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Research Papers: Energy Systems Analysis

Performance Analysis of a Solar-Powered Multi-Effect Refrigeration System PUBLIC ACCESS

[+] Author and Article Information
Ayman J. Alazazmeh

Mechanical Engineering Department,
College of Engineering,
King Fahd University of Petroleum and
Minerals (KFUPM),
P. O. Box 279,
Dhahran 31261, Saudi Arabia
e-mail: g201204580@kfupm.edu.sa

Esmail M. A. Mokheimer

Mem. ASME
Mechanical Engineering Department,
College of Engineering,
King Fahd University of Petroleum and
Minerals (KFUPM),
Dhahran 31261, Saudi Arabia;
Center of Research Excellence in Energy Efficiency (CEEE),
King Fahd University of Petroleum and
Minerals (KFUPM),
Dhahran 31261, Saudi Arabia;
Center of Research Excellence in Renewable Energy (CoRe-RE),
King Fahd University of Petroleum and
Minerals (KFUPM),
P. O. Box 279,
Dhahran 31261, Saudi Arabia
e-mail: esmailm@kfupm.edu.sa

Abdul Khaliq

Mechanical Engineering Department,
College of Engineering,
King Fahd University of Petroleum and
Minerals (KFUPM),
P. O. Box 279,
Dhahran 31261, Saudi Arabia
e-mail: khaliqsb@gmail.com

Bilal A. Qureshi

Mechanical Engineering Department,
College of Engineering,
King Fahd University of Petroleum and
Minerals (KFUPM),
P. O. Box 567,
Dhahran 31261, Saudi Arabia
e-mail: bqureshi@kfupm.edu.sa

1Corresponding author.

Contributed by the Advanced Energy Systems Division of ASME for publication in the JOURNAL OF ENERGY RESOURCES TECHNOLOGY. Manuscript received October 17, 2018; final manuscript received December 5, 2018; published online January 9, 2019. Editor: Hameed Metghalchi.

J. Energy Resour. Technol 141(7), 072001 (Jan 09, 2019) (13 pages) Paper No: JERT-18-1790; doi: 10.1115/1.4042240 History: Received October 17, 2018; Revised December 05, 2018

The main objective of the current work is to investigate the thermodynamic performance of a novel solar powered multi-effect refrigeration system. The proposed cycle consists of a solar tower system with a heliostat field and central receiver (CR) that has molten salt as the heat transfer fluid, an absorption refrigeration cycle (ARC), an ejector refrigeration cycle (ERC), and a cascade refrigeration cycle (CRC). Energy and exergy analyses were carried out to measure the thermodynamic performance of the proposed cycle, using Dhahran weather data and operating conditions. The largest contribution to cycle irreversibility was found to be from the CR system (52.5%), followed by the heliostat field (25%). The first and second-law efficiencies improved due to the increase in the following parameters: ejector evaporator temperature, turbine inlet and exit pressures, and cascade evaporator temperature. Parametric analysis showed that the compressor delivery pressure, turbine inlet and exit pressures, hot molten salt outlet temperature, and ejector evaporator temperature significantly affect the refrigeration output.

Solar cooling is a very attractive technology due to the coincident peak cooling demand and peak solar energy availability, so that more solar energy can produce more cooling. Air-conditioning systems are responsible for more than 50% of the electricity consumption in residential and commercial buildings. Therefore, researchers are focused on finding solutions that can provide the required refrigeration effect with minimal power from main grid [1]. Solar cooling is a cost-effective and clean technology and offers environmental benefits such as reducing harmful emissions produced from burning fossil fuel to produce electricity and reduce greenhouse gases. Solar cooling offers strategic benefit such as reducing dependence on the main grid demand to provide power for air-conditioning systems, which will shift the load during peak usage and reduce the requirement of new power plants.

Many applications require cooling at different temperatures and refrigeration capacities. Through the use of solar energy, multi-effect refrigeration systems are a solution for these requirements [2]. The continuous availability of solar energy is its most attractive feature and that it does not cause air pollution or produce harmful emissions. Also, it is easy to collect since it is available abundantly in the Arabian Gulf.

