Energy Conversion and Management 49 (2008) 2130–2141

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Energy Conversion and Management

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Analysis of combustion turbine inlet air cooling systems applied to an operatingcogeneration power plant

R. Chacartegui *, F. Jiménez-Espadafor, D. Sánchez, T. SánchezThermal Power Group (GMTS), Department of Energy Engineering, University of Seville, Camino de los Descubrimientos s/n, 41092 Sevilla, Spain

a r t i c l e i n f o a b s t r a c t

Article history:Received 20 June 2007Accepted 20 February 2008Available online 14 April 2008

Keywords:Combustion turbine inlet air coolingCogenerationPower plant modelling

0196-8904/$ - see front matter � 2008 Elsevier Ltd. Adoi:10.1016/j.enconman.2008.02.023

* Corresponding author. Tel.: +34 954 48 72 42; faxE-mail address: ricardo@esi.us.es (R. Chacartegui).

In this work, combustion turbine inlet air cooling (CTIAC) systems are analyzed from an economic out-look, their effects on the global performance parameters and the economic results of the power plant.The study has been carried out on a combined cogeneration system, composed of a General Electric PG6541 gas turbine and a heat recovery steam generator. The work has been divided into three parts. First,a revision of the present CTIAC technologies is shown, their effects on power plant performance and eval-uation of the associated investment and maintenance costs. In a second phase of the work, the cogene-ration plant was modelled with the objective of evaluating the power increase and the effects on thegenerated steam and the thermal oil. The cogeneration power plant model was developed, departingfrom the recorded operational data of the plant in 2005 and the gas turbine model offered by GeneralElectric, to take into consideration that, in 2000, the gas turbine had been remodelled and the originalperformance curves should be corrected. The final objective of this model was to express the power plantmain variables as a function of the gas turbine intake temperature, pressure and relative humidity.Finally, this model was applied to analyze the economic interest of different intake cooling systems, indifferent operative ranges and with different cooling capacities.

� 2008 Elsevier Ltd. All rights reserved.

1. Introduction

A combustion turbine inlet air cooling system (CTIAC) is a de-vice, or a group of devices, that cools the gas turbine air intake. Agas turbine’s performance is highly dependent on the intake flowconditions. A gas turbine loses approximately 7% of its nominalpower when the intake temperature increases from 15 �C, ISO con-ditions, to 25 �C, and in cases such as the power plant under consid-eration, when in summer, the ambient temperature increases abovethe 25 �C, the losses are still bigger, reaching even 15% of the powerrating with 36 �C. Other effects of the higher intake air temperatureare the increase of the heat rate (HR) and decrease of the compres-sion ratio, as well as the increase of the gas turbine exhaust temper-ature and decrease of the exhaust gases mass flow. Therefore, theintroduction of CTIAC systems will modify the heat delivery charac-teristics to the systems installed in the gas turbine exhaust, a steamgenerator in the case under study, and necessitate analysis of the ef-fect on the performance of the whole power plant [1–3].

The introduction of an intake air cooler in an operating powerplant allows producing a higher power than the nominal ratingpower by cooling the air below ISO conditions, depending on theconditions outside the plant [4]. This intake air cooling has limits;

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on the one hand, are the gas turbine mechanical characteristics andalternator power limitation, while on the other, ice crystals mayappear in the gas turbine intake, which can generate erosion andwear of the intake vanes. So, in practice, intake temperature doesnot go below 43–45 �F (6.1–7.2 �C) [1,3,5].

Installation of CTIAC systems in operating plants presents someassociated inconveniences, as they require investment in equip-ment and maintenance costs. In some cases, it even requiresremodelling of some gas turbine sections. Besides, the equipmentsinstalled at the turbine air intake sections create additional pres-sure drops that will depend on the CTIAC technology installed.Also, the profitability of introducing a CTIAC system depends onmany external factors that are not controllable, like climatology,electricity tariffs and fuel price. So, it requires a particular viabilityanalysis for each power plant.

2. Combustion turbine inlet air cooling systems

In summary, the main CTIAC technologies are the following[6–9]:

2.1. Evaporative coolers [10–13]

They use water evaporation to cool the air. The water evapo-rates as it absorbs heat from the incoming air and reduces the

Table 1Estimated investment costs and parasitic loads for evaporative systems

MH FOG

Investment cost (€/kW added) 25–60 30–70Parasitic load (% extra power generated) 0.3–0.5 0.5–0.7

Nomenclature

Cp constant pressure specific heatCH chillerCDS percentage of cooling demand satisfiedCOP coefficient of performanceCPI consumer price indexCTIAC combustion turbine inlet air coolingEA absorption chillersFOG fogging cooling systemIGV inlet guide vanesGN natural gasHR heat rateHRSG heat recovery steam generatorHp lower calorific valueIIR internal rate of returnm mass flow (kg/s)MH evaporative media cooling systemNPV net present valueOMEL electricity market operatorP pricer compression ratioRT refrigeration tons (1 RT equals 12,000 Btu/h or 3.52 kW)SH hybrid systemsT temperatureTES thermal energy storage systemsTIT turbine inlet temperatureW powerWc compression workw humidity

