Thermal energy storage system for energy conservation and water desalination in power plants

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Thermal energy storage system for energy conservation and waterdesalination in power plants

Venkataramana Gadhamshetty a, Veera Gnaneswar Gude b,*, Nagamany Nirmalakhandan c

a Environmental and Civil Engineering Department, Florida Gulf Coast University, United StatesbCivil and Environmental Engineering Department, Mississippi State University, Mississippi, United StatescCivil Engineering Department, New Mexico State University, United States

a r t i c l e i n f o

Article history:Received 23 July 2013Received in revised form6 December 2013Accepted 12 January 2014Available online 16 February 2014

Keywords:Thermal energy storageDesalinationDry coolingWater conservationAir-cooled condensersPower plants

a b s t r a c t

This paper presents a novel application of a sensible TES (thermal energy storage) system for simulta-neous energy conservation and water desalination in power plants. First, the TES mitigates negativeeffects of high ambient temperatures on the performance of ACC (air-cooled condenser) that cools a500 MW CCPP (combined cycle power plant); next, the same TES satisfies the cooling requirements in a0.25 mgd (950 m3/d) MED (multi-effect evaporation desalination) plant. Stack gases from CCPP are usedto drive an ARS (absorption refrigeration system) which maintains the chilled water temperature in a TEStank. A process model integrating CCPP, ARS, TES, and MED has been developed to optimize the volumeof the TES. Preliminary analysis showed that a tank volume of 2950 m3 was adequate in meeting thecooling requirements of ACC and MED in both hot and cold seasons. The proposed TES has a potential tosave 2.5% of the power loss in a CCPP on a hot summer day. Further, a desalination capacity of 0.25e0.43 mgd (950e1600 m3/d) can be achieved with top brine temperatures between 100 �C and 70 �C ofMED.

� 2014 Elsevier Ltd. All rights reserved.

1. Introduction

Freshwater and energy are two essential commodities for sus-taining human life on earth [1]. The ever-increasing demands forenergy and water, and the inseparable connection between waterand energy production have been well-recognized in the last fewdecades [2,3]. For instance, high-quality energy is required to treatand transport drinking water and wastewater in both urban andrural areas. The United States of America spends nearly 4480 MW todrive 21,594 WWTPs (wastewater treatment plants) on a daily basis[4]. The energy consumption in drinking water sector is also signif-icant [5]. At the same time, the refineries and power plants consumevoluminous amount of clean water for extracting oil and producingelectricity. Hence, it is important to coordinate the nexus betweenenergy and water sectors to conserve energy, reduce water con-sumption, and minimize the negative environmental impacts [5].

Steam power plants reject a major percentage of the inputenergy as waste heat that is directly discharged to atmosphere.

This heat-rejection is accomplished with wet-cooling towers thatdemand voluminous amount of freshwater. For example, a 500 MWwet-cooled steam power plant consumesw4million gallons per dayof freshwater. The steam-electric plants accounts for nearly 40% ofthe total freshwaterwithdrawals in the United States [6]. Dry-coolingtechnology is an attractive alternative towet-cooling technology thatallows power plants to reject the heat to air, directly or indirectly,without any loss of water [6]. The eminent advantages of dry-cooledpower plants include: i) minimal freshwater consumption [7]; ii)flexibility in plant siting that enables power plants to be locatedcloser to load centers rather than to cooling water resources [8]; iii)minimized risk for Legionella health risks [9]; and iv) reduced rates ofplume formation and brine disposal [10,11]. The major disadvantageof dry-cooling technology is that they use ACCs (air-cooled con-densers) whose performance declines with increasing ambient airtemperatures, and thereby induces cycle penalties in the associatedpower plants [12]. The effectiveness of ACCs has been reported toreduce by at least 10% due to high ambient temperatures in summer[6].

The waste heat from the power plants has been found as asuitable source of energy in thermal desalination processes [13,14].It is feasible to engineer high-rate desalination systems by incor-porating desalination systems in power plants. Such dual-purpose

* Corresponding author.E-mail addresses: [email protected] (V. Gadhamshetty), gudevg@gmail.

com, [email protected] (V.G. Gude).

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Energy 66 (2014) 938e949


power plants have been reported to deliver substantial cost savingsduring its entire life-cycle. For instance, the MSF (multi-stage flash)desalination integrated in a steam power plant was able to reducefuel consumption by 37%; further, such plants were able to reducewater consumption by 45% compared to awater-onlyMSF unit [15].A detailed LCA (life-cycle assessment) study revealed that a dual-purpose power plants reduces environmental impact by 75%compared to a thermal desalination technology [16].

TES (thermal energy storage) systems offer a viable and feasibleapproach to store thermal energy when the supply is more than thedemand, and to release it when the demand is more than thesupply [17]. For example TES systems can be designed to store solarheat energy during sunlight hours of the day and release it for spaceheating during colder periods of the day, and similarly, for fresh-water production in desalination systems powered bywaste heat orsolar energy [18,19]. The viability of TES systems has beendemonstrated in large-scale applications including HVAC (heating,ventilating, and air conditioning) systems, gas turbines[20,21,22,23], district cooling systems [24], and district heatingsystems [25].

This paper presents a novel application of sensible TES (thermalenergy storage) to integrate MED (multi-effect evaporation desa-lination) system in a CCPP (combined cycle power plant) forsimultaneous energy conservation and water desalination. Ourobjectives in this study are two-fold. First, wewill demonstrate thatthe TESmitigates the negative effects of high ambient temperatureson the performance of ACC (air-cooled condenser); next, we willdemonstrate a new method for obtaining 0.25 mgd (950 m3/d)MED (multi-effect evaporation desalination) using waste heat inCCPP and the chilled water from TES.

