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INTERNATIONAL JOURNAL OF ENERGY RESEARCHInt. J. Energy Res. 2007; 31:1292–1306Published online 13 February 2007 in Wiley InterScience(www.interscience.wiley.com) DOI: 10.1002/er.1302

Brayton refrigeration cycle for gas turbine inlet air cooling

Galal M. Zaki1,*,y, Rahim K. Jassim2 and Majed M. Alhazmy1

1Department of Thermal Engineering and Desalination Technology, King Abdulaziz University, P. O. Box 80204,Jeddah 21 587, Saudi Arabia

2Department of Mechanical Engineering Technology, Yanbu Industrial College, P. O. Box 30436,Yanbu Industrial City, Saudi Arabia

SUMMARY

In this paper, a new approach to enhance the performance of gas turbines operating in hot climates isinvestigated. Cooling the intake air at the compressor bell mouth is achieved by an air Brayton refrigerator(reverse Joule Brayton cycle) driven by the gas turbine and uses air as the working fluid. Fraction of the airis extracted from the compressor at an intermediate pressure, cooled and then expands to obtain a cold airstream, which mixes with the ambient intake. Mass and energy balance analysis of the gas turbine and thecoupled Brayton refrigerator are performed. Relationships are derived for a simple open gas turbinecoupled to Brayton refrigeration cycle, the heat rejected from the cooling cycle can be utilized by anindustrial process such as a desalination plant. The performance improvement in terms of power gain ratio(PGR) and thermal efficiency change (TEC) factor is calculated. The results show that for fixed pressureratio and ambient conditions, power and efficiency improvements are functions of the extraction pressureratio and the fraction of mass extracted from the air compressor.

The performance improvement is calculated for ambient temperature of 458C and 43.4% relativehumidity. The results indicated that the intake temperature could be lowered below the ISO standard with

power increase up to 19.58% and appreciable decrease in the thermal efficiency (5.76% of the site value).Additionally, the present approach improved both power gain and thermal efficiency factors if air isextracted at 2 bar which is unlike all other mechanical chilling methods. Copyright# 2007 John Wiley &Sons, Ltd.

KEY WORDS: gas turbine; Brayton cycle; cooling; reverse Brayton; power enhancement

1. INTRODUCTION

Gas turbine (GT) units are used extensively as prime movers in the power production, oil fields

and industrial applications. The compactness of GT units and short installation time as well as

the high thermal efficiency of the combined cycles encouraged many utilities to consider GT for

*Correspondence to: Galal M. Zaki, Department of Thermal Engineering and Desalination Technology, KingAbdulaziz University, P. O. Box 80204, Jeddah 21 587, Saudi Arabia.

yE-mail: [email protected]

Received 25 September 2006Revised 21 December 2006

Accepted 26 December 2006Copyright # 2007 John Wiley & Sons, Ltd.



their future programs. Generally, GT performance is affected by the weather conditions; in hot

arid areas, the warm air at the compressor intake decreases the air density and reduces the net

output power below the ISO standard. The net power decreases by 5–9% of the ISO-rated

power for every 108C increase above the 158C ISO standard. The effect of the inlet air

temperature on the Brayton cycle performance is a fundamental aspect of GT cyclethermodynamics (Saravanamutto et al ., 2001). Furthermore, in hot summer days the situation

gets worse due to the extensive use of air conditioning that increases the demand on power at

peak periods. The drive to boost the power or degrade the weather effect has motivated the

interest to explore methods for chilling the intake air.

There are few methods to achieve air cooling at the bell mouth of GT air compressors:

evaporative cooling via either water spray or fogging and mechanical methods employing

refrigeration technology. The relative merits and key considerations for each of the methods are

compared in Erickson (2003). Each of these methods has its own advantages and offsetting

disadvantages. The cooling techniques can be broadly classified into direct and indirect

methods.

