21 Bassily - Absopt x Evaporative

Performance improvements of the intercooledreheat recuperated gas-turbine cycleusing absorption inlet-cooling and

evaporative after-cooling

A.M. Bassily*

Gharian University, P.O. Box 64735, Gharian, Libya

Received 19 December 2002; accepted 11 January 2003

Abstract

Inlet air-cooling improves both efficiency and power of gas-turbine cycles. An absorption

inlet-cooling system is introduced to the intercooled reheat recuperated gas-turbine cycle (Icycle). The exhaust gas of the cycle is used to run the system, which cools the inlet air to thelow-pressure compressor and high-pressure compressor using two stages of cooling in the

intercooler. Five different layouts of the I cycle are presented. Those layouts include the effectsof absorption inlet cooling, evaporative inlet cooling, evaporative aftercooling, and absorp-tion inlet cooling with evaporative aftercooling. A parametric study of the effect of pressureratio, ambient temperature, ambient relative-humidity, turbine’s inlet-temperature (TIT), and

the effectiveness of the recuperated heat-exchanger ("HE1) on the performance of all cycles iscarried out. The results indicate that using two stages of cooling in the intercooler could boostthe gain in efficiency, because of applying evaporative inlet cooling, by up to 1.55%. Applying

absorption inlet-cooling could increase the efficiency of the I cycle by up to 6.6% comparedwith 3.9% for applying evaporative inlet cooling. Applying absorption inlet-cooling withevaporative aftercooling could increase the optimum efficiency of the I cycle by 3.5% and its

maximum power by more than 50%. Increasing TIT increases the capacity of the recuperatedheat-exchanger and absorption cooling system and raises the gain in efficiency because ofincreasing "HE1.# 2003 Published by Elsevier Ltd.

Keywords: Absorption; Evaporative; Inlet-cooling; Aftercooling; Gas turbine; Optimum pressure-ratio;

Pressure ratio; Evaporative inlet-cooling; Air-inlet-cooling

Applied Energy 77 (2004) 249–272

www.elsevier.com/locate/apenergy

0306-2619/03/$ - see front matter # 2003 Published by Elsevier Ltd.

doi:10.1016/S0306-2619(03)00099-0

* Corresponding author at 31 Ahmed El Barrad Street, 11241 Cairo, Egypt. Fax. +218-41-633-268.

E-mail address: [email protected] (A.M. Bassily).

http://www.sciencedirect.com



http://www.elsevier.com/locate/apenergy/a4.3d

mailto:[email protected]

Nomenclature

cp specific heat at constant pressure (kJ/kg K)h enthalpy (kJ/kg)m:

mass-flow rate (kg/s)P pressure (Pascal)Q:

heat rate (kW)r pressure ratio (dimensionless)rc total pressure-ratio (dimensionless)RH relative humidity (dimensionless)T temperature (K)w humidity ratio (dimensionless)X concentration (dimensionless)

Subscriptsa airAI1 the first cooling stage of the air intercoolerAI2 the second cooling stage of the air intercoolerc compressorg gashi highLIBR lithium-bromide solutionlo lowo ambientr refrigerantru used refrigerants strong solutiont turbine inlet or turbinew weak solutionwa water

Greek letters" heat-exchanger effectiveness (dimensionless)

AbbreviationsAC air coolerAFC air aftercoolerAI air intercoolerCC1 combustion chamber 1CC2 combustion chamber 2HE1 the air–gas heat exchangerHE2 the absorption system heat-exchanger

250 A.M. Bassily / Applied Energy 77 (2004) 249–272

1. Introduction

Inlet air-cooling improves both power and efficiency of gas-turbine cycles. Eva-porative inlet cooling, absorption cooling, and mechanical compression cooling arethe most common methods for inlet air-cooling of gas-turbine cycles. Inlet air-cool-ing has been the subject of many investigations. Giourof [1] discussed the benefits ofinlet air-cooling. De Lucia et al. [2] reported that evaporative inlet-cooling is eco-nomical and simple, but suitable for only dry hot climates. De Lucia et al. [3] con-cluded that evaporative inlet-cooling could enhance power by 2–4% a yeardepending on the weather. Ait-Ali [4] optimized the power of a regenerative gas-turbine cycle with compressor inlet air refrigeration. Bassily [5,6] presented theeffects of the turbine’s inlet-temperature, ambient temperature, and relative humid-ity on the performance of the recuperated gas-turbine cycle with evaporative inlet-cooling and the intercooled reheat regenerative gas-turbine cycle with indirect eva-porative inlet cooling. He showed that evaporative inlet-cooling could boost theefficiency of the recuperated cycle by up to 3.2%. He also showed that indirect eva-porative inlet cooling could increase the efficiency of the intercooled reheat regen-erative cycle by up to 3%. Although the advantages of inlet air-cooling, there is apractical limiting factor on inlet-cooling as the inlet air temperature decreases below280 K. In such a case, ice crystals could form on the compressor blades and affectthe compressor’s performance as the air temperature drops as much as 6 K as airenters the compressor and its velocity increases [7]. One of the promising inlet-cooling methods is absorption cooling.