2. Proposed TES for integrating MED in power plants

Fig. 1 shows a conceptual schematic of the TES (thermal energystorage) that integrates CCPP (combined cycle power plant) withARS (absorption refrigeration system) and a MED (multi-effectevaporation desalination) system. A TES tank is used to maintainthe cooling efficiency of ACCs during hot days, and enable CCPPoperation at rated power, instead of suffering power loss due toincreased steam turbine back pressure. The waste heat in the CCPPis used to drive an ARS (absorption refrigeration system) that coolschilled water in a TES tank. The chilled water is then used to precoolthe inlet air to the air-cooled condenser and improve the perfor-mance of the CCPP during hot summer. The TES meets the coolingneeds of a 0.25 MGD MED (multi-effect evaporation desalination)plant. Further, the MED is solely driven by the waste heat availablein stack gases. The state points for CCPP, ACC (air-cooled

condenser), ARS, TES, and MED are shown in Fig. 2. The specifictechnical details for MED are shown in Fig. 3.

2.1. CCPP (combined cycle power plant)

The CCPP is comprised of a COMP (compressor), CC (combustionchamber), GT (gas turbine), and HRSG (heat recovery steamgenerator) (Brayton cycle in Fig. 2). The heat discharged from HRSGis used to drive a steam power plant consisting of a ST (steamturbine), OFWH (open feed water heater) and the ACC (RankineCycle in Fig. 2). Atmospheric air is drawn into the compressor (8)and fed to the combustion chamber operating at constant pressure.The high temperature gases enter the gas turbine (10) that pro-ducing net-work, WGT. The exhaust gases from the gas turbine (11)are used to produce superheated steam in HRSG (5). The steamturbine produces net-work, WST. A fraction of the steam turbineexhaust (6) is diverted to the OFWH, where it is mixed with feedwater (2) pumped from the ACC. The remaining fraction of thesteam turbine exhaust (7) is returned to the ACC to obtain feedwater for reuse (1). Saturated liquid (3) at heater pressure ispumped by pump 2, to HRSG to generate superheated steam (5) tocomplete the cycle.

2.2. ARS (absorption refrigeration system)

The ARS (absorption refrigeration system) is based on lithiumbromide e water medium operating within a pressure range of0.87 kPae8.65 kPa. The design details of the ARS have been re-ported in our earlier study [6]. In essence, the GEN (generator) inARS system is powered by the waste heat extracted from the stackgases (12e13); the heat input to the EVAP (evaporator) is extractedfrom the TES (Fig. 2). The heat rejected from the ABS (absorber) wasestimated as 62.9 MW while that in the COND (condenser) was49.4 MW. The rate of heat input from TES to evaporator is estimatedas 46.2 MW.

2.3. TES (thermal energy storage)

The TES is designed to meet the cooling requirements of boththe ACC and MED systems. The working fluid in the TES system iswater that is maintained at 5 �C by the evaporator of the ARS. Wechose water as working fluid as it provides high energy capacity,transfers heat to air at higher efficiency, and offers safety in long-term operations [26]. A CW (chilled water) pump circulates waterfrom the TES through two cooling coils, one installed at the air inletto the ACC (Fig. 2), and another, in the final condenser in the MEDsystem (Fig. 3). The CW pump supplies cooling water to ACC systemonly when the ambient temperature exceeds the ACC design airinlet temperature, T*. The cooling water runs through the finalcondenser of MED system to remove the latent heat from the vaporgenerated in the last effect. The flow rate of the chilled water iscontrolled to maintain inlet air temperature of ACC at the designvalue of T*, and temperature of COND (Condenser) in MED at Tcw.The net external energy in the proposed system are due to the: i)pump in the ARS to transfer the dilute solution from the absorber tothe generator, running 24 h/day; ii) pump to circulate the chilledwater from the TES through the cooling coils (running 6e8 h/day);iii) fans for the air-cooled condenser and absorber of the ARS; andiv) pump to circulate the cooling water from the TES to MED finalcondenser.

2.4. MED (multi-effect evaporation desalination) system

The MED (multi-effect evaporation desalination) systems areknown for high thermal energy efficiency, lower capital costs, and

Fig. 1. TES integrated dual-purpose power plant. Note 1: TES, thermal energy storage;ARS, absorption refrigeration system; MED, multi-effect evaporation desalinationsystem; Note 2: The waste heat in power plant is used in ARS and MED; Note 3: ARScools the TES; Note 4: TES meets cooling requirements in dry-cooling tower and MED.

V. Gadhamshetty et al. / Energy 66 (2014) 938e949 939


reduced operational costs. The steam in the first effect of MED issupplied by the waste heat extracted from the stack gases (12e13)in CCPP (Fig. 2). The heat from condenser of MED is rejected tocooling water stream from the TES (Fig. 2). The stack gases in theCCPP are available in the temperature range of 120e150 �C (Fig. 2),and we therefore considered a MED unit at a top heat source(steam) temperature of 100 �C. The MED operating at higher tem-peratures (about 100 �C) offer following pre-eminent advantagesover MSF [28]: 1) high thermal efficiency with a lower number ofeffects, 2) high heat transfer coefficient, 3) relatively low specificinvestment cost, 4) low pumping power, and 5) operating flexibilitywith short start-up period and 6) demonstrated capability formatching production volume with water demand and energysupply.

3. Modeling studies

3.1. TES (thermal energy storage) system

The proposed TES meets the cooling requirements in i) ACCs(air-cooled condensers) of CCPP and ii) MED (multi-effect evapo-ration desalination) system. The volume of the TES system can beestimated from a heat balance around the TES system described inthe following equation:

VsrwCpwd½Ts�dt

¼ � Q�:evap þ V

�aCparaðTamb � T�Þ

þ V�fcCpwrwðTcw � TsÞ þ KðTamb � TsÞ

(1)

Fig. 3. Multi-effect evaporation desalination system driven by waste heat from stack gases of CCPP. Note 1: The TES in Fig. 2 meets the cooling requirements in MED.

