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.

REFERENCES

1. Imhamed M. Saleh, “Gas Turbine operation with inlet water injection in Libyan

conditions”, Master Diploma Work, Warsaw Uni. of Technology, Warsaw, 2005.

B. Saleh

1194

2. Heppenstall, T. “Advanced gas turbine cycles for power generation: a critical

review”, Applied Thermal Engineering, 18, 837-846, 1998.

3. Koleva, N., Schaberb, K., and Koleva, D. “A new type of a gas-steam turbine cycle

with increased efficiency”, Applied Thermal Engineering, 21, 391-405, 2001.

4. Shin, J.Y., Jeon, Y.J., Maeng, D.J., Kim, J.S., and Ro, S.T. “Analysis of the

dynamic review” Applied Thermal Engineering, 18, 837-846, 1998.

5. Manuel Vales, and Jose L. Rapun “Optimization of heat recovery steam generators

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6. Alessandro Franco, and Alessandro Russo “Combined cycle plant efficiency

increase based on the optimization of the heat recovery steam generator operating

parameters”, International Journal of Thermal Sciences, 41, 843–859, 2002.

7. Polyzakis, A. L., Koroneos, C., and Xydis, G. “Optimum gas turbine cycle for

combined cycle power plant”, Energy Conversion and Management, 49, 551–563,

2008.

8. Boonnasa, S., Namprakai, P., and Muangnapoh, T. “Performance improvement of

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Energy, 31, 2036–2046, 2006.

9. Felipe R. Ponce Arrieta, and Electo E. Silva Lora “Influence of ambient

temperature on combined-cycle power-plant performance”, Applied Energy, 80,

261–272, 2005.

10. Yousef S.H. Najjar “Efficient use of energy by utilizing gas turbine combined

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11. Holman, J. P. “Heat Transfer”, Ninth Edition, McGraw-Hill, 2002.

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SIMULATION OF GAS-STEAM TURBINE COMBINED CYCLE …

1195

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