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INTERNATIONAL JOURNAL OF ENERGY RESEARCH Int. J. Energy Res. 2005; 29:000–000 Published online in Wiley InterScience (www.interscience.wiley.com). DOI: 10.1002/er.1184

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Performance enhancement of gas turbines by inlet air-cooling in hot and humid climates

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Majed M. Alhazmy1,z, Rahim K. Jassim2,n,y and Galal M. Zaki1,} 1

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Faculty of Engineering, King Abdulaziz University, P.O. Box 80204, Jeddah 21 587, Saudi Arabia 2 YIC, Royal Commission for Yanbu, P.O. Box 30436, Yanbu El Sinaiyah, Saudi Arabia

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KEY WORDS:

gas turbine; air-cooling; power enhancement; spray cooler; vapour compression refrigeration

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1. INTRODUCTION

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The electric power generation sectors in many countries face two real problems, the continuous increase in fuel prices and the incessant growth in energy demand. Utilities focus considerable

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In this paper, a model to study the effect of inlet air-cooling on gas turbines power and efficiency is developed for two different cooling techniques, direct mechanical refrigeration and an evaporative water spray cooler. Energy analysis is used to present the performance improvement in terms of power gain ratio and thermal efficiency change factors. Relationships are derived for an open gas turbine cycle with irreversible compression and expansion processes coupled to air-cooling systems. The obtained results show that the power and efficiency improvements are functions of the ambient conditions and the gas turbine pressure ratio. The performance improvement is calculated for, ambient temperatures from 30 to 508C, the whole range of humidity ratio (10–100%) and pressure ratio from 8 to 12. For direct mechanical refrigeration air-cooling, the power improvement is associated with appreciable drop in the thermal efficiency. The maximum power gain can be obtained if the air temperature is reduced to its lowest limit that is the refrigerant evaporation temperature plus the evaporator design temperature difference. Water spray cooling process is sensitive to the ambient relative humidity and is suitable for dry air conditions. The power gain and efficiency enhancement are limited by the wet bulb temperature. The performance of spray evaporative cooler is presented in a dimensionless working graph. The daily performance of the cooling methods is examined for an ABB-11D5 gas turbine operating under the hot humid conditions of Jeddah, Saudi Arabia. The results indicate that the direct mechanical refrigeration increased the daily power output by 6.77% versus 2.57% for the spray air-cooling. Copyright # 2005 John Wiley & Sons, Ltd.

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SUMMARY

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Correspondence to: Rahim K. Jassim, YIC, Royal Commission for Yanbu, P.O. Box 30436, Yanbu El Sinaiyah, Saudi Arabia. E-mail: [email protected] z E-mail: [email protected] } E-mail: [email protected] y

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Copyright # 2005 John Wiley & Sons, Ltd.

Received 13 April 2005 Revised 24 September 2005 Accepted 19 October 2005

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attention on methods and techniques to improve the performance of their generation units. For gas turbine (GT) power plants, cooling the intake air is one of the proven technologies. In hot climate countries, as summer temperature rises, GT performance decreases. At the same time and because of the reliance on space air conditioning, the demand for power increases. Cooling the air at the compressor intake increases the air density and help boosting the power output. In addition, the ambient air humidity ratio plays an important role on selection of the cooling process. Several techniques are in use such as evaporative cooling, mechanical, absorption chillers and/or ice thermal storage (Dincer and Rosen, 2002). Each of these methods has its advantages and limits. In this paper the limits of air-cooling by two different methods, direct mechanical refrigeration and direct evaporative cooling are investigated. The components of a simple GT cycle are the compressor, combustion chamber and the turbine-generator set. The effect of the inlet air temperature on the cycle is well known and can be found in thermodynamics and gas turbine literature (Chengel and Bolos, 2003; Saravanamutto et al., 2001). The air temperature and humidity ratio at the inlet of the compressor affect the net efficiency and the plant heat rate. Several techniques are being used for inlet air-cooling, they can be broadly classified into direct and indirect methods. The direct cooling is achieved by the use of direct water sprays, foggier, or traditional evaporative coolers at the compressor intake. These systems are set to lower the temperature to a degree close to the ambient wet bulb temperature that may limit the applicability at coastal areas of humid weather. Johnson (1988) discussed the use of evaporative cooling technique for GT installations. Calculation procedure, installation as well as operation details were presented. Ameri et al. (2004) applied a fog type air cooling system where fog nozzles inject water at high pressures (above 70 bars) generating micro-fine droplets with sizes between 10 and 40 mm. Performance test results showed that the power output of the units has increased by 13% of the generated power, while the efficiency improvement was less than 1%. Since operation of the nozzles is critical to the fogging system, Meher et al. (2002) investigated the effect of nozzles type and droplets size on the performance of GT engines. Typically, the nozzles are stainless steel-316 of diameters less than 0.18 mm diameter. Because of the limited size of droplets produced by these nozzles (greater than 4 mm) humid air cannot exceed 90–95% of the saturation relative humidity (Bettocchi et al., 1995; Cyrus and Mee, 1999). Retrofitting of direct evaporative air coolers requires usually large ducts as the evaporation process requires low velocities. If the air velocity is high, water carry over may affect the compressor blades. For these reasons, use of direct evaporative cooling is limited and works better in dry air locations. Direct mechanical refrigeration can reduce the air temperature by circulating refrigerant fluid directly in a heat exchange installed in the inlet air ducting. This heat exchanger thus becomes the evaporator of the refrigeration cycle. In the indirect air chilling process, a refrigeration machine is used to lower the air inlet temperature through a secondary fluid (water or glycol). For these methods, a surface heat exchanger is mounted in the air intake manifold. The indirect method can achieve low temperatures regardless of the ambient relative humidity. The systems most readily utilized are the mechanical chilling, where an electric driven refrigeration machine or an absorption machine driven by waste heat. Ondrays et al. (1991) analysed the economics of employing mechanical, absorption chiller or thermal energy storage to boost the power in cogeneration plants. Kakaras et al. (2004) presented a simulation model for integration of a waste heat driven absorption machine, capable of reducing the air intake temperature to 58C. Generally, air cooling systems limit the inlet air temperature to about 3–48C. At lower temperatures, with the air at or near