HPC high-pressure compressorHPT high-pressure turbineI cycle the intercooled reheat recuperated gas-turbine cycleIAI cycle the intercooled reheat recuperated gas-turbine cycle with

absorption inlet coolingIAIEA cycle the intercooled reheat recuperated gas-turbine cycle with

absorption inlet cooling and evaporative after coolingIEA cycle the intercooled reheat recuperated gas-turbine cycle with

evaporative aftercoolingIEI cycle the intercooled reheat recuperated gas-turbine cycle with

evaporative inlet coolingLPC low-pressure compressorLPT low-pressure turbinethe I cycles I, IAI, and IEI cyclesTIT turbine’s inlet-temperatureV valveWC water coolerWH water heater

A.M. Bassily / Applied Energy 77 (2004) 249–272 251

An absorption-cooling system uses a low-grade thermal energy source to drive thesystem and generate the cooling effect. Absorption inlet-cooling can employ some ofthe rejected heat of the intercooled reheat recuperated gas-turbine cycle to improveboth efficiency and power. Many investigators [7–9] showed that absorption inlet-cooling is a promising technology and that it is the most suitable and economical forhot and humid weather. Najjar [10] reported improving the simple cycle efficiency andpower by cooling the inlet air using an absorption system. Bartolini and Salvi [11]showed an 8% increase in power and 4% increase in thermal efficiency of steam-injec-ted gas-turbines when cooling the inlet air using absorption chillers. Bassily [12] intro-duced an absorption inlet-cooling system to the recuperated gas-turbine cycle. Thesystem was driven by the exhaust of the recuperated heat-exchanger. He reported thatsuch a system could boost the efficiency by up to 4% compared with 2.2% for eva-porative inlet cooling. The most common absorption cooling systems are the lithium-bromide and the ammonia systems. The ammonia absorption system requires a largespace and extensive safety precautions because ammonia is toxic [1]. The lithium-bro-mide systems can provide chilled water to cool the inlet air to the required temperaturefor optimum operation at 280 K. Despite the numerous works on inlet air-cooling, aliterature review has shown the following for the intercooled reheat recuperated gas-turbine cycle:

1. The effect of absorption inlet-cooling on the intercooled reheat recuperatedgas-turbine cycle has not been investigated.

2. The combined effect of absorption inlet-cooling and evaporative aftercoolinghas not been presented.

3. A parametric study of the effect of ambient temperature, ambient relativehumidity, and the effectiveness of the recuperated heat-exchanger on the effec-tiveness of absorption inlet cooling has not been carried out.

In this paper, a lithium-bromide absorption inlet-cooling system is introduced tothe intercooled reheat recuperated gas-turbine cycle, with and without evaporativeaftercooling. The system is used to cool the inlet air to the LPC and HPC using twostages of cooling in the intercooler. The exhaust of the recuperated heat-exchanger(a low-grade energy source) is used to drive the absorption system and generate thecooling effect. The effects of ambient temperature, ambient relative humidity, tur-bine inlet-temperature, and the effectiveness of the recuperated heat-exchanger onthe effectiveness of inlet cooling and cycle performance is investigated. The effects ofevaporative inlet cooling, evaporative aftercooling, absorption inlet-cooling, andabsorption inlet-cooling with evaporative aftercooling on cycle efficiency and powerare presented and discussed.

2. Description of the cycles

Fig. 1 shows the intercooled reheat recuperated gas-turbine cycle with absorptioninlet cooling and evaporative aftercooling. This cycle is denoted by IAIEA cycle.Ambient air at 1 is cooled in the air cooler (AC) before it enters the low-pressure


compressor (LPC) at 2. The outlet air of the LPC at 3 is cooled in the air intercooler(AI) in two stages. Ambient air is used to cool the air at 3 in the first stage of coolingto a lower temperature at 4. Cool air out of the AC at 2 is used to cool the air at 4further in the second stage to a lower temperature at 5. The outlet air of the AI at 5is compressed in the high-pressure compressor (HPC) to a higher pressure and tem-perature at 6. The outlet air of the HPC at 6 is cooled in the air aftercooler (AFC)using water-injection. The injected water is warmed using the exhaust of the recup-erated heat exchanger (HE1) at 13 in the water heater (WH). The humid cooler airat 7 enters the HE1 where it counter passes the exhaust gas out of the low-pressureturbine (LPT) at 12. The inlet air is heated to a higher temperature at 8 where itenters the first combustion chamber (CC1) where fuel is burned, producing hot gasat 9. Hot gas at 9 enters the high-pressure turbine (HPT) where it expands to a lowerpressure and temperature at 10 before it is re-combusted in the second combustion-chamber (CC2). More fuel is added in the CC2, producing hot gas at 11. Hot gas at11 enters the LPT where it expands to a lower pressure and temperature at 12. Theabsorption system is shown inside the dotted box. The expanded low-pressure

Fig. 1. A schematic diagram of the intercooled reheat recuperated gas-turbine cycle with absorption inlet-

cooling and evaporative aftercooling.


