Thermodynamic performance analysis of a vapor compression–absorption cascaded refrigeration system Vaibhav Jain a,⇑ , S.S. Kachhwaha b , Gulshan Sachdeva a a Department of Mechanical Engineering, National Institute of Technology, Kurukshetra, India b Department of Mechanical Engineering, School of Technology, Pt. Deendayal Petroleum University, Gandhinagar, India article info Article history: Received 14 February 2013 Accepted 5 August 2013 Keywords: Vapor compression Absorption Cascaded refrigeration system First law Second law analysis abstract In the present study, a thermodynamic model for cascaded vapor compression–absorption system (CVCAS) has been developed which consists of a vapor compression refrigeration system (VCRS) coupled with single effect vapor absorption refrigeration system (VARS). Based on first and second laws, a compar- ative performance analysis of CVCAS and an independent VCRS has been carried out for a design capacity of 66.67 kW. The results show that the electric power consumption in CVCAS is reduced by 61% and COP of compression section is improved by 155% with respect to the corresponding values pertaining to a conven- tional VCRS. However there is a trade-off between these parameters and the rational efficiency which is found to decrease to half of that for a VCRS. The effect of various operating parameters, i.e., superheating, subcooling, cooling capacity, inlet temperature and the product of effectiveness and heat capacitance of external fluids are extensively studied on the COP, total irreversibility and rational efficiency of the CVCAS. Besides, the performance of environment friendly refrigerants such as R410A, R407C and R134A is found to be almost at par with that of R22. Hence, all the alternative refrigerants selected herein can serve as potential substitutes for R22. Furthermore, it has been found that reducing the irreversibility rate of the condenser by one unit due to decrease in condenser temperature depicted approximately 3.8 times greater reduction in the total irreversibility rate of the CVCAS, whereas unit reduction in the evaporator’s irrevers- ibility rate due to increase in evaporator temperature reduced total irreversibility rate by 3.4 times for the same system. Since the changes in the inlet temperatures of external fluid in the condenser and the evap- orator contribute significant changes in system’s overall irreversibility, due consideration is required in condenser and evaporator temperatures to improve the system performance. Ó 2013 Elsevier Ltd. All rights reserved. 1. Introduction Refrigeration is the process of maintaining the system temper- ature to a value lower than that of surrounding [1] by means of a refrigeration cycle. Vapor compression refrigeration system is widely used in domestic as well as industrial refrigerating and air-conditioning equipments such as in restaurants, hotels, hospi- tals and theatres. It is also used for manufacturing of ice, dehydra- tion of gases, lubricating oil purification, low temperature reactions, and separation of volatile hydrocarbons, etc. [1]. Many developing countries like India currently suffer from a major shortage of electricity. The demand for electricity in India is very high both in terms of base load energy and peak availability. During the year 2010–2011, base load requirement was 861,591MU against the availability of 788,355 MU – an 8.5% defi- cit. During peak loads, the demand was for 122 GW against the availability of 110 GW, a 9.8% shortfall. Out of the 1.4 billion people in the whole world with no access to electricity, India accounts for over 300 million [2]. The refrigerating and air conditioning equipment based on VCRS consume a considerable amount of electric power. In India, 56% of the total electrical capacity is generated using coal [2]. Not only does it exacerbate the depletion of fossil fuel, but also re- sults in the production of harmful gases such as CO 2 , nitrogen oxi- des and sulfur oxides etc. It is well known that these gases cause green house effect and deteriorate the environment. In various industrial sectors like liquid milk processing, chilled ready meals, frozen food, cold storage etc. the electricity employed for refriger- ation is 25%, 50%, 60% and 85% respectively of their total energy consumption [3]. However, due to poor availability of electric power, the total installed refrigeration capacity is inadequate to meet the refrigeration requirements. Nearly 30% of fresh fruits and vegetables produced in India are wasted due to lack of refrig- eration technology [4]. One of the alternatives to reduce the dependence on electrically powered VCRS is the use of VARS in which the compressor is re- placed by absorber, pump and generator [1]. VARS demands a 0196-8904/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.enconman.2013.08.024 ⇑ Corresponding author. Tel.: +91 9868004700. E-mail address: [email protected](V. Jain). Energy Conversion and Management 75 (2013) 685–700 Contents lists available at ScienceDirect Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
Transcript
Energy Conversion and Management 75 (2013) 685–700
Vaibhav Jain a,⇑, S.S. Kachhwaha b, Gulshan Sachdeva a
a Department of Mechanical Engineering, National Institute of Technology, Kurukshetra, Indiab Department of Mechanical Engineering, School of Technology, Pt. Deendayal Petroleum University, Gandhinagar, India
a r t i c l e i n f o a b s t r a c t
Article history:Received 14 February 2013Accepted 5 August 2013
Keywords:Vapor compressionAbsorptionCascaded refrigeration systemFirst lawSecond law analysis
In the present study, a thermodynamic model for cascaded vapor compression–absorption system(CVCAS) has been developed which consists of a vapor compression refrigeration system (VCRS) coupledwith single effect vapor absorption refrigeration system (VARS). Based on first and second laws, a compar-ative performance analysis of CVCAS and an independent VCRS has been carried out for a design capacityof 66.67 kW. The results show that the electric power consumption in CVCAS is reduced by 61% and COP ofcompression section is improved by 155% with respect to the corresponding values pertaining to a conven-tional VCRS. However there is a trade-off between these parameters and the rational efficiency which isfound to decrease to half of that for a VCRS. The effect of various operating parameters, i.e., superheating,subcooling, cooling capacity, inlet temperature and the product of effectiveness and heat capacitance ofexternal fluids are extensively studied on the COP, total irreversibility and rational efficiency of the CVCAS.Besides, the performance of environment friendly refrigerants such as R410A, R407C and R134A is foundto be almost at par with that of R22. Hence, all the alternative refrigerants selected herein can serve aspotential substitutes for R22. Furthermore, it has been found that reducing the irreversibility rate of thecondenser by one unit due to decrease in condenser temperature depicted approximately 3.8 times greaterreduction in the total irreversibility rate of the CVCAS, whereas unit reduction in the evaporator’s irrevers-ibility rate due to increase in evaporator temperature reduced total irreversibility rate by 3.4 times for thesame system. Since the changes in the inlet temperatures of external fluid in the condenser and the evap-orator contribute significant changes in system’s overall irreversibility, due consideration is required incondenser and evaporator temperatures to improve the system performance.
� 2013 Elsevier Ltd. All rights reserved.
1. Introduction
Refrigeration is the process of maintaining the system temper-ature to a value lower than that of surrounding [1] by means of arefrigeration cycle. Vapor compression refrigeration system iswidely used in domestic as well as industrial refrigerating andair-conditioning equipments such as in restaurants, hotels, hospi-tals and theatres. It is also used for manufacturing of ice, dehydra-tion of gases, lubricating oil purification, low temperaturereactions, and separation of volatile hydrocarbons, etc. [1].
Many developing countries like India currently suffer from amajor shortage of electricity. The demand for electricity in Indiais very high both in terms of base load energy and peak availability.During the year 2010–2011, base load requirement was861,591MU against the availability of 788,355 MU – an 8.5% defi-cit. During peak loads, the demand was for 122 GW against theavailability of 110 GW, a 9.8% shortfall. Out of the 1.4 billion people
in the whole world with no access to electricity, India accounts forover 300 million [2].