Subscripts01 stagnation inleta airamb ambient temperaturec compressorcc combustion chambercs single effect absorption cyclecd double effect absorption cycledisp availableelec electric motor driven chillerf fuelg gas turbine exhausth high pressure steami initial conditionsice ice storage systemmg gas engine driven chillersmed medium pressure steamt turbinetv steam turbine driven chillerw cold water storage system

Greek symbolsc specific heat ratiog efficiency

R. Chacartegui et al. / Energy Conversion and Management 49 (2008) 2130–2141 2131

dry bulb temperature. Evaporative systems can be classified as di-rect evaporative systems and indirect evaporative systems. In indi-rect evaporative systems, water is not brought in direct contactwith the incoming air, water evaporation and the decrease in drybulb temperature is done within an intermediate air stream flow,and it cools the compressor intake air in an additional heat exchan-ger [10]. In direct evaporative systems, water is brought in contactwith the incoming air, and it evolves with the compressor intakestream flow.

The main direct evaporative cooling systems are the evapora-tive media (MH) [11] and fogging systems (FOG). The minimumtemperature is limited to 10 �C to avoid the risk of freezing atthe intake of the compressor. They are capable of cooling inlet airto within 85–90% of saturation. In these evaporative systems, theevolution of fine water drops can produce an increase of extrapower [14]. In addition to increasing the incoming mass flow,evaporation of drops will absorb latent heat, reducing the temper-ature of the mass that passes through the compressor [2], reducingthe compression work that depends on the intake flow tempera-ture, Eq. (1) [10]

Wc ¼ CpT01ðrðc�1Þ=cc � 1Þ ð1Þ

Reduced NOx emissions with this CTIAC are attributed to the in-creased thermal heat capacity because of the water and the de-creased compressor discharge temperature [5].

In evaporative media, the air passes over surfaces where watercirculates. MH systems have the lowest installation costs of all theCTIAC systems. Also, they have a very low parasitic load (powerconsumed by the cooling system and auxiliary systems). Anotheradditional advantage is that MH systems use potable water (cheap-er than deionised water). Their main drawbacks are a high depen-dence on ambient conditions (temperature and relative humidityevolution), and the intake pressure drop even takes place whenthe cooling system is not working.

In fogging systems, high pressure water is injected [12,13].These systems can produce nearly saturated air due to the verysmall drop size, even with reduced flow residence times, and so,they have a very high effectiveness, about 0.95–1. The grill of injec-tors also does not disturb the flow, and the pressures drops arevery low. These systems have the possibility of overspray opera-tion, or high fogging, with higher mass flow than would be neededto saturate the intake air, to force an ‘‘intercooling” of the air pass-ing through the compressor [15]. Fogging systems have the secondlowest acquisition costs of all CTIAC systems. Also, they have a lowparasitic load. As drawbacks, these systems have a limited coolingcapacity, down to the wet bulb temperature, and a high ambientdependence, they consume deionised water, are more expensiveand present corrosion problems.

Estimated installation costs and parasitic loads for evaporativesystems are shown in Table 1.

2.2. Mechanical cooling compression or chillers (CH)

They use a classical mechanical compression cycle, with anintermediate refrigerant fluid, which cools the intake air passingthrough heat exchangers located in the gas turbine inlet [7,16].These systems can produce a stronger cooling effect than the evap-orative systems, cooling the air down to 45 �F (7.2 �C). Estimatedinstallation costs and mean parasitic load for these systems arecompared in Table 2, differing by the form of driving the compres-sor among chillers with electric motor drive (CHelec), chillers with

Table 2Estimated investment costs, parasitic loads and COP for mechanical and absorptionchillers

CHelec CHmg CHtv EAcs EAcd

Investment cost (€/RT) 870 1.225 1.300 1.000 1.130Parasitic load (kW/RT) 0.81 0.187 0.195 0.31 0.265COP 2 and 6 1 and 2.5 1 and 2.5 0.7 1.2

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gas engine drive (CHmg) and chillers with steam turbine drive(CHtv). Values are given as a function of the refrigeration capacityinstalled in RT (refrigeration tons).

2.3. Absorption chillers (EA)

They use an absorption refrigeration cycle with an absorbentfluid and a refrigerating fluid. The working fluids used are water/ammonia [17] or lithium bromide/water [3]. Although the ammo-nia presents better refrigerating properties than water, for toxicityproblems, the technology of lithium bromide/water [6,18–20] isused. The evaporation temperature limits the inlet air cooling to10 �C, higher than the temperature reached with mechanical chill-ers. Estimated mean installation costs and parasitic loads for thesingle effect LiBr–H2O absorption refrigeration cycle (EAcs) andthe double effect LiBr–H2O absorption refrigeration cycle (EAcd)are shown in Table 2. The double effect LiBr–H2O absorption refrig-eration cycle improves the performance of the single effect cycle.Also, different heat sources have been considered, direct heatsource in an auxiliary boiler or taking advantage of the residualheat of the plant, for the single effect cycle with hot water at115 �C and for the double effect cycle, medium pressure steam(8.5–10 bar).

