steam or flue gases,
C : heat capacity rate,
rC : heat capacity rate ratio = max
min
C
C,
: effectiveness; actual heat transfer rate by maximum possible heat transfer
rate (max
q ),
maxminmaxTCq
,
maxT : the maximum temperature possible difference; the difference between
inlet temperatures of both hot and cold fluids,
NTU: number of transfer units = UA/Cmin.
SIMULATION OF GAS-STEAM TURBINE COMBINED CYCLE … 1187
The variations of ambient temperature and relative humidity are used in the
developed steam properties software to get the wet bulb temperature and hence, the
cooling tower performance parameters. The cooling water inlet to the steam condenser
is then used to give an estimate for the condenser saturation temperature and pressure.
The ST output power is calculated at each operating point then the CC performance
parameters are obtained.
All the required steam thermodynamic properties in the present simulation
process are obtained from a computer code developed by the author. The
thermodynamic properties code is based on the equation of state at superheat, saturated
gas and saturated liquid regions as in [13]. Also, all ambient psychorometric properties
are obtained from both steam and air equations of state included in the present code.
The code has been validated with the traditional steam tables and psychorometry charts
properties. This code can be easily extended to get the thermodynamics properties for
any working fluid upon implementing the appropriate equation of state.
The GT data is taken from [1]. Their simulation of GT was based on
parametric investigation of the effects of ambient conditions. Some of these conditions
are taken from the environmental data such as ambient temperature and relative
humidity. Other conditions are deliberately used to get benefits out of dry weather by
water injection at compressor inlet. The mechanism of water injection has two
improvements. Primarily, it decreases the inlet air temperature due water evaporation
according to air psychometric properties which enhance the GT cycle efficiency.
Secondarily, the decrease of compressor inlet temperature (CIT) increases the GT cycle
mass flow rate according to the compressor operating characteristics which in turn,
increases the net output power. The effects of the fore mentioned parameters had been
investigated under two different operating conditions of GT plant. These operating
conditions are; constant turbine inlet temperature (TIT) and constant fuel consumption.
In the present code the following relations are applied:
GT
GTnet
GT
Q
W
,
(5)
., compGTGTnet WWW
(6)
HVmQ fGT
(7)
GT
STnetGTnet
CC
Q
WW
,,
(8)
ST
STnet
ST
Q
W
,
(9)
pumpSTSTnet WWW
, (10)
41 ggpgST
TTcmQ
(11)
Where:
GT : GT cycle thermal efficiency,
GTnetW ,
: net GT cycle output power
B. Saleh 1188
GTQ
: rate of heat added to GT cycle
GTW
: GT power,
.compW
: compressor power,
ST : ST cycle thermal efficiency,
STnetW ,
: net ST cycle output power,
STQ
: rate of heat added to ST cycle
STW
: ST power,
.pumpW
: pump power,
gm
: flue gases flow rate,
pc : flue gases specific heat,
41, gg TT : flue gases temperatures inlet to and exit from HRSG (Fig. 2)
fm : fuel flow rate,
HV : fuel heating value,
CC : CC thermal efficiency.
The present investigation uses the previous GT data of [1], in which the data
are divided into six groups:
In the first group of data, the effect of variation the amount of water injection
on the GT cycle performance at constant TIT is studied. The ambient air enters the
fogging chamber with ambient air condition of 40 C, 20% RH and flow rate ranges
from 17.64 kg/s to 18.37kg/s. The flow rate of water injection ranges from 0 to 0.14
kg/s. The injected water increases the relative humidity from that of the ambient up to
96% and decreases the inlet air temperature down to 23 C. Practically, using a higher
humid air ranges may damage the compressor due to formation of dense water
droplets. A relative humidity of 85% could be assumed as maximum limit after which
the inceptions of water droplets may occur.
In the second group of data, the same parameters are investigated as in the first
group but with constant fuel GT operation.
In the third, fourth and fifth groups of data, a study of the effect of inlet air
temperature on the GT cycle performance at different constant relative humidity, at the
conditions of constant TIT and without water injection, are considered. A wide range
of ambient conditions is selected to investigate the CC performance. The ambient air
enters the compressor with temperature ranged from 5 to 50 °C. The ambient relative
humidity in the three groups is 20, 40, and 80% respectively.