The refrigerating and air conditioning equipment based onVCRS consume a considerable amount of electric power. In India,56% of the total electrical capacity is generated using coal [2].Not only does it exacerbate the depletion of fossil fuel, but also re-sults in the production of harmful gases such as CO2, nitrogen oxi-des and sulfur oxides etc. It is well known that these gases causegreen house effect and deteriorate the environment. In variousindustrial sectors like liquid milk processing, chilled ready meals,frozen food, cold storage etc. the electricity employed for refriger-ation is 25%, 50%, 60% and 85% respectively of their total energyconsumption [3]. However, due to poor availability of electricpower, the total installed refrigeration capacity is inadequate tomeet the refrigeration requirements. Nearly 30% of fresh fruitsand vegetables produced in India are wasted due to lack of refrig-eration technology [4].
One of the alternatives to reduce the dependence on electricallypowered VCRS is the use of VARS in which the compressor is re-placed by absorber, pump and generator [1]. VARS demands a
BEP breakeven point (years)c concentration of LiBr solutionC heat capacitance rate of external fluid (kW/K)cp specific heat at constant pressure (kJ/kg K)CIC capital investment cost (Rs.)COP coefficient of performanceE exergy (kW)EC electric cost (Rs.)EEUF electrical energy utilization factorf circulation ratioFESR fuel energy saving ratioh specific enthalpy (kJ/kg)HUF heat utilization factorI irreversibility rate (kW)m mass flow rate (kg/s)P pressure (kPa)PP payback period (years)Q heat transfer (kW)s specific entropy (kJ/kg K)T temperature (�C)U uncertaintyW power input (kW)
Greek symbole effectiveness of heat exchangerg efficiencyd efficiency defectw specific exergy (kJ/kg)h Carnot factorr coefficient of structural bondq density of LiBr solution (kg/m3)
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substantial amount of heat energy for the generator; however, heatis low grade energy and its demand can be fulfilled using non-con-ventional sources such as solar, geothermal and waste heat. Theelectrical energy consumed by this system is merely 10% of the to-tal energy requirements [5]. Several researchers [6–11] have stud-ied the performance of VARS considering H2O–LiBr as a workingpair. This system is generally used for air conditioning purposesas it can maintain evaporator temperature up to 5 �C [5]. For lowtemperature applications i.e. below 5 �C, working pair of NH3–H2O can be used in vapor absorption system [12]. But NH3–H2Odoes not form an ideal pair for absorption system because the dif-ference in their Normal Boiling Points (N.B.Ps) is not large enough(138 �C). There should be a sufficiently large difference in theN.B.Ps of the two substances (at least 200 �C) [1] so that the absor-bent exerts negligible vapor pressure at the generator temperature.This system produces ammonia mixed with water vapor at the exitof the generator. Water in the refrigerant stream can cause opera-tional problems in the evaporator of the system. Thus H2O–LiBrpair is suitable from the view point of solubility and boiling pointrequirements but it cannot be used for low temperature refrigera-tion [1].
The application of cascaded refrigeration system maintains theadvantages of both vapor compression and vapor absorption sys-tems while minimizing the limitations of both simultaneously.The main advantage of CVCAS over VCRS is that it saves consider-able amount of high grade energy (electrical energy). While thestructure of the cascaded system is more complex and bulky, theoverall operating cost is relatively lower because of simultaneoususage of electricity and heat energy for refrigeration. Furthermore,the non-conventional sources of energy such as solar, geothermal
etc. can be used to supply heat energy for this system. The litera-ture on CVCAS is summarized in Table 1 with key findings.
Many researchers [13,21,22,26] have considered NH3–H2O pairof working fluid in absorption system which does not form an idealpair [1]. Unlike water, ammonia is both toxic and flammable. Theprevious work [13,21–26] reported on CVCAS is based on firstlaw analysis (energy conservation). In the present study, a CVCASconsisting of VCRS coupled with VARS (H2O–LiBr as working pair)has been proposed as an alternative to reciprocating vapor com-pression chiller [27]. Besides comparative energy as well as exergyanalysis supported with preliminary economic analysis has alsobeen performed. The concepts of coefficient of structural bond(CSB), heat utilization factor (HUF), electrical energy utilizationfactor (EEUF), fuel energy saving ratio (FESR) and heat to work ratiohave been applied for better understanding of performance of thesystem with wide range of cooling capacities. The effect of alterna-tive refrigerants (R134A, R410A and R407C), superheating, subco-oling, and size of heat exchanger are also investigated in detail.
2. Theoretical formulation of vapor compression–absorptioncascaded refrigeration system
2.1. System selection
Fig. 1 shows a simple variable speed VCRS described by Khanand Zubair [27] used for water cooling and the corresponding P-h diagram is plotted in Fig. 2. Saturated refrigerant vapor leavesthe evaporator at state 1 and saturated liquid exits the condenserat state 3. The refrigerant then flows through the expansion valveto the evaporator [28].
Table 1Summary of literature survey of CVCAS.
Author System selection Refrigerants Salient features
Cimsit and Ozturk [13] Cascade cycle H2O–LiBr and NH3–H2O pairs in absorptionsection and R134A, R410A and NH3 in vaporcompression section
� Electrical energy consumption in the cascaderefrigeration cycle is 48–51% lower than classical vaporcompression refrigeration cycle� Heat energy requirement at generator is reduced by 35%and overall COP of system is improved by 33% with H2O–LiBr fluid pair instead of NH3–H2O pair� Categorized hybrid vapor compression and absorptionsystem into two groups called as combined cycle andcascaded cycle. Many researchers [5,14–20] had workedon combined cycle.
Kairouani and Nehdi [21] Cascade cycle powered bygeothermal energy inabsorption section andcombined cycle
NH3–H2O pair in absorption section and R717,R22 and R134A in vapor compression section
� COP of proposed cascade system is reported to be 5.5which is superior to that presented by Ayala et al. [5] forcombined system. Ayala et al. [5] indicated a COP of 3.1with NH3–H2O as pair of refrigerant� COP of cascaded cycle was 37–54% higher than thevapor compression cycle.
Fernandez-Seara et al. [22] Cascade cycle powered bycogeneration system
CO2 and NH3 in compression section and NH3–H2O working pair in the absorption section
� COP of compression section is reported to be 2.602 and2.463 with CO2 and NH3 respectively for low evaporatortemperature of �45 �C� Intermediate temperature defined as the condensationtemperature of the compression system is an importantdesign parameter that determines the overall COP ofsystem
Seyfouri and Ameri [23] Cascade system in which thecompressor was powered bya microturbine
R22 in compression section and H2O-LiBrworking pair in the absorption section
� Four configurations of the integrated refrigerationsystem were considered and highest energy consumptionwas reported to reduce up to 133%� Use of absorption system improves the COP ofcompression chiller by 440%
Wang et al. [24] Solar assisted cascadedrefrigeration system
R134A in compression section and H2O–LiBrworking pair in the absorption section
� COP of the system can be attained up to 6.1 with thesolar intensity of 700 W/m2
� Power consumption was reported to be lowered by 50%in comparison with conventional mechanical vaporcompression systems
Garimella et al. [25] Cascade refrigeration systemfor naval ship application
CO2 in compression section and H2O–LiBrworking pair in the absorption section
� 31% electricity demand was found to be reduced whencompared to an equivalent vapor compression system
Chinnappa et al. [26] Solar operated cascadesystem for air conditioning
R22 in compression section and NH3–H2Oworking pair in the absorption section
� Considerable savings in electrical energy consumptionby the compression system
3
1
1’
2
Q loss, suc
Expansion valve
7 8
Water cooled evaporator
3’
4
Aircooled Condenser
5 6Q loss, liq
Compressor
Fig. 1. Schematic diagram of a simple VCRS.