Solar cooling may be divided into three major categories: thermal, electrical, and combined power and cooling cycles. Solar cooling systems may be classified into three main parts: the application, refrigeration cycles, and collecting element. Alazazmeh and Mokheimer [3] provided an intensive review of solar cooling technologies. They presented descriptions, applications, and temperature ranges for each technology. Furthermore, they explained the working principle of multi-effect refrigeration cycle, combined power and cooling cycles, and individual cycles such as absorption, adsorption, ejector, and conventional combined Rankine cycle with ejector and vapor compression cycle. Lu and Goswami [4] presented an energy analysis for a power and refrigeration cogeneration system that included a solar collector heat supply at 90 °C. The result of the thermodynamic analysis showed that the first law efficiency varied from 10.5% to 15.7% for an ambient temperature variation of 7 °C–27 °C. Wali [5] studied the safety requirement of the refrigerant and disclosed that halocarbon and fluorinated compounds meet these requirements. Kane et al. [6] presented an energy analysis of an organic-Rankine cycle, which included a hermetic-scroll expansion device with a generator and solar energy unit.

Many applications require high refrigeration capacity, which can be achieved by combining Rankine and vapor compression cycles. It has disadvantages such as high installation cost, more complex maintenance requirements due to many moving parts of this system. Agrawal and Karimi [7] presented a triple effect refrigeration cycle where waste energy from a power plant would be utilized to supply thermal energy to the generator of a vapor absorption refrigeration system. It would also provide a sufficient amount of thermal energy for a steam turbine to produce power in order to run the compressor of a cascade refrigeration cycle (CRC) while the remaining thermal energy went to an ejector refrigeration cycle (ERC). The analysis of first and second law was performed for all the components of these cycles. It highlighted the importance and effect of some influential operating parameter on these cycles. Some of these parameters were ejector and cascade cycle evaporator temperature, gas temperature of waste energy, inlet and outlet turbine pressure, and discharge pressure of the compressor. Higher GWP of N2O compared to other refrigerants and its toxic nature were the major drawback in this novel cycle. Fan et al. [8] proposed a combine cycle of ejector and absorption refrigeration systems. The cycle analysis showed that the proposed system has the ability to meet the cooling and refrigeration requirements without affecting the environment.

Khaliq et al. [9] studied a triple refrigeration-effect cycle and conducted by an analysis of first and second law for all the components. The results showed that the exergetic analysis was more important than the first law analysis for the purpose of pointing out thermodynamic imperfections in the cycle components. Improving the components with the highest exergy destruction would result in enhanced system performance. Agrawal et al. [10] conducted a thermodynamic investigation of a solar driven multi-effect refrigeration cycle. The central receiver (CR) and heliostat field were found to have the highest level of imperfection. First and second-law efficiencies calculated in this study were 11.5% and 2%, respectively.

In this paper, the newly developed solar powered multi-effect refrigeration cycle is studied based on the first and second law of thermodynamics. A complete description of the proposed cycle is provided in Sec. 2. Section 3 briefly describes the mathematical framework for the system. In Sec. 4, the results from these investigations are presented and discussed. Finally, Sec. 5 provides some conclusions based on the current study.

The solar-driven triple effect refrigeration cycle involves a solar tower with a central receiver connected to a steam Rankine cycle (SRC) providing thermal energy to an absorption refrigeration cycle (ARC), a CRC, and an ERC. A schematic of the complete cycle is shown in Fig. 1.

The aperture of the solar-tower CR receives solar radiation that is reflected from the mirror of a heliostat field where molten salt (a 60/40% mixture of NaNO3/KNO3 by weight) absorbs it. The key advantage in using molten salt is that it is cheaper, has good thermal storage, and can work effectively for 24-h operation. The molten salt absorbs thermal energy at the CR, which raises its temperature up to 560 °C. It then goes through the heat recovery vapor generator (HRVG) (1–2). Heat is transferred in the SRC to the water, which produces power in a steam turbine. The steam then passes through the ERC and a water-cooled condenser providing refrigeration in the ejector cycle.