2.4. Hybrid systems

They are combinations of two or more of the previously men-tioned systems. Their main purpose is to give operating flexibilityto the cooling system to cover the demand, avoiding high parasiticload (electricity consumption) in periods with high electricity tar-iffs. The most usual combinations for hybrid systems are mechan-ical chillers (mainly electric motor driven chillers) in combinationwith absorption chillers. The use of evaporative coolers combinedwith mechanical chillers is not effective.

2.5. Thermal energy storage systems

A thermal energy storage system (TES) allows storing cold(heat) so that a delay exists between the cold fluid storage(charging) and its consumption (discharging). Their feature isdouble, the main one is to avoid parasitic load consumption in‘‘on-peak” periods, and the second one is to be able to cover cool-ing demand peaks with smaller cooling systems [13]. The use ofcontinuous cooling is of interest when cooling the intake isneeded for any one period greater than six hours [6]. They areusually installed with electric chillers, although they can also beused with hybrid systems. According to their storage form, thesesystems will be classified as cooled water storage systems (TESw),ice storage systems (TESice) and eutectic salt mixtures storagesystems [21–24].

Table 3Estimated investment costs for thermal storage systems

TESw TESice

€/RT h 157.5–2.75 � 10-3 RT h for 2000 < RT h < 20,000 145 €/RT h

Cooled water storage systems store sensible heat energy; theyare easy to install and have quick response on demand. They aremore profitable for large storage sizes. Ice storage systems storeenergy in form of latent heat energy for which they are as muchas six times smaller than a cooled water storage system of thesame heat storage capacity. They work at a lower temperatureof evaporation, and this penalizes the COP of the plant. A widevariety of types exists, but one of quicker response is the IceHarvester.

The estimated costs for each storage system are given in Table 3.

3. Power plant description

The power plant under study is located in the south of Spain, inCádiz, next to the sea. The cogeneration plant is inside a chemicalplant, and it is composed of a General Electric gas turbine, modelPG6541B, installed in 1995, and a heat recovery boiler. The turbineworks in an open cycle and is composed of an axial compressorwith 17 stages and a compression ratio of 1:11, and an axial tur-bine with 3 stages with a turbine inlet temperature (TIT) of1100 �C. This gas turbine can be fuelled with natural gas or withnaphtha. Their nominal characteristics values are

Nominal Power ðGNÞ ¼ 38:340 kWNominal Heat Rate ðGNÞ ¼ 11:460 kJ=kW h

The gas turbine is coupled to a natural circulation heat recoverysteam generator (HRSG), without post combustion, which takesadvantage of the residual heat of the exhaust gases. In the HRSG,superheated steam is generated at two pressures (10 and 64 bars)and thermal oil is heated for its use in the chemical processes ofthe plant (290–345 �C). The HRSG generated mass flows andstream flow temperatures are given in Table 4.

Of the electric power generated in the gas turbine, part is con-sumed in the chemical plant, and the remainder is exported tothe electric network. Also, from the generated steam, part is usedin their own chemical plant, and the remainder is exported to anadjacent refinery, while all the thermal oil is consumed in thechemical plant. The thermal oil mass flow and temperature are im-posed by the demands of the chemical external process.

In this part of the work, a model of the power plant was devel-oped to be used to evaluate the economic effect of inserting differ-ent CTIAC systems for this power plant [7,3,25].

In 2000, were remodelled the gas turbine, changing the vanes ofthe second and third stages and the IGV (inlet guide vanes). As aconsequence, neither the original curves of the gas turbine, GEPG6541B, nor those corresponding to a similar model after theremodelling, GE PG6551, were a valid solution for direct applica-tion in the power plant model [24]. After the modification, GE pro-vided two correlations for the power and the heat rate, but they didnot include the other necessary variables for study of the cogene-ration power plant performance, like the exhaust flow gases andthe exhaust temperature.

On the other hand, for the HRSG, isolated and insufficient infor-mation was provided for its model. So, a model was generatedthrough 2005 registered operational data. It was generated usinga filtrate of the data, discarding anomalous or not valid data forthe model that was desired. The data corresponding to operative

Table 4HRSG generated mass flows and stream flow temperatures

Production (T/h) Temperature (�C)

Steam 64 bar 40.7 480Steam 10 bar 13.7 187Thermal Oil 395 290 and 345

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conditions in which the values of power delivery by the gas turbinethat indicated start conditions or stopping conditions, that showoperating with washing on line or that with strongly partial oper-ating conditions were discarded. Also, the data of generation ofsteam that indicated a heating or stopping process in the HRSGor a stack diverter by pass were discarded. After this filtration,about 85% of the 2005 time registers were valid data to preparethe model.