The direct cooling is accomplished by spraying water at the compressor inlet either through

flexuous media (cellulose fibre) or fogging (droplets size in the order of 20 mm) into the airstream, Ameri et al . (2004). All direct cooling systems lower the intake temperature close to the

ambient wet bulb temperature. Ameri et al . (2004) applied fog-type air-cooling system for a GT

plant where the climatic conditions (T dp ¼ 31–398C and relative humidity between 5 and 15%)

are suitable. For these conditions, 13% power improvement was reported. Johnson (1988)

discussed the use of evaporative cooling technique for GT installations. Meher-Homji et al .

(2002) investigated the effect of nozzles type and droplets size on the performance of GT

engines. Moreover, Bettocchi et al . (1995) and Meher-Homji and Mee (1999) studied the effect

of nozzle size on the humidity ratio levels attainable using fogging systems. Although

evaporative cooling systems have moderate installation, maintenance and operational costs,

they are accompanied with offsetting disadvantages as the low efficiency and high water content

in the combustion air. Problems of water carryover, Tillman et al . (2005), which are hazardous

for compressor blades, are among the reasons to barricade the use of evaporative coolers for GTplants in humid coastal areas.

The problem of humidity is eliminated by using mechanical refrigeration approach that

can reduce the air temperature to any desirable value regardless of the ambient relative

humidity. There are two common approaches for mechanical air chilling: (a) use of refrigeration

units via chilled water coils supplied from thermal storage tanks; and (b) use of exhaust

heat-powered absorption machines. Generally, application of the mechanical air

cooling increases the net power on the expense of the thermal efficiency. For GE LM 6000

GT, an increase of 1% in the power output could be achieved for every 1.688C drop in the

air inlet temperature, Elliot (2001). The economics using absorption machines was examined for

inlet air cooling of cogeneration plants, Ondrays et al. (1991). Similarly, Kakaras et al.

(2004) presented a simulation model for NH3 waste heat-driven absorption machine for cooling

the air intake. Erickson (2003, 2005) presented a study on the aqua-absorption approach andsuggested the combination of the waste-driven absorption cooling with water injection into

the combustion air for power boosting; the concept is termed the ‘ power fogger cycle’.

The drawback of the mechanical chilling is the risk of ice formation either as ice crystals in the

air or as solidified layer on surfaces, such as the bell mouth or inlet guide vanes (Stewart and

Patrick, 2000).

BRAYTON REFRIGERATION CYCLE 1293

Copyright # 2007 John Wiley & Sons, Ltd. Int. J. Energy Res. 2007; 31:1292–1306

DOI: 10.1002/er



There is considerable literature base on studies that compare both evaporative and

mechanical cooling approaches. Mercer (2002) reported that evaporative cooling has increased

the GT power by 10–15%, while the improvement for refrigeration chillers could reach 25%.

Alhazmy and Najjar (2004) concluded that for spray coolers the drop in air temperature by 3–

158C increased the power by 1–7%, while cooling coils improved the net power by 10–18%.Furthermore, Alhazmy et al. (2006) performed analysis for the performance improvement of

open GT units using spray and mechanical cooling methods. They introduced two generic

terms, power gain ratio (PGR) and thermal efficiency change (TEC) factor, for the evaluation of

intake air-cooling approaches. They presented the results in general dimensionless working

charts covering a wide range of working conditions. A paper by Zadpoor and Golshan (2006) is

more specifically concerned with the discussion of the effect of using desiccant-based

evaporative cooling on GT power output. Their study where a computer program was

developed to simulate the GT cycle and the NOx emission showed that the power output could

be increased by 2.1% which agrees with the result of Alhazmy and Najjar (2004). Extensive

overview on the current inlet air-cooling technology and its economic impact on the energy

market can be found in Cortes and Willems (2003), and Darmadhikari and Andrepon (2004).