In the sixth group of data, a study of the effect of relative humidity on the GT
cycle performance at constant TIT is considered. The ambient air enters the
compressor with 40 °C and the relative humidity ranged from 10 to 90% without water
injection.
SIMULATION OF GAS-STEAM TURBINE COMBINED CYCLE … 1189
3. RESULTS AND DISCUSSION
In the CC simulation process, the GT cycle operational parameters are taken from [1].
While in the ST cycle, the inlet conditions of steam to the steam turbine are fixed at
500 °C and 12.5 MPa. Hence, the upper part of steam cycle is fixed while the
condenser pressure varies according to the ambient conditions. Event the present
results are computational, the curves are used with symbols to differentiate between the
multiple curves plot in all figures.
In groups 1 and 2 of data the only change of CIT is due to water injection
system in which the inlet temperature changes from 40 °C to 23 °C. In groups 3, 4, and
5 of dada the CIT is varied as a simulation parameter from 5 °C to 50 °C. Therefore,
owing to the compressor map characteristics, this result in narrow range of mass flow
rate in groups 1 and 2 compared with that of other three groups. In the sixth group of
data the CIT is kept unchanged at 40 °C and hence the air mass flow is almost kept
constant while the RH varies.
Figure 3 shows the relation between the mass flow rate of flue gases from the
GT and the flue gases temperature for the first five groups. The data of groups 1 and 2
are plotted as dashed lines with the upper extended x-axis. The remaining three groups
are plotted as continuous lines with the lower x-axis.
As shown in Fig. 3, the flue gases temperature decrease as the mass flow
increase for all groups. This is basically due to the controlling parameters of the
compressor according to its map as in [1], in which the compressor operating
parameters are controlled such that the compressor efficiency remains constant. This
way of compressor control increases the outlet pressure with the increase of air mass
flow rate through the system. While the GT isentropic efficiency and TIT are kept
fixed, the increase of turbine inlet pressure will lead to a reduction of turbine outlet
temperature. In case of constant fuel the flue gases outlet temperature decrease even
further than CIT operation. The constant fuel operation reduces the TIT because of
increase of air mass flow rate due to operation of compressor at constant efficiency.
In the case of constant GT isentropic efficiency with fixed TIT operation
(group 1), due to reduction of CIT the air mass flow rate through the compressor
increased from 17.64 to 18.37 kg/s. This leads to increase the turbine pressure from
9.41 to 9.86 MPa and hence the flue gases temperature slightly decreases (8°C
reduction) from 767 to 759 °C. While in the case of constant GT isentropic efficiency
with constant fuel operation (group 2), due to reduction of CIT the air mass flow rate
increased from 17.64 to 18.31 kg/s. This leads to increase the turbine pressure from
9.40 to 9.93 MPa while the air mass flow rate increase leads to a reduction in TIT from
1303 to 1242 °C due to constant fuel, and hence a considerable decrease occurs for the
flue gases temperature (54°C reduction) from 769 to 715 °C. By the same reasoning,
the flue gas temperature decreases as well in all remaining groups. It is observed that
the amount of reduction of flue gases temperature at constant TIT operation (group-1)
is much smaller than in the case of constant fuel operation (group-2). Thus, the main
fact results in this figure is as the mass increase the exhaust temperature decrease. In
general, the turbine exhaust temperature decreases at off-design operation in all
investigated data.
While the specific heat is fairly unchanged, the heat content of flow gases is
proportional to the product of both flue gas temperature and mass flow rate. This heat
B. Saleh 1190
content reaches the maximum value at design point and decreases when the system is
operated away from the design point.