Fig. 1. Registered power data as fu

30

32

34

36

38

40

42

44

0 5 10 15

Intake Tem

Pow

er (

MW

)

Fig. 2. Power calculated correlation from measured data and GE p

Fig. 3. Heat rate obtained from 2005 registere

The model of the gas turbine was prepared for base load oper-ation, about 95% of the annual performance of the plant. The usedexpressions to model power, heat rate and exhaust temperature inthe plant as functions of the intake temperature, relative humidity,nature of fuel and pressure drops are given in Appendix 1. Figs. 1–4represent the values for these three parameters obtained from theoperation registers as well as those obtained from the correlationsgiven by the manufacturer.

nction of intake temperature.

20 25 30 35 40

perature (ºC)

Original manufacturer correlation

Correlation from measured power

Manufacturer correlation after gasturbine modification

ower correlations before and after gas turbine modifications.

d data as function of intake temperature.

10500

11000

11500

12000

12500

13000

0 10 20 30 40

Intake Temperature (ºC)

Hea

t Rat

e (k

J/kW

h)

Original Heat Rate correlation

Heat Rate from measured data

Heat Rate correlation after gas turbine modification

Fig. 4. Heat rate calculated correlation from measured data and GE HR correlations before and after gas turbine modifications.

Fig. 5. Exhaust gases temperature correlation obtained from measured data.

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With the measured values, or those estimated from them, theexpressions for the model of the gas turbine were adjusted, havingthe gas turbine inlet temperature, relative humidity, fuel and pres-sure drops, as independent variables and obtaining the power andheat rate as a result of the inlet conditions from the expressions.The dispersion of data can be due to differences in the compositionof the fuel from the one expressed in the correlations, pollution offilters, uncertainty in measurements, imperfect fit of the correla-tions given by the manufacturer, or safety regulation controls thatcontinuously vary the gas turbine performance around a fixedoperating point (gas generator speed, TIT, exhaust gas temperatureetc.).

For the joint model of the gas turbine–HRSG, the registered ex-haust temperature was used, as is shown in Fig. 5, developing acorrelation with a 95% confidence interval.

Thus, obtaining the correlation given in Eq. (2) where the lastterm shows the confidence interval between both signs

Tg ð�CÞ ¼ 0:722Tamb þ 537:49� 5:72 ð2Þ

The exhaust gases flow was correlated with the intake conditionsthrough the global energy balance in the gas turbine, Eq. (3). Thecorrelation obtained is shown in Eq. (5). The original correlationfor the exhaust gases flow and the new correlation obtained areshown in Fig. 6

Fig. 6. Exhaust gases mass flow, original correlation an

maCpaTamb þmf Hpgcc ¼Wgtþ ðma þmf ÞCpgTg ð3Þ

where the efficiency of the turbine was correlated by theexpression

d new correlation obtained from measured data.

0

5

10

15

20

25

30

35

40

45

50

10000 11000 12000 13000 14000 15000 16000

Thermal oil power consumption (kW)

Ste

am m

ass

flow

(T

n/h)

High pressure steamMedium pressure steam

Fig. 9. Relationship among steam mass flows and thermal oil heat demand.

R. Chacartegui et al. / Energy Conversion and Management 49 (2008) 2130–2141 2135

gt ¼ 0:0008W þ 0:9203 ð4Þmg ¼ �0:4489Tamb þ 138:29� 1:48 ð5Þ

With these expressions for the exhaust temperature and ex-haust mass flow, the model of the HRSG and the power plant withbase load operation was constructed with the independent vari-ables being the mass flow and exhaust temperature, related di-rectly with the gas turbine intake conditions [26]. As results, themodel gives the steam production and the thermal oil productionas a function of a unique variable that includes, simultaneously,the mass flow and the exhaust temperature in the form of availablepower in the exhaust gases, Fig. 7.

The obtained correlation for this variable is shown in Eq. (6).Starting from the steam and thermal oil productions expressed asa function of available power, Fig. 7, as well as the relationshipamong the production of each one of the flows, Fig. 8, conditionedby the HRSG design, Fig. 9, and the heat demand of the chemicalplant where the power plant is placed, the expressions shown inAppendix 1 were obtained. With this model, the economic studyof the different CTIAC systems, as functions of the prices of fueland electricity, as well as the costs of the equipments (investment,maintenance, others) can be conducted

WdispðkWÞ ¼ �179:4 � Tamb þ 87080 ð6Þ

80000

81000

82000

83000

84000

85000

86000

87000

88000

0 5 10 15

IntakeTem

Ava

ilabl

e po

wer

inex

haus

t gas

es (

kW)

Fig. 7. Available exhaust gases heat powe

0

10000

20000

30000

40000

50000

60000

70000

80000 81000 82000 83000 84

Exhaust gases avail

HR

SG

Hea

t Bal

ance

(kW

)

Fig. 8. HRSG he

The relationship among the power supplied with the thermal oilflow and the steam flows generated is shown in Fig. 10 where, dueto the thermal oil temperature ranges, the thermal oil economizeris situated between the medium pressure steam generator and thehigh pressure steam generator. The HRSG control system adjusts

20 25 30 35 40

perature (ºC)

r as function of intake temperature.

