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Case Studies in Thermal Engineering

Case Studies in Thermal Engineering 7 (2016) 25–35

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Performance evaluation of combined ejector LiBr/H2Oabsorption cooling cycle

Hasan Sh. MajdiAl-Mustaqbal University College, Babylon, Iraq

a r t i c l e i n f o

Article history:Received 16 November 2015Received in revised form4 January 2016Accepted 12 January 2016Available online 12 January 2016

Keywords:LiBr/H2O absorption systemEnhance COPImproved refrigerant qualityCombined ejector absorption cycle

x.doi.org/10.1016/j.csite.2016.01.0037X/& 2016 The Authors. Published by Elsevireativecommons.org/licenses/by-nc-nd/4.0/)

ail addresses: Hasanshker1@gmail.com, info@

a b s t r a c t

The objective of this work is to develop a computer simulation program to evaluate theperformance of solar-assited combined ejector absorption (single-effect) cooling systemusing LiBr/H2O as a working fluid and operating under steady-state conditions. The ejectorpossess no moving parts and is simple and reliable, which makes it attractive for com-bination with single-stage absorption cycle for further improvement to the system'sperformance. In this research, improvement to the system is achieved by utilizing thepotential kinetic energy of the ejector to enhance refrigeration efficiency. The effects ofthe entrainment ratio of the ejector, operating temperature, on the thermal loads, andsystem performance have been investigated. The results showed that the evaporator andcondenser loads, post-addition of the ejector, is found to be permanently higher than thatin the basic cycle, which indicates a significant enhancement of the proposed cycle andthe cooling capacity of the system increasing with the increase in evaporator temperatureand entrainment ratio. The COP of the modified cycle is improved by up to 60 % comparedwith that of the basic cycle at the given condition. This process stabilizes the refrigerationsystem, enhanced its function, and enabled the system to work under higher condensertemperatures.& 2016 The Authors. Published by Elsevier Ltd. This is an open access article under the CC

BY-NC-ND license (http://creativecommons.org/licenses/by-nc-nd/4.0/).

1. Introduction

Many researches are recently interested in applications of solar-assisted absorption cooling/refrigeration systems and theimprovement of their corresponding performance [1,2], as it saves energy and is environmentally friendly. Solar energy isavailable in most areas and regarded as a good source of thermal energy. For many solar absorption cooling systems, LiBr/Water and Ammonia/Water is a major working fluid pair in the context of these systems. The absorption performance of acooling system is critically dependent on the chemical and thermodynamic properties of the working fluid [3]. A funda-mental requirement of absorbent/refrigerant combination is that, in its liquid phase, they must have a margin of miscibilitywithin the operating temperature range of the cycle. The mixture should also be chemically stable, non-toxic, and non-explosive. In solar applications, the LiBr–H2O system is superior to the NH3–H2O system, due to its simpler design andoperation and low cost. Moreover, it is functional under low generator temperature and perform better than that ofNH3–H2O [4]. The LiBr/water system has been widely used for many years and their properties are well established. Manytypes of absorption cycles have been developed, however, the system's complexities increased over a conventional single-effect absorption system. The double-effect absorption systems using lithium bromide/water seem to be a high performance

er Ltd. This is an open access article under the CC BY-NC-ND license.

mustaqbal-college.com

Nomenclature

ARS Absorption refrigeration systemAt Cross sectional area at nozzle throat (m2)Ak Cross sectional area at diffuser-inlet (m2)COP Coefficient of performanceh Enthalpy (kJ/kg)m Mass flow rate (kg/s)M Mach numberP Pressure (kPa)Q Thermal load (kJ/kg)SHE Solution heat exchangerT Temperature (°C)Tabs Absorber temperature (°C)Tcond Condenser temperature (°C)Tevp Evaporator temperature (°C)Tgen Generator temperature (°C)v Specific volume (m /3 kg)X Solution concentration