The impact of variation of the flue gases flow rate and temperature, leads to
increase the flow rate of steam up to the design point and decrease after worth as
shown in Fig. 4 for groups 3-5. Operating at off-design condition, the steam mass flow
is characterized by the three heat exchangers performance, which grantee the pinch
point requirement and the functional operation for each heat exchanger to provide the
fixed ST inlet conditions. The steam mass flow rate depends proportionally on the
product of flue gas mass flow rate and temperature. As mentioned before, as the
product decreases away from the design point, this will make steam flow rate attains a
peak value at design and deteriorates away from it. It should be noted that in Fig. 4 the
design points for all groups have a fixed air mass flow rate, this leads a marginal
change of flue gases mass flow rate in case of constant TIT. In the simulation for
groups 1 and 2, the mass flow rate of flue gases starts from that of design point and up
while for the other three groups the simulation starts with mass flow rate less than that
of design point and up beyond design point. For this reason, as shown in Fig .4, the
steam mass flow rate for groups 1 and 2 is maximum at design point and decreases as
the flue gas mass flow rate away from the design point. For the other three groups the
steam mass flow rate increases with the increase of flue gases until it reaches its
maximum value at design point and decreases after that. It is observed that the amount
of reduction of steam mass flow rate at constant TIT operation is much smaller than in
the case of constant fuel operation due to off-design operation of HRSG, which
associated the reduction of flue gases temperature.
The influence of water injection at constant TIT (data group-1) on both cycle
performance (power and efficiency) are displayed in Fig. 5 for an injected water flow
rate range from 0 - 0.14 kg/s. Increasing the amount of water injected leads to decrease
in the CIT from design point temperature (40 °C) down to 23 °C which in turn reduces
the flue gases temperature and hence rising the GT cycle power and efficiency. As
shown at design points, the GT efficiency is in the mid twenty and ST efficiency in
mid thirty while the CC efficiency in mid forty values. The decrease of CIT due to
water injection actually has two opposite folds. Firstly, increases the gas cycle
efficiency from 26% at design point (CIT = 40 °C) to 27% at maximum water injection
0.14 kg/s (CIT = 23°C). Secondly, reduces the steam cycle efficiency from 35% (Tg1 =
767 °C) to 33% (Tg1 = 759 °C) due to the HRSG off-design performance
characteristics, which results a nearly constant combined efficiency. On the other hand
for the generated power, there is a quiet small increase in GT power from 6.2 to 7.0
MW due the reduction of CIT, and a marginal decrease in steam power from 4.25 to
4.19 MW due to the reduction of flue gases temperature. The resultant is increasing of
the CC power by about 6.7% (0.7 MW).
Figure 6 illustrates the results of variation water injection with constant fuel
supply to GT combustor (group-2). In this case the TIT reduces with water injection by
about 61 C from the design point of 1303 C to 1242 C. This descending in the TIT
values is substantially due to the lower air temperatures entering the combustion
chamber that resulting in lower temperatures of the flue gases by about 54 C from the
design point of 769 C to 715 C. As shown from the figure that the decrease of CIT
due to water injection, primarily improves the gas cycle performance in terms of both
SIMULATION OF GAS-STEAM TURBINE COMBINED CYCLE … 1191
power and efficiency by about 2.7 %. But, as mentioned before the reduction of TIT
reduces the GT exhaust temperature which in turn deteriorates the steam cycle power
by about 16.3% and efficiency by about 11.4%. The resultant overall performance of
CC slightly reduced by about 3.8% for power and 4.3% for efficiency. It is observed
from Figs. 5 and 6 that the amount of reduction of steam cycle power and efficiency at
constant TIT operation is smaller than in the case of constant fuel operation due to off-
design operation of HRSG, which associated the reduction of flue gases temperature.
Figures 7-9 display the effect of ambient temperature on CC performance at
selected ambient relative humidity of 20, 40, and 80% (data groups 3-5). There is a
little increase in GT output power as the ambient relative humidity increase at the same
ambient temperature, while the GT efficiency is almost unchanged. This is a
consequence of the change of inlet humid air composition, which results in a change of
the equivalent molecular weight of air mixture. This will alter the compressor
operation to a new operating point in view of its map. On the other hand, the steam
cycle power and efficiency decrease as the RH increases at the same ambient
temperature due to the reduction of cooling tower cooling capacity because of higher
ambient humidity. Other conclusion can be drawn from Figs. 7-9 that the steam cycle
performance (power and efficiency) peaks at the design point [(40 C, 20% RH), (40
C, 40% RH), and (40 C, 80% RH)] for each case and deteriorates away at off-design
operation. This trend is due to the performance of HRSG at design and off-design
operation. On the other hand, the gas cycle performance decreases as CIT increases at
constant RH. In this case, there is no peak at the design point because the gas cycle
does not affect with the HRSG performance.