000 85000 86000 87000 88000

able heat power (kW)

Total HeatHigh pressure steam

Thermal oilMedium pressure steam

at balance.

Fig. 10. HRSG scheme. Stack gases temperature (�C), high pressure steam flow and medium pressure steam flow evaporators and thermal oil economizer.

2136 R. Chacartegui et al. / Energy Conversion and Management 49 (2008) 2130–2141

the thermal oil temperature, and when the thermal oil demand de-creases, the high pressure sections absorb more heat and the dis-posal heat to the medium pressure sections of the steamgenerator decrease.

4. CTIAC systems modelling

In this section, the treatment given to the different cooling sys-tems, as well as the main hypotheses used, is discussed.

Depending on the inlet air conditions, the cooling evolution inthe chillers will be different. Different cooling possibilities areshown in the psychometric chart, Fig. 11. Three zones can be dis-tinguished in this psychometric chart, delimited by two lines,one horizontal and one vertical, one line of dry bulb constant tem-perature and another line of constant relative humidity. The first isdetermined by the cooling temperature limit of the chiller, 7.2 �C inthe case of mechanical chillers and 10 �C with absorption chillers,while the second line corresponds to constant absolute humiditywith the value of absolute humidity given by the intersection pointwith the limit dry bulb temperature.

Zone 1 corresponds to all the points with dry bulb temperatureless than the limit temperature; therefore, the final conditions ofany point with these initial conditions within this zone are thesame as the initial conditions of temperature for any humidity.Zones 2 and 3 are those in which the temperature of the air isabove the limit temperature, and therefore, they can be cooledby the chiller. The difference resides in the absolute humidity ofthe points of the zones, the first has a lower humidity than the suit-able one and the second ha a greater humidity. The points withinzone 2 evolve according to an absolute constant humidity line;they consume sensible heat but not latent heat, since they cannever arrive at the saturation. Their limit is exactly the consignedtemperature limit, although if the cooling capacity of the chilleris not enough, the final temperature will be greater than the limittemperature of the chiller. For the points belonging to zone 3, twodifferent evolution forms exist, one for the initial conditionsmarked in the figure as type 1, in which the cooling capacity ofthe chiller is not able to cool the air sufficiently to take it to the sat-

uration line, and the other for the initial conditions marked in thefigure as type 2, in which the cooling capacity of the chiller is ableto take the initial point to the saturation and to continue cooling it,following the saturation line.

Given the initial conditions (Ti,Wi) and the cooling capacity ofthe selected chiller, it is possible to calculate the maximum coolingthat could be provided for any initial point and the final tempera-ture that could be reached, depending upon which of the threeways of cooling is chosen.

The intake pressure drop penalties in power and heat rate havebeen considered, and in the case of inserting a heat exchanger inthe exhaust, as in the case of absorption chillers, the additionalpressure drops have been included and the power and heat ratecorrected using the expressions shown in Appendix 1.

In the case of absorption chillers, the maximum heat that can berecovered from the stack is conditioned by the stack exhaust tem-perature that can be reached without corrosion problems due tocondensation on the tubes. In the case of a Li–Br single effectabsorption chiller, the maximum cooling capacity that can be used,using hot water, will be 300 RT, while for a Li–Br double effectabsorption chiller, using steam, it can reach a maximum capacityof 500 RT.

Hourly electricity tariffs have been considered. In Fig. 12, thehourly prices evolution of representative days taken for Januaryand July are shown. In the viability analysis of hybrid systems,and later of the TES systems, this treatment of the electricity pricesthroughout the day is fundamental to the analysis of the chillerthat will operate in determined periods to minimize the operatingcosts of the group, obtaining the maximum benefits of the powerplant.

For thermal energy storage systems, a daily storage system waschosen, developing a typical day model for each month, with theassociated hourly electricity tariffs, taking a series of representa-tive days such that the amount of cooling reached 80% of the cool-ing demand of the month. In Fig. 13, the mean chiller coolingcapacity and minimum storage capacity obtained for the yearusing these representative days to cover the whole cooling demandare shown.

Fig. 11. Chiller cooling treatment as function of the initial conditions and the chiller cooling capacity.

Fig. 12. Hourly electricity prices.

Fig. 13. Evaluated monthly chiller mean cooling power and related minimum st-orage capacity.