Pres
sure
,P
Enthalpy, h
1’
2
4 1
33’
Fig. 2. Pressure–enthalpy diagram for simple VCRS.
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Fig. 3 shows a CVCAS in which VCRS [27] and VARS [6,13] arecoupled together by a heat exchanger called the cascade con-denser, where the refrigerant vapor (R22) of VCRS is condensed,
by rejecting heat to the refrigerant (water) of VARS. The main com-ponents of CVCAS are evaporator, compressor, cascade condenser,VCRS expansion valve, absorber, pump, generator, solution heatexchanger, pressure reducing valve, condenser and VARS expan-sion valve [13]. The heat absorbed by the evaporator and the workdone to the compressor of VCRS is theoretically equal to the heatabsorbed by the cascade condenser of the VARS [13]. The con-denser in this cascaded refrigeration system rejects heat to its
15 16
12
Pump
Absorber
11
8
9
10
Pressure reducing valve
Expansion valve 1
6
Generator
18
17
19
20
1
2 3
4
22 21
1’
5 14
Cascade condenser
Expansion valve 2 Compressor
7
Q loss, suc
Q loss, liq
13
13’
Evaporator
Condenser
SHX
Fig. 3. Schematic diagram of CVCAS.
688 V. Jain et al. / Energy Conversion and Management 75 (2013) 685–700
warm coolant (air). The condensate (water) supplied through theexpansion valve of VARS converts into low pressure vapor byabsorbing the heat rejected in cascade condenser. This low pres-sure cold vapor enters the absorber, where it get mixed and ab-sorbed with a hot solution of LiBr. The heat generated inabsorber is carried away by circulating cooling air. The solutionleaving the absorber, being rich in refrigerant vapor, is a weak solu-tion of salt in vapor. This weak solution of LiBr is pumped to thegenerator through a heat exchanger [13]. The pump work is verysmall compared to the compressor work of VCRS as the specificvolume of liquid is extremely small compared to that of vapor.Mainly, the energy consumption in VARS takes place only withinthe generator in the form of low grade energy. Water (refrigerant)gets boiled off in the generator due to heat transfer. Since the saltdoes not exert any vapor pressure, the vapor leaving the generatoris a pure refrigerant (water vapor) [1]. This high pressure water va-por is condensed over cooling coils of air cooled condenser. Theabsorption section using H2O–LiBr as working pair operates undervery low pressures (high vacuum) and the limiting evaporatortemperature of this section is 5 �C [5]. Hence proper care is re-quired to avoid air leakage into the system and moreover, thereare chances of crystallization of LiBr salt under high concentrations[1]. The intermediate temperature between the two circuits is alsoan important design parameter that influences the COP of the en-tire system [22].
The high and low pressure cut outs, fusible plug, pressure reliefvalve, antifreeze thermostat, oil safety switch, capacity controlswitch, fluid flow control devices, evaporator pressure regulatingvalve, rupture disc in generator, safety control to prevent crystalli-
zation of LiBr salt and cavitation in the pump etc. are the essentialcontrols used in individual VCRS and VARS [1]. The similar controlsare also applicable to CVCAS [Uppal Refrigeration & Eng. Co., India,personal communication, May 23, 2013].
2.2. Thermodynamic modeling of vapor compression–absorptioncascaded refrigeration system
Following assumptions are made in thermodynamic modelingof VCRS and CVCAS [13,27].
1. The system is in steady state.2. All the pressure losses in different components of system are
neglected.3. Superheating and subcooling take place due to heat gain and
heat loss in suction and liquid lines respectively. Heat lossesin suction and liquid lines are not taken into consideration atdesign point.
4. Refrigerant at the exit of evaporator, cascade condenser andcondenser is in saturation state.
5. Isentropic efficiency of compressor is constant.6. Product of effectiveness of heat exchanger and heat capacitance
rate of external fluid (eC) is constant with respect to coolingcapacity.
7. The process occurring in expansion valves are isenthalpic.
Exergy analysis of a complex system can be performed by ana-lyzing the components of the system separately. This describes alllosses both in the various components of the system and in the
Table 2Governing equations for different components of CVCAS.
S.No.
Component First law equations Second law equations
V. Jain et al. / Energy Conversion and Management 75 (2013) 685–700 689
system as a whole [29]. With the help of this analysis, the magni-tude of these losses or irreversibility is evaluated and their order ofimportance can be understood. Thus exergy analysis can indicatethe possibilities of thermodynamic improvement for the processesof present CVCAS under consideration.
Thermodynamic modeling includes following set of governingequations [30,31] for a particular system component expressed as:
(i) Mass balance,P
m ¼ 0;P
cm ¼ 0.(ii) Energy balance,
PQ þ
PW þ
Pmh ¼ 0.
(iii) Exergy balance, I ¼ Ein � Eout þ EQ þ EW .
where I is the total irreversibility or exergy destruction. EQ and EW
are the exegy flow associated with heat transfer and work transferrespectively.
Ein and Eout are the exergies of matter at the inlet and outlet ofcontrol volume and are given by, E = mw where w is the specificexergy of matter that includes chemical, kinetic, potential andphysical exergies. Chemical, kinetic and potential exergies are ig-nored in this work. The physical specific exergy is given by,w ¼ ðh� TosÞ � ðho � TosoÞ.
Applying the above three fundamental equations to all the com-ponents of CVCAS, the governing equations for different compo-nents of CVCAS are given in Table 2. The state points are definedrelative to Fig. 3.
For complete cycle, the change in internal energy is zero.
Efficiency defect (dk) of kth component of the system may be ex-pressed as fractions of input which are lost through irreversibility.Efficiency defect (dk) of kth component of the system may be ex-pressed as,
dk ¼ Ik=ðW þ Qghcarnot þWpÞ ð7Þ
where Carnot factor,
hcarnot ¼ 1� To=Tg
From the exergy analysis irreversibility of the components andthe cycle are determined. Since the irreversibility do not indicatehow to improve the cycle, the structural analysis using the coeffi-cients of structural bonds (CSB) has been applied. The coefficient ofstructural bonds (CSB) is defined as [31],
rk;i ¼@It
@xi
� ��@Ik
@xi
� �ð8Þ
where Ik is the irreversibility of kth component of system underconsideration and xi the parameter of system which produces thechanges.
The effect of a change in xi on the system would be to alter therate of exergy input while leaving the output constant. This accep-tance confirms to the usual practice of specifying a plant in termsof its output rather than its input. From the exergy balance of thesystem,
Ein ¼ Eout þ It ð9Þ
But, Eout = constant, thus
DEin ¼ DIt ð10Þ
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As seen from Eq. (10), changes in the irreversibility of the sys-tem are equivalent to changes in the exergy input. In general, theratio of the rates of exergy output to exergy input is less than unity.This ratio denotes the degree of thermodynamic perfection of theprocess and is called the rational efficiency (gR).
gR ¼ Eout=Ein ð11Þ
System rational efficiency [29],
gR ¼ 1�X
dk ð12Þ
The concept of heat utilization factor (HUF), electrical energyutilization factor (EEUF) and fuel energy saving ratio (FESR) are ta-ken from [32] and modified for CVCAS accordingly.
Heat utilization factor (HUF) for the present system is theamount of heat required at generator of VARS for unit refrigerationeffect.
HUF ¼ Q g=Qevap ð13Þ
Electrical energy utilization factor (EEUF) is described as theamount of total electrical energy required at compressor and pumpof cascaded system for unit refrigeration effect.