Fig. 2. Schematic of proposed TES (thermal energy storage) for energy conservation and water desalination in CCPP (combined cycle power plants). Note 1: The CCPP is based ondry-cooling technology (i.e. Air Cooled Condensers); Note 2: The waste heat from CCPP is used to drive ARS (absorption refrigeration system) and MED (multi-effect evaporationdesalination) system; Note 3: The chilled water in TES meets cooling requirements in ACC and MED.

V. Gadhamshetty et al. / Energy 66 (2014) 938e949940


where _Qevap ¼ heat required by the evaporator of ARS(kW);Ts ¼ TES temperature in �C; Tcw ¼ temperature of coolingwater from the final condenser (in desalination) in �C. Other no-tations for the variables in Eq (1) are described in nomenclature.

V�w > 0 only if Tamb > T*, otherwise V

�w ¼ 0.

The first term on the right hand side of the Eq (1) represents theheat removed by the ARS; second term is a heat input to TES, theheat removed from the inlet air to ACC; third term is another heatinput to the TES, the heat rejected by the condenser in MED (multi-effect evaporation desalination). The last term in Eq (1) representsthe heat gained from the surroundings. The ambient temperature(Tamb) is a time-dependent that is largely influenced by the localmeteorological conditions.


The specific number of stages (n) in MED depends on the topbrine temperature and the allowable temperature drop betweenthe stages [27]. Fig. 3 provides the state points for various com-ponents in the proposed MED. The heat source in the first effectpromotes the evaporation of incoming preheated saline water(feed). The temperature of the vapor in the first effect (Tv1) is lessthan the boiling temperature (T1) by the magnitude of the boilingpoint elevation (BPE1). The vapor in the first effect is transported tothe second effect through a demister. The latent heat of conden-sation (D1) is exploited for further evaporation in the second effect.The unevaporated brine in the first effect (Mf�D1) flows throughthe second effect that operates at a lower pressure. The vaporformed in the second effect can be attributed to the two mecha-nisms. First, the vapor is generated by boiling the liquid over theheat transfer surfaces D2. Second mechanism can be attributed toflashing, i.e. free boiling, within the bulk liquid (d2). The tempera-ture of the vapor due to flashing (Tv2) differs from the boilingtemperature (T2) by the magnitude of boiling point elevation(BPE2). Small quantity of vapor (d2) condenses in the flashing boxdue to the flashing of distillate condensed in the second effect (D1).The processes occurring in the second effect are repeated in eachsubsequent effect all the way down to the last [28].

Fig. 3 shows that the amount of steam generated (by evapora-tion) in each effect is less than that in the previous effect. Thisphenomenon is due to the higher latent heat of vaporization cor-responding to the temperature in the lower effect. Consequently,the amount of vapor in an evaporator (by boiling) is less than theamount of condensing steam used for heating. The unevaporatedbrine flows through the subsequent effects until the last effect toreach a final concentration of Xn. The remaining brine (Mb) isrejected to the source from the last effect. The vapor formed byboiling and flashing in the last effect (Dn þ dn) and the end flashingbox (dn) passes to the final condenser.

3.3. Case study

The primary objective of this research is to size the TES systemthat restores the performance of ACC (air-cooled condenser) andmaintains the power output of CCPP during hotter days of the year.The second objective is to design a MED system that operates solelyon the waste heat energy from power plants, and the cooling loadprovided by TES system. The utility of TES is illustrated for a hy-pothetical 500 MW CCPP to be located in Las Vegas, NV (McCarranStation). This site represents temperature profile of arid regions insouthwestern United States. It is assumed that this geographicalsite is capable of providing 0.25 million gallons of brackish water tooperate a MED (multi-effect evaporation desalination) system on a

daily basis. The CCPP and ARS were analyzed with EngineeringEquation Solver� (F-Chart Software). MATLAB software was used toobtain numerical solution to the TES system modeled with a non-linear differential equation in Eq. (1).

3.3.1. Analysis of CCPPThe operating details of a 500 MW CCPP were described in our

earlier publication [6]. The rate of heat input to CCPP is 1023 MW.The net output of the gas turbine is 320.1 MW while the steamturbine is rated at 170.9 MW. The rate of heat rejected by thecondenser is 316.4 MW. The waste heat in the stack gases isavailable at a rate of 82.6 MW, and 86% of the waste heat isextracted from the stack gases (¼ 71 MW), out of which 66 MW istransferred to the generator of the ARS and another 4.96 MW isused in MED desalination. The standard design procedures wereused to rate individual components of combined cycle power plants(Eqs. (2)e(22)). The notations for the model parameters aredescribed in Table 1.

3.3.1.1. Design equations for combined cycle power plants.

wg comp ¼ h9 � h8 (2)

qin ¼ h10 � h9 (3)

wg turb ¼ h11 � h10 (4)

wpump1 ¼ h2 � h1 (5)

wpump2 ¼ h4 � h3 (6)

wsteam pumps ¼ ð1� yÞ �wpump1 þwpump2 (7)

wsteam turb ¼ h5 � y� h6 � ð1� yÞ � h7 (8)

qout ¼ ð1� yÞðh7 � h1Þ (9)

W�net ¼ m

�gas �

�wg turb �wg comp

�þm

�steam

� �wsteam turb �wsteam pumps

� (10)

Table 1Model parameters for CCPP and ARS.