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


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Int. J. Energy Res. 2005; 29:000–000

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PERFORMANCE ENHANCEMENT OF GAS TURBINES BY INLET AIR-COOLING

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saturation, a risk of ice formation exists either as ice crystals in the air or forming on surfaces, such as on the bellmouth or inlet guide vanes (Stewart and Patrick, 2000). Application for a simple GT cycle showed an increase in the power and decrease in plant thermal efficiency. Mercer (2002) investigated the use of different cooling techniques, reporting that evaporative cooling has increased the GT power by 10–15%, while chillers with energy storage systems have improved the power by 25% during peak periods. Elliot (2001) reported that for a GE LM 6000 turbine type, an increase of 1% in the power output could be achieved for every 1.68C drop in the air inlet temperature using water chiller. In the present study, two types of air coolers are considered; a direct mechanical refrigeration cycle and a direct spray cooling system. Both systems are analysed and the study focuses on ascertain the limits of each cooling process. The objective is to boost the power output and enhance the thermal efficiency of gas turbines operating for long periods in a hot and humid climate. The performance of the two systems is compared for different operation modes under actual climatic conditions, where the relative humidity and air temperature are time dependent.

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Condenser

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. i1, ma,To . Po, mv,o, o

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Figure 1. (a) A simple open type gas turbine with a direct mechanical refrigeration air-cooling unit; and (b) a direct evaporative cooler. Copyright # 2005 John Wiley & Sons, Ltd.

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Combustion Chamber

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. Wel, ref

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In this study, a simple open type gas turbine cycle is considered, Figure 1. The cycle performance can be improved by cooling the compressor intake air. Two types of air-cooling systems are considered, a direct mechanical refrigeration or a direct evaporative cooling. For the first, a refrigeration machine is used and the evaporator coil is placed in the intake ducts as seen in Figure 1(a). The direct cooling is achieved by evaporation of water spray in an evaporative cooler installed ahead of the compressor inlet manifolds as seen in Figure 1(b). The cold air

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2. ANALYSIS

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enters the compressor at state 1. To investigate the effect of the different parameters each of the systems is analysed separately.

3 2.1. Gas turbine cycle analysis

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’ el;cs is the electrical power consumed by the cooling system and can be either the where W ’ el;ref ) or the pumping power to circulate required power to drive the refrigeration compressor (W ’ el;pump ). the water inside the spray cooler (W Applying the first law of thermodynamics to the gas turbine (neglect the potential and kinetic energy terms), the power produced by the turbine is ’t ¼ m ð3Þ W ’ t cpg Zt ðT3 T4s Þ where m ’ t is the total gases mass flow rate at the turbine inlet given as m ’aþm ’vþm ’f ¼m ’ a ð1 þ o1 þ f Þ ’t ¼m

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ð4Þ

and o1 is the humidity ratio at state 1, Figure 1 and the fuel air ratio f ¼ m ’ f =m ’ a: Substituting for T4s and m ’ t from Equations (1) and (4), respectively, in Equation (2) yields 1 ’t ¼ m W ð1 þ o þ f Þc Z T 1 ð5Þ ’a 1 pg t 3 PRðk1Þ=k

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2.1.1. Turbine. If a cooling system (refrigeration machine) extracts its power from the turbine output as shown in Figure 1, the thermal efficiency of the cycle is ’ comp;air þ W ’ el;cs Þ ’ net W ’ t ðW W Zcy ¼ ¼ ð2Þ ’h ’h Q Q

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where PR is the pressure ratio and k is the specific heats ratio.

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Consider an irreversible gas turbine 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. For the isentropic processes 1–2s and 3–4s, we have ðk1Þ=k T2s T3 P2 ¼ ¼ ¼ PRðk1Þ=k ð1Þ T1 T4s P1

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Figure 2. T-s diagram of an open type gas turbine cycle. Copyright # 2005 John Wiley & Sons, Ltd.