EEUF ¼ ðW þWpÞ=Q evap ð14Þ
Fuel Energy saving ratio (FESR) for the present case is the netamount of energy consumed for unit refrigeration effect.
FESR ¼ ðQ g þW þWp � Q evapÞ=Q evap ð15Þ
P
0.7
0.8
0.9
2.3. Model validation
In the present study, the thermodynamic model equations arehighly nonlinear in nature and are solved using Engineering Equa-tion Solver (EES). Many researchers [23,25] working in the field ofrefrigeration employed EES for thermodynamic modeling of refrig-eration systems. EES provides many built-in mathematical andthermophysical property functions useful for engineeringcalculations.
Previous work [13,21–26] on CVCAS is based on numericalstudy. In the present work, vapor compression and vapor absorp-tion systems are individually validated by comparing the systemperformance predictions with experimental results. The designdata and simulation range used for CVCAS is within the close prox-imity of reference data used for model validation of subsystems(i.e. VCRS and VARS). The variables for validation of models of VCRSand VARS are set depending on the reference data. EES automati-cally identifies and groups equations that must be solved simulta-neously and solves all these simultaneous equations by applying avariant of Newton’s method. It works on the principle of number ofequations and variables which must be same to solve the simulta-neous equations. The subject model of VCRS is validated with theexperimental data presented by Stoker and Jones [28] with R22as the refrigerant for the following set of input variables: Tin,
evap = 10 �C, Tin, cond = 35 �C, Qevap = 83.09 kW, (eC)cond = 9.39 kW/K,(eC)evap = 8.65 kW/K and gisen = 0.65. The comparison of experi-
Table 3Comparison of present VCRS model with Stoecker and Jones [28].
mental values with model prediction is summarized in Table 3.The results show that all the parameters are predicted with-in ±0.35% error.
Besides, experimental data obtained by Xuan and Chen [33] forVCRS using R404A as refrigerant are also considered for validationof the present model for the following set of inputs:Qevap = 1152.3 W, Tcond = 43 �C, Tevap = �23 �C. The percentage er-rors in the computation of compressor work, COP and dischargetemperature (temperature at the exit of compressor) are 0.2%,0.0% and 4.2% respectively. The large difference between the pre-dicted and experimental value of discharge temperature is attrib-uted to the uncertainty in the measurement of dischargetemperature which is reported to be ±1% [33].
In order to validate the present model of VARS, the simulationresults have been compared with the theoretical results presentedby Aphornratana and Sriveerakul [34] for the input parameters:Tevap = 5 �C, Tcond = 29.8 �C, Ta = 35 �C, T10 = 50 �C, m9 = 0.82 kg/minand gp = 0.9. The maximum difference between the COP values ob-tained using the present model and the corresponding theoreticalresults [34], shown in Fig. 4, is merely ± 0.7%. The variation in the-oretical and experimental results is caused due to unwanted heatloss and heat gain in different components of absorption system[34].
The results obtained using the complete model for CVCAS arealso compared with the corresponding numerical results pre-sented by Cimsit and Ozturk [13]. The following set of input datais used for comparison: Tcond = 40 �C, Ta = 40 �C, Tg = 90 �C,Tevap = 10 �C, Qevap = 83 kW, gp = 0.90, gisen = 0.80 and e = 0.6.H2O–LiBr is taken as the working pair in absorption section andR134A is considered in compression section. The maximum errorin the calculated values is found to be ±1.62% (Table 4). Discrep-ancies in the results are because of different correlations used infinding the thermophysical properties of H2O–LiBr solution. In thepresent work, the properties are calculated using the built-infunction of EES. The COP of CVCAS is predicted with an accuracyof ±0.51%.
3. Results and discussion
3.1. Initial conditions
The thermodynamic model described in Section 2.2 is applied toevaluate the performance of CVCAS shown in Fig. 3. For VCRS, theformulation is modified accordingly. The value of inputs at the de-sign point is given in Table 5.
Tg (°C)60 65 70 75 80 85 90
CO
0.3
0.4
0.5
0.6
COP_present modelCOP_theortical:Aphornratana and Sriveerakul [34]COP_experimental:Aphornratana and Sriveerakul [34]
Fig. 4. Performance curve of present VARS model and Aphornratana andSriveerakul [34].
Table 5Value of inputs at design point.
Parameter Value
Evaporator coolant inlet temperature (Tin, evap in �C) 4Evaporator coolant outlet temperature (Tout, evap in �C) 1Condenser coolant inlet temperature (Tin, cond in �C) 35Condenser coolant outlet temperature (Tout, cond in �C) 40Generator coolant inlet temperature (Tin, g in �C) 100Generator coolant outlet temperature (Tout, g in �C) 95Generator temperature (Tg in �C) 90Absorber coolant inlet temperature (Tin, a in �C) 35Absorber coolant outlet temperature (Tout, a in �C) 38Absorber temperature (Ta in �C) 40Rate of heat absorbed by evaporator (Qevap in kW) 66.67Product of condenser effectiveness and capacitance rate of external
fluid [(eC)cond in kW/K]9.39
Product of evaporator effectiveness and capacitance rate ofexternal fluid [(eC)evap in kW/K]
8.2
Temperature at exit of cascade condenser (T5 in �C) 10Degree of superheat in suction line (Tsup in �C) 0Degree of subcooling in liquid line (Tsub in �C) 0Effectiveness of SHX (eSHX) 0.6Isentropic efficiency of compressors (gisen) 0.65Electrical efficiency of pump (gp) 0.9Degree of overlap in cascade condenser (Toverlap in �C) 8Environmental temperature (To in �C) 25Atmospheric pressure (Po in kPa) 101.325
Table 4Comparison of present CVCAS model with Cimsit and Ozturk [13].
Parameter Present model Cimsit and Ozturk [13] Difference (%)
V. Jain et al. / Energy Conversion and Management 75 (2013) 685–700 691
Tables 6 and 7 represent the thermodynamic properties at inletand outlet of each component of VCRS and CVCAS respectively atthe design point. Temperature, pressure and mass flow rate ofR22, water vapor and LiBr are the operating parameters of theworking fluids, whereas, enthalpy, entropy and physical exergyare the calculated parameters.
A comparison of Tables 6 and 7 shows that refrigerant (R22)mass flow rate in CVCAS is reduced by 18%. The present thermody-namic model predicts the condenser and evaporator temperaturesfor both VCRS and CVCAS. Based on cooling capacity and effective-ness of evaporator, the evaporator temperature for the systems is�4.1 �C. The predicted condenser temperatures are 44.6 �C and43.6 �C for VCRS and CVCAS respectively. The reduction in con-denser temperature also causes 11% lower mass flow rate of airrequirement for condenser cooling. Table 8 represents the compar-
Table 6Value of thermodynamic properties at design point for VCRS.
State point T (�C) P (kPa) m (R22) (kg/s) m (H2O) (
ison of performance parameters for VCRS and CVCAS. Based on thedata of low and high grade energies of different components, theenergy balance (first law) condition is satisfied.
The heat transfer rate is found to be maximum (104.90 kW) inthe generator. On the other hand, the heat transfer rates in the con-denser and absorber are 80.96 kW and 100.10 kW respectively. Theeffect of pump on the total energy input is found to be negligible.