Parameter Value Parameter Value

Model parameters for combined cycle power plants (CCPP)Net power output 500 MW Pressure ratio of gas

turbine cycle14

Isentropic efficiency ofcompressor

82% Isentropic efficiencyof gas turbine

86%

Gas turbine inlet temperature 1400 K Isentropic efficiencyof pump

95%

Steam turbine inlet temperature 823 K Steam turbine inletpressure

8000 kPa

Isentropic efficiency of steamturbine

86% Air inlet temperatureof compressor

300 K

Air inlet temperature ofcompressor

300 K Condenser pressure 20 kPa

Open feed water heater pressure 600 kPa CCPP StackTemperature

460 KHeat exchanger gas exit

temperature395 K

Model parameters for absorption refrigeration system (ARS)Evaporator temperature 278 K Absorber

temperature311 K

Condenser temperature 313 K Generatortemperature

363 KHeat exchanger efficiency 70%



hth ¼ W�net

Q�in

(11)

W�net steam ¼ m

�steam � �

wsteam turb �wsteam pumps�

(12)

W�net gas ¼ m

�gas �

�wg turb �wg comp

�(13)

Q�gen ¼ 0:86�m

�gas � ðh12 � h13Þ (14)

3.3.1.2. Design equations for absorption refrigeration system.

Q�c ¼ m

�2’ � ðh2’ � h4’Þ (15)

Q�a ¼

�h5’ �m

�5’

�þ�h9’ �m

�9’

��h1’ �m

�1’

�(16)

Q�gen ¼

�h2’ �m

�2’

�þ�h3’ �m

�3’

��h6’ �m

�6’

�(17)

Q�evap ¼ m

�2’ � ðh5’ � h4’Þ (18)

Pwork ¼ m�1’ � v1’ � ðHP� LPÞ (19)

hhx ¼ ðT3’ � T8’ Þ=ðT3’ � T7’Þ (20)

Q�hx ¼ m

�3’ � ðh3’ � h8’Þ (21)

COP ¼ Q�evap

Q�gen

(22)

3.3.1.3. Design equations for air cooled condensers.

Q�out ¼ V

�a � Cpa � ra � ðT� � ToutÞ (23)

3.3.2. Analysis of MED (multi-effect evaporation desalination)system

The multi-effect evaporation was simulated with heat sourcetemperatures ranging from 70 �C to 100 �C. The heat source isavailable at temperatures (steam from stack gases) suitable for theoperation of a high temperature MED at a rate of 5 MW. We havestudied the MED configuration by evaluating the effect of heatsource temperature on the heat transfer areas, condenser areas, andthe desalination capacity. Towards this end, the heating and coolingrequirements were estimated under a range of brine temperaturesand for different number of stages. Temperature differentialsranging from 2.5 �C to 10 �C were used to estimate the coolingwater flow rate requirements for MED operation, at differentnumber of stages and evaporation temperatures. The desalinationcapacity for a fixed amount of energy extracted from the stack gaseswas evaluated at low and high temperature (70 �C to 100 �C)operation of MED system. The design calculations were performedusing Eq. (24) and (37) [29]. The relevant design parameters andprofiles for the MED are compiled in Table 2 and Table 3.

Temperature drop across all effects is obtained from thefollowing relation:

DTt ¼ Ts � ðn� 1ÞDTl � Tbn(24)

Temperature drop in first effect is obtained by

DT1 ¼ DT1U1

Pni¼1

1Ui

(25)

Similarly, the temperature drop in effects 2 to n is obtained by

DTi ¼ DT1U1

Ui(26)

Brine temperature in first effect is obtained from the relation

Tb1¼ Ts � DT1 (27)

Brine temperature in effects 2 to n

Tbi¼ Tbi�1 � DT1

U1

Ui� DTl (28)

Distillate flow rate in the first is given by

D1 ¼ Md

lv1

�1lv1

þ 1lv2

þ ::::::þ 1lvn�1

þ 1lvn

� (29)

Distillate flow rate in effects 2 to n

Di ¼ D1lv1

lvi

(30)

Brine flow rate in effects 1 to n

Bi ¼XcwDi

ðXbi � XcwÞ (31)

Feed flow rate in effects 1 to n

Fi ¼ Di þ Bi (32)

Heat transfer area in the first effect

A1 ¼ D1lv1U1

�Ts � Tb1

� (33)

Heat transfer area in effects 2 to n

Ai ¼Dilvi

Ui�Tvi�1 � Tbi

� (34)

Heat steam flow rate, Ms

Table 2Design parameters of the MED system.

Parameter Value

ne 6e10Md (m3/d) 3785Xf (ppm) 35,000DTev (�C) 2Ts (�C) 70e95Tbn (�C) >30Tf (�C) 30Tcw (�C) 5Cp (kJ/kg) 4.2Uc (kW/m2-K) 2dt (diameter of tube) 0.03Lt (length of tube) 10



Ms ¼ D1lv1ls

(35)

Heat transfer area of the condenser

Ac ¼ DnlvnUcðLMTDÞc

(36)

Flow rate of cooling water (Mcw)

Dnlvn ¼ ðMcwÞCp�Tf � Tcw

�(37)

3.3.3. Analysis of TESEq 1 was solved numerically using the heat transfer value

determined from the ARS analysis, Tamb obtained from hourlyweather records, and cooling requirements from MED. Eq (1) wassolved by trial-and-error to obtain the optimal volume of the TESsystem, while ensuring that the TES temperature returned to itsinitial value at the end of each 365-day cycle. This analysis yieldedthe profiles of the TES temperature and the ACC inlet air temper-ature after pre-cooling. The value of T* was set at 20 �C.

