Performance improvement of gas turbines of Fars (Iran) combined cycle power plant by intake air...

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Energy Conversion and Management 48 (2007) 1055–1064

Performance improvement of gas turbines of Fars (Iran) combinedcycle power plant by intake air cooling using a media evaporative cooler

R. Hosseini a,*, A. Beshkani a, M. Soltani b

a Mechanical Engineering Department, Amirkabir University of Technology, 424 Hafez Avenue, P.O. Box 15875-4413 Tehran, Iranb Mechanical Systems Department, Niroo Research Institute (NRI), End of Pounak Bakhtari Boulevard, P.O. Box 14665-517, Shahrak Gharb, Tehran, Iran

Received 24 April 2006; accepted 29 October 2006Available online 13 December 2006

Abstract

In this paper, a media evaporative cooling system installed in the gas turbines of the Fars (Iran) combined cycle power plant are mod-eled and evaluated. In this model, different design parameters such as inlet air velocity, geometric shape and sizes and depth of the mediahave been considered. Analysis of the results shows that at a constant Prandtl number, the cooler effectiveness decreases and pressuredrop increases, as incoming air velocity increases. The water evaporation rate and temperature drop of the cooler have been calculated,and the effects of temperature, relative humidity and pressure of the ambient have been investigated as well. The practical and analyticalresults show that the output of a gas turbine of the Fars combined cycle power plant at 38 �C ambient temperature and 8% relativehumidity is 11 MW more, and the temperature drop of the inlet air is about 19 �C with media evaporative cooling installations. Becauseof annual power augmentation of 5280 MWh and considering 2.5 Cents/kWh selling price for excess electricity generated, the paybackperiod obtained is about four years.� 2006 Elsevier Ltd. All rights reserved.

Keywords: Media evaporative cooler; Power augmentation; Fars power plant; Gas turbine

1. Introduction

The gas turbine is designed to operate with a constantair volume flow in the compressor. When the ambient airtemperature increases, its specific mass is reduced, so thatthe mass flow rate entering the turbine is accordinglydecreased. This would, in turn, decrease the power outputof the gas turbine. Strictly speaking, the power output isdirectly dependent on the mass flow rate of the air in thecycle. For each degree celsius increase of the air tempera-ture, 0.5–0.9% and 0.27% of the power output of the gasturbine and the combined cycle is reduced, respectively [1].

The simple open cycle gas turbine and the T–S diagramfor this cycle are shown schematically in Fig. 1. As is evi-dent from the diagram, the constant pressure lines in thecompressor (1–4) move upward when the ambient temper-

0196-8904/$ - see front matter � 2006 Elsevier Ltd. All rights reserved.

doi:10.1016/j.enconman.2006.10.015

* Corresponding author. Tel.: +98 21 66405844; fax: +98 21 66419736.E-mail address: [email protected] (R. Hosseini).

ature increases. Because of the divergence of the constantpressure lines, the compressor work would be increased.Conversely, by increasing the environmental temperature,the output pressure of the compressor would decrease,and this would cause a reduction of the cycle efficiency.Also, because of the reduction of the air density due tothe increasing temperature, the heat rate would decrease,and hence, the specific fuel consumption would beincreased. From the T–S diagram, one could easily con-clude that on a hot summer day and for a maximum con-stant temperature of the cycle (the maximum temperatureof the gas entering the turbine is fixed by metallurgical con-siderations), the pressure ratio would decrease. Conse-quently, the turbine work would be reduced.

In Fig. 2, the variation of the mass flow rate of air,power output and heat rate of a gas turbine with ISOconditions (15 �C, 60% relative humidity) with temperatureof the air entering the compressor is shown. As is clearfrom this figure, when the inlet air temperature increases

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Nomenclature

a0, b0, w0 parameters to define corrugated surfaceB barometric pressure, kPaD dimensionless depth, d/sD+ hydrodynamic dimensionless depthd depth of media, mmE evaporation rate of water, L/minQ gas turbine air volume flow rateP dimensionless pressurep pressure, Paps saturated water vapor pressurepv moist air water vapor pressurePe Peclet numberPr Prandtl numberR gas constant, 0.287 kJ/kgKRH relative humiditys mean plate space (mm), entropyT dimensionless temperaturet temperature, �C~V dimensionless velocity vector~v velocity vector

u,v,w x,y,z components of velocity

Greek symbols

a thermal diffusivitye cooler effectiveness (e = T2)�e e/ecf

m kinematics viscosityq densityx humidity ration, g, f independent variables for transformed spaceDP dimensionless pressure dropDp pressure dropDP DP/DPcf