The prescribed temperature (T5) for refrigerant (water vapor) anddegree of overlap are 10 �C and 8 �C respectively in cascade con-denser [13]. Thus the refrigerant (R22) has a temperature of 18 �Cat the exit of cascade condenser. Due to reduction in mass flow rateof refrigerant and low temperature at the compressor exit in CVCASas compared to that in VCRS, the power consumption in the com-pressor is reduced by 61%. Besides, the COP of CVCAS is found tobe 0.583, which is quite low as compared to that for VCRS becauseit utilizes low grade energy. The COP of vapor compression and vaporabsorption subsystems of CVCAS are 7.082 and 0.725 respectively.The value of COP is higher for vapor compression refrigeration sub-system due to the low refrigerant temperature at compressor exitwhich leads to a reduction in the electricity requirement of compres-sor. Similar trends are also reported by other researchers[13,21,24,25]. The heat rejection in the condenser of CVCAS is re-duced by 11% as compared to that for the case of VCRS.
Table 8 also shows the irreversibility rate and efficiency defectspertaining to each system component. Further, the most sensitivecomponents are compressor in VCRS and generator in CVCAS asfar as irreversibility rate and efficiency defect are the primary out-put variables of concern. The irreversibility rate of the generatorconstitutes a large fraction of the total irreversibility rate in thesystem primarily due to the temperature difference between theheat source and the working fluid. The total irreversibility rateand efficiency defect for the entire CVCAS is 22.81 kW and0.8086 respectively. The rational efficiency of CVCAS is almost halfof that for VCRS. The total fixed cost of the CVCAS is higher as com-pared to VCRS due to addition of VARS components; however, therunning cost will decrease due to small electricity consumption incompressor and utilization of low grade energy in generator.
3.2. Performance study of cascaded system
In the present study, the cascaded system developed is used tostudy its performance with wide range of cooling capacities whichis varied by varying the mass flow rate of refrigerant. The effect ofparameters such as superheating, subcooling, generator tempera-ture, degree of overlap, size of heat exchangers and inlet tempera-ture of external fluids are also investigated. System performance isalso compared for alternative refrigerants to explore substitutes ofR22 in CVCAS.
3.2.1. Effect of variation of cooling capacityThe effect of variation of cooling capacity on the performance
characteristics of CVCAS are shown in Figs. 5–14. The evaporator
5. Other parameters f 10.26 –HUF 1.570 –EEUF 0.141 –FESR 0.715 –
692 V. Jain et al. / Energy Conversion and Management 75 (2013) 685–700
1/Qevap (kW-1)
0.010 0.012 0.014 0.016 0.018 0.020
Hea
t rej
ecte
d in
con
dens
er, Q
cond
(kW
)
40
60
80
100
120
140
160
% re
duct
ion
in h
eat r
ejec
tion
8
9
10
11
12
13
14
15
16Qcond, VCRSQcond, CVCAS% reduction in Q
Design point
Fig. 6. Variation of heat rejected in condenser and % reduction in heat rejectionwith inverse of cooling capacity.
1/Qevap (kW -1)
0.010 0.012 0.014 0.016 0.018 0.020
1/C
OP
0.0
0.5
1.0
1.5
2.0
2.5
% re
duct
ion
in 1
/CO
PC
VCAS
59.5
60.0
60.5
61.0
61.5
62.0
62.5
63.0
63.51/COPVCRS1/COPVCRS, CVCAS1/COPVARS,CVCAS1/COPCVCAS% reduction in 1/COPCVCAS
Design point
Fig. 7. Variation of 1/COP and % change in 1/COP with inverse of cooling capacity.
UF
0.15
0.16
0.17
0.18
UF1.70
1.75
1.80
1.85
EEUFHUFDesign point
V. Jain et al. / Energy Conversion and Management 75 (2013) 685–700 693
cooling capacity is varied by varying the compressor speed. As thecooling capacity is varied from 50 to 100 kW, the electric power(compressor work) required in both VCRS and CVCAS increases(Fig. 5). The electric power needed for the pump pertaining tothe absorption section is very small as compared to that for thecompressor of compression section. The total electric power re-quired in CVCAS is always less than the electric power requiredfor VCRS for the entire range of refrigeration. At the design point,the electric power consumption in CVCAS is 61% less than that inVCRS for the same cooling capacity. Moreover the percentagereduction in compressor work increases with increase in coolingcapacity. The reduction in electric power consumption in CVCASis 63% and 60%, respectively for cooling capacities of 100 kW and50 kW.
Fig. 6 shows the plot between the inverse of cooling capacityand heat rejection in condenser. The total heat rejection quantityin CVCAS is smaller as compared to VCRS system for the entirerange of cooling capacity. The heat rejected in the condenser ofCVCAS is 11% lower than that for the case of VCRS at design point.The percentage reduction in cooling capacity increases with in-crease in cooling capacity. The trend is obvious due to reductionin compressor work (Fig. 5). Therefore, the size of the CVCAS con-denser will be relatively small as compared to that for VCRSsystem.
Fig. 7 depicts that the COP reduces with increase in coolingcapacity for all the cases considered here. The COP of independentVCRS decreases from 3.174 to 2.127 as the cooling capacity in-creases from 50 kW to 100 kW. For CVCAS, the COP of compressionand absorption sections is found to decrease from 7.911 to 5.793and 0.741 to 0.654 respectively for the entire range of coolingcapacity. The COP of absorption section is less than that for thecompression section of CVCAS because it uses a large amount oflow grade heat energy. Hence, the total COP of CVCAS is always lessthan VCRS for the entire range of cooling capacity. At design point,the COPs of CVCAS, compression section and absorption section are0.583, 7.082 and 0.725 respectively. The COP of compression sec-tion is 155% higher than VCRS due to reduction in compressorwork, whereas, the COP of CVCAS is 79% lower than VCRS at designpoint.
Fig. 8 shows the variation of HUF and EEUF with cooling capac-ity. As the cooling capacity of CVCAS changes from 50 kW to100 kW, the amount of heat required in generator increases from75.93 kW to 179.20 kW. At design condition the heat required inthe generator is 104.90 kW. It can be seen that HUF and EEUF varyfrom 1.519 to 1.792 and from 0.1264 to 0.1726 respectively withincreasing cooling capacity. It can be seen from Fig. 8 that the
1/Qevap (kW-1)
0.010 0.012 0.014 0.016 0.018 0.020
Com
pres
sor w
ork,
W (k
W)
0
10
20
30
40
50
% re
duct
ion
in c
ompr
esso
r wor
k
59.5
60.0
60.5
61.0
61.5
62.0
62.5
63.0
63.5WVCRSWVCRS,CVCAS % reduction in W
Design point
Fig. 5. Variation of total electric power and% reduction in electric power withinverse of cooling capacity.
1/Qevap (kW-1)0.010 0.012 0.014 0.016 0.018 0.020
EE
0.12
0.13
0.14
H
1.50
1.55
1.60
1.65
Fig. 8. Variation of EEUF and HUF with inverse of cooling capacity.
HUF is always greater than 1 for the entire range of cooling capac-ity. This implies that the amount of heat required in the CVCASgenerator is always greater than its cooling capacity. The total elec-tric power requirement in CVCAS increases from 6.31 kW to17.26 kW as the cooling capacity increases from 50 to 100 kW. Itcan be seen that EEUF is always less than 1 due to reduction incompressor work pertaining to the cascaded system and increasesslightly with increase in the cooling capacity of CVCAS.
1/Qevap (kW -1)
0.010 0.012 0.014 0.016 0.018 0.020
Qg
/ W
10.2
10.4
10.6
10.8
11.0
11.2
11.4
11.6
11.8
12.0
12.2
Qg / WDesign point
Fig. 9. Variation of Qg/W with inverse of cooling capacity.