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ER :1184 PERFORMANCE ENHANCEMENT OF GAS TURBINES BY INLET AIR-COOLING

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The turbine isentropic efficiency can be estimated using the practical relations recommended by Korakianities and Wilson (1994) as PR 1 Zt ¼ 1 0:03 þ ð6Þ 180 The gas specific heat (cpg) is evaluated as (Alhazmy and Najjar, 2004) cpg ¼ 1:0887572 103 1:4158834 101 T þ 1:9160159 103 T 2

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1:2400934 106 T 3 þ 3:0669459 1010 T 4 2:6117109 1014 T 5

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2.1.2. Air compressor. For humid air, the compression power can be estimated from ’ comp;air ¼ m W ’ a cpa ðT2 T1 Þ þ m ’ v ðig2 ig1 Þ

where ig2 and ig1 are the enthalpies of saturated water vapour at the compressor exit and inlet states, respectively, m ’ v is the mass of water vapour ¼ m ’ a o1 : Relating the compressor isentropic efficiency to the changes in temperature of the dry air and assuming that the compression of water vapour behaves as an ideal gas then T2s T1 Zc ¼ ð9Þ T2 T1 from which T2 is expressed in terms of T1 and the pressure ratio PR as " # PRðk1Þ=k 1 þ1 T2 ¼ T1 Zc Substituting for T2 into Equation (8) gives the actual compressor power as T1 ðk1Þ=k ’ Wcomp;air ¼ m 1Þ þ o1 ðig2 ig1 Þ ’ a cpa ðPR Zc where Zc can be evaluated using the following empirical relation (Korakianities, 1994): PR 1 Zc ¼ 1 0:04 þ 150


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where ra is the moist air density which is a function of the temperature T1 and the humidity ratio o1 and can be calculated using the engineering equation solver (EES) software (Klein and Alvarado, 2004). The effect of the air pressure drop across cooling coils is small and can be neglected, hence P1 ffi P0 : The air density will vary significantly with humidity ratio change oo ! o1 and decrease in the air temperature To ! T1 :

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The gas turbine is almost constant volume machine at a specific rotating speed, and then the inlet air volumetric flow rate, V’ a is fixed regardless of the ambient air conditions. As the air temperature rises in hot summer days, its density falls but the volumetric flow rate remains constant. Therefore, the mass flow rate reduces and consequently the power output decreases (Ameri et al., 2004). Equation (5) can be written in terms of the volumetric flow rate at the compressor inlet state as 1 ’ t ¼ V’ a ra ð1 þ o1 þ f Þcpg Zt T3 1 W ð7Þ PRðk1Þ=k

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Substituting for T2 in terms of T1 from Equation (10) gives the cycle heat rate as " ! # T3 PRðk1Þ=k 1 o1 ’ Qh ¼ m cpa þ1 þ ðiv3 iv2 Þ ’ a T1 ð1 þ f Þ cpg Zc T1 T1 where f, as expressed in Alhazmy and Najjar (2004), is cpg ðT3 298Þ cpa ðT2 298Þ þ o1 ðiv3 iv2 Þ f ¼ NCV cpg ðT3 298Þ

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’ comp;air and ’ t; W It is seen that the three terms of the gas turbine efficiency in Equation (2) (W ’ h ) depend on the air temperature and relative humidity at the compressor inlet, whose values Q are affected by the type and performance of the cooling system. Analysis of a mechanical chilling arrangement as seen in Figure 1 follows.

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2.2.1. Refrigeration evaporator. A schematic of the evaporator-finned coil is shown in Figure 3(a). The temperature and humidity of the air at the outlet of the evaporator depends on the refrigerant evaporating temperature (Te) and the ambient air inlet states ðTo ; oo Þ: When the evaporator outer surface temperature falls below the dew point (corresponding to the partial pressure of the water vapour) the water vapour condensates and leaves the air stream. The refrigerant evaporator cooling capacity can be adjusted to control the air intake temperature. Steady-state heat balance on the evaporator coil gives the cooling load as ’ e;a ¼ m Q ð17Þ ’ a ðio i1 Þ m ’ w iw where m ’ w is the rate of water condensation

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In this study, it is suggested to place the evaporator coil of a refrigeration machine at the GT compressor intake. In this scheme, the air-cooling is achieved without the need to a secondary cooling loop with a thermal fluid. In addition, it is possible to cool the air to temperatures below the ISO standard and is somewhat simpler and less costly than other cooling systems.

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iw is the specific enthalpy of water, evaluated at the condensate water exit temperature. Equation (17) represents the total amount of heat transfer from the moist air. The last term on the RHS is usually small compared to the first and can be neglected, (McQuiston et al., 2000). ’ s and latent heat This cooling and dehumidifying process involves two terms the sensible Q ’ transfer Ql : The sensible heat rate is associated with the decrease in dry bulb temperature and Copyright # 2005 John Wiley & Sons, Ltd.

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2.1.3. Combustion chamber. Heat balance on the combustion chamber (see Figure 1) gives the heat rate supplied to the GT cycle as ’h ¼m Q ð13Þ ’ f NCV ¼ ðm ’aþm ’ f Þcpg T3 m ’ a cpa T2 þ m ’ v ðiv3 iv2 Þ

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where iw,v is the specific enthalpy of water vapour in the air, which can be estimated using the following relation (Dossat, 1997): iw;v ¼ 2501:3 þ 1:8723ðTo Tdp Þ ð21Þ Tdp is the dew point temperature of the water vapour corresponding to its partial pressure. The specific enthalpy of the air leaving the cooling coil at state 1, Figure 3(a) is calculated from i1 ¼ io CFðio is Þ

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CF is the contact factor of the evaporator coil, defined as the ratio between the actual air temperature drop to the maximum at which the air leaves at the evaporator surface temperature Copyright # 2005 John Wiley & Sons, Ltd.

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the latent is associated with the decrease in humidity ratio as seen in Figure 3(b). These quantities may be expressed as (McQuiston et al., 2000) ’s ¼m Q ð19Þ ’ a cp;a ðTo T1 Þ ¼ m ’ a ðil i1 Þ

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Figure 3. (a) Evaporator with continuous fins on an array tubes; and (b) cooling and dehumidifying process on the psychrometric chart.