4. Results and discussion

The TES tank volume was determined as 2950 m3. Simulationresults obtained from themodel were used to develop performancecurves to aid in preliminary evaluations and conceptual designs forACC and MED.

4.1. Ambient temperature profiles

Fig. 4 shows the typical temperature profile of Mccarran Stationin Nevada. As shown, the typical ambient temperature rangesfrom �10 to 40 �C, with summer temperatures reaching as high as39 �C. In general, the high ambient temperature (>T*) reduces thetheoretical minimum attainable temperature of air-cooled con-densers (in CCPP), and thereby reduces cooling efficiency of ACC,and ultimately incur cycle penalties in a steam turbine of CCPP.From the frequency profiles shown in Figs. 4 and 5, the perfor-mance of the ACC deteriorates for nearly w50% of the total oper-ating hours throughout the year, especially, when the ambient airtemperature (Ti) is higher than that of air inlet design temperature(T* ¼ 20 �C) (Fig. 6a).

4.2. Inlet air temperature profiles of ACC with proposed TES

The benefit of the TES system in maintaining the steam turbineoutput at various ambient temperatures is illustrated in Fig. 5. Asshown in Fig. 5, the performance of CCPP decreases with increasingambient temperatures due to reduced cooling capacity of ACC

systems. For instance, the power output of CCPP drops by 4.6MWatdry bulb temperature of 32 �C, and nearly 10MWat 35 �C. The cyclepenalties in CCPP can be mitigated by maintaining Ti closer to T* onhotter days of the year. Fig. 6b shows the temporal profiles of ACCinlet temperature, and Fig. 6c depicts the TES tank temperature. Theresults in Fig. 6 demonstrate the utility of TES system in pre-coolingthe ambient air and maintaining its temperature at T*. The mostimportant finding is that the chilled water temperature in TESreturns to its original value at end of annual cycle. Therefore, theproposed TES system can be used tomaintain cooling requirementsin both ACC and MED throughout the year.

4.3. Power penalty imposed by ACCs in CCPP during hotter days

The current practice of turbine back pressure control results inexplicit loss of power in the steam power cycle of CCPP due tonegative effects of high ambient temperatures on the performanceof ACCs. Fig. 5 demonstrates the utility of TES in improving thecooling efficiency of ACC, and reducing the cycle penalties of CCPPduring high ambient temperatures.

Fig. 5 shows the simulated power losses at ambient tempera-tures ranging from 5 �C to 40 �C for the two systems. Withincreasing ambient temperatures, the loss of power with backpressure control is significantly higher than the loss associatedwiththe proposed system. For example, at a dry-bulb temperature of32 �C, the efficiency of cooling is restored to such an extent that thepower loss of 4.2 MW is eliminated. The minimum power losses inCCPP with proposed TES/ACC system is due to chilled water pumpand the refrigerant transfer pump (Curve 2, Fig. 5). In real appli-cations, the energy consumption for pumps (i.e. chilled waterpump, refrigerant pump) may be significantly higher due toinherent frictional losses in the fluid systems. Curve 3 in Fig. 5

Table 3Profiles of the MED system (heat source temperature ¼ 100 �C and n ¼ 10).

Effect 1 2 3 4 5 6 7 8 9 10

U (kW/m2-K) 2.75 2.67 2.59 2.51 2.43 2.36 2.29 2.22 2.16 2.09DT (⁰C) 4.52 4.66 4.80 4.95 5.10 5.26 5.42 5.59 5.76 5.94Tb (⁰C) 95.5 88.8 82.0 75.1 68.0 60.7 53.3 45.7 37.9 30.0Tv (⁰C) 93.5 86.8 80.0 73.1 66.0 58.7 51.3 43.7 35.9 28.0lv (kJ/kg) 2280 2296 2312 2329 2346 2363 2381 2399 2417 2436D (kg/s) 1.2 1.1 1.1 1.1 1.1 1.1 1.1 1.1 1.1 1.1A (m2) 212.6 212.6 212.6 212.6 212.6 212.6 212.6 212.6 212.6 212.6B (kg/s) 31.9 30.8 29.6 28.5 27.4 26.3 25.2 24.0 23.0 21.9F (kg/s) 33.1 31.9 30.8 29.6 28.5 27.4 26.3 25.2 24.0 23.0Xb (ppm) 36,172 37,519 38,961 40,507 42,168 43,958 45,892 47,988 50,267 52,752

-10 0 10 20 30 40 50 600

100

200

300

400

Ambient dry bulb temperature ( OC)

Tim

e (h

ours

/ann

um)

Fig. 4. Frequency profile of ambient air dry-bulb temperature.



shows the effect of frictional losses (assuming w100% of base en-ergy consumption) on the net power output with proposed ACC. Itis clear that the proposed ACC can achieve significant power sav-ings in real time applications for typical temperature rangesencountered on a hot summer day (Curve 5, Fig. 5).


4.4.1. Effect of heat source temperatureFig. 7A shows the thermal energy requirements of MED as the

function of heat-source temperature (70 �C to 100 �C) and thenumber-of-stages (n ¼ 6e10). Fig. 7A shows that the energy re-quirements for desalination increases with increasing heat sourcetemperatures. Further, lowest thermal energy requirements areobserved at operating conditions imposed by low heat sourcetemperature and highest number of stages (i.e. 70 �C; n ¼ 10). Theenergy requirements in MED (w2.87e4.85 MW) can be obtainedfrom the stack gases of CCPP (Fig. 2).