water–vapour mixture at 3c enters the water cooler (WC) to cool the counter-passwater at 12c to a lower temperature at 13c. The WC is also the evaporator of theabsorption system. The cooler water at 13c counter passes the inlet air to the AC at1 and cools it to a cooler temperature at the inlet of the LPC at 2. When the coolingload in the air cooler increases beyond the capacity of the WC, valves V2 and V3 areused to supply additional cool water from the storage tank. The additional coolwater enters the AC at 13c. The warmer water at 14c returns to the supply tank. Thevalves V1, V2, and V3 are used to control the amount of water flowing to the AC.The extra amount of water is stored in the storage tank. The water–vapour mixturethat enters the evaporator at 3c is boiled and exits the evaporator in a saturatedstate at 8c. The saturated steam at 8c enters the absorber, where it mixes with ahigh-concentration water/lithium bromide mixture at 9c, generating heat that hasto be dissipated to increase the efficiency of the mixing process. Ambient water canbe used to dissipate the generated heat. The mixing process results in a low-con-centration water/lithium bromide mixture that exits the absorber at 10c and ispumped to the upper pressure of the cycle at 5c. The high-pressure low-concen-tration mixture at 5c is heated to a higher temperature at 7c in the heat exchanger(HE2) using the counter-pass high-pressure high-concentration mixture at 6c. Thecooler high-concentration mixture exits the HE2 at 4c and is expanded in theexpansion valve (EV2) resulting in a low-pressure high-concentration mixture at 9c.The high-pressure low-concentration mixture at 7c is heated using the exhaust gas at14 in the desorber, resulting in high-pressure superheated steam at 1c and a high-pressure high-concentration water/lithium bromide mixture at 6c. The high-pressuresuperheated steam at 1c is condensed in the condenser, resulting in high-pressurewater at 2c. The condensed water at 2c is expanded in the expansion valve (EV1),resulting in a low-pressure water–vapour mixture at 3c. The storage system can pro-vide an additional cooling effect whenever the cooling load of the inlet air increasesbeyond the capacity of the water cooler. Fig. 2 shows a representation of the cycle onthe temperature-entropy diagram with no pressure drops in the heating devices.The second cycle under investigation is the intercooled reheat recuperated gas-

turbine cycle with absorption inlet cooling (IAI cycle). This cycle is similar to theIAIEA cycle except that there is no evaporative aftercooling. The third cycle is theintercooled reheat recuperated gas-turbine cycle with evaporative aftercooling. Thiscycle is similar to the IAIEA cycle, except that there is no cooling of the inlet air byany means. The fourth cycle is the intercooled reheat recuperated gas-turbine cyclewith evaporative inlet cooling, denoted by the IEI cycle. The fifth cycle is the inter-cooled reheat recuperated gas-turbine cycle, denoted by the I cycle (see Appendix).

3. Results and discussion

3.1. Effect of inlet cooling and aftercooling

Fig. 3 shows the effect of pressure ratio on the efficiencies of all cycles. For thatfigure TIT=1400 K, To=315 K, and RHo=60%. The optimum pressure-ratio is


affected by many factors such as compressor work and turbine’s output-power.Compressor work is a function of air temperature at the compressor inlet. The tur-bine’s output power is a function of TIT. It is clear from the figure that the optimumpressure ratio for IAI cycle is higher than that for IEI cycle, which is higher than theoptimum pressure ratio for I cycle. That indicates that the optimum pressure ratiocould be a function of the air temperatures at the inlets of the LPC and the HPC aswell as TIT. It can be noticed from the figure that applying evaporative aftercoolingincreases the optimum pressure-ratio. Applying evaporative aftercooling increasesthe capacity of the recuperated heat-exchanger by reducing the inlet-air temperatureand increasing the mass flow rates of air and gas through the heat exchanger.Increasing the capacity of the recuperated heat-exchanger increases energy utiliza-tion, cycle efficiency, and the optimum pressure-ratios for cycles with evaporativeaftercooling. The optimum pressure-ratio for the IAIEA cycle is higher than that forthe IEA cycle. Evaporative and absorption inlet-cooling are also applied at the inletof the HPC, so that, as rc increases, the gain in efficiency because of applying eva-porative inlet-cooling or absorption inlet-cooling increases. As rc increases the airtemperature at the outlet of the HPC increases; thus, raising the amount of injectedwater to reach saturation at the AFC, the mass flow rates through gas-turbines,cycle power, and the gain in efficiency because of applying evaporative aftercooling.It is clear from Fig. 3 that the gain in efficiency because of applying evaporative

Fig. 2. Representation of the cycle shown in Fig. 1 on temperature-entropy diagram with no pressure-

drops in the heating devices.


inlet-cooling could reach 3.9% compared with about 6.6% for applying absorptioninlet-cooling. The optimum efficiency of the IAIEA cycle is 1.5% higher than that ofthe IEA cycle and 3.5% higher than the optimum efficiency of the I cycle.Fig. 4 shows the characteristic curves for all cycles. For that figure, TIT=1400 K,

To=315 K, and RHo=60%. It is clear that applying evaporative inlet-cooling orabsorption inlet-cooling improves both power and efficiency of the I cycle. The effectof applying evaporative aftercooling on the intercooled reheat regenerative cycle isdiscussed in Ref. [6]. The maximum power of the IEI cycle is about 11% higher thanthat of the I cycle and almost the same as that of the IAI cycle. Applying evapora-tive aftercooling could increase the maximum power of the I cycle by more than65%. Applying evaporative aftercooling with absorption inlet cooling could increasethe maximum power of the I cycle by more than 50%.Fig. 5 is similar to Fig. 3 except that the TIT is at a constant value of 1700 K

instead of 1400 K. As the TIT increases, the average temperature of heat receptionincreases, thus, raising the Carnot cycle efficiency and cycle efficiency. Increasing theTIT from 1400 to 1700 K increased the optimum pressure ratios of all cycles by 30–45%. Comparing Figs. 3 and 5, it can be noticed that increasing the TIT reduces thegain in efficiency because of applying evaporative inlet cooling, absorption inletcooling, or evaporative aftercooling at the same value of rc.Fig. 6 is similar to Fig. 4 except that the TIT is at a constant value of 1700 K

instead of 1400 K. It is clear that increasing TIT improves both power and efficiencyof all cycles. Increasing the TIT from 1400 to 1700 K increased the maximum power

Fig. 3. Effect of pressure ratio on the efficiencies of the five cycles at a TIT of 1400 K.