Y Vapor solution concentrationε Effectiveness of heat exchangerω Flow entrainment ratio, secondary stream to

primary stream

Subscripts

a Absorberp Primaryc Condensere Evaporatorg Generatorm Mixind flowv Vapor solutionl Liquid solutions SeconderyBC Basic cycle1–10 State points

H.Sh. Majdi / Case Studies in Thermal Engineering 7 (2016) 25–3526

system, which is available commercially. The (single-effect) cooling system using LiBr/H2O as its working fluid system canprovide COP as high as a double-effect system at minimal increase of system complexity. Thus, research has focused onimproving these absorption cooling systems by enhancing the coefficient of performance (COP) of these systems withrespect to both heating and cooling applications [5,6]. Ventas et al. [7] numerically studied single-effect absorption cyclesand utilized ammonia-lithium nitrate solution as its working pair to determine the effect of mass flow rate recirculatingthrough the absorber, as well as system performance. Numerous works also aim to optimize the operating parameters toenhance performance and minimize energy consumption in ARS [8–16].

Kaynakli and Kilic [17] numerically investigate the effect of operating temperature and the effectiveness of the heatexchangers in the ARS on the thermal loads of the components, the coefficient of performance (COP, COPc), and theefficiency ratio (η) with water/lithium bromide. In their study, SHE affected the investigated parameters more than therefrigerant heat exchanger (RHE); SHE increased the COP to a maximum rate of 44%, whereas the RHE increased this valueby only 2.8%. Moreover, ejector cooling systems can operate given refrigerants with low boiling points as working fluids.These systems are among the most promising devices for solar cooling/refrigeration applications. Previous studies indicatethat the COP values of absorption and ejector refrigerators falls within 0.5–0.85, whereas those of the basic absorptioncycle range from 0.2 to 0.6 [3,18–20]. The advantage of using an ejector refrigerator cycle can utilize thermal energy attemperature levels upward from 333 K [19]. Heat energy at these temperatures is available from the absorber of theabsorption refrigerator. Arbel and Sokolov [21] proposed combined systems, using the booster sub-cycle as a compressionenhanced approach of increasing the COP of ERS. Yu et al. [22] evaluated a new ejector refrigeration system with anadditional liquid–vapor jet pump. The jet pump was used to decrease the backpressure of the ejector, and then theentrainment ratio and the coefficient of performance (COP) of the new system could be increased. Abdulateef et al. [23]developed a model by adding an ejector between the generator and the condenser to improve the system's performance.Sirwan et al. [24,25] showed that the performance of a combined absorption–ejector cooling cycle can be further enhancedby adding a flash tank to the cycle in the presence of the ejector. Abed et al. [26] modified the combined ejector-flash tankabsorption cycle by utilizing heat recovery via the rearrangement of the streamlines at the solution heat exchanger and theaddition of RHE. Recently, Abed et al. [27] reported that the ejector efficiency could be optimized by removing the boosterfrom the refrigerant streamline side and allowing the secondary flow of the ejector to work only under the intermediatepressure of the flash tank with NH3–H2O as its working fluid.

This work conceptualizes and analyze the single effect absorption refrigeration system combined with the ejector and anupgraded system that considerably lower the energy footprint. This paper aims to improve the system's performance byusing the potential kinetic energy from the primary streamlines to drive the low pressure secondary streamlines to producea quieter flow with more thrust and lower temperature. The proposed cycle will be evaluated via the entrainment ratio andthe effect of the operation temperature on the thermal loads and the system's performance. A general description of theproposed system will be presented, and further analysis on its performance is conducted based on a correspondingmathematical model.

Fig. 1. System schematic of single effect combined absorption–ejector cooling system.