Fig. 7B shows the top-brine temperatures as function of heatingsteam temperatures and the number-of-stages (n¼ 6e10). The top-brine temperature for 10 stages is 95.5 �C and 92 �C for 6 stages. Thesame are 68.1 �C and 66.6 �C for corresponding stages of 10 and 6.The lowest top-brine temperature value corresponds to 10 stageswith a top-brine temperature of 68.1 �C and the highest value is for6 stages with a top-brine temperature of 92 �C. For low number-of-stages, the temperature difference is higher and the top-brinetemperature in the first stage is lower. This is reflected in theheat transfer areas required for evaporation (Fig. 7C). The heattransfer areas for higher top-brine temperatures are smallercompared to that for low top-brine temperatures. The specific heattransfer areas at 100 �C heat source temperature are 42% and 39% ofthat required at 70 �C when the number of effects are 6 and 10.These ratios resemble the observations reported by El-Dessoukyet al. [28]. This suggests that high temperature operation of MEDresults in lower heat transfer areas, and lower capital costs, how-ever, at the expense of higher energy costs for the heat source.

4.4.2. Effect of top brine temperature and performance ratioThe PR (performance ratio) of MED is defined as the kg of

distillate produced by 2300 KJ of heat input [30]. Fig. 7D establishesthe relationship between the PR and the number of stages. Ingeneral, the PR decreases with increasing top-brine temperature.

For example, when the number of stages (n) is 10, the PR is 9.6 and9.8 while the top-brine temperatures were 95.5 �C and 68.1 �C.Similarly, for n ¼ 6e9, the PR varies between 5.8 and 5.9; 6.7e6.9;7.7e7.8; and 8.7e8.8 for heat source temperatures between 70 �Cand 100 �C, high end PR representing low temperature operation.However, it can be noted that this difference is minimal andwill notadversely affect the performance of the desalination plant [31].

4.4.3. Number of stages vs. condenser heat transfer areaFig. 8A shows the relationship between the cooling water flow

rate (kg/s) and the number of stages (n). The requirements forcooling water decreases with increasing values of ‘n’. This is due tolower evaporating temperatures in the final condenser with highervalues of ‘n’. Fig. 8A shows the cooling water requirements (kg/s) asa function of available temperature differential (DT), i.e., 2.5e10 �C.The cooling water requirements correspond to the return coolingwater temperatures. Assuming the cooling water inlet temperatureto be 5 �C, the flow rate increases when DT ranges from 2.5 to 10 �C(Fig. 8A). At higher values of DT, the cooling water flow rates aresignificantly lower and vice versa. Fig. 8B shows the condenser heattransfer areas required for ‘n’ ranging from 6 to 10. Again, with theavailability of temperature differential between the cooling waterstream and condenser, the heat transfer areas changes with varyingvalues of ‘n’. The higher the DT, the smaller the condenser heattransfer area and vice versa. Fig. 8C shows the relationship betweenthe brine temperatures in the last effect and the cooling water flowrates (kg/s). As expected, lower the value of ‘n’, higher the evapo-rating temperatures in the last effect, and higher the cooling waterflow rates. As shown in Fig. 8C, they change with the availabletemperature differentials. Cooling water temperature differentialsbetween 2.5 and 10 �C are required to extract the latent heat fromthe final stage condenser with temperatures between 28 and 36� C.Since TES unit chilled water temperatures vary between 5 and20 �C, a constant cooling water stream can be provided for all finalcondenser temperatures. Fig. 8D shows the brine outlet tempera-tures for different values of ‘n’. Lower number of effects results inhigher brine discharge temperatures, which in practice, areexchanged with the incoming seawater or feed water. This can bean issue in the Gulf and Mediterranean coastal areas since theincoming seawater temperatures can be as high as 35 �C. Coolingwater stream from TES enables MED operation with lower finalcondenser temperatures and lower brine discharge temperatures.

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Fig. 5. Enhanced performance of CCPP due to ACC modified with TES system.



4.4.4. Outlook e TES cooling system in desalination plantsFig. 9 shows the relationship between the number of stages and

the additional energy required for high-temperature operation forthe same desalination capacity of 0.25 mgd. The lowest energy(heat) requirements of 2.87 MW were estimated for low temper-ature operation at 70 �C and n ¼ 10. The heat energy requirementscontinue to increase by 4% for 100 �C and n¼ 10; 15% for 100 �C andn ¼ 9; 30% for 100 �C and n ¼ 8; 48% for n ¼ 7; and 73% for n ¼ 6.Similarly, the desalination capacity for low temperature MEDoperation (70 �C and n¼ 10) increases for the corresponding energyextracted from the stack gases of the power plant by 4% (0.26mgd);15% (0.275 mgd); 30% (0.325 mgd); 48% (0.37 mgd); and 73%(0.43 mgd). This suggests that low-temperature operation of MEDresults in an increased desalination capacity (with two trains in thiscase) albeit with increased requirements for heat transfer areas.While, high-temperature MED is better than MSF process, low-temperature MED has the potential to offer higher desalinationcapacities with lower operation/maintenance and energy costs. Inaddition, low-temperature MEDs facilitate the use of low-cost heat

transfer materials which is not possible with high-temperatureMED or MSF.

In the existing desalination plants, the majority of the coolingwater is discharged into the oceans since only about 8e30%freshwater is recovered in thermal desalination [29,32]. MSF(multi-stage flash) releases brines 10e15 �C warmer than oceanicintake waters leading to ecological and environmental imbalancesin coastal water systems [32,33]. The cooling water stream from theproposed TES will mitigate the ecological and environmental issuesin coastal seawaters.