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(ts), which is assumed to be equal to the refrigerant evaporating temperature te, and 100% RH (see Figure 3(b)). In addition is, is the specific enthalpy of the process and evaluated for saturated air at (ts). CF ¼

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All values are evaluated to the accuracy of the EES software (Klein and Alvarado, 2004).

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2.2.2. Refrigeration compressor. The refrigerant air chiller is composed of a compressor, a condenser, an expansion device and an evaporator as shown in Figure 1(a). Figure 4 shows the thermodynamics cycle. The theoretical energy utilized by the compressor is the amount of energy that must be imparted to the refrigerant vapour as it follows an isentropic process. In practice, more power is required due to mechanical transmission losses, inefficiency in the drive motor converting electrical to mechanical energy and the volumetric efficiency (Dossat, 1997; Cleland et al., 2000). For an actual compressor the electric power is

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’ ref ðib ia Þref m ’ ref ðib ia Þref ’ comp;ref ¼ m W ¼ Zm Zel Zvo Zeu

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where Zm ; Zel and Zvo are the compressor mechanical, electrical and volumetric efficiency, respectively. The compressor energy use efficiency, Zeu is normally determined by manufacturers and depends on the pressure ratio, the application and type of the compressor. Cleland et al. (2000) proposed an approximate method for calculating the electric power usage rate for all refrigerants as

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where a is an empirical constant that depends on the type of refrigerant and x is the quality at state d, Figure 4. The empirical constant is 0.77 for R22 which used as the working fluid in the refrigeration machine (Cleland et al., 2000). The empirical constant n depends on the number of compression and expansion stages. In this study n=1 for a simple refrigeration cycle with single stage compressor. ’ e;ref ; should be capable to remove the required cooling load from The evaporator capacity, Q ’ ’ ’ e;a : the moist air, Qe;a : Hence Qe;ref ¼ Q

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’ e;ref Q m ’ ref ðia id Þref ðTc Te Þ ¼ Te ð1 axÞn Zeu COPc ð1 axÞn Zeu

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Substituting Equation (17) into Equation (25) gives the electric power extracted from the turbine output to operate the chiller machine in terms of the airflow rate and states o and 1, Figure 1; ’ a ðio i1 ÞðTc Te Þ ’ el;ref ¼ m W Te ð1 axÞn Zeu

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In evaporative cooling water and intake air are brought into direct contact where the warm air stream transfers heat to sprayed water as seen in Figure 5(a). During the air–water heat exchange process, part of the liquid water evaporates causing the temperature of the air to decrease adiabatically, line o–1 in Figure 5(b). The air humidity ratio increases from oo to o1 approaching the saturation condition. As the air approaches the saturation limit, the evaporation process takes more time where the air cannot carry more water and further water injection is not utilized. Therefore, the direct evaporative cooling of air is limited by the temperature difference (T1To). In practice to cool the air to the saturation state, requires water over spraying that may initiate carry over of droplets causing fouling of compressor blades

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Figure 5. (a) Schematic of adiabatic spray cooler; and (b) adiabatic saturation process on the psychrometric chart. Copyright # 2005 John Wiley & Sons, Ltd.

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and/or rust of the entrance ducts. Controlling the parameters of evaporative coolers is an important key to the successful seasonal operation of coolers. The effectiveness of an evaporative air-cooler ðeev;c Þ is defined as the ratio between the actual dry bulb temperature decrease to the theoretical temperature difference if the air leaves the cooler at saturation state (wet bulb temperature). Typical evaporative cooler effectiveness range is 0.8–0.9 (Cortes and Willems, 2003). Figure 5(a) shows a schematic of a spray cooler where the ambient air at To, oo and Po enters the spray chamber and leaves at T1, o1 and P1. The spray cooler is assumed to operate in a steady adiabatic process such that the ambient moist air enters at To and RHo and leaves at state 1. Adequate quantity of water is added to the air stream to raise its moisture content close to that corresponding to 100% relative humidity and decrease its temperature as seen in Figure 5(b). Applying energy balance yields (McQuiston et al., 2000) oo ðivo iw Þ ¼ cpa ðT1 To Þ þ o1 ðifg1 Þ

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where ivo and iw are saturated water vapour enthalpy at To and saturated water liquid enthalpy at T1. ifg1 is the latent heat of vapourization at state 1. oo is evaluated at the ambient conditions using EES software (Klein and Avarado, 2004).

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3. GAS TURBINE COUPLED TO COOLING SYSTEMS

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In order to evaluate the feasibility of a cooling system coupled to a GT plant, the performance of the plant is examined with and without the cooling system. In general the net power output of a complete system is ’ net ¼ W ’ t ðW ’ comp;air þ W ’ el;cs Þ W ð28Þ

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The PGR is a generic term that takes into account all the parameters of the GT and the associated cooling system irrespective of the cooling process. For a stand-alone GT under specific climatic conditions PGR=0. If a cooling system is used, the PGR increases with the reduction of the intake temperature but this increase is restricted by the physical limits of the cooling process. However, the PGR gives the percentage enhancement in power generation, and the thermal efficiency of a coupled system is an important parameter to describe the input output relation. Let us define another factor that physically relates the thermal efficiency of a stand-alone GT to that coupled to a cooling system as the thermal efficiency change (TEC) Zcy;with cooling Zcy;without cooling TEC ¼ 100% ð30Þ Zcy;without cooling

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The three terms of Equation (28) are functions of the air properties at the compressor intake (T1 and o1 ), which in turn depends on the performance of the cooling system. Let us define a dimensionless term that gives the advantage of using any cooling system as the power gain ratio (PGR) ’ net;with cooling W ’ net;without cooling W PGR ¼ 100% ð29Þ ’ net;without cooling W

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While, the PGR is always positive TEC can be negative, which means that the efficiency of the coupled system is less than that of a stand-alone GT even at low intake temperatures. Both PGR and TEC provide dimensionless parameters that can be easily employed and interpreted. Copyright # 2005 John Wiley & Sons, Ltd.