The objective of the present analysis is to investigate the potential of boosting the poweroutput of GT plants. A novel approach is considered, where a reverse Joule-Brayton air cycle

(some times referred to as air refrigeration cycle or Brayton refrigeration cycle) is used to reduce

the air temperature at the compressor inlet. Portion of the air is extracted from the compressor

at an intermediate pressure and temperature and cooled in an isobaric process rejecting its heat

to a process heat sink. Then, the air irreversibly expands to near atmospheric pressure where the

temperature drops significantly. Mixing the intake ambient air with this cooled stream produces

the required cooling effect at the compressor inlet as seen in Figure 1.

m

G

Fuel,m f

C o m p

T u r b i n e

Combustion

chamber W & el

W &comp W &net

2

3

Qh

&

&

4

Heat exchanger

Expander

6Qout &

7

1

8

o

m&1

T 7 = T

o + ∆T m&o

Mixing

chamber

mo ,T

o ,

1 m

1 ,T

1 ,&

Qo

= 0&

To desalination

unit

Humidity eliminator

W exp

&

&m1 ,T

8 ,

8

m1 = m

o+& & &m

1

&

mo

= (ma+ m

v& &&

o

)

Figure 1. A schematic diagram of a simple gas turbine coupled to Brayton refrigeration cycle.

G. M. ZAKI, R. K. JASSIM AND M. M. ALHAZMY1294


DOI: 10.1002/er



Although the air cycle has low efficiency compared to vapour compression cycles, it has high

reliability and low maintenance cost. In addition, using the air cycle for air conditioning is a

well-developed technology and extensively in use for trains in United Kingdom and Germany,

Butler (2001). At present, the air refrigeration cycle is the backbone of aircraft cabins’ air

conditioning.The performance of a GT power cycle coupled to Brayton refrigeration air cooler is compared

to the basic GT cycle without air cooling at different operation modes.

2. ANALYSIS

Figure 1 shows a schematic diagram of a simple open GT cycle ‘Brayton cycle’ coupled to an air

refrigeration cycle. The power cycle is represented by states 1–2–3–4 and consists of a

compressor, combustion chamber and a turbine. The reverse Brayton refrigeration cycle is

represented by states 1–6–7–8 and consists of a cooling coil and an expansion device. The two

cycles use the same compressor, where the working fluid is divided between the two cycles.

Portion of the compressed air a ’m1 at pressure P6 is extracted from the mainstream cooled in a

heat exchanger to T 7 then expands to the atmospheric pressure and T 8, Figure 1. The hot

ambient intake air at T 0 mixes with the cold stream at T 8 before entering the compressor.

During the operation without cooling, the intake air is at T 0 and states 1 and 0 are identical.

Because of mixing cold air at T 8 with that at T 0 the temperature at the compressor inlet drops by

DT air ¼ T 0 T 1: Therefore, the temperature at state 1 is a function of T 8, which depends on the

extraction pressure at state 6 and the mass flow rate through the cooling cycle, a ’m1: Figure 2 is T – s

presentation of the power cycle without cooling 0– % 2–3–4, the refrigeration cycle 1–6–7–8 and the

power cycle with cooling 1–2–3–4. It can be observed, Figure 2, that the overall net plant power

output of the GT increases by the difference between the two areas 0– % 2–2–1–0 and 1–6–7–8–1.

s

T

4

1

2s2

4s

3

o

2s2

7

6s

8s8

Po

P3

P6

6

Figure 2. T – s diagram for the proposed cycle.



DOI: 10.1002/er



Ambient air at T 0 and o0 enters the mixing chamber with mass flow rate of ’m0; where it is

mixed with the cold air stream having mass flow rate of a ’m1 at temperature T 8. The mixing

chamber delivers the combined stream to the compressor at T 1.