1/Qevap (kW-1)
0.010 0.012 0.014 0.016 0.018 0.020
Tem
pera
ture
, T (
o C)
-20
-10
0
10
20
30
40
50
60
Tin, evapTevapTin, condTcondT3T5
Design point
Fig. 10. Variation of system temperatures with inverse of cooling capacity.
1/Qevap (kW-1)
0.010 0.012 0.014 0.016 0.018 0.020
mas
s flo
w ra
te, m
(kg/
s)
0.0
0.2
0.4
0.6
0.8
circ
ulat
ion
ratio
, f
6
8
10
12
14
16
18
20mref, VCRSmref, VCRS, CVCASmref, VARS, CVCASf
Design point
Fig. 11. Variation of mass flow rate of refrigerant and circulation ratio with inverseof cooling capacity.
Fig. 12. Variation of irreversibility rate with inverse of cooling capacity.
1/Qevap (kW-1)
0.010 0.012 0.014 0.016 0.018 0.020
Rat
iona
l effi
cien
cy (%
)
10
15
20
25
30
35
40
45CVCASVCRS Design point
Fig. 13. Variation of rational efficiency with inverse of cooling capacity.
694 V. Jain et al. / Energy Conversion and Management 75 (2013) 685–700
The variation of heat to work ratio with cooling capacity isshown in Fig. 9. The plot shows that the low grade heat require-ment in generator is 10–12 times higher in magnitude as com-pared to compressor work and the heat to work ratio variesinversely with respect to the cooling capacity. Thus CVCAS mainlydemands low grade heat energy with low electric power for its
operation. The heat to work ratio is 11.14 at the design point andit varies from 10.38 to 12.01 as the cooling capacity decreases from100 to 50 kW.
Fig. 10 shows the effect of variation of inverse of cooling capac-ity on system component temperatures. Inlet temperature of exter-nal fluids at evaporator and condenser, temperature difference incascade condenser and temperature of evaporator of absorptionsection are kept constant and hence independent of refrigerationcapacity. Fig. 11 shows that the refrigerants mass flow rate in com-pression and absorption sections of CVCAS increases with increasein refrigeration capacity and, thus, the temperature difference inthe heat exchangers (evaporator and condenser of CVCAS) is high.The temperature gradient inside the condenser and evaporator ofthe cascaded system increases with increase in cooling capacity.Therefore, the losses due to finite-temperature difference in theseheat exchangers are also high and, hence, the COP is reduced.
The circulation ratio in absorption section of CVCAS is also highat high cooling capacity. The mass flow rate of refrigerant (R22) inthe compression section of CVCAS is 18% lower than that in VCRS atthe design point. Thus the reduction in mass flow rate coupledwith lower exit temperature of compressor significantly reducesthe compressor work in the CVCAS as compared to VCRS.
The effect of variation of refrigeration capacity on irreversibilityof systems and their respective components are shown in Fig. 12.For the system as well as its components, the irreversibility in-creases with increase in cooling capacity. The overall irreversibility
1/Q evap (kW -1)
CO
P
0.50
0.52
0.54
0.56
0.58
0.60
R22R407CR410AR134A
Design point
(a)
1/Q evap (kW -1)
Tota
l irre
vers
ibilit
y ra
te (k
W)
15
20
25
30
35
40
45
R22R407CR410AR134A
Design point
(b)
1/Qevap (kW-1)
0.010 0.012 0.014 0.016 0.018 0.020
Rat
iona
l effi
cien
cy (%
)
12
14
16
18
20
22
24
R22R407CR410AR134A
Design point
(c)
0.010 0.012 0.014 0.016 0.018 0.020
0.010 0.012 0.014 0.016 0.018 0.020
Fig. 14. Effect of cooling capacity on (a) COP, (b) total irreversibility, and (c) rationalefficiency.
V. Jain et al. / Energy Conversion and Management 75 (2013) 685–700 695
for CVCAS is higher than that for the VCRS, whereas, the irrevers-ibility of CVCAS components such as compressor, condenser andexpansion valve is significantly lower as compared to their counterparts in VCRS for the entire range of cooling capacity. This is be-cause of the fact that irreversibility rate in generator, absorberand cascade condenser is significantly higher. At design point,the total irreversibility for CVCAS and VCRS is 22.81 kW and15.06 kW respectively.
Rational efficiency denotes the degree of thermodynamic per-fection of a process. As shown in Fig. 13, the rational efficiencyfor CVCAS and VCRS decreases with increase in cooling capacity.
It follows from this figure that the rational efficiency for cascadedsystem varies from 14% to 22% for the selected range of coolingcapacity.
Several refrigerants have emerged as substitutes for R22, themost widely used fluorocarbon refrigerants in VCRS across theglobe. These include the environmental friendly hydrocarbon(HFC) refrigerants R134A, R410A and R407C used by world manu-facturers to meet the challenges of higher efficiency and environ-mental responsibility while keeping the system affordable.Fig. 14 shows the effect of cooling capacity on (a) COP, (b) totalirreversibility and (c) rational efficiency of CVCAS consideringalternative refrigerants in comparison to R22. The results showthat the performance of all the selected alternative refrigerants isnearly the same as R22. Hence, it is well justified to conclude thatthese refrigerants have the potential to substitute R22.
3.2.2. Effect of superheating and subcoolingThe superheating of refrigerant (after exiting the evaporator
and before entering the compressor) may occur owing to the heatgain in the suction line of compressor. This heat gain process isshown from state 1 to 10 in Fig. 3. It is obvious that the specific vol-ume of refrigerant vapor is increased owing to superheating, thusreducing the mass-flow rate through the fixed-displacement com-pressor. On the other hand, subcooling of refrigerant (after exitingthe condenser and before entering the expansion valve-1) may oc-cur due to the heat loss in the liquid line of condenser. It is ex-pected that subcooling increases the system performancebecause the specific refrigeration capacity increases withsubcooling.
Fig. 15 shows the separate effect of superheating and subcool-ing. As expected, the superheating degrades the performance ofthe system, while subcooling improves the system performance.
When equal amounts of superheating and subcooling are con-sidered, COP and rational efficiency of system decreases, whereas,the total irreversibility rate of system increases. Therefore, asshown in Fig. 15, for the given operating condition, the effect ofsuperheating has more influence on the system overallperformance.
3.2.3. Effect of generator temperatureFig. 16 shows the variation of system COP, total irreversibility
and rational efficiency with generator temperature. System COPimproves slightly with increase in generator temperature becausethermal capacity needed for generator decreases. Total irreversibil-ity of system decreases with increase in generator temperature andhence, the system rational efficiency is improved. The irreversibleloss in generator contributes a major fraction of the total irrevers-ibility of the system mainly due to temperature difference betweenthe heat source and the working fluid. Therefore, in order to de-crease the total irreversibility of the system, generator should beoperated at high temperature.