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Equation (31) for an ideal air reversible cycle, where Zc ¼ Zt ¼ 1; fuel air ratio f=0, cpg ¼ cpa and the inlet air humidity ratio o1 ¼ 0 gives the standard expression for an open Brayton cycle, where the efficiency is only pressure ratio dependent Zcy;rev ¼ 1 1=PRðk1Þ=k : For a stand-alone (without cooling) GT the third term of the nominator vanishes and the inlet conditions T1 and o1 are replaced by To and oo :

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3.2. Gas turbine with direct evaporative water spray cooler

’ el;cs is the pumping power to For spray cooling the power consumed by the cooling system, W circulate the water inside the chamber (see Figure 5(a)). This power is small compared to the other terms in Equation (28) and can be ignored. Therefore, the PGR can be calculated from Equations (5), (11), (28) and (29). The coupled system thermal efficiency is derived in a similar way as Equation (31) to give T3 1 1 o1 ðk1Þ=k Zcy;spray ¼ ð1 þ o1 þ f Þcpg Zt 1 1Þ þ ðig2 ig1 Þ cpa ðPR Zc T1 T1 PRðk1Þ=k ," ! # T3 PRðk1Þ=k 1 o1 ð1 þ f Þcpg cpa þ1 þ ðiv3 iv2 Þ ð32Þ Zc T1 T1

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Notably, the nominator presents the net power for a GT cycle with humid air as the working fluid, the inlet conditions for the spray evaporative cooling would be entirely different from those for mechanical refrigeration system.

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The power gain ratio for a gas turbine with air cooler is obtained by substituting Equations (5), (11), (26) and (28) into Equation (29). From Equations (2), (5), (11) and (28) the cycle thermal efficiency with a cooling system, Zcy;cs in terms of the air properties at the compressor intake, the fuel air ratio and the refrigeration machine characteristics (the third term of the nominator Equation (31)) is T3 1 1 o1 ðk1Þ=k Zcy;ref ¼ ð1 þ o1 þ f Þcpg Zt 1 1Þ þ ðig2 ig1 Þ cpa ðPR Zc T1 T1 PRðk1Þ=k ! ," ðio i1 ÞðTc Te Þ T3 PRðk1Þ=k 1 cpa þ1 ð1 þ f Þcpg T1 ðTe þ 273:15Þð1 axÞn Zeu Zc T1 o1 þ ðiv3 iv2 Þ ð31Þ T1

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4. RESULTS AND DISCUSSION

In order to investigate the performance of the direct spray and refrigeration machine air-cooling methods on the GT power and efficiency, a computer program has been developed to calculate the PGR and TEC for different operation conditions. The thermophysical properties were determined to the accuracy of the EES software. In particular, the specific heats of air and


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combustion gases are both temperature dependent. Table I shows the range of the different parameters. The selected maximum air temperature to,max is based on meteorological data recorded during the past few years in Saudi Arabia, which was close to 508C. An average value of 44.5% for the relative humidity is selected as base data (daily average in coastal areas). Even though, the performance of cooling systems under actual weather conditions is considered in application section. Applying Equations (29) and (30) for a refrigeration cooling system, the variation of the PGR and TEC with the ratio xT ¼ t1 =to is obtained and presented in Figure 6. The factor xT presents physically the ratio between the air temperatures at the cooler exit and the ambient temperature in Celsius (cooling process attainment). For a constant pressure ratio, the PGR increases with decreasing xT and a power gain of 16.3 % can be obtained by cooling the air from 50 to 158C for PR of 12. The TEC decreases with decreasing of xT which is attributed to the power consumed by the refrigeration machine and the increase in fuel air ratio required to overcome the drop in the intake air temperature. For the significant increase in power of 16.3% the drop in efficiency reaches 3%. The theoretical limit for power enhancement, using the chilling refrigeration method is controlled by the lowest value of xT : This limit depends on the ambient air temperature and relative humidity and can be as low as the refrigerant evaporating temperature plus the evaporator design temperature difference (TDe). For the range of parameters shown in Figure 6, there is gain in power but notable drop in the overall efficiency. Therefore, the feasibility of cooling the intake air to boost the power cannot be assessed only on basis of power gain but other aspects as cost estimates and practical size of the cooling coils are important factors. Figure 6 shows also that within the considered values (PR=8 ! 12), cooling the air for

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25 Parameter

Ambient air Max. ambient air temperature, To,max Relative humidity, RHo Volumetric air flow rate Net calorific value, NCV

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Table I. Range of parameters for the present analysis.