The mass and energy balance about the mixing chamber gives

’m1 ¼ ð ’m

0 þ a ’m

1Þ ð1Þ

For humid air, substitute for ’m0 ¼ ð ’ma þ ’mvÞ and make use of the humidity ratio o ¼ ’mv= ’ma in

Equation (1) to get

’m1 ¼ ’ma

1 að1 þ o1Þ ð2Þ

where o1 is the humidity ratio at state 1 that depends on the conditions of the ambient air and

the cold air stream at state 8. Mass balance of the water vapour gives

o1 ¼ o0 aðo0 o8Þ ð3Þ

If the design employs moisture eliminator system before the expander, Figure 1, then o1

depends on the effectiveness of the moisture removal process and for ideal case all the water

vapour is removed leaving o8 ¼ 0:From the energy balance, we get an expression for the enthalpy of the mixture at the

compressor intake as

h1 ¼ h0 aðh0 h8Þ ð4Þ

In general, the enthalpy of moist air at any state is expressed as

h ¼ ha þ ohv ¼ c paT þ oðhfg þ c pvT Þ ð5Þ

Substituting for the enthalpy at states 0, 1 and 8 using Equation (5), into Equation (4) gives the

air inlet temperature to the compressor as

T 1 ¼ð1 aÞc p0T 0 þ ac p8T 8

c p1

ð6Þ

Air leaves the compressor at two different states, bleed off air at P6 flowing through the Braytonrefrigerator and the rest at P2 as the working fluid for the power cycle. The two pressures are

given as follows:

P6 ¼ xP1 ð7aÞ

P2 ¼ rP1 ð7bÞ

where x is the extraction pressure ratio and r is the pressure ratio.

The temperature of the air leaving the compressor at states 6 and 2 can be estimated assuming

irreversible compression processes between states 1–2 and 1–6 as

T 6 ¼ T 1 þT 1

Zcx

ðxðg1Þ=g 1Þ ð8Þ

T 2 ¼ T 1 þT 1

Zcr

ðrðg1Þ=g 1Þ ð9Þ

where Zc is the compressor polytropic efficiency, it can be described as a function of the

compression ratio, as given by Korakianitis and Wilson (1994) with a value at any pressure



DOI: 10.1002/er



ratio z as follows:

Zcz ¼ 1 0:04 þz 1

150

where z is either x or r ð10Þ

2.1. Brayton refrigeration cycle analysis

The cooling cycle uses bled off air from the main air compressor at P6 and T 6 as seen in Figures

1 and 2. The hot compressed air at P6 and T 6 rejects its heat through a heat exchanger to cooling

water. For an ideal condition, state 7 will have the same pressure as P6 and temperature that

depends on the cooling process. Since many of the desalination plants in the Gulf area are using

combined GT for dual-purpose plants, where waste heat boilers provide energy for desalination

plants, it is suggested here to utilize the rejected heat ( ’Qout; Figure 1) for brine heating. In

general, the temperature at state 7 can be controlled according to the requirements of any

industrial process that requires low-grade heat. In general, the lower limit for T 7 is determined

by the ambient temperature.

Air at state 7 expands in an irreversible process as seen in Figure 1 furnishing the conditions at

the entrance of the mixing chamber. Irreversible expansion process between 7 and 8 yields

T 8 ¼ T 7 T 7Ze 1 1

x

ðg1Þ=g" #

ð11Þ

where the expansion efficiency is Ze. It is worth mentioning that the extraction pressure ratio x is

the main parameter that determines the final cold air temperature T 8 attainable by the cooling

cycle. The cold stream flow rate a ’m1 proceeds to the mixing chamber to cool down the ambient

air entering the compressor. The mixture temperature T 1 depends on the mass flow rate and the

temperature of each stream as seen in Equation (6).

For the Brayton refrigerator, the power out of the expander, ’W exp; and the heat to the joined

process, ’Qout; are computed as

’

W exp ¼ a ’m1ðh7 h8Þ ð12Þ

and’Qout ¼ a ’m1ðh7 h6Þ ð13Þ

In Equations (12) and (13), the enthalpy term is calculated according to Equation (5).

2.2. Gas turbine cycle analysis

Consider an irreversible GT cycle as shown in Figure 2, processes 1–2 and 3–4 are irreversible

and processes 2–3 and 4–1 are isobaric heat addition and rejection, respectively. Processes 1–2s

and 3–4s are isentropic, presenting the process in an ideal cycle. The different components of the

power cycle are considered in the following.