Subscripts

1 incoming air to cooler2 leaving air from coolercf channel flow (b0 = 0.0)db dry bulbwb wet bulb

Compressor inlet air temperature (˚C)

Co

rrec

tio

n (

%)

-20 -10 0 10 20 30 40 5075

80

85

90

95

100

105

110

115

120

PowerAir FlowHeat Rate

Fig. 2. Variation of power output, flow rate, and heat rate withcompressor inlet air temperature.

1056 R. Hosseini et al. / Energy Conversion and Management 48 (2007) 1055–1064

from 15 �C to 32 �C, the power output decreases by about10%.

During the summers, with long and hot days, the gasturbine power generators’ power output goes down con-siderably, relative to the regular operating temperatureof cold or mild spring and fall seasons. The requiredpower on these days increases because of operating airconditioning and cooling systems for residences, offices,and marketplace as well as factories. Strictly speaking,the production and consumption go in opposite direc-tions. Installation of new or use of reserved power gener-ation units for this period is one way to overcome thisproblem. Another way is to recover part of the lostpower due to the hot weather by cooling the air enteringthe gas turbine.

There are different methods of cooling the inlet air.These methods could be differentiated as:

s

t

1

2

3

4

p=constantp=constant

a b

Fig. 1. The simple open cycle gas turbine (a), and the corresponding T–S diagram (b).

R. Hosseini et al. / Energy Conversion and Management 48 (2007) 1055–1064 1057

• Evaporative cooling using media evaporative cooling orwater spraying to the inlet air (fogging).

• Cooling by use of thermal energy storage, chilled waterstorage or ice harvesting.

• Cooling the inlet air by use of mechanical chiller orabsorption chiller.

According to the technical and economic study forinstallation of the different possible methods for the envi-ronmental condition of Iran (37� North), evaporative cool-ing (fog and media) is more advantageous than the otheroptions. This is especially true for those areas with warmand dry conditions [1]. The advantages of media evapora-tive cooling include:

• Increasing the power output up to 14% [2].• Small operation and maintenance costs (3–5% of initial

installation cost) [2].• Low investment cost (approximately $50 for each kWh

additional power production) [3].• Simplicity of the system and equipment.• Reduction of NOx (0.8–1.5% for each degree Celsius of

air cooling).• Low water consumption (0.6–1.2 m3 for each MW addi-

tional power production).• Low power consumption for operating pumps (less than

0.5% of additional power generation) [1].

There are also some disadvantages:

• Limitation of air velocity: The velocity of air should below to prevent high pressure drop when flowing throughthe media and to prevent water carry over.

• The use of the system in regions with high humidity dur-ing summer has the limitation of increasing poweroutput.

• The shutdown period for the installation of the system(10 days) is longer than the installation of the fog system[1].

• Pressure drop is relatively high (approximately 0.5% ofthe power output of the gas turbine) [1].

Many researchers have studied the performanceimprovement of combined cycle power plants and gas tur-bine units by intake air cooling in recent decades.

Johnson [4,5] presented a discussion of the theory andoperation of evaporative coolers for industrial gas turbineinstallations. Calculations of the parameters to predict theperformance of an evaporative cooler were included inaddition to installation, operation, feed water quality andthe causes and prevention of water carry over.

A comparison was made between using two differenttypes of air coolers, namely water spraying system andcooling coil, to improve the performance of gas turbinepower plants, by Alhazmy and Najjar [6]. The performancecharacteristics have been examined for a set of design andoperational parameters including ambient temperature, rel-

ative humidity, turbine inlet temperature, and pressureratio.

It has been reported that directing a controlled spray ofwater into the compressor inlet will result in a 10–20%increase in power output [7,8]. Water fogging is anothercooling method. It is reported that a power enhancementby 10% [8] to 20% [9] would be reached.