H.Sh. Majdi / Case Studies in Thermal Engineering 7 (2016) 25–35 27

2. Materials and methods

The binary mixture LiBr/H2O were used in the proposed system. The detailed thermodynamic property of LiBr/H2O werecoded with the help of correlations proposed by Patek and Klomfar [28], and Engineering Equation Software [29]. Theseproperties (i.e., pressure,temperature, concentration, enthalpy, and density) are necessary for the simulation to calculatethe heat and mass balance for the proposed cycle. A computer simulation program has been developed to evaluate theperformance of the system. The mass balances, energy balances, and the equations of state for LiBr/H2O solution, andrefrigerant at each component involved in the cycle are required for system simulation.

3. Description of absorption cycle

3.1. Description of solar single effect absorption cycle

The basic cycle consists of generator, absorber, a condenser, an evaporator, solution heat exchanger, circulating pumps,and a solar collector. The cycle works between two pressure levels: low pressure at the evaporator-absorber, and highpressure at condenser-generator. The emitted vapors in the generator consist of pure water, while LiBr sal remained in thesolution.

3.2. Description of modified single effect combined absorption–ejector cooling system

The main part of the single effect absorption cooling system is shown schematically in Fig. 1. The solution heat exchangeris advantageous, because it cools down the solution coming from the generator, which then heats up the solution enteringthe generator [30]. An ejector was added between the generator and the condenser. The primary high-pressure water vaporfrom the generator enters the ejector, then from the secondary inlet of the ejector, the entrainment low pressure watervapor from the evaporator is mixed with the primary flow at the mixing chamber, passing the diffuser, then entering thecondenser.

In this new design, the lithium bromide mixture leaves the absorber (state 1) in the form of a saturated solution at lowpressure. It is pumped to the system at high pressure (state 3). The generator operates from a high temperture source toseparate the binary solution of water and Lithium bromide (strong solution comes from absorber). This two-phase mixtureis separated, and the weak liquid flows through SHE (state 4 to state 5) than throttled to the low pressure system andsprayed into the absorber (state 6). On the ejector, the secondary flow (water vapor from evaporator (state 10A)) and theprimary flow of water vapor from the generator are mixed and passed to the condenser (state 7ʹ).

H.Sh. Majdi / Case Studies in Thermal Engineering 7 (2016) 25–3528

4. Simulation and analysis of proposed cycle

For the purpose of simulation and analysis, the following assumptions are made:

1. The system operates under a steady state condition.2. The refrigerant leaving the condenser and evaporator is saturated (state points 8 and 10).3. LiBr/H2O solution in the generator, solution heat exchanger, and absorber are assumed to be in the equilibrium state at

their respective pressure and temperature and assumed saturated state.4. The frictional pressure drop in the cycle is neglected except through the expansion device.5. The flow inside the ejector is steady and one-dimensional. The ejector walls are adiabatic.6. The primary flow and the secondary flow are saturated and their velocities are negligible before entering the ejector

(states 7 and 10B in Fig. 1 respectively). The velocity of the mixed flow leaving the ejector (at state 7ʹ) is also neglected.

In order to calculate the heat and mass balance for the proposed cycle, the thermodynamic properties (pressure,temperature, concentration, enthalpy and density) are necessary for the simulation. The binary mixture of LiBr/H2O andpure H2O are used in the proposed system. The detailed thermodynamic property equations of LiBr/H2O are found byEngineering Equation Software [29].

4.1. Thermodynamic analysis

The coefficient of performance (COP) is used to measure the system performance:

=+ ( )P

COPQ

Q 4.1e

g work

In order to use Eq. (4.1), mass and energy conservation should be determined at each component.

– For the generator – SHE – Absorber loop

The mass and energy balances around the generator

( ) = + ( )m m m total mass balance 4.23 4 7

( = + ) ( )m x m x m x H Omass balance 4.33 3 4 4 7 7 2

(where x3¼x1 and x4¼x6)

= + − ( )Q m h m h m h 4.4g 4 4 7 3 3

The fluid properties in this loop can be derived and developed as:The liquid weak solution at state (4)T4¼Tgen, P4¼Pgen, x4¼xs (Tgen, Pgen), and h4 ¼ hS(Tgen,x4)

– Heat exchangerSHE performance is expressed in terms of effectiveness ϵshe.The solution and refrigerant heat exchanger performance, expressed in terms of an effectiveness ϵshe.