The temperature of the molten salt decreased to approximately 315 °C, and then, it flows to the generator (2–12) to provide the required heat energy to separate LiBr and water [1116].

The molten salt reaches its minimum operating temperature, which is around 290 °C before freezing. Then, it flows back to CR to absorb the solar radiation and passes again to the HRVG.

The water refrigerant in the ejector cycle will become superheated vapor (4) after heat transfer takes place in the HRVG with molten salt. Then, the superheated vapor expands in a turbine to produce work that will be used to run the cascade cycle.

The steam exiting the turbine (5) passes through the ejector. This steam will be at a very high speed, which creates high vacuum to entraining secondary vapor (11) from the evaporator-1 (E-1) of the ejector cycle.

In the ejector, the primary (5) and secondary fluid (11) are mixed. This fluid (6) then flows to the condenser-1 (C1) where condensation takes place as the condenser is cooled by means of cold water from the cooling tower, and the temperature of the mixed fluid reduced.

The mixed stream leaving the condenser is a saturated liquid (7) and is separated into two streams (8, 9). The first stream (9) flows through the expansion valve-1 (TV1), which drops its pressure to that of the evaporator (10), and then, it goes to E-1. Pump-1(P1) is used to send the second stream (8) to the HRVG.

The superheated water vapor (13) leaves the generator after the separation process, and it will get cooled in condenser-2 (C2) by means of cold water from the cooling tower. The water refrigerant leaving the condenser is a saturated liquid (14). It then goes into the throttle valve-2 (TV2) to produce low-pressure liquid (15), which passes to evaporator-2 (E2). The water refrigerant leaving the evaporator-2 (E2) will be vapor saturated (16), and then, it will flow into the absorber. The weak LiBr-H2O solution (20) flows into the heat exchanger from the absorber to the generator. Now, due to heat transfer in the solution heat exchanger, the strong solution will be cooled to (21) and the weak solution will be heated. Then, the strong solution passes through throttle valve-3 (TV3) back to the absorber, and the weak solution will pass to the generator.

The absorber temperature will be maintained at 35 °C and the same pressure of evaporator-2. At the absorber, the saturated vapor (16) leaving evaporator-2 (E2) and the strong solution (22) are mixed to create a new mixture (17). This flows into the second pump (P2), following the solution heat-exchanger (SHX), and then finally passes into the generator (19) for the separation process.

The turbine power produced is used to run both the CRC compressors (COMP-1 and COMP-2). Low-temperature refrigeration is provided by using nitrous oxide (N2O) as the working fluid in both the low and high-temperature cycles. The temperature and pressure of superheated N2O vapor (29) are increased through compression (30) in compressor-2 and is then cooled in the cascaded heat exchanger, to become saturated liquid (31).

The nitrous oxide then flows through the internal heat exchanger-2 (IHX 2) (32) and cooled further. After that, the nitrous oxide passes to the throttle valve-5(TV5) (33). The nitrous oxide evaporates from state 33 to state 34 to get a cooling effect in the evaporator-3 (E3). N2O vapor (23) is first compressed in compressor-1 (24), and then, it is cooled in condenser-3(C3) (state 24 to state 25). N2O goes through the internal heat-exchanger-1 (IHX 1) and is cooled further. N2O (26) flows through the throttling valve-4 (TV4) and heats up (27–28) in the cascade heat exchanger.

In this paper, a multimode thermodynamic cycle is presented for solar power cooling that can provide a refrigeration effect at various capacities simultaneously at various temperature levels.

The cycle can meet the cooling demand for air conditioning (15–22 °C), refrigeration for food preservation (2–8 °C), and deep freezing for vaccine preservation and pharmaceutical plants (−50 °C to −80 °C). This cycle employs solar tower technology, which will use molten salt as a heat transfer fluid and integrates a heliostat field and central receiver with the SRC, ERC, ARC, and CRC. Table 1 shows the working fluids used in each cycle.