Ameri and Hejazi [10] studied gas turbine power aug-mentation using an absorption chiller. The study showedthat the power output could be increased by about 11.3%in a typical hot day, and the estimated payback periodwas 4.2 years.

The influence of ambient temperature on combinedcycle power plant performance was investigated by Arrietaand Lora [11]. They showed that a drop in the ambienttemperature increases both the electric power generatedand efficiency in the plant, and the rise in the temperatureof the gas after the supplementary firing increases the gen-erated electric power, but reduces the efficiency.

Several cases of chilled water and ice storage systemshave been studied for the capacity enhancement in the Kish(Iran) power plant based on the variation of the coolingload. The best option was recommended and selected aftereconomical analysis. This study shows that the combustionturbine capacity enhancement will be around 13.6% byusing the best selected method [12].

Power augmentation of a typical gas turbine cycle byusing a desiccant based evaporative cooling system wasstudied by Zadpoor and Golshan [13]. The techniquerequires a desiccant based dehumidifying process to directthe air through an evaporative cooler, which could beeither media based or spray type. The performanceimprovement of the gas turbine cycle for each of thesemethods, direct evaporative cooler, indirect and directevaporative cooler and desiccant based evaporative cooler,was studied for different climatic conditions.

A steam absorption chiller was used to cool the intakeair to the desired temperature level. The power output ofthe gas turbine increased by about 10.6% and that of thecombined cycle power plant by around 6.24% annually.Economic analysis shows that the payback period will beabout 3.81 years [14].

Dai and Sumathy [15] applied an integration method tosolve the governing equations for a cross flow direct evap-orative cooler and assumed channels with constant spac-ing. They used honeycomb paper as packing material andinvestigated the effect of sizes on cooler effectiveness.

Using a numerical simulation, the effectiveness and pres-sure drop of a media evaporative cooler that is equippedwith corrugated papers as a wetted medium have been cal-culated by Beshkani and Hosseini [16]. They showed thatthe corrugation shape and media sizes have a significanteffect on the cooler performance.

In the present study, using the above mentioned numer-ical simulation [16], the model is improved using differentboundary conditions in the height direction of the humidair channel. The economic aspects of cooling the intake


air of gas turbines of the Fars power plant are studied, andthe temperature drop, evaporation rate and power aug-mentation are investigated for various ambient tempera-tures and relative humidities.

2. Description of the system

Evaporative coolers are used in gas turbine intake aircoolers and residential applications. The Fig. 3 shows aschematic arrangement of the main parts of a media evap-orative cooling system. Water is sprayed upward from aheader pipe into the top of an inverted half pipe and isdeflected downward onto a distribution pad on top of themedia. The distribution pad facilitates the water entry tothe media by gravitational force and wets a very large sur-face area formed by several layers of the media. The excesswater (not absorbed by the media and not evaporated bythe hot air) is collected in the bottom part of the coolerand forwarded to a reservoir tank. The hot air passesthrough the media and evaporates the water up to the sat-uration point before entering the compressor. In this pro-cess, the temperature of the air drops by both sensibleheat transfer because of the temperature differences

Fig. 3. Schematic of the media evaporative cooler.

Fig. 4. (a) Configuration of corrugated papers, (

between the water and air and the latent heat ofevaporation.

The air velocity entering the media should not be morethan a certain value because of the possibility of watercarry over of droplets to the compressor. These dropletswould damage the compressor blades. The water ispumped from the reservoir tank to the header pipe. Inorder to prevent the water from becoming hardened by thisprocess, a certain amount of fresh water is always injectedinto the reservoir tank, while part of the water retunedfrom the media is drained. Hence, the overall water con-sumption of the cooler consists of the amount of evapora-tion of water and the blow down value. The value of blowdown depends on the hardness of the make up water,capacity of the reservoir, and evaporation rate [5]. In orderto eliminate water carry over, a mist eliminator is ofteninstalled on the downstream side of the media.

The position of installation of the media is also veryimportant. The position is chosen by the designer of thesystem and is normally placed after the air filters, whichhas the advantages of preventing dust and other particlesfrom the air to adhere to the media, keeping them clean,and preventing the increase of pressure drop.