ϵ =( − ) ( − ) ( )T T5 T T4 / 4 2 4.5she

( )= − ( − ) ( )C m h h T T/ 4 5 4.6hot 4 4 5

( )= − ( − ) ( )C m h h T T/ 3 2 4.7cold 2 3 2

( )= − ( )Q m h h 4.8hx 1 3 2

( )= − ( )Q m h h 4.9hx 4 4 5

where,

( ) = = = = = =m m T T T h h T x x x0. 05kgs

, , ,S1 2 1 abs 2 5 5 5 5 4

– Solution expansion valve model

= = =h h m m x x,5 6 5 6, 5 6

– Pump calculation

= + ( )h h P m/ 4.102 1 work 1

( )= – ( )P m v P P /1000 4.11work 1 1 high low

= ( ) ( )v v T x, 4.121 S 1 1

H.Sh. Majdi / Case Studies in Thermal Engineering 7 (2016) 25–35 29

– Absorber

= + − ( )m m mQ h h h 4.13a 10 10 6 6 1 1

– Ejector–Condenser – evaporator loopWith regard to the cycle layout, the refrigeration vapor produced from the generator induces the primary ejector flow. Itthen entrains the secondary vapor from the evaporator. Given the properties above, the ejector entrainment ratio, whichis expressed as ω= m /m10b 7, can be determined based on the following relationship:

⎛⎝⎜

⎞⎠⎟ω=

( )ʹf P , T , P , T , PAA 4.14

gen 7 10A 10A 7t

k

To calculate the entrainment ratio, we apply the performance evolution procedure for the ejector. We determine theentrainment ratio through iterative mathematical calculation, which defines ejector capacity and performance. Once weobtain ω, we derive the properties at states (10B) and (7) from the following mixing process:

= ω ( )m m 4.1510A 7 =( + ω) ( )ʹm 1 m 4.167 7

= +ω+ω ( )ʹh

h h1 4.177

7 10B

= = = + = h h h , m m m , m m10 10A 10B 10 10B 10A 10 9= ( = )h T Xh WATER, 10, 110 V

– Condenser

( ) ( )= − = = ( )′ ′Q m h h h h T X, WATER, 8, 0 4.18c 7 7 8 8 v

– Refrigerant valve

= =x xh h ,8 9 8 9

– Evaporator

= ( − ) ( )Q m h h 4.19Q 9 10 9

4.2. Ejector analysis

A one-dimensoional mathematical model for prediction of ejectoer performnce of Huang et al. [31] is used to analyze theeffect of mixing chamber and enterinment ratio on the performnce of system. The schematic diagram of the ejector used inthe present study is shown in Fig. 2.

4.3. Nozzle equations

For a givan pressure Pg ¼ P7, tempertature =T Tg 7 and = m mp 7, the nozzle throat area required for choking condition followsthe gas dynamic equation:

( )=

η ( )+

( + )( − )

A

P

m T

4.20

tp g

g pkR

2k 1

k 1k 1

where ηp is a coefficint relationg to the isentropic efficincy of the compressible flow in the nozzle. The relations betweenthe Mach number at the exit of nozzle Mp1 and the exit corss section area AP1 and pressure Pp1 are, using isentropic relationsas showing in the following eqautions:

Fig. 2. schematic diagram of the ejector.