R. Chacartegui et al. / Energy Conversion and Management 49 (2008) 2130–2141 2137

From Fig. 13, two main conclusions can be extracted. The first isobvious; the required mean cooling power is increased in summer,and the second is that the associated cooling storage minimumcapacity does not follow the demand shape exactly, reaching itsmaximum values in months in which the relative time delay be-tween the cooling demand and cooling supply is higher.

5. Economic considerations of the analysis

The main aspects and hypotheses of the economic analysis areshown in this section. The phases of the economic analysis aresummarized in the following steps:

– Election of a cooling system with a refrigeration capacity.– Evaluation of the benefits of the group before and after implan-

tation of the cooling system.– Differential evaluation of the costs and benefits of operation

(electricity, steam, thermal oil, natural gas, deionised water).– Evaluation of the maintenance costs and investment.– Economic analysis of the solution.– Return to the first point, varying the capacity of the equipment.

Prices of steam, thermal oil and natural gas have been takenfrom 2005 data. Also, the electricity tariffs have been taken fromthe OMEL. The annual evaluation of costs and benefits has beendivided into four periods, corresponding to the quarterly revisionsof the rates of natural gas.

A power plant operation factor of 0.94 has been considered overthe total annual time. It affects the benefits, the operating costs andthe maintenance costs. As for other economic aspects: the capitalcost has been considered as 7.77%. The refund period and the

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considered redemption period have been taken as 10 years.Although this value is high for evaporative media, it is adjustedfor the other cooling equipments, and it has been taken with thesame time horizon for all them. A null salvage value and a yearlyforeseen CPI of 2.9% have been considered . The tax rate on the eco-nomic activity is 33%; it affects the difference between the rawbenefit and the depreciation of the installation.

While the costs associated with the fuel are the most importantoperating cost, there are other operating costs that have been keptin mind also; among them are the natural water cost, deionisedwater cost and the cost of parasitic electricity consumption inthe cooling equipment. The cost of natural water only has to beconsidered in wet media evaporative systems because, for theother systems, this consumption is null. For deionised water, nec-essary for the additional production of steam in the HRSG and foroperation of the fogging system, its price is not only considerable,but in addition, it is necessary to keep in mind the maximumcapacity of deionised water production in the plant because havingto install new process lines for deionised water excessively dimin-ishes the productiveness of the fogging system. In this case, thecompany has enough capacity to cover the demand of the foggingsystems considered in this work with the increase in steam pro-duction, so the price of deionised water is taken as about 1.52 €/Tm.

In this plant, the increase of steam production can become animportant part of the benefits (it reaches a value of 20% of the rea-lised profit for electricity). The cost of using steam will take thesame value as its selling cost. The price of the additional produc-tion of steam of high and medium pressure is given by Eqs. (7)and (8)

Psmed ð€=TmÞ ¼ 1:108Pf ð664� 15Þ0:86 � 0:9

þ A ð7Þ

Psh ð€=TmÞ ¼ 1:108Pf ð670� 15Þ0:86 � 0:9

þ A ð8Þ

where Pf is the price of fuel (€/kW h), A is the price of demineralisedwater (€/Tm).

Fig. 14. Equipment investment for the best econ

6. Results

As a result of this work, an extensive economic analysis was ob-tained for the cogeneration plant considered, showing the effectsof the different CTIAC systems in the power plant performance.In Figs. 14–17, the results obtained for the best configuration stud-ied for each CTIAC system are shown. In Fig. 14, the investments forthe best configurations of each CTIAC system are compared; inFig. 15, the annual benefits are shown for these configurations,and the internal rates of return (IIR) are shown in Fig. 16. Thenet present values (NPV) of the configurations are shown in Fig. 17.

The analysis of evaporative cooling systems showed that, as acertain number of stages in the fogging system is surpassed, theraw annual benefit begins to diminish as well as the NPV, and then,the increase of stages is not translated into an increase of benefits.In these cases, the cooling system would be unnecessarily large.The most interesting evaporation system for this power plant isthe fogging system with two stages.

The analysis of mechanically driven chillers showed that steamturbine driven chillers have a very low productiveness. This is dueto the appraised price of the steam in this installation being high,being more interesting in its direct sale than its use in a chillerto obtain extra additional power. With electrically driven chillersor gas engine driven chillers, we found the best results with refrig-eration capacities around 750 RT for the electrically driven chillerand 700 RT for the gas engine driven chiller. With higher refriger-ation capacities, although the percentage of cooling demand cov-ered was increased, the annual incomes were not increased inthe same manner, which indicates that most of the time, the CTIACsystem was over designed. There is a better solution with interme-diate refrigeration capacities due to two opposite effects; on onehand, as the cooling capacity of the chiller increases, a great partof this power is used in cooling saturated air, which gives a low ra-tio between sensible cooling and power consumption, and on theother hand, is the fact that not reaching the minimum temperaturelimit is a loss of additional electricity production and additionalsteam production.

omic configuration of every CTIAC system.

Fig. 15. Annual benefits for the best economic configuration of every CTIAC system.