The effects of ambient relative humidity on the CC performance are shown in
Fig. 10, with fixed ambient temperature at 40 °C and constant TIT without fogging
system. As shown in the figure, the GT power increases little bit from 6.24 MW to 6.46
MW (3.5%), while the efficiency is almost unaffected by the increase of ambient
relative humidity from 10% to 90%. On the other hand, the steam cycle performance
deteriorates, where both the power and efficiency decrease by about 3.6% and 4.7%
respectively as the ambient RH increases from 10% to 90%. This may be attributed to
the deterioration of cooling tower performance. This results in a decrease of the CC
efficiency from 43 % to 42 % and little increase of CC power from 10.37 MW to 10.44
MW. It is concluded from the previous analysis that the operation of the combined
system become more efficient at the lower humidity condition of air.
In fact, using the fogging system in the gas cycle gave two main functions. The
first is the reduction of air CIT and the second is the raising the RH at compressor inlet.
To illustrate these effects, the simulation modeling of the CC is conducted one with
using the fogging system and the other without fogging system. Where the two systems
operate at the same conditions of ambient air temperature and RH.
Figure 11 displays the performance data along with CIT against the
compressor inlet air relative humidity due to fogging system at which the ambient
conditions are fixed at 20% and 40 °C. The X symbols on the plot show the same
parameters against relative humidity due to the change of ambient relative humidity
without fogging system at the same CIT at three selected point of operations (20, 40,
and 80 % RH). The small deviation at some points is due to the interpolation between
the data points. From the Figure, it can be withdrawn that the only effect of fogging
B. Saleh 1192
system is the decrease of CIT and the accompanied increase in relative humidity does
not affect the GT performance as concluded from figure 10. While the ST power and
efficiency decrease with ambient RH because of deterioration of cooling tower
effectiveness as ambient RH increase.
24
28
32
36
40
44
Eff
icie
ncy,
,
%
2
4
6
8
10
12
Pow
er (
MW
)
0 10 20 30 40 50
Compressor inlet temperature (oC)
Combined power & efficiency
Steam power & efficiency
Gas power & efficiency
Fig. 7. Combined cycle performance at
ambient 20% RH without fogging
system and constant TIT
24
28
32
36
40
44
Eff
icie
ncy,
,
%
2
4
6
8
10
12
Pow
er (
MW
)
0 10 20 30 40 50
Compressor inlet temperature (oC)
Combined power & efficiency
Steam power & efficiency
Gas power & efficiency
Fig. 8. Combined cycle performance at
ambient 40% RH without fogging
system and constant TIT
24
28
32
36
40
44
Eff
icie
ncy,
,
%
2
4
6
8
10
12
Pow
er (
MW
)
0 10 20 30 40 50
Compressor inlet temperature (oC)
Combined power & efficiency
Steam power & efficiency
Gas power & efficiency
Fig. 9. Combined cycle performance at
ambient 80% RH without fogging
system and constant TIT
24
28
32
36
40
44
Eff
icie
ncy,
,
%
2
4
6
8
10
12
Pow
er (
MW
)
0 20 40 60 80 100
Ambient relative humidity, %
Combined power & efficiency
Steam power & efficiency
Gas power & efficiency
Fig. 10. Combined cycle performance at 40 oC
ambient temperature without fogging
systemat and constant TIT
Gas & steam design point Gas & steam design point
Gas & steam design point Gas & steam design point
SIMULATION OF GAS-STEAM TURBINE COMBINED CYCLE …
1193
24
28
32
36
40
44
Eff
icie
ncy,
,
%
2
4
6
8
10
12
Pow
er (
MW
)
0 20 40 60 80 100
Relative humidity, %
20
24
28
32
36
40
44
Com
p. i
nlet
tpm
p. (
o C)
Compressor inlet temp.