3. Modeling and assumptions

The active parts of the evaporative cooling media are themedia themselves. The media consists of several wavy thinlayers of special corrugated paper or cellulose corrugatedpaper packed together. The layers are alternately arranged,so that the direction of the corrugation in one layer isapproximately perpendicular to that of the adjacent layer.The surface area of each corrugated layer (depending, ofcourse, on the frequency and amplitude of the cosine func-tion of the wave) is approximately 2.5 times a similar sizeflat plate layer. This means a much greater surface areafor contact between the air and water. The combinationof several of these paper layers provides many narrowchannels for the air and water flow. The schematic arrange-ment of these corrugated plates is shown in Fig. 4a. Forprevention of water carry over and better contact betweenthe air and water as well as increasing the efficiency of the

b) System coordinates and physical domain.


cooling process by increasing the evaporation, these corru-gated paper layers are arranged with different angles suchas 15� and 45� (not considered in this model).

In the model presented here, the channel is regarded aslayers with a flat plate on one side and a vertically corru-gated plate on the opposite side. In Fig. 4b, the physicalmodel and the system coordinates are shown. The watersprays perpendicularly along the z direction, while the airflows in the x direction. The depth of the air channel is d

and the average distance between two adjacent plates(mean plate space) is s, which is equal to its value for achannel flow.

4. Governing equations

To simplify the model, the air–water interface tempera-ture (channel walls temperature) is assumed to be the wetbulb temperature of the air. This allows solving the govern-ing equations of air in a period of cosine waves. Symmetricboundary conditions have been used to solve the governingequations in a half period of the cosine waves instead of theperiodic boundary conditions that have been used by Bes-hkani and Hosseini [16] (Fig. 4a).

For an incompressible flow of air, the governing equa-tions, including continuity and conservation of momentumand energy, can be written in the following forms:

ðaÞ r � ~V ¼ 0

ðbÞ ð~V � rÞ~V þrP ¼ 1

Rer2~V

ðcÞ ð~V � rÞT ¼ 1

Per2T

ð1Þ

All dimensional lengths are converted to non-dimensionallengths by use of s, and the following non-dimensionalparameters are employed:

P ¼ pqu2

1

; ~V ¼ ~vu1

; T ¼ t � tdb1

twb � tdb1

Re ¼ u1sm; Pr ¼ m

a; Pe ¼ RePr

ð2Þ

In the problem at hand, certain boundary conditions areapplied:

1. On the channel walls, no-slip and isothermal conditionsare employed.

2. In the inlet, uniform velocity and temperature areconsidered.

3. In the outlet, fully developed conditions are assumed fortemperature and velocity.

4. On the upper and lower boundaries, symmetric bound-ary conditions are used.

5. Because of use of projection scheme to solve the govern-ing equations, Neumann boundary conditions are usedfor pressure at all the boundaries.

The dimensionless form of the cosine function of thecorrugated layer is regarded as Eq. (3).

a0 þ b0 cosðw0ZÞ � Y ¼ 0 ð3ÞUnless otherwise specified, a0 = 1.0, b0 = 0.5, w0 = p,Pr = 0.7 and s = 6.7 mm have been used. Introducing thefollowing algebraic relations will transform the non-rectan-gular physical domain into a rectangular domain to imple-ment the finite difference method.

n ¼ X

g ¼ Ya0þb0 cosðw0ZÞ

f ¼ Z

8><>:

9>=>; ð4Þ

5. Results and discussion

To solve the dimensionless form of the governing equa-tions, we need the Reynolds and Prandtl numbers, non-dimensional depth of the channel (D) and non-dimensionalform of the cosine wave for the media (a0,b0,w0).

Almost all of the available experimental results for pres-sure drop and effectiveness of the media give the variationof these two parameters with incoming air velocity (u1) anddepth of the media (d). Since the calculated pressure dropwith the present model is non-dimensional, to make a com-parison of different results, the pressure drop will be givenin a dimensional corrected form (in Pa). From the differenttypes, sizes and materials commercially available for mediapackages, those with cellulose material and commercialidentity of CELdek 7060-15 with average distance betweenadjacent layers of 6.7 mm have been chosen for the presentproject that has been installed in the Fars power plant [17].