Fig. 16. Internal rate of return for the best economic configuration of every CTIAC system.

Fig. 17. Net present value for the best economic configuration of every CTIAC system.

R. Chacartegui et al. / Energy Conversion and Management 49 (2008) 2130–2141 2139

Table 5Economic results obtained for the best configuration of each CTIAC system

System Investment(€)

Incomes(€/year)

Expenses(€//year)

Benefits(€/year)

Payback(years)

IIR (%) NPV (€) CDS (%)

Evaporative media 74,000 192,416 151,318 41,098 3 40.5 146,642 100Fogging of two stages 30,666 139,764 109,873 29,891 1.5 69 124,311 100Electric Chiller 850 RT 850,425 904,645 596,317 308,328 4.5 21.22 875,020 73.5Gas chiller 700 RT 943,250 996,634 704,838 291,796 5 16.42 722,526 64.1Steam chiller 250 RT 357,500 443,490 344,633 98,857 6 13.32 215,845 25.8Absorption chiller simple effect 300 RT 330,000 438,853 267,406 171,447 3 34.2 595,532 37.8Absorption chiller double effect 250 RT 352,000 420,562 322,764 97,798 5.5 13.45 214,809 32.2Absorption chiller simple effect and Electric

chiller 850 RT906,683 933,403 603,838 329,565 4.5 21.30 937,055 73.5

Electric chiller – gas chiller 575–275 RT 974,225 984,701 660,909 323,792 4.5 18.53 857,006 73.5Electric chiller – gas chiller 250–500 RT 951,591 994,084 677,191 316,893 4.5 18.59 840,186 67.4Electric chiller – gas chiller 425–425 RT 1,027,837 1,027,830 697,031 330,779 5 17.57 850,753 73.5Electric chiller 600 RT – TES 500 RT h 680,050 768,467 482,819 285,648 3.5 30.64 892,887 73.5

2140 R. Chacartegui et al. / Energy Conversion and Management 49 (2008) 2130–2141

With absorption chillers, the double effect absorption chiller(direct fired) was shown not to be appropriate for this application,and it cannot compete with the other chillers of similar capacities.On the other hand, the single effect absorption chiller is a veryinteresting solution, although it is desirable that its cooling capac-ity would be bigger, being limited in its maximum capacity by theavailable heat in the stack exhaust compatible with the steam gen-eration and thermal oil demand. This is due to its low parasiticload.

For hybrid systems, two different combinations were checked,the first one was composed of an electric motor driven chillerand a single effect absorption chiller, which works with hot waterproduced with the HRSG stack exhaust gases. Given the high pro-ductiveness of the single effect absorption chiller, due mainly to itsvery low operating costs, in this hybrid system setup, the single ef-fect absorption chiller will always work and the electric chiller willact when the demand requires it.

The other combination for hybrid systems checked was com-posed of an electric motor driven chiller and a gas engine drivenchiller, the operation modes of both are similar and simply varythe form of work. The main philosophy of this hybrid group wasto use the gas engine driven chiller whenever the price of electric-ity was high to diminish the parasitic load electricity consumptionand to maximize the electricity sale. From the results obtainedwith this configuration, the electric motor driven chiller alonegives better economic performance, and it is not worthwhileinstalling a hybrid system with this chillers combination. This isdue fundamentally to the investment needed with the gas enginedriven chiller, higher than that of an electric motor driven chiller,and the operating costs for natural gas, so the savings obtainedusing gas engine driven chillers, because of its very low parasiticload, is counteracted by the other two factors.

Finally, the results of CTIAC systems composed of thermal en-ergy storage combined with electric driven chillers were analyzed.The results obtained showed that the configurations with lowerstorage capacities were better than those with greater storagecapacities due to the fact that in the summery periods, the TES sys-tem does not provide an advantage because the demand is higherthan the chiller cooling capacity, while in some months, the de-mand is always below the cooling capacity of the chiller and,although the consumption is the same, the redistribution of para-sitic load at lower electricity tariff hours does not compensate forthe investment increase. Finally, in those periods in which an oscil-lation of demand occurs above and below the chiller cooling capac-ity, it was shown that with a 500 RT h storage capacity, it is enoughto cover acceptably the cooling demand peaks. In fact, when thedemand is satisfied with the TES system, to raise the storage sizeaffects only the hourly redistribution of the parasitic load, dimin-ishing it at the hours with higher electricity tariff.

In Table 5, the economic results obtained for the best configura-tion of each CTIAC system are shown. The values presented areinvestment, incomes, expenses, benefits, payback, IIR, NPV andpercentage of cooling demand satisfied (CDS) by the cooling sys-tem installed.

7. Conclusions

From this work the following conclusions are derived.The main features and costs of the different CTIAC systems have

been identified, and this has allowed identifying the most feasibletypes of CTIAC systems to be introduced in this cogenerationpower plant, ranging from very profitable systems with low initialinvestment, as the evaporative systems are, to others with higherinvestments but more profitable at the end of their capital refundperiod, as the hybrid systems with thermal storage are.