Combined power & efficiency
Steam power & efficiency
Gas power & efficiency
Gas & steam design point
CC performance at three climate RH
Fig. 11. Combined cycle performance with fogging system at ambient conditions 40oC,
20% RH and constant TIT
4. CONCLUSION
The combined cycle power plants take advantages of both gas turbine and steam
turbine benefits. The main disadvantage of gas turbine is the major deterioration of its
performance due to the increase of climate ambient temperature. In hot and dry climate
environment the use fogging system at the compressor inlet alleviate this drawback.
The current study presents a simulation model to evaluate the performance of gas-
steam turbine combined cycle with water injection at compressor inlet.
From the present result the sole effect of the fogging system is the reduction of
compressor inlet temperature and hence increases gas turbine power and efficiency. It
is clear that this scheme is more efficient in dry than relatively humid climate.
The use of steam turbine combined with gas turbine get advantages from
fogging system as well. However, the enhancement is marginal because of the
deterioration of the heat recovery steam generator at off-design operation.
The design point of heat recovery steam generator is a critical issue in the
operation of combined cycle with fogging system. If the demand of power plant vary
extensively, the loss of efficiency is quit pronounced. In this operation, a multiple heat
recovery steam generator in parallel arrangements is recommended, which enable
multiple design point operations.
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الدخول للضاغطالمركبة مع حقن ماء عند يةوالبخار يةالغاز التربين دورةل نظرية محاكاة
هذذ الاحث ذذثل محذذالم ظرذذظبلستخ ذذ لثظمذذوتراللاحرمث ذذزوخلحذذرزخالاحكذذظولزاحثتذذظخلاحمخرثذذ لزاحممذذوترم ل ذذ لاحث ذذرا ل ا لزلثظحوذظح لوو ذرلرذاللوذ ولل ق لاحهزاءلاحراتالح ضظغطلثظحمذظءلحوهرةذ لرخجذ ل خاخلل له بلاحرزخالاحجزلاح ظخلاحجظف.
رمذظلرزخالاحوذخث لاحكظو ذ لحذم ظرذظبللمذالاحلوذلل.احمو صذالل هذظلمذ لاحذرزخارخالقذاحلاحرفظءالاح خاخ لح رزخالزصظ لم احذذخثطلثذذذ للوذذذلل.ةحذذذ لرزخالاحوذذخث لاحكظو ذذ ثظمذذوتراللاحثتذذظخلرزخالخاسرذذ لاحق ظمذذذ لحذذللوذذللةضذذذظ لل1احمخجذذرلخ ذذذلل ذذ ل
حذذ ثلل ذذ لاءلاحجذذز ,,ثذذر ملمذذ لطخرهذذظلح هذذزللامذذوتراللغذذظوا لاح ذذظرللمذذ لرزخالاحوذذخث لاحكظو ذذ لاحذذرزخو للذذ لطخ ذذ احح ح لاح خاخ لاحمثظر للوصم لولللثظح خاخالاح وم لحهظ.مرارلرزخالاحوخث لاحثتظخ لإلل لل لاحوزاح خاخ ل مثظر
لذذررلز ذذرا لاسوقذذظالاح ذذخاخالز حذذللح صذذزالل ذذ لأراءلاحمثذذظر لاحسذذظءلاح مذذالث ذذرامللذذ لتذذخزفل-طخ قذذ لاحرفذذظءبثموخ لح هزاءلولل مظثهظلل لطخ لثخسظمجلرمث زوخلوللوصم م لثزامذط لزلاحوصم ل.لتزاصلاحثتظخلزاحتزاصلاحم رخل
ا لرزخالاحوذذخث لاحثتظخ ذذ لو طذذ ل ذذرخالاحسوذذظةجللاتهذذخ وذذلللمذذالثخسذذظمجلرمث ذذزوخلحم ظرذذظالاحذذرزخالاحمخرثذذ .للاحثظ ذذث.رفذظءالاحذرزخاللا ضذظملا لاحسوذظةجزرذ حللاتهذخ لوقظخبلسصفل مذ لاحقذرخالاحمو صذالل هذظلمذ لرزخالاحوذخث لاحكظو ذ .ل
احوصم م .تخزفلاحوشك اللاح مالو ل%للسر54 زاح لاحمخرث ل