H.Sh. Majdi / Case Studies in Thermal Engineering 7 (2016) 25–3530

⎡⎣⎢

⎛⎝⎜

⎞⎠⎟

⎤⎦⎥( ) =

++ ( −

( )

( + )( − )A

A M kk

M1 2

11

12 4.21

P

t pp

kk1

112

11

⎛⎝⎜

⎞⎠⎟= +

( − )

( )

( − )PP

k M1

12 4.22

g

p

pk

k

1

12 1

4.4. Mixing section equations

The Mach number Mpy of the primary flow at the y–y section follows the isentropic relations as an approximation

( )( )

=+ (( − ) )

+ (( − ) ) ( )

( − )

( − )

PP

k M

k M

1 1 /2

1 1 /2 4.23

py

p

p

kk

py

kk1

12 1

2 1

and the primary flow core at the y–y section, can expressed in terms of isentropic relation

⎛

⎝

⎜⎜⎜⎜⎜⎜

⎞

⎠

⎟⎟⎟⎟⎟⎟

⎛⎝⎜⎜

⎞⎠⎟⎟⎡⎣⎢

⎤⎦⎥

⎛⎝⎜⎜

⎞⎠⎟⎟⎡⎣⎢

⎤⎦⎥

( )( )

( )

( )=

+ (( − ) )

+ (( − ) )( )

∅( + )

+( ( − ))

( + )

+( ( − ))

AA

k M

k M

1 1 /2

1 1 /24.24

Py

p

M k py

kk

M k p

kk1

21

21

2 1

1 21 1

21

2 1

p

py

py

The arbiterary coefficint ∅p is included to account for the loss of the primary flow from section 1–1 to y–y.The entrained flow reaches chocking condition at the y–y section, i.e. =M 1sy . similarty to the primary nozzle, the

equivalent form of Mach number the secondry fluid at the nozzle exit plane is given as:

⎛⎝⎜

⎞⎠⎟= + −

( )( − )P

Pk

M11

2 4.25low

sysy

kk2 1

The entrained flow rate at choking condition follows

⎛⎝⎜

⎞⎠⎟ η=

+ ) ( )

( + ) ( − )

mP A

TkR k

21 4.26

slow sy

low

k k

s

1 / 1

where ηs is the coefficinet related to the isentropic efficincy of the entrained flow.The geometrical cross sectional area at section y–y is A3 that is the sum of the areas for the primary flow APy and for the

entrained flow Asy. That is,

+ = ( )A A A 4.27Py sy 3

Temperture and Mach number at section y–yThe temperture and the Mach number of the two stream at section y–y can be derived as

= + −( )

TT

kM1

12 4.28py

py2 2

= + −( )

TT

1k 1

2M

4.29low

sysy

2

4.4.1. Mixed flow at section m–m before the shockTwo streams starts to mix from section y–y. Ashock then takes place with a sharp pressure rise at section s–s. A mo-

mentum balance relation thus obtained from:

⎡⎣ ⎤⎦∅ + =( + ) ( )m V m V m m V 4.30m p py s sy p s m

∅ = +( )

Where, 1.037 0.02857AA 4.31m

3

t

where Vm is the velocity of the mixed flow and ∅m is the coefficient accounting for the frictional loss [32]. Similarly, anenergy balance relation can be derived as

H.Sh. Majdi / Case Studies in Thermal Engineering 7 (2016) 25–35 31

⎛⎝⎜⎜

⎞⎠⎟⎟

⎛⎝⎜⎜

⎞⎠⎟⎟

⎛⎝⎜

⎞⎠⎟ + + + =( + ) +

( )m C T

Vm C T

Vm m C T

V2 2 2 4.32

p p pypy

s p sysy

p s p mm

2 2 2

where Vpyand Vsy are the vapor velocity of the primary and secondery flow at the section y–y.