3.2.4. Effect of degree of overlapThe temperature difference between the evaporating fluid of
absorption section and condensing fluid of compression sectionin cascade condenser known as degree of overlap or approach isa design parameter that plays an important role in deciding theCOP of the overall system. Fig. 17 shows the variation of COP, totalirreversibility and rational efficiency with degree of overlap. It canbe seen that the larger the temperature difference, the lower theCOP and rational efficiency of the system because of higher totalsystem irreversibility. The COP of the system is maximum whenthe temperature difference in cascade heat exchanger is zerowhich is true only for ideal case. The assumption of a single tem-perature for both the fluids is not practically feasible as it woulddemand infinite area for the cascade heat exchanger. The optimum
ΔΤ (oC)
CO
P
0.55
0.56
0.57
0.58
0.59
0.60
SuperheatingSubcoolingSuperheating and subcooling
(a)
ΔΤ (oC)
Tota
l irrv
ersi
bilit
y ra
te (k
W)
22.0
22.5
23.0
23.5
24.0
24.5
SuperheatingSubcoolingSuperheating and subcooling
(b)
ΔΤ (oC)
0 2 4 6 8 10
0 2 4 6 8 10
0 2 4 6 8 10
Rat
iona
l effi
cien
cy (%
)
17.5
18.0
18.5
19.0
19.5
20.0
SuperheatingSubcoolingSuperheating and subcooling
(c) Fig. 15. Effect of superheating and subcooling on (a) COP, (b) total irreversibilityrate, and (c) rational efficiency.
Generator temperature (oC)86 87 88 89 90 91 92
CO
P
0.50
0.52
0.54
0.56
0.58
0.60
Tota
l irre
vers
ibilit
y ra
te (k
W) a
nd ra
tiona
l
14
16
18
20
22
24
COPTotal irreversibility rateRational efficiency
Design point
effic
ienc
y (%
)
Fig. 16. Effect of generator temperature.
CO
P
0.55
0.60
0.65
0.70
irrev
ersi
bilit
y ra
te (k
W) a
nd ra
tiona
l
18
20
22
24
26
COPTotal irreversibility rateRational efficiency
Design point
effic
ienc
y (%
)
696 V. Jain et al. / Energy Conversion and Management 75 (2013) 685–700
temperature difference between the two fluids depends not onlyon the heat transfer characteristics of the refrigerants in the twocircuits but also on the economics of the design (operating versuscapital cost).
Degree of overlap (oC)0 2 4 6 8 10 12
0.50
Tota
l
16
Fig. 17. Effect of degree of overlap.
3.2.5. Effect of (eC) of external fluid in evaporator and condenserThe factor (eC) represents multiplication of effectiveness of heat
exchanger and heat carrying capacity of external fluid. The effect ofvariation of (eC) of external fluid on COP, total irreversibility and
rational efficiency are shown in Fig. 18. As the (eC) of external fluidin evaporator or condenser increases, COP and rational efficiency ofthe system increase, whereas, the total irreversibility of system de-creases. The design of CVCAS system involves proper selection ofexternal fluid capacitance rates and heat exchanger sizes. The heatexchanger effectiveness strongly depends on the heat exchangersurface area. The product (eC) for all of the heat exchangers ofthe system is an expensive commodity. Increasing the size of heatexchanger increases the overall performance of system; however,it also increases the system cost. Hence a compromise is to seek be-tween the system fixed cost and operating cost.
3.2.6. Effect of inlet temperature of external fluids in evaporator andcondenser
Fig. 19 shows the variation of system COP, total irreversibilityand rational efficiency with inlet temperature of external fluid inevaporator and condenser. It can be observed that as the tempera-ture of external fluid at evaporator inlet decreases, the COP and ra-tional efficiency of system decreases due to increase in evaporatorirreversibility and hence the total irreversibility of system. Withthe increase in evaporator temperature of external fluid, the corre-sponding evaporator temperature also increases due to which spe-cific volume of refrigerant at the compressor inlet decreases thusdecreasing the specific work of compression. When the evapora-tion temperature increases, the heat supplied to the generator ofCVCAS is also reduced due to lower refrigeration load on cascadecondenser.
It can also be observed that as the temperature of external fluidat condenser inlet increases, the COP and rational efficiency of sys-
Inlet temperature of external fluid at evaporator (oC)
2345678
CO
P
0.50
0.55
0.60
0.65
0.70
18
19
20
21
22
23
24
COPTotal irreversibility rateRational efficiency
Design point
(a)
CO
P
0.54
0.56
0.58
0.60
vers
ibilit
y ra
te (k
W) a
nd ra
tiona
l
20
22
24
COPTotal irreversibility rateRational efficiency
Design point
effic
ienc
y (%
)To
tal i
rreve
rsib
ility
rate
(kW
) and
ratio
nal
effic
ienc
y (%
)
V. Jain et al. / Energy Conversion and Management 75 (2013) 685–700 697
tem decreases due to increase in condenser and the total systemirreversibility. With the increase in condenser temperature ofexternal fluid, the corresponding condenser refrigerant tempera-ture also increases which causes enhanced heat supply in genera-tor and reduction in COP. However, there is no significant changein mass flow rate of the refrigerant and compressor work in com-pression section.
The rational efficiency and the COP decreases with increase incondenser temperature and decrease in evaporator temperaturerespectively. The result indicates that available part of input en-ergy (exergy) continuously decreases with increase in condensertemperature and decrease in evaporator temperature. The presentcalculation shows that 6 �C increase in condenser temperature ofexternal fluid causes 14% decrease in rational efficiency and 9% de-crease in COP whereas 6 �C decrease in evaporator temperature ofexternal fluid causes 1% decrease in rational efficiency and 6% de-crease in COP.
Fig. 20(a) shows the variation of total irreversibility rate withevaporator irreversibility rate at different inlet evaporator temper-ature of external fluid (inlet condenser temperature of externalfluid is kept constant). As the inlet evaporator temperature de-creases, the total system irreversibility rate is proportional to evap-orator irreversibility. For a unit change in evaporator’sirreversibility, the total irreversibility of the system increases by3.4 times (CSB value for evaporator).
Fig. 20(b) shows the variation of total irreversibility rate withcondenser’s irreversibility rate for different inlet condenser tem-perature of external fluid (inlet evaporator temperature of external
(εC)evap of external fluid (kW/K)6 7 8 9 10 11
CO
P
0.50
0.52
0.54
0.56
0.58
0.60
17
18
19
20
21
22
23
24
25
COPTotal irreversibility rateRational efficiency
Design point
(a)
(εC)cond of external fluid (kW/K)7 8 9 10 11 12
CO
P
0.50
0.52
0.54
0.56
0.58
0.60
Tota
l irre
vers
ibilit
y ra
te (k
W) a
nd ra
tiona
l
16
18
20
22
24
26
COPTotal irreversibility rateRational efficiency
Design point
(b)
effic
ienc
y (%
)To
tal i
rreve
rsib
ility
rate
(kW
) and
ratio
nal
effic
ienc
y (%
)
Fig. 18. Effect of (eC) of external fluid in (a) evaporator and (b) condenser.
Inlet temperature of external fluid at condenser (oC)33 34 35 36 37 38 39
0.50
0.52
Tota
l irre18
(b) Fig. 19. Effect of inlet temperature of external fluid in (a) evaporator and (b)condenser.
fluid is kept constant). As the inlet condenser temperature in-creases, the condenser’s irreversibility rate increases proportion-ally with total system irreversibility. For a unit change incondenser’s irreversibility, the total irreversibility of the system in-creases by 3.8 times (CSB value for condenser).
Thus values of CSB for condenser and evaporator are 3.8 and 3.4respectively. Based on these values, it can be concluded that con-denser is relatively more sensitive component as compare toevaporator.
3.3. Economic analysis
A fair preliminary comparison of two alternative refrigerationsystems i.e., VCRS and CVCAS is done for water chiller whereinthe capacity of both the systems is same. The other operating con-ditions for both the systems are also same (Table 5).