Fig. 4. The characteristics curves of the five cycles at a TIT of 1400 K.

Fig. 5. Effect of pressure ratio on the efficiencies of the five cycles at a TIT of 1700 K.


of the I cycles by about 47–52%. Applying evaporative inlet-cooling increased themaximum power by 10% compared with a 7% increase for applying absorptioninlet cooling. Lines of constant rc show that applying absorption inlet-cooling forthe IEA cycle reduces the output power slightly.

3.2. Effect of ambient temperature and turbine’s inlet-temperature

Fig. 7 shows the effect of ambient temperature on the efficiencies of all cycles. Forthat figure, TIT=1400 K, rc=20, RHo=60%, and "HE1=85%. It is clear from thefigure that increasing the ambient temperature increases the gain in efficiencybecause of applying evaporative inlet cooling or evaporative aftercooling. As theambient temperature increases, the specific work of the compressors increases [5],thus reducing cycle efficiency for the I, IEI, and IEA cycles. For the IAI cycle, up toTo of 312 K, the heat available for the absorption cooling system was more than orequal to that required to cool the inlet air to the optimum inlet-temperature. Theabsorption cooling system maintained the air temperature at the inlet of the LPCclose to the optimum inlet temperature of 280 K. As To increases beyond 312 K (adiscontinuity point), the required heat exceeds the heat available for the absorptioncooling system and the inlet temperature to the LPC increases, thus reducing theefficiency of the IAI cycle. For the IAIEA cycle, applying evaporative aftercoolingincreases the capacity of the recuperated-heat exchanger, energy utilization of thecycle, and reduces the heat available for the absorption cooling-system. Therefore,

Fig. 6. The characteristics curves of the five cycles at a TIT of 1700 K.


up to To of 304 K, the absorption inlet cooling-system maintained the inlet tem-perature to the LPC close to its optimum value of 280 K. Increasing To beyond 304K (the discontinuity point) increases the load on the inlet cooling system, so that therequired heat exceeds the heat available for the cooling system, the temperature atthe inlet of the LPC increases, and cycle efficiency decreases. It can be noticed fromFig. 7 that the gain in efficiency because of applying evaporative inlet-coolingincreases as To increases and could reach 2.6%. For the IAI cycle, the gain in effi-ciency because of applying absorption inlet cooling increases as To increases up tothe discontinuity point, reaching about 4.1%. Increasing To beyond that pointreduces the gain in efficiency. The same can be said for the IAIEA cycle. The gain inefficiency because of applying evaporative aftercooling increases as To increases andcould reach 3.6% (at To=325 K).Fig. 8 is similar to Fig. 7 except that the TIT is at a constant value of 1700 K

instead of 1400 K. Comparing Figs. 7 and 8, it can be noticed that increasing TITincreases the efficiencies of all cycles and slightly reduces the gain in efficiencybecause of applying evaporative inlet-cooling or absorption inlet-cooling. The gainin efficiency because of applying evaporative aftercooling increased slightly as TITincreased. Increasing TIT increased the heat available for the absorption coolingsystem, thus raising its capacity. For the IAI cycle, up to To of 317 (instead of 312 Kfor TIT of 1400 K), the absorption inlet-cooling system maintained the tem-perature at the inlet of the LPC close to its optimum value of 280 K. For theIAIEA cycle, up to To of 310 K (instead of 304 K for TIT of 1400 K), the

Fig. 7. Effect of ambient temperature on the efficiencies of the five cycles at a TIT of 1400 K and "HE1 of

85%.


absorption inlet-cooling system maintained the inlet temperature of the LPC closeto its optimum value. Therefore, increasing the TIT increases the range of themaximum gain in efficiency for the IAI and IAIEA cycles.

3.3. Effect of ambient relative-humidity and turbine’s inlet-temperature

Fig. 9 shows the effect of ambient relative-humidity on the cycle efficiency. Forthat figure, TIT=1400 K, To=315 K, rc=20 and "HE1=85%. It is clear thatincreasing RHo has a negligible effect on the efficiency of the I and IEA cycles. AsRHo increases, the wet-bulb temperature increases, thus, reducing the inlet air tem-perature of the LPC and HPC and there is a gain in efficiency because of applyingevaporative inlet-cooling as shown in the figure. At RHo of 0.98, the wet and drybulb temperatures are almost equal. The difference in efficiency between I and IEIcycles at RHo of 0.98 (a gain of 1.55%) is mainly attributed to using two stages tocool the air at the outlet of the LPC in the case of the IEI cycle. Using two stagesreduced the air temperature at the inlet of the HPC by about 30 K, decreased thework of the HPC, and increased the cycle efficiency by 1.55%. As RHo increases, theload on the absorption inlet cooling system increases. Up to a RHo of 0.51, the heatavailable to generate the cooling effect of the absorption system was more than orequal to the heat required for optimum cooling and the absorption inlet-coolingsystem maintained the temperature at the inlet of the LPC close to its optimumvalue and did not affect the efficiency of the IAI cycle. As RHo increases beyond

Fig. 8. Effect of ambient temperature on the efficiencies of the five cycles at a TIT of 1700 K and "HE1 of

85%.