The corrosion of underlying materials is another major problemwith high-temperature desalination plants. The cost of materials inthe MEDs represents 25% of the total plant costs. It is thereforeimportant to consider low-cost materials without reducing theoverall heat transfer coefficients [27]. Low-temperature MEDoperation provides superior thermodynamic efficiency, and lowpressure drops at high volumetric vapor flows, as prevailing in lowtemperature operation at a maximum brine temperature of 70 �C.Higher heat transfer areas in low-temperature operation are oftenviewed as negative parameter; however, they facilitate lowtemperature-drop per effect (1.5e2.5 �C), enabling incorporation ofa large number of effects (10e16) even with a maximum brinetemperature as low as 70 �C, and consequently results in higheconomy ratios (product to steam) [34]. Overall, the low-temperature MED operation i) enables the use of low-cost heattransfer materials, ii) reduces the scaling rates as the operatingtemperatures are maintained below the saturation limits of prob-lematic scalants found in sea and most ground waters [35,36] andiii) facilitates the use of low-grade waste heat from flue gases[37,38].

4.5. Sensitivity analysis

A major concern with the proposed TES system is the largevolume requirements of the storage tank. The volume of the TESincreases with an increase in net plant output for any value of T*and the volume of the TES decreases with an increase in T* at agiven CCPP output [6]. Further, the TES volume is influenced by thecooling requirements in MED system. The TES volume may bereduced by thermal insulation and enhancing desired heat transferby adjusting such parameters as fluid flow rates and heat-exchanger materials [6]. Sensitivity analysis was performed toevaluate the sensitivity of storage temperature as a function of TESvolume in the range of 2534e4492 m3. As demonstrated in Fig. 10,the storage temperature, Ts, increases with decreasing TES volume.It can also be noted that the sensitivity of Ts would depend on thevariations in ambient temperature. Therefore, ambient air tem-perature variation patterns play an important role in a preliminaryanalysis of the TES. Sensitivity analysis demonstrated a strongfunctionality of storage temperature on TES volume. Volume re-quirements of the TES can be significant which warrant preliminaryevaluation and site specific economic analysis for determining theoptimum TES volume.

The availability of the waste heat determines the size of the ARS(absorption refrigeration system), the TES (thermal energy storage),both of which influences the performance of the ACC (air-cooledcondenser) and desalination system. In this study, the heat avail-able in the generator (Qg) depends on the percentage of heat ex-change (hse) between the stack gases and the generator of the ARS.Fig. 11a shows that the lower values of hse tend to reduce the ca-pacity of ARS (Qgen and Qevap) and increase the size of TES tank.Future studies are warranted to investigate the effect of in-efficiencies in individual components (e.g. open feed water heater)in combined cycle power plants. Further, Fig. 11b demonstrates theinfluence of hse on the time required to restore tank temperature

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Fig. 6. Temperature profile of ambient air, ACC inlet, and TES.



Fig. 8. A) Cooling water flow rates vs. number of stages; B) condenser heat transfer area vs. number of stages; C) cooling water flow rates vs. brine temperature in last effect; D)brine outlet temperature vs. number of stages.

Fig. 7. A) Heat requirements vs. source temperature; B) top brine temperature vs. source temperature; C) evaporator heat transfer area vs. source temperature; D) Performance ratiovs. number of stages.



(Ts) after each annual cycle. Fig. 11b shows that in cases of reducedhse, the TES performance can be recovered by increasing the size ofTES tank. The analysis presented in this study provides a usefulframework to assess the practical feasibility of TES in enhancingperformance of dry cooling and desalinating the water.

4.6. Economic analysis

Based on the estimates from Nevada Power Company, theelectricity costs (per kWh) for different seasons of the year are asfollows: Summer onepeak, $ 0.08017; Summer midepeak, $0.06252; Summer offepeak, $ 0.04502; and winter, $ 0.05095. Theaverage power cost for state of NV can be estimated as $0.06/kwh[39]. Therefore, in this study, the annual savings can be estimated as$91,980 due to the power savings of 1533 MWH (Fig. 5). If theannual operations & maintenance costs constitute 10% of the totalrevenue, the net savings reduces to w$82,870/year.

The base load on thermal energy storage (TES) is 52MWhwhichis primarily imposed by evaporatormodule in ARS system. Based onthe NREL’s cost estimate [2] for individual components of sensiblethermal energy storage (Table 4), the capital expenditure for pro-posed TES can be estimated as $997,360. It should be noted thatTable 1 were based on the following assumptions: i) Electrical andinstrumentation costs ¼ 7% of total tank costs ii) Piping, valves, andfittings costs ¼ 3% of total tank costs iii) Sales tax ¼ 4% and iv)Contingency costs ¼ 7% [40]. Therefore, the payback period for theproposed TES system can be estimated as 22 years with thefollowing parameters: interest rate ¼ 6.125%; Loanamount ¼ $997,360; annual installment ¼ $82,878.

For the multi-effect desalination system, the fuel savings can berealized in the form of waste heat that is available from the powerplant. The electricity requirements for MED system ranges from 2.5to 3 kWh/m3, and that for the RO (reverse osmosis) are in the rangeof 6e10 kWh/m3 [41]. For combined cycle seawater desalination(cogeneration), the costs were reported to range from $0.70 to 0.96/m3 for RO, and from $1.10 to 1.71/m3 for MED for a combined cyclepower plant driven by fossil fuel and nuclear energy sources [42].Ettouney and El-Dessouky [41] have reported a wide range of costsfor multi-effect evaporation desalination plants ranging from 0.35to 2.34 $/m3 which depends on the capacity and design configu-ration of the desalination plant. Table 5 shows the power and watercosts for a combined cycle power plant integrated with twodifferent desalination processes, i.e., with MED and RO. It can benoted that MED process integrated with the proposed configura-tion has less water costs (0.67/m3) compared to the RO process(0.96/m3) since the heat requirements for MED process are sup-plied by the waste steam generated from the power plant while theRO process requires high quality prime electricity only. Thus theproposed configuration can reduce the desalination costs byincorporating a MED desalination system to utilize waste heat fromthe power plant. Another advantage with the proposed configura-tion is that when the fossil fuel prices increase and the carbon taxcarried by the greenhouse emissions is enforced in the future, the

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Fig. 10. Sensitivity of Storage temperature to TES Volume.