To calculate the pressure drop, Eq. (5) is used

Dp ¼ 1:1qu21DPDP cf

DP cf ¼1

2Dþ

� �6:88ffiffiffiffiffiffiffi

Dþp þ

24þ 0:674Dþ � 6:88ffiffiffiffiffi

Dþp

1þ 4:64� 10�4ðDþÞ�2

!

Dþ ¼ DRe¼ dm

u1s2

ð5Þ

In the above equation, DPcf is the non-dimensional pres-sure drop of a channel flow with constant spacing of s

and length of d [18], 1.1 is a correction factor to considerthat the walls are completely wetted [17], and DP is the ra-tio of the calculated pressure drop by the model at b0 = 0.5to its value at b0 = 0.0 (channel flow).

The left side of the channel that is assumed flat here ishorizontally corrugated in the actual media evaporativecooler. As it is clear, the pressure drop increases whenthe left side is considered a horizontally corrugated plate.We considered the same effect on pressure drop for the hor-izontally corrugated plate and the vertically corrugatedplate. So, the calculated pressure drops are multiplied bya factor of DP . The value of DP is estimated from Fig. 5and is approximately 2.0, so the pressure drop is as follows:

Dp ¼ ð1:1DP Þ2qu21; DP cf ¼ 4:84qu2

1DP cf ð6ÞThe above mentioned points show that the pressure drop ofthe media evaporative cooler is approximately five times

Re200 400 600 800

1.25

1.5

1.75

2 ΔP

D=15

D=22

D=30

ε

Fig. 5. Percentage increase in pressure drop and effectiveness for b0 = 0.5in comparison to b0 = 0.0.


the pressure drop of a dry channel flow. Using Eqs. (5) and(6), the pressure drop of a media evaporative cooler can becalculated from the following equation:

Dp ¼ 2:42Dþqu21

6:88ffiffiffiffiffiffiffiDþp þ

24þ 0:674Dþ � 6:88ffiffiffiffiffi

Dþp

1þ 4:64� 10�4ðDþÞ�2

!ð7Þ

5.1. Model validation

Fig. 6 shows a comparison between the results obtainedby the model and available experimental data. The figureshows the result of the implemented symmetric and periodicboundary conditions in this study and [16], respectively andthe experimental results. As is evident from the figure, theexperimental and model results are in good agreement.

5.2. Effect of corrugation shape

In this part, the effect of b0 is investigated, which havenot been considered in the available experimental results.

u1 (m/s)

ε

0 0.5 1 1.5 2 2.5 3

50

60

70

80

90

(%)

[16]

[16]

Pres.

Pres.

[17]

[17]

d=150 mm

d=200 mm

a b

Fig. 6. Model and experiment comparison:

Variation of b0 (the parameter defining the corrugated sur-face) does not change the inlet area of the channel, so whenthe value of b0 changes, the mass flow rate of the channelremains in a constant velocity condition.

The effect of b0 on effectiveness and pressure drop isshown in Fig. 7 for a non-dimensional depth of 15. Thisfigure shows that increasing b0 would increase the effec-tiveness, and the effect is more pronounced at higher Rey-nolds numbers. At low Reynolds numbers, b0 wouldapparently not play an important role. However, forgas turbine power generators, such low Reynolds num-bers are not practical. Evidently, the minimum pressuredrop would be for b0 = 0.0, with minimum effectivenessas well. In other applications of the media evaporativecooler when there is no lower limit on the air velocity,the plate layer of the media with constant distance of afew millimeters may be more effective because of thelower pressure drop and reasonable effectiveness. Forthe present case and for practical Reynolds numbers of700–800, the use of paper with a flat plate layer in themedia means that the effectiveness is less than 30% at anon-dimensional depth of 15. Strictly speaking, althoughthe pressure drop is less, the media would not provideenough cooling of air.