0.51, the air temperatures at the inlets of the LPC and HPC increase; thus, raisingthe work of the compressors and reducing cycle efficiency. For the IAIEA cycle,applying evaporative aftercooling increases the capacity of the recuperated heat-exchanger and reduces the heat available for the absorption inlet-cooling system sothat the maximum cooling effect of the system was reached at a lower value of RHo

than in the case of the IAI cycle. For the IAIEA cycle, up to a RHo value of 0.23,the absorption inlet-cooling system maintained the air temperature at the inlet of theLPC close to its optimum value and the increase in RHo did not affect the cycleefficiency. As RHo increases beyond a RHo of 0.23, the heat required to generate theoptimum cooling effect exceeds the heat available so that air temperatures at theinlets of the LPC and HPC increase; thus, raising the work of the compressors andreducing the cycle efficiency. It can be noticed from Fig. 9 that the gain in efficiencybecause of applying evaporative inlet-cooling could reach 3.2% compared with4.3% for applying absorption inlet-cooling. The gain in efficiency because of apply-ing evaporative aftercooling and absorption inlet-cooling could reach 5.7%.Fig. 10 is similar to Fig. 9 except that the TIT is at a constant value of 1700 K

instead of 1400 K. Comparing Figs. 9 and 10, it can be noticed that increasing theTIT increases the efficiencies of all cycles and reduces the gain in efficiency becauseof applying evaporative inlet cooling or absorption inlet cooling. Increasing TITincreases the heat available for the absorption cooling system, raising its maximumcooling effect, and the range of RHo in which the maximum gain in efficiency takesplace for the IAI and IAIEA cycles as shown by comparing Figs. 9 and 10.

Fig. 9. Effect of ambient relative-humidity on the efficiencies of the five cycles at a TIT of 1400 K and a

"HE1 of 85%.


Fig. 10. Effect of ambient relative-humidity on the efficiencies of the five cycles at a TIT of 1700 K and a

"HE1 of 85%.

Fig. 11. Effect of ambient temperature on the efficiencies of the five cycles at a TIT of 1400 K and a "HE1

of 80%.


3.4. Effect of the effectiveness of the recuperated heat-exchanger

Fig. 11 is similar to Fig. 7 except that the "HE1 has at a constant value of 80%instead of 85%. Increasing the effectiveness of the recuperated heat-exchangerincreases energy utilization of the cycle and cycle efficiency. Comparing Figs. 7 and11, it can be noticed that increasing "HE1 from 80 to 85% increased the efficiencies ofall cycles by about 1.3–1.8%. Increasing "HE1 reduces the heat available for theabsorption inlet cooling system and the range of the maximum gain in efficiency forthe IAI and IAIEA cycles as shown by comparing Figs. 7 and 11. Fig. 12 is similarto Fig. 8 except that "HE1 is at a constant value of 80%. Comparing Figs. 8 and 12, itcan be noticed that increasing "HE1 from 80 to 85% increased the efficiencies of allcycles by about 1.8–2.1%. Increasing TIT increases the capacity of the recuperatedheat-exchanger and the gain in efficiency because of increasing "HE1. ComparingFigs. 8 and 12 with Figs. 7 and 11, it can be noticed that increasing TIT hasincreased the gain in efficiency because of the increasing "HE1.Fig. 13 is similar to Fig. 9 except that "HE1 is at a constant value of 80%

instead of 85%. Comparing Figs. 9 and 13, it can be noticed that increasing"HE1 from 80% to 85% increased the efficiencies of all cycles by about 1.2–1.8%. Fig. 14 is similar to Fig. 10 except that "HE1 is at a constant value of80% instead of 85%. Comparing Figs. 10 and 14, it can be noticed thatincreasing "HE1 from 80 to 85% increased the efficiencies of all cycles by about1.7–2.1%. Comparing Figs. 9 and 13 with Figs. 10 and 14, one can draw the

Fig. 12. Effect of ambient temperature on the efficiencies of the five cycles at a TIT of 1700 K and a "HE1

of 80%.


Fig. 13. Effect of ambient relative-humidity on the efficiencies of the five cycles at a TIT of 1400 K and

"HE1 of 80%.

Fig. 14. Effect of ambient relative-humidity on the efficiencies of the five cycles at a TIT of 1700 K and

"HE1 of 80%.


same conclusion, which is increasing the TIT increases the capacity of the recup-erated heat exchanger and the gain in efficiency is because of the increasing "HE1.

4. Conclusions

An absorption inlet-cooling system has been introduced to the intercooled reheatrecuperated gas-turbine cycle. Such a system has been driven by the exhaust of therecuperated heat-exchanger. The system has been used to cool the inlet air to theLPC and HPC using two stages of cooling in the intercooler. A parametric study ofthe effects of pressure ratio, ambient temperature, ambient relative humidity, TIT,and the effectiveness of the recuperated heat exchanger on the performance of thefive different lay-outs of the intercooled reheat recuperated gas-turbine cycle leads tothe following conclusions:

. Applying absorption inlet-cooling with evaporative aftercooling, couldincrease the optimum efficiency of the I cycle by 3.5% and its maximum powerby more than 50%.