Fig. 11. HE (Heat exchanger) efficiency and TES performance. Note: Heat exchanger forstack gases to generator in ARS.

Table 4Component costs for sensible thermal energy storage.

Component Materials($/kWh)

Installation($/kWh)

Total ($/kWh)

Carbon steel tank 1.3 1.84 7.04Tank supports 1.1 1.55 2.65Storage medium 11.74 0.36 12.1Electricity & instrumentation 0.47 0.43 0.9Pipes, valves, and fittings 0.20 0.18 0.39

Total 23.08

Fig. 9. Energy requirements and desalination capacity vs. number of stages (forevaporating temperatures 100 �C and 70 �C).



cost and energy (steam) advantage for MED desalinationwill not beaffectedwhereas it may severely affect the RO process by increasingthe water production costs significantly [42]. It was also reportedthat the combined cycle power plants integrated with desalinationsystems are more profitable than other gas turbine based dualpurpose power and desalination plants [43] and the integratedconfigurations (cogeneration) are proven to be more thermody-namically efficient and economically feasible than single purposepower generation and water production plants [44].

5. Conclusions

We evaluated the feasibility of using waste heat to drive the ARS(absorption refrigeration system) and multi-effect evaporationdesalination system in power plants. The ARS configuration wasfound to be effective in maintaining the temperature of chilledwater in TES (thermal energy storage) which catered the coolingrequirements in both MED and ACCs (air-cooled condensers). TheTES system configuration was able to mitigate the disadvantagesassociated with use of ACCs in combined cycle power plants,especially during hot summers in arid regions of Nevada. Theavailability of waste heat from stack gas determines the sizing ofARS which in turn influences the size of storage tank. Availability ofwaste heat is a function of specific power plant configuration. Thesizing and feasibility of MED system is again dependent upon thequantity and quality of brackish waters present at the site of a givenpower plant. The outcomes from this study provide a strongimpetus to develop TES-based ACCs in power plants to reap thebenefits of enhanced power production and low-cost desalinationsystem.

Acknowledgments

Dr. Gadhamshetty acknowledges funding support from FGCUMultidisciplinary Research Initiative Award by Office of Research &Graduate Studies at Florida Gulf Coast University. The fundingsource had no involvement in study design; in the collection,analysis and interpretation of data; in the writing of the report; andin the decision to submit the paper for publication.

Veera Gnaneswar Gude acknowledges the support provided bythe Office of Research and Economic Development (ORED), theBagley College of Engineering (BCoE), and the Department of Civiland Environmental Engineering (CEE) at Mississippi StateUniversity.

Nomenclature

Notations for CCPP, ARS, and TES

SymbolsABS absorberCC combustion chamberCOMP gas compressor

COND condenserCOP coefficient of performanceCW chilled waterE, EVAP evaporatorG volumetric flow rate (m3/s)GT gas turbineGEN generatorHE heat exchangerHP high pressure in ARS (kPa)h specific enthalpy (kJ/kg)K thermal conductivity, (W/m-K)LP low pressure in ARS (kPa)m mass flow rate at point i (kg/s)OFWH open feed water heaterP pumpQ net heat transfer rate (MW)q specific net heat (kJ/kg)ST steam turbineT TES temperature, (�C)T ambient temperature, (�C)T* design temperature for air inlet to ACC, (�C)t time (s)V volume (m3)V�

volumetric flow rate (m3/s)W power output (MW)w specific net work (kJ/kg)y mass fractionh efficiencyr density (kg/m3)v specific volume (m3/kg)

Subscriptsa absorberc condensere evaporatorg, gen generatorg_comp gas compressorg_turb gas turbinehx heat exchangerin, i inletout outlets storagesteam_turb steam turbineth thermal efficiencyw water

Superscripts� indicates rate

Notations for MEDA area, m2

B brine flow rate from each evaporation effect, kg/sCp specific heat at constant pressure, kJ/kg ECCR conversion ratio, CR ¼ Md/Mf, dimensionlessD amount of vapor formed in each flashing stage or

evaporation effect, kg/sF feed flow rate to each evaporation effect, kg/sLMTD logarithmic mean temperature differenceM mass flow rate, kg/sn number of tubes, flashing stages, or evaporation effectsP pressure, kPaPR performance ratio, PR ¼ Md/Ms, dimensionlessT temperature, �CDT temperature drop, ECDTl temperature losses in each evaporation effect, �C

Table 5Cost comparison between the MED and RO processes with combined cycle powerplants.

Specific power costs ($/kWh) Specific water costs ($/m3) MED RO

Fixed charge cost 0.008 Fixed charge cost 0.328 0.278Fuel cost 0.075 Heat cost e e

O&M cost 0.006 Plant electricity 0.204 0.224Decommissioning

costN/A Purchased electricity cost 0.000 0.019

O&M cost 0.139 0.175Levelized electricity

cost0.088 Total specific water cost 0.671 0.696



U overall heat transfer coefficient, kW/m2�CV specific volume, m3/kgX salinity, ppm

Subscriptsl latent heat for evaporation, kJ/kgb brinebh brine/feed preheaterc condenser or condensatecw intake seawaterd distillatee evaporatorf feedh brine heaterj heat rejection section in MSFo outer diameter or outlet temperaturen last flashing stage or last evaporation effectr heat recovery section in MSFv vapor

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Thermal energy storage system for energy conservation and water desalination in power plants

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