The initial cost of the VCRS includes high efficiency reciprocat-ing chiller, air fan and chilled water pump where as the cost ofCVCAS includes the cost of VCRS (includes high efficiency recipro-cating chiller, cascade condenser and chilled water pump) andVARS (includes the cost of single effect absorption system, air fanand solar system to supply heat to the generator). The operatingcosts for both the systems include the cost of electricity for runningthe compressor. The running costs of auxiliary items, maintenancecosts, replacement cost, salvage cost etc. are not considered in thispreliminary study [35]. The economic analysis data for VCRS is gi-ven below,
(b) Fig. 20. Plot for determining the CSB of the (a) evaporator with Tin, evap as a variable(Tin, cond = 35 �C), and (b) condenser with Tin, cond as a variable (Tin, evap = 4 �C).
698 V. Jain et al. / Energy Conversion and Management 75 (2013) 685–700
� Capital investment cost, CICVCRS = Rs. 300,000 [Uppal Refrigera-tion & Eng. Co., India, personal communication, May 23, 2013](Rs. 58 = $1).� Compressor power, WVCRS = 24.04 kW.� Let the compressor works 15 h/day for 26 days in a month [35].� Total compressor power per day = 15 � 24.04 � 361 units.� Commercial electrical charge per unit [35] = Rs. 5.� Total electric cost per day for compressor = 361 � 5 = Rs. 1805.� The cost electric charge for compressor per month = Rs.
1805 � 26 = Rs. 46,930.� The annual electricity cost for compressor, ECVCRS = Rs. 563,160.
The economic data for CVCAS is given below,
� Capital investment cost, CICCVCAS = Rs. 1,500,000 [Uppal Refrig-eration & Eng. Co., India, personal communication, May 23,2013].� Compressor power, WCVCAS = 9.41 kW.� Total compressor power per day = 15 � 9.41 � 142 units.� Total electric cost per day for compressor = 142 � 5 = Rs. 710.� The cost electric charge for compressor per month = Rs.
710 � 26 = Rs. 18,460.� The annual electricity cost for compressor, ECCVCAS = Rs. 221520.� Net annual savings = Annual power cost for the VCRS � Annual
power cost for proposed CVCAS = Rs. 563,160 � Rs.221,520 = Rs. 341,640.
The cost of VCRS and CVCAS quoted above is applicable only inIndia and it may vary from region to region. Payback period is the
period during which the initial investment of the proposed systemcan be recovered. Firms expect to receive a pay back of their capitalexpenditure in a minimum period of time.
Payback period;PP ¼ Total cost of CVCAS=Annual savings fromthe proposed CVCAS � 4:5years ð16Þ
Using the techno-economic analysis of the proposed CVCAS, itcan be concluded that, even though the space required and initialinstallation cost for the proposed system is high, the shorter pay-back period makes it a commercially viable option. Payback periodcan be further reduced by reduction in fixed cost of CVCAS systemthrough mass production.
Hence after 3.5 years, CVCAS will be more beneficial as compareto VCRS.
3.4. Uncertainty analysis
The present work is based on thermodynamic modeling of va-por compression–absorption cascaded refrigeration cycle in EES.The errors in the results are introduced because of the equationof state or correlations used for the property calculations whichare inbuilt in EES. The correlations are based on the experimentaldata having percentage deviation from the experimental data.The property correlations of H2O–LiBr refrigerant absorbent pairinbuilt in EES are referred from Patek and Klomfar [36]. The errorsreported by Patek and Klomfar [36] in calculation of properties ofH2O–LiBr refrigerant absorbent pair, are specified in Table 9.
Similarly there are uncertainties associated with the correla-tions of water and steam. These are based on highly accurate ther-modynamic properties of water substance with the 1995formulation for the thermodynamic properties of ordinary watersubstance for general and scientific use, issued by the InternationalAssociation for the Properties of Water and Steam (IAPWS) [37].These uncertainties for the operating region of vapor absorptionsystem are specified in Table 10.
The uncertainty associated with the properties of R22 (inbuilt inEES software) is taken from Wagner et al. [38]. The uncertaintiesfor the operating region of vapor compression system are specifiedin Table 11.
To compute the overall uncertainty due to the combined effectof uncertainties associated with different variables, let us considerthe following equation in most general form.
y ¼ f ðx1; x2; . . . xi; . . . xnÞ ð18Þ
where y is a parameter that depends on the independent variables,x1, x2,. . ., xi,. . ., xn.
Thus Uy is the uncertainty in y due to the combined effect of dif-ferent variables as given in Refs. [39,40]
Uy2 ¼ �fðdy=dx1Þ2ðUx1Þ2 þ ð@y=@x2Þ2ðUx2Þ2
þ :::ð@y=@xiÞ2ðUxiÞ2 þ :::ð@y=@xnÞ2ðUxnÞ2g ð19Þ
Based on the above equation, the uncertainties in the values ofmref,VCRS, mref,VARS, m8 and m9 are calculated at design point usingthe mass and material balance equations. The uncertainty valuesfor masses are given below.
Table 9Error/uncertainty in various property correlations of H2O–LiBr pair [36].
S. No. Property Error/uncertainty in the property
1. Density ±0.5%2. Pressure ±2.1%3. Isobaric heat capacity ±2%4. Enthalpy ±10 kJ /kg5. Entropy ±0.03 kJ/kg K
Table 10Error/uncertainty in various property correlations of water/steam [37].
S. No. Property Error/uncertainty in the property
1. Specific enthalpy ±1 kJ/kg2. Density ±0.03%3. Pressure ±0.025%
Table 11Error/uncertainty in various property correlations of R22 [38].
S. No. Property Error/uncertainty in the property
1. Specific enthalpy ±0.25 kJ/kg2. Density ±0.1%3. Heat capacity ±1%
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Using the above values, the uncertainty in COPVCRS of compres-sion section is found to be ±0.1537. Similarly, the uncertainty anal-ysis for other parameters of CVCAS is also carried out and it isfound that the uncertainty in COPCVCAS, COPVARS, W, Qa, Qg and Qcond
are ±0.02542, ±0.03426, ±0.2033 kW, ±0.4.968 kW, ±4.97 kWand ±0.2567 kW respectively. Using EES for thermodynamic mod-eling of vapor absorption refrigeration system with H2O–LiBrworking pair, an uncertainty of ±0.036 was reported [41] in the va-lue of COP. This is different from the uncertainty value obtainedherein as the operating conditions used in the present work arequite different.
4. Conclusions
In this communication, an extensive thermodynamic study ofvapor compression–absorption cascaded refrigeration system hasbeen presented. From the comparative performance study of VCRSand CVCAS, the following conclusions are drawn:
1. Electric power requirement in VCRS is reduced substantially bycascading it with absorption system. However, the total size ofthe VCRS will increase when it is cascaded, but the running costis expected to decrease due to the utilization of waste heatavailable at lower cost. In addition, the COP of vapor compres-sion section will also increase because of low electric powerrequirement.
2. Subcooling improves system performance, whereas, superheat-ing is found to degrade the system performance.
3. The larger the temperature difference in cascade heat exchan-ger, the lower the COP of the system; however, a lower temper-ature difference will lead to increased heat exchanger size andcost.
4. Increasing the size of heat exchanger increases the overall per-formance of system, but it also increases the system cost.
5. Using the technique of CSB, the thermodynamic performance ofcondenser is found to be more sensitive to external fluid tem-perature as compared to evaporator.
6. From the comparative performance of various refrigerants, itcan be concluded that the performance of all the selected alter-native refrigerants is nearly same as R22. Hence, it may be con-cluded that the refrigerants considered here have the potentialto substitute R22.
7. The initial installation cost of the proposed CVCAS is high butthe shorter payback period makes it a commercially viableoption.
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