2.2.1. Air compressor. The compression power between states 1 and 2s with extraction at state6, separating the effects of the dry air and water vapour can be written as

’W c ¼ ’mac paðT 2 T 1Þ þ ’mvðhv2 hv1Þ þ a’ma

1 a

c paðT 6 T 1Þ þ a

’mv

1 a

ðhv6 hv1Þ ð14Þ

where hv is the enthalpy of the saturated water vapour at the state’s pressure, Equation (5).



DOI: 10.1002/er



Relating the compressor isentropic efficiency to the changes in temperature of the dry air and

assuming that the water vapour behaves as an ideal gas then

Zcr ¼T 2s T 1

T 2 T 1ð15Þ

Substituting for T 2 and T 6 in terms of T 1, the pressure ratio r and the pressure extraction ratio x

from Equations (8) and (9) in Equation (14) gives the actual compressor power as

’W c ¼ ’ma c pa

T 1

Zcr

ðrðg1Þ=g 1Þ þ o1ðhv2 hv1Þ þ a

1 a

c pa

T 1

Zcx

ðxðg1Þ=g 1Þ þ o1ðhv6 hv1Þ

ð16Þ

Equation (16) is a general expression for the compressor work that considers the effects of the

air bleeding and humidity. If the cooling system is not in operation ‘off’, just substitute a ¼ 0:Further, the equation takes care of the air relative humidity, which can be replaced by zero for

dry air.

2.2.2. Combustion chamber. Heat balance on the combustion chamber (see Figure 1) gives the

heat rate supplied to the integrated cycle as

’Qh ¼ ’mf NCV ¼ ð ’ma þ ’mf Þc pgT 3 ’mac paT 2 þ ’mvðhv3 hv2Þ ð17Þ

where hv2 and hv3 are the enthalpies of water vapour at the combustion chamber inlet and exit

states, respectively.

Substituting for T 2 from Equation (5) gives the input heat rate as

’Qh ¼ ’maT 1 ð1 þ f Þc pg

T 3

T 1 c pa

rðg1Þ=g 1

Zcr

þ 1

þ

o1

T 1ðhv3 hv2Þ

ð18Þ

where f is the fuel to air ratio f ¼ ’mf = ’ma (related to the dry air rate) and has been expressed by

Alhazmy and Najjar (2004) as

f ¼c pgðT 3 298Þ c paðT 2 298Þ þ o1ðhv3 hv2Þ

NCV c pgðT 3 298Þ ð19Þ

2.2.3. Turbine. Applying the first law of thermodynamics to the GT (neglect the potential and

kinetic energy terms) with the assumption of perfect gas behaviour, the power produced by the

turbine is

’W t ¼ ’mtc pgZtðT 3 T 4sÞ ð20Þ

where ’mt is the total gases mass flow rate at the turbine inlet given by

’mt ¼

’ma þ

’mv þ

’mf ¼

’mað1 þ o1 þ f Þ ð21Þ

Substituting for T 4s (assuming isentropic expansion) and ’mt from Equation (21) into Equation

(20) yields

’W t ¼ ’mað1 þ o1 þ f Þc pgZtT 3 1 1

rðg1Þ=g

ð22Þ



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The turbine isentropic efficiency can be estimated using the practical relation recommended

by Korakianities and Wilson (1994) as

Zt ¼ 1 0:03 þr 1

180 ð23Þ

Since, the GT is almost constant volume machine at a specific rotating speed, the

inlet air volumetric flow rate, ’V a; is fixed regardless of the intake air conditions. Equation

(22) can be written in terms of the volumetric flow rate at the compressor inlet state by replacing

’ma by ra ’V a: The moist air density ra is a function of T 1 and the humidity ratio o1 and

can be calculated using the Engineering Equation Solver (EES) software (Klein and Alvarado,

2004).