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Refrigeration machine Refrigerant Evaporating temperature, te Superheat Condensing temperature, tc Condenser design temperature difference TDc Evaporator design temperature difference TDe Subcooling Energy use efficiency, Zeu


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Range 323.15 K 34 ! 100% 1 m3 s1 42 500 kJ kg1 8 ! 12 1373.15 K 0.91 ! 0.93 0.88 ! 0.91 R22 08C 58K to+TDc 108C 68C 3K 0.75 Int. J. Energy Res. 2005; 29:000–000


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Figure 6. Variation of the power gain ratio and thermal efficiency change with the temperature ratio for a gas turbine with mechanical refrigeration air-cooling.

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a gas turbine of PR=8 from 50 to 408C boosts the power by 3.85% and decreases the thermal efficiency by 1.037%. For the same conditions the improvement at PR=12 is 4.23% and the thermal efficiency reduction is 0.882%. This result simply indicates that the effect of the air compressor pressure ratio on the power gain and thermal efficiency change are significant and should be considered in the design and selection of the air cooling systems. A spray air cooler is designed to lower the air temperature to a degree close to the ambient wet bulb temperature that limits its cooling capability in humid climate areas. Applying Equations ’ el;cs ¼ 0 for spray cooler, PGR and TEC are computed for (29)–(31) with the assumption that W fixed ambient conditions as seen in Figure 7. For 508C and 44.5% relative humidity, the air temperature can be theoretically reduced in an adiabatic process to 37.128C (wet bulb temperature) for which xT ¼ 0:76: From Figure 7 at xT ¼ 0:76 and PR of 10, the power gain reaches 4.8 % and the thermal efficiency change is 0.23%. However, the value is small but substantiates an improvement in efficiency as compared to that of the mechanical refrigeration air cooler system (PGR=5.3% and TEC=1.16%). This result shows that for spray cooling there is gain in power and improvement in thermal efficiency. This advantage is soon offset by the fact that the system cannot provide any further improvement beyond the limiting state at specified ambient conditions. Figure 8 shows the PGR and TEC for different air temperatures (50, 40 and 308C) and the whole range of RH up to 100%. The curves on the figure present the variation of the RH with xT at the maximum gain, which equals twb =to : Therefore, these curves determine the limiting condition for spray cooling once the ambient temperature and relative humidity are prescribed. For example to determine the maximum power gain ratio and thermal efficiency change for ambient conditions of 508C and 60% RHo; first draw a horizontal line from 60% relative humidity until it intersects the cooling limit of to=508C point A on the figure. At this point xT determines the lowest air temperature at the compressor inlet, intersection with the lines of PGR and TEC (at to=508C) gives the maximum values of the PGR and TEC as 3.5 and 0.18%,

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Figure 7. Variation of the power gain ratio, thermal efficiency change and the temperature ratio for a gas turbine with direct inlet air spray cooling.

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Figure 8. The cooling limit of air spray cooler.

respectively. Further, the results in Figure 8 are presented for an ideal spray process effectiveness eev;c ¼ 100%; where (Figure 5(b))

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To T10 To T1

ð33Þ

To include the effectiveness in the analysis, just replace T1 by T10 ¼ To eev;c ðTo T1 Þ in Equations (27) and (31). Copyright # 2005 John Wiley & Sons, Ltd.

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The relative humidity at the air intake is a key parameter for evaporative cooling systems. For illustration, let us consider air at constant dry bulb temperature of 408C and 60% RHo, for which the maximum PGR and efficiency change are 3.075 and 0.154%, respectively. For this case xT ¼ 0:8125; which means that the air temperature can only be reduced from 40 to 32.58C. For an intermediate level of humidity, 40%, and the same dry bulb temperature the corresponding temperature ratio xT is 0.695, for which the temperature limit is 27.88C. The corresponding power gain is 5.114%. The detailed results for the effect of the RH on the spray cooling capacity are summarized in Table II, which shows a direct inverse proportionality. To illustrate the difference in capability between the direct spray cooling and the mechanical refrigeration cooling method, let us operate a mechanical refrigeration system at its limit, where the cooling process reaches the 100% RH curve and the 158C set by the ISO standard. The results (Table II) show that for the same initial conditions, the power gain is much higher for the refrigeration cooling than that for spray cooling. The power gain reaches 11.36% for dry air versus 7.7% for spray cooling process, under the same initial conditions. The refrigeration process is actually less affected by the intake air RH but is associated with an appreciable drop in efficiency. Furthermore, Table II illustrates that the required refrigeration machine capacity depends not only on the ambient temperature but also on the relative humidity. For instance, cooling air at a temperature of 408C and RH of 60% to 158C and saturation condition needs 52% more power than that required for air at 40% RH. Therefore, selection of the unit capacity is a concern for design engineers; the problem arises from the dependency of the capacity on the rational climatic parameters. As a guide for selection of machines’ capacity the cooling load per kg of dry air for different ambient conditions is calculated and presented in Figure 9 for an outlet at temperature of 158C.

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Figure 9. Dependence of the cooling load of a direct mechanical refrigeration on ambient conditions (t1=158C). Copyright # 2005 John Wiley & Sons, Ltd.