. Using two stages of cooling in the air intercooler could raise the gain in effi-ciency because of applying evaporative inlet-cooling by up to 1.55%.

. Applying absorption inlet-cooling could increase the efficiency of the I cycle byup to 6.6% and its maximum power by up to 10%.

. Applying evaporative inlet-cooling could increase the efficiency of the I cycleby up to 3.9% and its maximum power by up to 11%.

. Increasing To increases the gain in efficiency because of applying evaporativeinlet-cooling or evaporative aftercooling. Increasing To could not impact theefficiency of the IAI cycle or the IAIEA cycle as long as the absorption inletcooling system can maintain the air temperature at the inlet of the LPC close toits optimum value.

. Increasing RHo has a negligible effect on the efficiency of the I cycle and IEAcycle. Increasing RHo reduces the gain in efficiency because of applying eva-porative inlet cooling. Increasing RHo could not impact the efficiency of theIAI cycle and the IAIEA cycle as long as the absorption inlet-cooling systemcan maintain the air temperature at the inlet of the LPC close to its optimumvalue.

. Increasing the TIT increases the heat available for the absorption cooling sys-tem, raising its maximum cooling effect, and the range of RHo or To in whichthe maximum gain in efficiency takes place for the IAI and IAIEA cycles.Increasing the TIT increases the capacity of the recuperated heat-exchangerand the gain in efficiency because of increasing "HE1.

. Increasing "HE1 increases the energy utilization and the efficiencies of all cyclesby 1.2–2.1%. Increasing �HE1 decreases the heat available for the absorptioncooling system and reduces its maximum cooling effect and the range of RHo

or To in which the maximum gain in efficiency takes place for the IAI andIAIEA cycles.


Appendix. Analysis of the cycles

Assumptions

1. The pressure drop for air in the air aftercooler (AFC) is 0.5%.

2. The pressure drop for gas in the water heater (WH) is 0.5%.

3. The pressure drop for air in the air cooler (AC) is 1%.

4. The pressure drop for air in the heat exchanger (HE1) is 3%.

5. The pressure drop for gas in the heat exchanger (HE1) is 3%.

6. The pressure drop for gas in the desorber of the absorption system is 1.5%.

7. The pressure drop for air in the air intercooler (AI) is 1.5%.

8. The pressure drops for air in the humidifiers and air evaporative cooler arenegligible.

9. The pressure losses and heat transfer losses to the surroundings in the absorptioncycle are ignored.

10. The condensing pressure of the refrigerant of the absorption system equals thepressure of condensing steam at a temperature of 10 K above ambient temperature.

11. The evaporator pressure of the absorption system is constant and equals 650Pascal. The corresponding evaporator temperature is 274 K. The evaporator pres-sure is chosen so that the inlet air can be cooled to a temperature that is close to therequired temperature for optimum operation of 280 K.

12. The pressure of the condensing water that exits the air cooler (AC) is atmo-spheric.

13. The temperature of the superheated steam at the outlet of the desorber at 1c is 15K lower than the temperature of the outlet high-concentration mixture at 6c.

14. The outlet temperature of gas in the desorber of the absorption system at 15 is20 K higher than the temperature of the strong solution at the outlet of the deso-rber at 6c.

15. The temperature of the condensing water in the outlet humid air at the outlet ofthe air cooler (point 2) equals the temperature of the outlet humid air.


16. The effectiveness of the air cooler (AC) is 80%.

17. The effectiveness of the water heater (WH) is 80%.

18. The effectiveness of the heat exchanger of the absorption system (HE2) is 80%.

19. The effectiveness of the heat exchangers of the air intercooler (AI) is 80%.

20. The effectiveness of the recuperated heat-exchanger (HE1) is 85% or as noted.

21. The compressor polytropic efficiency (�c) is 0.88.

22. The turbine efficiency is determined from the following relation [13]:

�t ¼ 1� 0:03þrt � 1

180

� �ðA1Þ

where rt is the pressure ratio across the turbine.

23. The relative humidity at the outlets of the evaporative cooler and evaporativeaftercooler is 98%.

24. The conditions at the outlets of the condenser and the evaporator of theabsorption system are saturated vapour and saturated water, respectively.

25. The concentrations of the strong and weak solutions of the absorption systemare 0.6 and 0.55, respectively. The assumed values are chosen within the operatingrange of the lithium-bromide absorption systems and are commonly used [14].

26. The specific heat at constant pressure for both weak and strong solutions are equal.

27. Adiabatic expansion in all throttling valves.

28. The pump work of the absorption system is negligible.

29. The fuel is methane that has a lower heating value of 50 016 kJ/kg.

30. Some applications may require combining the water cooler (WC) and the aircooler (AC) in one heat-exchanger. For calculation purposes, the water cooler (WC)and the air cooler (AC) are combined in one heat-exchanger that exchanges heatbetween the evaporated steam and the inlet air to the LPC.

31. The cooling process in the air cooler (AC) occurs at constant humidity-ratio ifthe point of relative humidity of 98% has not been reached. If the point of 98%relative humidity has been reached during the cooling process, the process continuesat a constant relative humidity of 98%.


The analysis presented here will be for the air intercooler (AI) and the absorptioncooling section of the cycle. Analysis of the other components of the intercooled reheatrecuperated gas turbine cycles with evaporative inlet and aftercooling is given in Ref. [6].