3. PERFORMANCE OF THE INTEGRATED CYCLE

For the proposed cycle the net power output and heat input can be easily calculated from

Equations (12), (16), (18) and (22). The performance advantage of the present proposed cycle isnot limited to cooling the inlet air but also includes the amount of useful heat used for the

desalination process. For a single-shaft machine, the expander power ’W exp is recovered by a

single shaft and the net power of the cycle may be expressed as

’W net; with cooling ¼ ’W t ð ’W c ’W expÞ ð24Þ

Equation (24) gives the net shaft power in kW, and if the energy utilized for the desalination

process is considered then the net input heat to the cycle can be reduced by the amount of energy

utilized for the desalination process. Therefore, the input useful energy is

’Quseful ¼ ’Qh ’Qout ð25Þ

Let us define a general term that combines the performance of the GT and the

Brayton refrigerator including the heat supplied to the desalination process as usefulefficiency

Zth;u ¼’W net

’Quseful

ð26Þ

The subscript u means that the GT plant is serving other industrial products, so that the

conventional thermal efficiency term is not applicable for this condition. The conventional

thermal cycle efficiency can be deducted from Equation (26) if the energy rejected is not utilized

for any industrial process, i.e. ð ’Qout ! 0Þ:The power generation industry is mainly concerned with the net power gain out of

introducing the cooling cycle. The net power for the GT unit without the cooling cycle is

obtained by introducing a ¼ 0 in Equation (16) to get ’W c; no cooling: The turbine power without

cooling ’W t; no cooling is obtained from Equation (22) using o1 ¼ o0; T 1 ¼ T 0 and f is calculatedusing T % 2 instead of T 2, Equation (19). Therefore,

’W net; no cooling ¼ j ’W t ’W cjno cooling ð27Þ

Alhazmy et al. (2006) has recently proposed generic parameters to evaluate the effectiveness

of GT inlet cooling methods. The PGR and TEC factor are terms that directly reflect the



DOI: 10.1002/er



significance of using air-cooling technology for GT. Let us extend the definition of the PGR to

include any form of useful energy. Then, a useful power gain ratio (PGRu) is defined as

PGRu ¼’W net; with cooling ’W net; no cooling

’W net; no cooling

ð28Þ

In the same way, let us generalize the definition of the TEC of Alhazmy (2006) to include

savings in fuel due to utilization of the reject heat as

TECu ¼Zth; u Zth; no cooling

Zth; no cooling

100% ð29Þ

For the present parametric analysis, let us focus on the gain that can be achieved by using

Brayton refrigerator for a simple open cycle GT plant, for this case the term ’Qout is eliminated in

Equation (25) and Equations (28) and (29) lead to

PGR ¼’W net; with cooling ’W net; without cooling

’W net; without cooling

100% ð30Þ

The change in thermal efficiency is due to cooling only, neglecting any use of the heat rejection

is presented by

TEC ¼Zcy; with cooling Zcy; without cooling

Zcy; without cooling

100% ð31Þ

For a stand-alone GT under specific climatic conditions, both PGR and TEC are zeros. If

Brayton refrigerator is used, PGR increases with the reduction of the intake temperature.

However, the PGR gives the percentage enhancement in power generation; the TEC of a

coupled system is an important parameter to describe the fuel deployment efficacy.

4. RESULTS AND DISCUSSION

In order to investigate the performance of the proposed Brayton refrigerator for intake air

cooling, a computer program has been developed using EES program. Therefore, all

thermophysical properties were determined to the accuracy of the EES software. In particular,

the specific heats of air were taken as temperature and pressure dependent. The calculation

procedure was first verified for the benchmark case of simple open cycle with dry air as the

working fluid, for which a ¼ 0; o0 ¼ 0; f ¼ 0 and assuming isentropic compression and

expansion processes, Equations (16), (18), (20) and (29) leads to the following expression for the

thermal efficiency of the air standard cycle:

Zth ¼ 1 1

rðg1Þ=gð32Þ

For the present analysis, the ambient air at Jeddah, Saudi Arabia (latitude 22.308N and

longitude 39.158E) a typical city with over 40 GT plants operating under severe weatherconditions was considered. On the basis of annual daily average, ambient temperature of T 0 ¼

458C and j0 ¼ 43:4% were selected. The cooling brine leaves the heat exchanger with 10 K

terminal temperature difference (i.e. T 7 ¼ 558C). The brine flow rate can be controlled to obtain

other outlet temperature suitable for the need of the industrial process. Table I shows the range

of the different parameters.