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t18C

wref (kJ kg1) 32.23 21.21 11.89

TEC% 1.844 1.04 0.359

26.33 30.95

t18C 0.01594 0.02759

oo (kg kg1 d.a) 0.0118 0.02569

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The performance of GT plants deteriorate with the high air temperature in summer days. The problem is more severe in locations where the ambient temperature is far from the ISO standard, where the output and efficiency degrade during all working hours. It is, therefore, of interest to investigate the daily performance of air cooling systems under actual operation conditions. This application is important for countries like Saudi Arabia, where 274 different GT units ranging from 10 to 63 MW capacities generate 46% of the total nations 30 GW power. The weather of the city of Jeddah, Saudi Arabia (228N latitude, 398-E longitude and 23 m altitude) is selected as a typical hot humid weather. The extreme daily air temperature and humidity variation is the 16th August, which is selected according to the statistical methodologies and presented in Figure 10. The temperature varies between 33 and 418C, the relative humidity reaches 100% during early morning hours, and it is above 60% for 10 h daily. These are inadequate conditions for gas turbine plants. For the present illustrative application, the ABB-Type 11D5 GT (ASEA Borwn Boveri Co.) is considered with the following characteristics:

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The power output of the unit at 358C and 60% RH drops to 57.1 MW, and to 51.5 MW at 508C (19% loss in power) and the same humidity. The unit output drops to 49.9 MW at 100% RH and 508C. The site MW is 45.9 MW. In order to compare the two cooling systems let us assume a fixed cooling effect, for which the energy removed from the intake air is the same and the coolers operate for 24 h. Applying

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45 Figure 10. 16 August RH and ambient temperature variations. Copyright # 2005 John Wiley & Sons, Ltd.

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Figure 11. (a) Dependence of gas turbine maximum PGR on RH of direct spray cooler; (b) dependence of gas turbine PGR on RH of direct mechanical refrigeration for the same temp. drop of direct spray cooler; (c) air coolers limits; and (d) dependence of gas turbine maximum PGR on RH of direct mechanical refrigeration. Copyright # 2005 John Wiley & Sons, Ltd.

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Equations (29) and (30), on hourly basis, using the ambient data of Figure 10 and assuming that both systems reduce the intake air temperature from To to T1 (Figure 11(c)) but in different processes. The variation of the PGR and TEC for the spray and refrigerant cooling are presented in Figures 11(a) and (b), respectively. For the refrigeration chiller, Figure 11(b) shows that during early morning hours with the 100% RH the power gain ratio PGR and thermal efficiency change TEC drop rapidly to reach zeros at 6 AM where xT approaches 1. At 11 AM with 40.18C air temperature and RH of 34.1%, the gain reaches 6.166% for the direct mechanical refrigeration and 5.82% for direct evaporative air cooling. The difference in the PGR for the same operating conditions and temperature drop is due to the difference in humidity ratio at the inlet of the compressor as seen in Table III. At relatively high (58.5–82%) RHo by nighttime 18–21 O’clock, the gain drops for both cooling systems as seen in Figures 11(a) and 12(b) to be as low as 1.277 and 1.136%. For the spray cooling, the TEC follows the same pattern of the PGR variation, it is zero at 6 AM, where cooling the air is not possible close to the 100% humidity ratio. The maximum improvement is at 11 AM, 0.284%, where the humidity is the least. Figure 11(b) shows negative values of TEC for the mechanical refrigeration air cooling system with a maximum drop of 0.557 % at 11 AM. The power of the gas turbine can be further improved by reducing the air temperature for the refrigeration chilling system. The limit is set at the condition for which RH is 100% (state b in

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Figure 11(c)). For this operation mode, the daily variation of the PGR and TEC is shown in Figure 11(d). The power gain is zero for saturated air, but the maximum gain at 11 AM reaches 16.41%. The data presented for the ABB gas turbine operating under real climatic conditions indicates that the power output is far from the rated design power. The site output of the unit is even less than the power at the sever climate of 508C and 100% RH. Let us assume that the base load for comparison is the site output of 46 MW and the daily net power gain is the time integral of the hourly variation as Z 24 Egain ¼ PGRðLsite Þ dt MW h ð34Þ 0

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Performing the integration for both cooling systems shows that the maximum daily energy gain for a spray evaporative cooler is 28.36 MW h, while that of the refrigeration cooling operating at the same cooling capacity is 31 MW h. The daily energy gain for the refrigeration cooling with the 100% RH limit is 75 MW h. The values show that the cooling system may be promising when operated under favourable conditions but the overall daily performance is what counts at the end. The practical illustrative application indicates that the spray system provides additional energy of 2.57% versus 6.77% for the mechanical refrigeration. For the present illustrative application, the superior performance of the mechanical refrigeration is attributed to the high level of the ambient relative humidity.

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The performance augmentation in this study was characterized by two dimensionless parameters the PGR and the TEC term; the results are presented in general dimensionless form. For both systems, the performance improvement showed strong dependency on the climatic condition and to some degree on the gas turbine pressure ratio. For a direct mechanical refrigeration cooling, an improvement up to 11% in output power is obtained for cooling air from 50 to 258C, RH of 44.5% and PR=11. On the other hand, the thermal efficiency decreased by nearly 2.25% due to the power consumed for the refrigeration machine. For the mechanical refrigeration cooling process, the maximum limit of power boosting is reached if the air compressor inlet temperature is cooled to the lowest degree, which equals the refrigerant evaporating temperature plus the design temperature difference (6oC in the present study). At this limiting condition, power gain is in the order of 16.4%. The direct mechanical refrigeration system is somewhat simpler and less costly than other cooling systems, but there may be risk of refrigerant leakage into gas turbine. Special alarm systems have to be installed to detect refrigerant leakage into the compressor and to shut down and evacuate the refrigerant. Spray cooling is quite efficient for dry air, for 408C and 20% RH ambient air, the maximum power gain and efficiency improvements are 7.7 and 0.37%, respectively. The direct spray cooling process is limited by the wet bulb temperature at which additional water spray would not contribute any further cooling effects. The performance of the spray cooling is presented in a general dimensionless working graph that directly relates the maximum power gain and efficiency variation to the ambient conditions. In addition to improving the thermal efficiency, spray cooler improves the environmental impact of the GT, since increasing water vapour in the