Air intercoolerThe cooling process in the air intercooler occurs in two stages or using two heat

exchangers for the IAI, IEI, and IAIEA cycles. The effectiveness of the first heatexchanger is defined as

"AI1 ¼m:

acpaT3 � T4ð Þ

m:

acpaT3 � Toð Þ

¼T3 � T4ð Þ

T3 � Toð ÞðA2Þ

where T3, T4, and To are air temperatures at the inlet and outlet of the first heatexchanger and ambient temperature, respectively. Analysis of air compressors canbe used to determine T3, so that T4 can be determined using Eq. (A2). The corre-sponding enthalpy h4 can be determined as a function of air temperature. Theeffectiveness of the second heat-exchanger is defined as

"AI2 ¼m:

acpaT4 � T5ð Þ

m:

acpaT4 � T2ð Þ

¼T4 � T5ð Þ

T4 � T2ð ÞðA3Þ

where T2 and T5 are air temperatures of the cooling stream at the inlet of the secondheat exchanger and at the inlet of the HPC, respectively (please view Fig. 1). Thetemperature, T2, can be determined using the analysis of the evaporator (SectionEvaporator) so that Eq. (A3) can be used to determine T5. The correspondingenthalpy h5 can be determined as a function of air temperature.

Desorber

Applying the energy equation for the desorber of the absorption system yields

m:

g h14 � h15ð Þ ¼ m:

sh6c þ m:

rh1c � m:

wh7c ðA4Þ

: : : :
where h15 is the enthalpy of gas exiting the desorber, mg, mw, mr, ms are the massflow rates of gas entering the desorber, the weak solution entering the desorber, thesteam (the refrigerant) exiting the desorber, and the strong solution exiting the des-orber, respectively; and h14, h7c, h1c, and h6c are the corresponding enthalpies,respectively. Applying the mass balance equation yields
m:

s þ m:

r ¼ m:

w ðA5Þ

Using the concentration–mass flow rate relation gives

m:

w ¼Xs

Xwm:

s ðA6Þ

where Xs and Xw are the concentrations of the strong and weak solutions, respec-tively. Both concentrations are known and equal to 0.6 and 0.55, respectively (please


see Assumptions 16–25, 28–31). By using Assumptions 1–11, the high pressure of thecycle (Phi) equals the condenser pressure, and the low pressure (Plo) of the cycleequals the evaporator pressure so that both pressures can be determined. The tem-perature of the outlet high-concentration mixture (T6c) can be determined as afunction of its concentration and pressure using a program called EngineeringEquation Solver (EES) [15] as follows

T6c ¼ TLIBR Phi;Xsð Þ ðA7Þ

A web site that has additional information on the program can be found atwww.fChart.com. The corresponding enthalpy of the high-concentration mixture(h6c) can be determined as a function of its concentration and temperature usingEES [15] so that

h6c ¼ hLIBR T6c;Xsð Þ ðA8Þ

An equation can be written for the temperature of gas at the desorber outlet inrelation to the temperature of the high-concentration mixture at 6c using Assump-tion 14.

T15 ¼ T6c þ 20 ðA9Þ

The gas pressure at the outlet of the desorber can be determined using Assumption6. The enthalpy of the outlet gas can be determined as a function of its pressure andtemperature using EES [15] as

h15 ¼ hAIR T15;P15ð Þ ðA10Þ

An equation can be written for the temperature of the outlet steam in relation to theoutlet high-concentration mixture at 6c using Assumption 13.

T1c ¼ T6c � 15 ðA11Þ

The enthalpy of the outlet steam can be determined as a function of its pressure andtemperature using EES [15] as

h1c ¼ hSTEAM Phi;T1cð Þ ðA12Þ

Heat exchanger

The heat-exchanger (HE2) effectiveness is defined as

"HE2 ¼m:

scps T6c � T4cð Þ

m:

scps T6c � T5cð Þ¼

T6c � T4c

T6c � T5cðA13Þ

where T4c and T5c are the temperatures of the outlet strong solution and inlet weaksolution respectively. Assumption 26 is used to simplify Eq. (A13). The temperature,T5c can be determined as a function of its pressure and concentration using EES [15] as

T5c ¼ TLIBR Phi;Xwð Þ ðA14Þ


lying the energy balance equation yields
App
m:

s h6c � h4cð Þ ¼ m:

w h7c � h5cð Þ ðA15Þ

where h4c and h5c are the enthalpies of the outlet strong solution and inlet weaksolution, respectively. The enthalpies, h4c and h5c can be determined as a function oftheir temperatures and concentrations using EES [15] as

h4c ¼ hLIBR T4c;Xsð Þ ðA16Þ

h5c ¼ hLIBR T5c;Xwð Þ ðA17Þ

Eqs. A4–A17 are solved simultaneously for the unknowns h1c, h4c, h5c, h6c, h7c, h15,m:

r, m:

s, m:

w, T1c, T4c, T5c, T6c, and T15 using EES [15].