= × = ( )V M a a kRT 4.33py py py pypy

= × = ( )V M a a kRT 4.34sy sy sy sysy

The Mach number of the mixed flow can be evaluted using the following relation:

= =( )

MVa

a kRT4.35m

m

mm m

4.4.2. Mixed flow across the shock from section m–m to section 3–3A sharp pressure rise occur at section s–s due to a supersonic shock will take place at this section in the constant area

mixing chamber. Assuming that the mixed flow after the shock undergoing an isentropic process, the flow inside theconstent area mixing chamber between the section m–m and section 3-3 has a uniform pressure P3. Therefore, the pressure

Fig. 3. (a–d) Comparison of influence of operation temperature on the COP values for basic and modified cycle.

H.Sh. Majdi / Case Studies in Thermal Engineering 7 (2016) 25–3532

ratio across the mixing chamber can be obtained from:

= ++

( )( )

PP

12k

k 1M -1

4.363

mm

2

( )( )=

+

( )M

1 M

kM - 4.3732

k-12 m

2

m2 k-1

2

4.5. Diffuser equations

4.5.1. Pressure ratio across the subsonic diffuserThe pressure at the exit of the diffuser folllows the relation, assuming isentropic process

⎛

⎝⎜⎜

⎞⎠⎟= +

( )

PP

1k-12

M4.38

c

332

kk-1

4.6. Optimization of ejector geometry

Finally, using the above 1-D model of the ejector, we can carry out the performnce analysis to determine the entraimnentratio ω and the required cross sectional area of the constant area section A3 and nozzle throat area At. The consideredejector's thermodynamic parameters are its entrainment ratio

ω= ( )

mm 4.39

s

p

The ejector geometry ratio can determined from the following equation

⎛⎝⎜

⎞⎠⎟

( )( )

=( +ω)( +ω )

× ×( ) ( )

( )

+

( )

+

AA

PP

1

1 1

PP

1-

1- 4.40

t

3

c

2TT

py

c

1k

P

P

k-1k

2k 1

1/ k-1 2k 1

e

g

py

c

Eqs. (4.20)–(4. 40) will be solved by iteration to optimize the ejector geometry.

5. Results and discussion

Engineering Equation Software [29] is used to evaluate the performance of the system. The operating temperatures of theproposed cycles were set to Tgen¼60–110°C, Tcond¼Tabs¼20–50 °C, Tevp¼0–15 °C. The effectiveness values of SHE is pre-sumably 0.64. Given the operating condition and the cooling load, the software predicts the COP for the system.

5.1. Comparison between the basic and the modified cycle

Fig. 3(a–d) depicts the COP of basic cycle and the modified absorption cycle under various operating temperatures.Comparison of COP values vs. generator temperature for (basic and modified cycles) are shown in Fig. 3(a). It is noticed thatfor the two absorption cycle, there is an optimum value of COP. This value of COP increases with generator temperature untilit reaches the optimum value. This value depends on the type of the cycle. Moreover, there is a low generator temperaturelimit, where the the cycle cannot operate at a generator lower than this. This is an important point for the utilization solarenergy, since the fluid temperature for solar collector are generally below 100 °C. Therefore, the simulation exhibited in thisfigure reveals that modified cycle has a higher COP value than the basic cycle. The effect of evaporator temperature on theCOP of the cycles is shown in Fig. 3(b). It is evident that the COP value of two cycles increases as the evaporation tem-perature increases, as shown in Fig. 3(b). It also can be seen that the highest value of the COP is obtained from the modifiedcycle. Fig. 3(c and d). illustrate the comparison of COP value vs. condenser and absorber temperature for two cycles. Increasecondenser and absorber temperature causes a decrease in COP for each cycle. For modified cycle operation under chokedconditions, the cycle is independent of condenser pressure when the condenser pressure is less than a certain critical value;however, the cycle falls to similar values to the basic cycle when the condenser pressure is greater than this critical value, asshown in Fig. 3(c). It is observed that the absorber temperature of 44 °C and above the performance of the system tend tolean towards zero. This is due to the fact that the the concentration of the liquid leaving the absorber has dropped to thesame concentration as the liquid coming back from the generator. From this figure, it could be determined that the COP ofthe modified cycle is higher than that of the conventional cycles at all simulated working conditions. This is due to the factthat the performance of a refrigeration system is influenced by adding an ejector to the basic cycle.