Due to the large number of components and multiple subsystems, only the most important equations will be mentioned below along with the main assumptions in practice [933].

Following are the main assumptions used in the current study for the cycle investigated:

  1. (1)All processes are taken as steady-state.
  2. (2)Pressure drop in the pipes is neglected.
  3. (3)The heat losses to the surrounding are neglected in the steam turbine, HRVG, condensers, and evaporators.
  4. (4)Chemical exergy for the solar heat source is considered negligible.
  5. (5)The saturated condition of the solution leaving the absorber and the generator.
  6. (6)The compression process is considered to be adiabatic.
  7. (7)Ejector walls are well-insulated, and the flow is considered as one-dimensional.
  8. (8)The saturated condition of the primary and secondary flow entering the ejector (states 5 and 11).

First-Law Analysis.

The efficiency of the solar assisted triple effect refrigeration cycle is as follows: Display Formula

(1)ηI=Q˙E1+Q˙E2+Q˙E3Q˙Sol

where Display Formula

(2)Q˙E1=m˙E1hahb=m˙sfh11h10
Display Formula
(3)Q˙E2=m˙E2hihj=m˙rh16h15
Display Formula
(4)Q˙E3=m˙E3hmhn=m˙N2O,2h34h33

In Eq. (1), the term in the denominator can be determined by understanding the optical and thermal losses from the heliostat and the CR (see Eq. (6)). Modeling related to the CR is dealt with in Secs. 3.1.13.1.6.

Central Receiver Energy Balance.

The delivered thermal power from the solar radiation can be determined from Display Formula

(5)Q˙Sol=Ahq

where Ah is the aperture area and q is the solar radiation per unit area Display Formula

(6)Q˙Sol=Q˙CR+Q˙lost,hs

where Display Formula

(7)Q˙CR=m˙msh1h12+Q˙lost,CR

Central Receiver Surface Temperature.

In order to calculate the heat losses from the CR, its surface temperature must be determined, which can be expressed as Display Formula

(8a)TCR,sur=Q˙CRACR,surdodihc,ms+do2ktlndodi+Tms
Display Formula
(8b)hc,ms=kmsdi0.023Rems0.8Prms0.4
Display Formula
(8c)Rems=diρmsumsμms
Display Formula
(8d)Prms=μmsCpmskms

where Eq. (8b) is the well-known Dittus–Boelter equation [30]. The required equations to determine properties of molten salt are given in the Appendix.

Central Receiver Emissive Heat Losses.

The average receiver surface temperature is used to determine the emissive heat losses. It is expressed by [31] Display Formula

(9a)Q˙CR,em=εavgσTCR,sur4To4ACR,sur
Display Formula
(9b)εavg=εwεw+1εwFr

where Fr is the view factor.

Central Receiver Convection Losses.

Forced convection heat losses were taken as forced convection over a flat plate and can be expressed by the following equation: Display Formula

(10a)Q˙CR,FC=hc,FC,aTCR,surT0ACR,sur
Display Formula
(10b)hc,FC,a=kaL0.0287Rea0.8Pra1/3

where L is the aperture height, and the convective-heat-transfer coefficient is determined using the correlation reported in Ref. [32]. The average of the receiver and ambient temperatures was used to determine the air properties (see the Appendix).

The natural convection heat losses inside the central receiver can be expressed as [32] Display Formula

(11a)Q˙CR,NC=hc,NC,aTCR,surT0ACR,sur
Display Formula
(11b)hc,NC,a=0.81TCR,surT00.426

Central Receiver Reflective Heat Losses.

Reflective heat losses considering surface reflectivity (ρr), and the view factor can be expressed as Display Formula

(12a)Q˙CR,ref=Q˙CRFrρr

where Display Formula

(12b)Fr=AhACR,sur

Central Receiver Conductive Heat Losses.