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inlet air tends to lower the amount of nitrogen oxides emissions as well as reduces the dust due to air washing. The performance of both systems was examined for an ABB-11D5 gas turbine operating in the hot humid conditions of Jeddah, Saudi Arabia. The actual climate on 16 August is selected as base data for comparison. However, the refrigeration chilling consumes part of the power the system increased the daily power by 6.77% and the thermal efficiency drops by 1.27%, while the power and thermal efficiency improvement were only 2.57 and 0.126%, respectively, for the direct spray cooling.

9

Greek symbols

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a Z Zeu eev;c xT

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=an empirical constant that depends on the type of refrigerant, Equation (25) =efficiency =compressor energy use efficiency =evaporator effectiveness =temperature ratio t1 =to

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=specific heat at constant pressure (kJ kg1 K1) =energy (MW h) =specific enthalpy (kJ kg1) =specific heats ratio =load (MW) =net calorific value (kJ kg1) =pressure (kPa) =pressure ratio=P2/P1 =power gain ratio, Equation (30) =heat rate (kW) =mechanical =mass flow rate (kg s1) =temperature (8C) =absolute temperature (K) =thermal efficiency change, Equation (31) =power (kW) =quality

cp E i k L NCV P PR PGR ’h Q mech m ’ t T TEC ’ W x

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=electricity =fuel =latent heat =heat =latent =empirical constant depends on the number of compression and expansion stages =refrigerant =reversible =sensible, surface =turbine =water vapour =volume =water =quality

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REFERENCES

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Alhazmy MM, Najjar YSH. 2004. Augmentation of gas turbine performance using air coolers. Applied Thermal Engineering 24:415–429. Ameri M, Nabati H, Keshtgar A. 2004. Gas turbine power augmentation using fog inlet cooling system. Proceedings ESDA04 7th Biennial Conference Engineering Systems Design and Analysis, Manchester, U.K., Paper ESDA200458101. Bettocchi R, Spina PR, Moberti F. 1995. Gas turbine inlet air cooling using non-adiabatic saturation process. ASME Cogen-Turbo Power Conference, Paper 95-CTP-49:1–10. Chengel YA, Bolos MA. 2003. Thermodynamics: An Engineering Approach (5th edn). McGraw-Hill: New York. Cleland AJ, Cleland DJ, White SD. 2000. Cost-Effective Refrigeration, Short Course Notes, Institute of Technology and Engineering, Massey University, New Zealand. Cortes CPE, Willems D. 2003. Gas turbine inlet cooling techniques: an overview of current technology. Proceedings Power GEN 2003, Las Vegas, Nevada, 9–11 December. Cyrus B, Mee RT. 1999. Gas turbine power augmentation by fogging of inlet air. Proceedings of the 28th Turbo Machinery Symposium. Dincer I, Rosen M. 2002, Thermal Energy Storage: Systems and Application. Wiley: New York. Dossat RJ. 1997. Principles of Refrigeration. Wiley: New York. Elliot J. 2001. Chilled air takes weather out of equation. Diesel and Gas Turbine Worldwide, October, 49–96. Johnson RS. 1988. The theory and operation of evaporative coolers for industrial gas turbine installations. GT and Aero-Engine Conference, Amsterdam, 6–9 June, Paper#88-GT-41. Kakaras E, Doukelis A, Karellas S. 2004. Compressor intake air cooling in gas turbine plants. Energy 29:2347–2358. Klein KA, Alvarado FL. 2004. EES-Engineering Equation Solver, Middleton, WI. Korakianities T, Wilson DG. 1994. Models for predicting the performance of Brayton-cycle engines. Journal of Engineering for Gas Turbine and Power 116:381–388. McQuiston FC, Parker JD, Spilter JD. 2000. Heating, Ventilating and Air Conditioning: Design and Analysis (5th edn). Wiley: New York. Meher H, Cyrus B, Mee RT, Thomas R. 2002. Inlet fogging of gas turbine engines, part B: droplet sizing analysis nozzle types, measurement and testing. Proceedings of the ASME Turbo Exo 2002, Amsterdam, Netherlands, June, Paper No. 2002 GT-30563. Mercer M. 2002. One stop shop for inlet cooling systems. J Diesel and Gas Turbine-Worldwide, 10–13 June. Ondrays IS, Wilson DA, Kawamoto N, Haub GL. 1991. Options in gas turbine power augmentation using inlet air chilling. Journal of Engineering for Gas Turbine and Power 113:203–211. Saravanamutto H, Rogers G, Cohen H. 2001. Gas Turbine Theory (5th edn). Prentice-Hall: Englewood Cliffs, NJ. Stewart W, Patrick A. 2000. Air temperature depression and potential icing at the inlet of stationary combustion turbines. ASHRAE Transactions 106:part 2.

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Author Queries For

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Details required

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Chengel and Bolos, 2003 or Cengel and Bolos, 2003 - Pl. check spelling.

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Ondrays et al., 1991 or Ondryas et al., 1991 - Pl. check spelling.

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