Evaporator

By making use of Assumptions 9, 24, and 27, the enthalpy at the inlet and outletof the evaporator can be determined. Applying the energy-balance equation yields

Q:

cooling ¼ 2m:

ah1 � 2 m:

a � m:

wað Þh2 � 2m:

wahwa ðA18Þ

Q:

cooling ¼ m:

ru h8c � h3cð Þ ð19Þ

where h3c and h8c are the enthalpies at the inlet and outlet of the evaporator,respectively; h1, h2, and hwa are the enthalpies of the inlet air to the air cooler, theoutlet air to the cooler, and the condensing water; m

:a, m

:wa, and m

:ru are the mass

flow rates of air, condensing water, and used refrigerant, respectively. The mass flowrate of the used refrigerant is the mass flow rate that is used to cool the inlet air to atemperature that is close to the optimum temperature of 280 K. In Eq. (A18), 50%of the mass flow rate of air enters the LPC and 50% is used to cool the inlet air tothe HPC as shown in Fig. 1. The humidity ratio at the air cooler outlet can bedetermined from the following relation

w2 ¼ w1 �m:

wa

m:

aðA20Þ

where w1 and w2 are the humidity ratios at the inlet and outlet of the air cooler. Theminimum temperature of air at the cooler outlet is assumed to be 280 K. If the heatavailable to heat the desorber of the absorption system is equal to or more than theheat required to cool the inlet air to its minimum temperature, the air temperature atthe inlet of the LPC will approach a value that is close to its minimum value of 280K (case 1). If the heat available is less than that required to cool the inlet air to avalue that is close to its minimum temperature m

:r ¼ m

:ruð Þ, the air temperature at the

inlet of the LPC will be higher than 280 K (case 2). Therefore, there are two cases tobe considered. If the humidity ratio at the air cooler inlet is less than the humidity


ratio at a temperature of 280 K and relative humidity of 98%, no condensation willtake place in the air cooler (w1=w2). By making use of Assumption 3, the pressureat the outlet of the air cooler can be determined. An equation for the humidity ratiocan be written as a function of its pressure, temperature, and relative humidity usingEES [15] so that

w2 ¼ wAIR T2;P2;RH2ð Þ ðA21Þ

The enthalpy of air at the cooler exit can be determined as a function of its humidityratio, pressure, and temperature using EES [15].

h2 ¼ hAIR w2;P2;T2ð Þ ðA22Þ

The enthalpy of the condensing water can be determined as a function of its tem-perature and pressure using EES [15] so that

hwa ¼ hWATER Pwa;Twað Þ ðA23Þ

By using Assumptions 12 and 15, the water pressure is known and its temperatureequals the air temperature at the cooler exit. In case 1, Eqs. (A18)–(A23) can besolved simultaneously for the unknowns m

:wa, h2, w2, Q

:cooling, m

:ru, and hwa. In case 2,

Eqs. (A18–(A.23) can be solved simultaneously for the unknowns m:

wa,Q:

cooling, h2,w2, hwa, and T2. Equations for energy and mass balance for other components of thecycle and for thermodynamic properties are added to form a system of non-linearequations. EES [15] is used to evaluate thermodynamic properties and solve thesystem of non-linear equations. Cycle efficiency was determined using two energybalances. The results were checked first by comparing cycle efficiency determinedusing the energy balance with the cycle efficiency determined using the second energybalance. Both efficiencies were matched in all runs. The results of a few runs werealso checked using hand calculations.

References

[1] Giourof A. Gas-turbine inlet-air cooling: you can almost pick your payback. Power 1995;139(5):56–8.

[2] De Lucia M, Carnevale E, Falchetti M, Tesei A. Performance improvements of a natural-gas injec-

tion station using gas-turbine inlet air cooling. In: Proceedings of the International Gas Turbine and

Aeroengine Congress and Exposition, Orlando, FL, 2–5 June 1997.

[3] De Lucia M, Lanfranchi C, Boggio V. Benefits of compressor inlet air cooling for gas-turbine

cogeneration plants. In: Proceedings of the International Gas Turbine and Aeroengine Congress and

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[4] Ait-Ali MA. Optimum power boosting of gas-turbine cycles with compressor inlet air refrigeration.

Journal of Engineering for Gas Turbines and Power 1997;119(1):124–9.

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Applied Thermal Engineering 2001;21(18):1875–90.

[6] Bassily AM. Performance improvements of the intercooled reheat regenerative gas-turbine cycles

using indirect evaporative cooling of the inlet air and evaporative cooling of the compressor dis-

charge. Proceedings of Institute of Mechanical Engineers Part A, Journal of Power and Energy 2001;

215(A5):545–57.

[7] Ondryas IS, Daniel F. Go beyond evaporative coolers to stretch gas-turbine output. Power magazine

1991;July:27–9.


[8] Hufford PE. Absorption chillers improve cogeneration energy efficiency. ASHRAE Journal 1992;

March:46–53.

[9] Ondryas IS, Wilson DA, Kawamoto M, Haub GL. Options in gas turbine power augmentation

using inlet air chilling. Journal of Engineering for gas turbines and power 1991;113(4):203–11.

[10] Najjar YS. Enhancement of performance of gas-turbine engines by inlet air cooling and cogeneration

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Orlando, Florida, 2–5 June 1997.

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cooling and evaporative aftercooling. Proceedings of Institute of Mechanical Engineers Part A,

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[13] Najjar YS, Zaamout MS. Enhancing gas-turbine engine performance by means of the evaporative

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[14] Herold KE, Radermacher R, Klein SA. Absorption chillers and heat pumps. 1st ed. New York: CRC

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21 Bassily - Absopt x Evaporative

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