5.3. Effect of Prandtl number

Fig. 8 shows the variation of the effectiveness of thecooler with Reynolds numbers for various Pr numbersand D = 22. Increasing Pr number decreases the effective-ness. The figure indicates the importance of the quality ofthe air entering the media and passing through it.Increasing the Pr number, resulting from air contamina-tion of dirt, oily particles or any substance causing thisincrease (also due to poor performance of filters) wouldreduce the media effectiveness, in addition to makingthem dirty, which, in turn, increases the pressure drop,as earlier stated.

u1 (m/s)

p(P

a)

0.5 1 1.5 2 2.5

5

10

15

20

25

30

35

Pres.[17]

Δ

d=150 mm

d=200 mm

(a) effectiveness, and (b) pressure drop.

u1 (m/s)0.5 1 1.5 2

1.3

1.4

1.5

1.6

1.7

1.8

1.9

2

2.1

2.2b0=0.4b0=0.5b0=0.6b0=0.7

u1 (m/s)

_

0.5 1 1.5 21

1.5

2

2.5

3

3.5

4

4.5

5

5.5

b0=0.4b0=0.5b0=0.6b0=0.7

ΔP

a b

Fig. 7. Effect of corrugation shape for D = 15 on: (a) effectiveness, and (b) pressure drop.

Re200 400 600 800

55

60

65

70

75

80

85

90

Pr=0.6Pr=0.7Pr=0.8Pr=0.9

ε (%

)

Fig. 8. Effect of Pr on effectiveness, D = 22.


6. Technical and economic details of Fars media installation

If the incoming air temperature (ambient dry bulbtemperature), relative humidity, evaporative cooler effec-tiveness and ambient pressure are known as input parame-ters, the density, the moisture of the incoming air and themoisture in the air leaving the evaporative cooler can becalculated. For this reason, to calculate first the wet bulbtemperature of the air at a desired barometric pressure,the following equations are solved simultaneously insteadof using a psychrometric chart (temperature in �C andpressure in Pa) [15,19].

ln ps ¼ 23:1964� 3816:44

t � 46:13

pv ¼ psw �1:8ðB� pswÞðtdb � twbÞ

2800� 1:8twb � 32

pv ¼ psd � RH

ð8Þ

Then, the dry bulb temperature of the air leaving theevaporative cooler and the water evaporation rate is calcu-lated. The results for e = 0.85, Q = 354 m3/s and B = 85kPa are shown in Fig. 9.

x ¼ 0:622ps

pt � ps

; q1 ¼B

Rðtdb1 þ 273Þtdb2 ¼ tdb1 � eðtdb1 � twbÞE ¼ 60q1Qðx2 � x1Þ

ð9Þ

The figure shows the water evaporation rate (L/min), out-put temperature, temperature drop and increase in relativehumidity against the inlet temperature of air and relativehumidity of the environment. These figures are for 85%effectiveness of the media at an air flow rate of 354 m3/sand ambient pressure of 85 kPa (at elevation of 1530 mfor Fars power plant), where the power plant is installed.From these figures, it is clear that at high humidity situa-tion, the media performance is weak. This evidently indi-cates that this method of air cooling is not suitable for ahot and humid atmosphere.

The effect of ambient pressure on temperature drop iscalculated from Eqs. (8) and (9), and the results forRH1 = 10 and RH1 = 40% and Q = 354 m3/s are shownin Fig. 10 (the cooler effectiveness assumed as 85%). Thefollowing correlations are estimated for the temperaturedrop for the ambient temperature between 10 �C and50 �C.

Dtð�CÞ � 0:383tdb1 þ ð4:282� 0:237iÞ RH 1 ¼ 10%

Dtð�CÞ � 0:188tdb1 þ ð3:231� 0:136iÞ RH 1 ¼ 40%

i ¼ 0; 1; 2; 3; 4 for B ¼ 100; 95; 90; 85; 80 ðkPaÞð10Þ

This indicates that the ambient pressure could have a con-siderable effect on the calculated temperature drop, andhence, if the barometric pressure was considered greaterthan its actual value, the power output of the gas turbinewould be underestimated. For example, if the ambientpressure was considered as 100 kPa instead of 80 kPa, the

t1 (˚C)

t 1-t 2

10 20 30 40 50 60

5

10

15

20

25RH1=10 %

20 %

30 %

40 %

50 %

60 %

70 %

t1 (˚C)

t 2(˚

C)

