er.1302] galal m. zaki; rahim k. jassim; majed m. alhazmy -- brayton refrigeration cycle for gas...
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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.
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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
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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
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DOI: 10.1002/er
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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.
BRAYTON REFRIGERATION CYCLE 1295
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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
G. M. ZAKI, R. K. JASSIM AND M. M. ALHAZMY1296
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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).
BRAYTON REFRIGERATION CYCLE 1297
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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
G. M. ZAKI, R. K. JASSIM AND M. M. ALHAZMY1298
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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
BRAYTON REFRIGERATION CYCLE 1299
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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.
G. M. ZAKI, R. K. JASSIM AND M. M. ALHAZMY1300
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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.
BRAYTON REFRIGERATION CYCLE 1301
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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.
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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.
BRAYTON REFRIGERATION CYCLE 1303
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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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DOI: 10.1002/er