With the model of the plant implemented, as developed fromregistered performance data, it has been possible to eliminatesome uncertainty in the previous information concerning thepower plant equipment. This model has allowed obtaining the glo-bal variables of interest for the economic model as functions of thegas turbine intake conditions. The statistical treatment of the oper-ation registers of the power plant obtained in 2005, with the corre-sponding filtration of operative conditions that were outside thestudy conditions, has generated a precise model of the power plantfor easy coupling of the gas turbine intake with the exit from dif-ferent cooling systems.

From the economic analysis, it is shown that installing a CTIACsystem can be a very profitable solution, which also improves thegas turbine efficiency, saves fuel and reduces the CO2 and NOx

emissions. From the data shown in Table 5, it is seen that the threemost interesting intake cooling systems for this power plant are anelectrically driven chiller of 850 RT cooling capacity, a hybrid sys-tem with an absorption chiller of 300 RT cooling capacity disposedsequentially with an electrically driven chiller of 450 RT coolingcapacity and an electrically driven chiller of 600 RT cooling capac-ity combined with an Ice Harvester thermal energy storage systemwith 500 RT h of storage capacity.

Of the comparison between the electric motor driven chiller andthe electric motor driven chiller with TES system, it is concludedthat both systems present very similar results with the exceptionthat the TES system has a bigger NPV with a lower investment,and so, the system with storage would be a more interestingoption.

Taking the criteria of NPV as the main economic reference, theCTIAC system chosen would be the hybrid system with a value ofthe NPV 45.000 € greater than the NPV of the TES system, but witha worse IIR and payback period. However, other factors should also

R. Chacartegui et al. / Energy Conversion and Management 49 (2008) 2130–2141 2141

be contemplated, the main ones of them being the simplicity andinstallation time. Under these considerations, the system with ther-mal storage is clearly favourable relative to the hybrid system. Thejoint chiller plus TES is not a compact group, but it is simpler thanthe hybrid system with two chillers and will present a reasonablevolume for an industrial plant; for a tank of 500 RT h, its volumewould be about 35.5 m3. The installation time of the hybrid coolingsystem is bigger, not only because it is necessary to install two chill-ers and their interconnections and controls, but also because one ofthem is an absorption chiller working with hot water generated inthe final section of the HRSG. All the necessary actions in the HRSGclearly exceed the installation time of the TES system. As we arestudying the enhancement of an existing operating plant, the avail-able time for the cooling system installation will be scheduled tocoincide with the power plant planned overhauls. With the hybridcooling system, the risk of not finishing the installation works intime is higher. The cost of one day with the chemical plant stoppedis about 50,000 €, greater than the difference of final benefit be-tween the two options. Also, the control of the plant is easier forthe TES system than for the hybrid system, since the control be-comes independent of the operation in the HRSG. Under all theseconsiderations, the option of the electrically driven chiller withice storage would be the CTIAC system chosen for this power plant.

Appendix 1. Power plant model equations

Gas turbine

Power

W ðkWÞ ¼ ð42487:878� 236:51Ta � 0:256T2aÞF

WwabsFW

Hf FWPc ðA:1Þ

Power corrections

FWwabs ¼ 1:00096� 0:1519wabs ðA:2Þ

FWHf ¼

100þ ð10:254� 3:466 log Hf þ 0:0837 log H2f Þ

100ðA:3Þ

FWpc ¼ 1� 0:015D inchH2O

4ðA:4Þ

Heat rate

HR ðkJ=kW hÞ ¼ ð2731:796þ 2:859Ta

þ 0:064T2aÞ4:18FHR

wabsFHRHf FHR

Pc ðA:5Þ

Heat rate corrections

FHRwabs ¼ 0:9977þ 0:35443wabs ðA:6Þ

FHRHf ¼

100þ ð�4:50628þ 1:31713 log Hf þ 0:01431 log H2f Þ

100ðA:7Þ

FHRpc ¼ 1þ 0:005D inchH2O

4ðA:8Þ

Exhaust gas temperature

Tg ð�CÞ ¼ 0:722Ta þ 537:49� 5:72 ðA:9Þ

Exhaust mass flow

mg ðkg=sÞ ¼ �0:4489Ta þ 138:29� 1:48 ðA:10Þ

HRSG

Exhaust gas available power

Wdisp ðkWÞ ¼ �179:4Ta þ 87080 ðA:11Þ

Total heat delivery to cogeneration

Wcons:tot ðkWÞ ¼ 0:5993Wdisp þ 6268:5 ðA:12Þ

Heat delivery to medium pressure steam

Wvapmed ðkWÞ ¼ 0:2339Wdisp � 11503 ðA:13Þ

Heat delivery to high pressure steam

Wvaphigh ðkWÞ ¼ 0:3654Wdisp þ 17771:5�W th oil ðA:14Þ

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