10 20 30 40 50 600

5

10

15

20

25

30

35

40

45

50

RH1=10 %

20 %

30 %40%50 %60 %70 %

t1 (˚C)

E(L

/min

)

10 20 30 40 50 60

20

40

60

80

100

120

140

160

180

RH1=10 %20 %

30 %

40 %

50 %

60 %

70 %

t1 (˚C)

RH

2-R

H1(

%)

10 20 30 40 50 6020

25

30

35

40

45

50

55

60

65

70

RH1=10 %

20 %

30 %

40 %

50 %

60 %

70 %

a b

c d

Fig. 9. Output temperature (a), temperature drop (b), evaporation rate (c), and increase of relative humidity (d) of cooler for various ambienttemperatures and relative humidity, B = 85 kPa.

t1 (˚C)

t 1-t

2

20 25 30 35 40

11

12

13

14

15

16

17

18

19

20

B=80 kP859095100

RH1=10 %

t1 (˚C)

t 1-t 2

20 25 30 35 40

7

8

9

10

11

B=80 kPa859095100

RH1=40 %

a b

Fig. 10. Effect of ambient pressure on temperature drop: (a) RH1 = 10%, and (b) RH1 = 40%.


temperature drop is underestimated by 1 �C forRH1 = 10%. This inaccuracy results in a 0.5 MW underes-timation in power augmentation.

The design condition of the media evaporative coolingfor the Fars combined cycle power plant is 35 �C and20% relative humidity [20].

Ambient temperature (˚C)

Ou

tpu

tin

crea

se(%

)

20 30 40 502

3

4

5

6

7

8

9

10

11

12

13

14

1560 % Relative Humidity50 %40 %30 %20 %10 %

Fig. 11. Power augmentation of Fars media installation for variousambient temperature and relative humidity.


Fig. 11 indicates the increase of power output of thepower plant with ambient temperature and relativehumidity for the installed system. From this figure, it isevident that the increase of power output would be moreat a lower relative humidity. Based on the data collectedfor evaluation of the system performance for units 1 and2 of the power plant at conditions of 38 �C and 8% rela-tive humidity, the percentage increase in power outputhas been 14.6% and 13.3% (11.2 MW and 10.8 MW),respectively.

The investment cost is $50 per kWh of additional elec-tricity generated, since $550,000 was invested for thisincrease of power generated during the hot season.

The gas turbine installation cost is $275 per kW inIran, so considering this 11 MW additional electricitygenerated, the investment for the same would be$30,25,000, and its economic justification is quite clear.During the summer season in the Fars province with480 hours an year of operation (4 h a day for 120 days),the electricity generated as a result of the installation ofthe media cooling would be 5280 MWh annually. Thecurrent cost of electricity in Iran is 2.5 Cents/kWh; hencethe payback period is calculated to be 4.2 years. With thecost of electricity at 3 Cents/kWh and 4 Cents/kW h, thepayback period would be 3.5 years and 2.6 years,respectively.

7. Conclusions

With the results and discussion presented above regard-ing the installation of media evaporative cooling for gasturbine electric energy generators, the following could beconcluded:

• Increasing the velocity of the inlet air would increase thepressure drop, while decreasing the effectiveness.

• Increasing the media depth would increase both effec-tiveness and pressure drop.

• With opposite effect of depth of the media and velocityof air on effectiveness, in application when the highvelocity of the entering air is inevitable (such as gas tur-bine with high flow rate requirement as a result of limi-tation of the passages) media with longer depth shouldbe used for better performance.

• Design of corrugation shape and size of the media for aminimum possible pressure drop and maximum possibleeffectiveness are important factors. Generally, the geo-metrical shapes of the media and the parameters incosine function of the corrugation have a profoundeffect on the effectiveness and pressure drop.

• It is important in design to regard the actual ambientpressure that has a considerable effect on the estimatedpower augmentation.

• The pressure drop of the media evaporative cooler isapproximately five times the pressure drop of a drychannel flow.

• Economic consideration of the system with practicalresults of installation show that the return of investmentwould be about four years with the sale price of addi-tional electricity generated at 2.5 Cents/kW h.

References

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