DOI: 10.1002/er



For the proposed cooling system, the temperature variation at the compressor intake with

both the extraction pressure and the amount of mass bled from the compressor is shown in

Figure 3. For a fixed extraction pressure (P6) increasing mass extraction rate (a) increases the

Table I. Range of parameters for the present analysis.

Parameter Range

Ambient airMaximum ambient air temperature (T 0) 318.15 KRelative humidity (j0) 43.4%Volumetric air flow rate ð ’V aÞ 1 m3 s1

Net calorific value (NCV) 42 500 kJ kg1

Gas turbinePressure ratio (P2/P1) 10Turbine inlet temperature (T 3) 1373.15 KSpecific heat ratio of gas (g) 1.333 kJ kg1 K1

Specific heat of gas ðc pgÞ 1.147kJ kg1

K1

Air compressorExtraction pressure ratio (Px/P1) 2 ! 9Extracted mass ratio (a) 0.1 ! 0.5Specific heat ratio of air (g) 1.4 kJ kg1 K1

Heat exchangerDT 10 K

1 2 3 4 5 6 7 8 9 10

260

270

280

290

300

310

320

x = P 6 / P

1

T 1 , K

α=0.4

α=0.1

α=0.2

α=0.3

288.15

(15 °C)

Figure 3. Air intake temperature variations with extraction pressure ratio and extraction mass flow rate.



DOI: 10.1002/er



chilling effect and T 1 decreases. From the figure, it is clear that it is possible to adjust the values

of a and the extraction pressure to operate the GT at the ISO standard (288.15 K). As the level

of the extraction pressure P6 approaches the compressor exit pressure P2 at constant a

(assuming constant T 7), the temperature at the entrance of the mixing chamber T 8 decreases and

hence the intake temperature drops. Introducing the air refrigeration cycle provides theadvantage of quite low temperatures close to the 08C and even lower. This opportunity is not

possible with other methods for intake air cooling as evaporative cooling or use of waste heat-

driven absorption machines.

The variation of the PGR and TEC is shown in Figure 4 for extraction ratios a up to 0.4 and

P6 from 2 to 9 bar. For constant extraction pressure, it is seen that the power gain increases with

the extracted mass rate a, which means enhanced chilling effect due to the large mass passing

through the air refrigeration cycle. Though the power is boosted, the thermal efficiency

decreases due to more consumption of fuel to substitute for the low intake air. The drop in the

TEC is quite large for high a and high extraction pressure P6. For example, at a ¼ 0:4 and

P6 ¼ 7 bar the power is boosted by 17.98% but the thermal efficiency decreases by 10.76%. This

result indicates that the selection of the operation conditions depends on the utility choice,

boosting on the power on penalty of the thermal efficiency. The results show that if moderatevalues are selected, such as a ¼ 0:2 and extraction pressure of 4 bar, the power increases by

9.11%, while the thermal efficiency drops by only 1.34%. In other words, passing 20% of the

intake air at 4 bar through a Brayton refrigeration cycle reduces the intake temperature to

300.5% as seen in Figure 3 and increases the net power by 9.11% with only reduction in the

thermal efficiency of 1.34%.

It has been established that mechanical air cooling at the compressor intake increases the

power and decreases the thermal efficiency. For the proposed integration of the Brayton

1 2 3 4 5 6 7 8 9 10

0

4

8

12

16

20

24

-20

-16

-12

-8

-4

0

4

x = P 6 /P

1

P G R %

α = 0.4

T E C %

TEC

PGR

α = 0.3α = 0.2α = 0.1

Figure 4. Pressure gain ratio and thermal efficiency change factors for a gas turbinecooled by air Brayton refrigerator.