Fig. 4. Variation of thermal loads and entrainment ratio vs.: (a) generator temperature, (b) evaporator temperature, and (c) condenser temperature.

H.Sh. Majdi / Case Studies in Thermal Engineering 7 (2016) 25–35 33

5.2. Effect of operating temperature on thermal loads

The effect of the generator, evaporator, and condenser temperature on the thermal loads of the condenser and theevaporator are shown in Fig. 4(a–c). It can be seen from Fig. 4(a and b) that when the generator and evaporator temperatureincreases, the thermal loads of the condenser and the evaporator increases for the two cycles. It is also shown that thegenerator temperature does not affect the entrainment ratio of ejector, while the entrainment ratio (ω) increases rapidlywith the evaporator's temperature, since an increase in the evaporator pressure causes an increase in the vapor flow ratefrom the evaporator entering the ejector. It can also be seen that the higher value of thermal loads is obtained with themodified cycle. The variations of the thermal loads with a condenser temperature is given in Fig. 4(c). It is shown that thethermal loads of the two cycle decrease with increasing condenser temperature. This is due to the fact that increasedpressure of the system will increase the amount of the saturated liquid enthalpies leaving the condenser. This leads to adecrease in the thermal load of the condenser. It is also shown that the entrainment ratio decreases, and hence the pressureor the ejector back pressure increases with increased secondary flow pressure at the evaporator, leading to low ejectorefficiency. The results also show that the new modified cycle has higher thermal loads than that of the basic cycle.

Fig. 5. Comparison of COP versus generator temperature at different operating conditions for modified and basic cycles.

H.Sh. Majdi / Case Studies in Thermal Engineering 7 (2016) 25–3534

5.3. Evaluate system performance at different operating temperature

The temperatures of the components that evaluate system performance falls within the following ranges: 70–110 °C forthe generator temperature; 25–45 °C for water-cooled condensers and absorbers; and 0–10 °C for evaporators. Thus, theideal cooling effect was obtained under the employed generator and ambient-condensation temperature ranges. Fig. 5(a–d)illustrates the comparison between the modified and the basic cycle at different condenser and evaporator temperatures.The proposed cycle enhances cycle peak under all operating conditions, therefore, the improvement in the modified cycleincreases the COP by 8–60 % under condenser temperature range, and by 30–85% under evaporator temperature range overthe basic cycle. This improvement in the overall COP is attributed to the ability of the ejector to achieve highest dischargepressure with the given amount of motive steam, thereby increasing the energy available to entrain and compress thesuction load.

6. Conclusions

In this study, an improved system of the single-stage absorption cycle operated with LiBr/H2O as its working fluid wasconducted. Thermodynamic analysis of ejector–absorption refrigeration system has been carried out, and the theoretical

H.Sh. Majdi / Case Studies in Thermal Engineering 7 (2016) 25–35 35

performance of the cycles were compared. The results show that the modified combined absorption cycle is superior to thebasic cycle over a wide range of operating conditions. The results also show that the thermal load of the condenser andthe evaporator increase as the generator and evaporator temperature increases, while it decreases as the condensertemperature increases. It is shown that the entrainment ratio of the ejector is unaffected by the generator temperature,while increases rapidly with evaporator temperatures, and decreases with condenser temperature, leading to a decrease inthe ejector’s efficiency. The results indicated that the overall COPs increments of the modified cycle was � 8–60 % at acondenser temperature of 25–45 °C, and by 30–85% at evaporator temperature of 0–10 °C over the basic cycle. Finally, thecontribution of the present study is the stimulation of wider interest in the technology of absorption refrigeration system,which can be useful for any newcomer in this field of technology, with the greatest benefits in hot climates.

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