DOI: 10.1002/er



refrigeration, it is possible to boost the power and slightly improve the thermal efficiency, as seen

in Figure 4. At 2 bar, extraction pressure the TEC is positive for all values of a, which indicates

improvements in the thermal efficiency or better utilization of the fuel. However, the

improvement is small but demonstrates an enhancement in efficiency as compared to all the

current-known mechanical air-cooling approaches. To further elaborate on this point, Figure 5shows the variation of PGR and TEC up to 4 bar extraction pressure with a from 0 to 0.5 (step

Figure 5. Gas turbine power and efficiency improvement at low Brayton refrigeration cycle pressure.

0 0.1 0.2 0.3 0.4 0.5

0.3

0.305

0.31

0.315

0.32

0.325

0.33

α

η t

h

x = 2

x = 3

x = 4

no cooling

Figure 6. Thermal efficiency variations with a.



DOI: 10.1002/er



0.05). It is clear that there is no power gain nor efficiency change at a ¼ 0; i.e. there is no

cooling. At P6 ¼ 2 bar with cooling, the slight increase of thermal efficiency is shown in

Figure 6.

Figure 4 shows that for fixed value of the extracted mass rate, the power gain increases with

the extraction pressure to reach a maximum then decreases. Before explaining this effect, it is tobe noted that the inlet temperature at the turbine is fixed by the control system and the air

temperature at the expander inlet T 7 is constant. The later temperature is set by the desalination

plant requirement or the effectiveness of the heat exchanger. Increasing the extraction pressure

P6 means that state 7 is pushed to the left on the T–s diagram, causing low temperature T 8 after

expander and hence low intake air temperature T 1. This trend tends to increase the PGR. On the

other hand, increasing the extraction pressure increases the term ð ’W c ’W expÞ in Equation (24)

that tend to reduce the PGR. Therefore, the ascending–descending pattern shown in Figure 4

for fixed value of a is expected.

5. CONCLUSIONS

A new method for improving the performance of gas turbine units and eliminate the warm

weather power degradation is proposed. In this method, fraction of the intake air is extracted

from the main compressor and used as the working fluid for a reverse Brayton cycle. The gas

turbine inlet temperature is reduced by mixing the chilled air from the Brayton refrigeration

cycle and the main intake air streams. The inlet temperature depends on the extracted mass rate

and the extraction pressure. Mass and energy analysis of the coupled Brayton–reverse Brayton

cycles showed that the intake air temperature could be reduced to the ISO standard (158C) and

the gas turbine performance can be improved to attain power increase up to 19.58% of the site

value.

The performance improvement of a GT irreversible cycle of 10 pressure ratio operating in hot

weather of 458C and 43.4% relative humidity is investigated for extraction pressures from 2 to

9 bar and extraction mass ratio from 0.1 to 0.5. The results showed that the power augmentation

due to low intake air temperature is associated with increase in fuel consumption rate. The

proposed integration of the Brayton refrigerator showed that both the power and thermal

efficiency can be improved, which is an advantage over all the present mechanical intake air-

cooling methods.

NOMENCLATURE

c p = specific heat at constant pressure (kJ kg1 K1)

f = fuel to air ratio

h = specific enthalpy (kJ kg1)

’m = mass flow rate (kg s1)NCV = net calorific value (kJ kg1)

P = pressure (kPa)

PGR = power gain ratio, Equation (30)

PGRu = useful power gain ratio, Equation (28)’Q = heat rate (kW)



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r = pressure ratio ¼ P2=P1

T = temperature (K)

TEC = thermal efficiency change factor, Equation (31)

TECu = useful thermal efficiency change factor, Equation (29)’

W = output power (kW)x = extraction pressure ratio, P6/P1

Greek symbols

a = fraction of air mass flowing through the cooler cycle

g = specific heats ratio

Z = efficiency

j = relative humidity

Subscripts

0 = ambienta = dry air

c = compressor

cy = cycle

f = fuel

t = turbine

u = useful

v = vapour

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