Conductive heat losses from the insulation layer can be expressed as follows: Display Formula

(13)Q˙CR,cd=kinsδinsTCR,surTins,wACR,sur

The heat transfer coefficient of the outer receiver insulation layer is calculated using the combined convective heat transfer. It can be expressed as

Display Formula

(14)kinsδinsTCR,surTins,wACR,sur=kaL0.0239Rea,o0.805(0.785Tins,wTO)0.21.167Pra,o0.45

Second-Law Analysis.

The second-law efficiency of the investigated cycle can be written as Display Formula

(15)ηII=ΔE˙E1+ΔE˙E2+ΔE˙E3E˙Sol

where ΔE˙E1 is the exergy change at the ejector evaporator of ERC,ΔE˙E2 is the exergy change at the evaporator of ARC, ΔE˙E3 is the exergy change at the evaporator of CRC, and E˙Sol is the exergy received through the solar radiation Display Formula

(16)ΔE˙E1=m˙sf[h10h11T0s10s11]
Display Formula
(17)ΔE˙E2=m˙r[h15h16T0s15s16]
Display Formula
(18)ΔE˙E3=m˙N2O,2[h33h34T0s33s34]
Display Formula
(19)E˙Sol=Q˙Sol1T0Ts

where Ts is the apparent sun temperature (equal to 4500 K [10]).

A part of the exergy gained by the heliostat is lost to the environment while the rest is supplied to the CR as shown below Display Formula

(20)E˙Sol=E˙CR+E˙lost,hs

where Display Formula

(21)E˙CR=m˙mscP,msT1T12T0lnT1T12+E˙D,CR

Exergy destruction in the ejector, generator, and compressors is, respectively, given by Display Formula

(22)E˙D,EJ=T0m˙6s6m˙pfs5m˙sfs11
Display Formula
(23)E˙D,Gen=T0m˙rs13s20+m˙ss20s19+m˙mss12s2
Display Formula
(24a)E˙D,cp,1=T0m˙N2O,1s24s23
Display Formula
(24b)E˙D,cp,2=T0m˙N2O,2s30s29

Exergy destruction in the above components can also be calculated by applying exergy balance on each component.

Before the results can be presented, the model needs to be validated. This is presented in Sec. 4.1.

Model Validation.

After modeling the investigated cycle, the next step is to verify the available models in the literature for each cycle individually. The present developed model has been validated against the previously published models in the literature, and excellent agreement with these models was revealed as reported hereunder. The present model of the investigated cycle was validated against the work of Agrawal et al. [10] and Jamel et al. [33].

Solar Central Receiver Model Validation.

The present solar central receiver model was compared with the model reported by Jamel et al. [33], and the comparative result was plotted to show the differences between both models. Figure 2 shows the variation of central receiver surface temperatures with the changing aperture area. It is observed that the difference between both models and the error percentage was less than 5%. Figures 3 and 4 depict the variation of the CR thermal efficiency and surface temperature with changing hot molten salt outlet temperature.

Absorption, Cascade and Ejector Refrigeration Cycles Validation.

The refrigeration output for ERC and ARC had been compared with the work of Agrawal et al. [10]. The following comparison shows the difference between the present model and the model reported by Agrawal et al. [10] varied with turbine inlet pressure and hot molten salt outlet temperature.

Figures 5 and 6 show excellent agreement between the present model and the model reported by Agrawal et al. [10] for the refrigeration output of ERC with variation of hot molten salt temperature and turbine inlet pressure. Table 2 shows the relevant percentage errors for the ERC, which is less than 5%.

Figures 7 and 8 show excellent agreement between the present model and the model reported by Agrawal et al. [10] for the refrigeration output of absorption refrigeration cycle with changing hot molten salt outlet temperature and turbine inlet pressure.

Table 3 shows the error percentage between the present model and the model reported by Agrawal et al. [10] for ARC, which is less than 4%.

Figures 9 and 10 show excellent agreement between the present model and the model reported by Agrawal et al. [10] for the refrigeration output of CRC with variation of hot molten salt temperature and turbine inlet pressure.

Table 4 shows the error percentage between the present model and the model reported by Agrawal et al. [10] of cascade refrigeration cycle, which is less than 5%.

Exergy Distribution.

A second-law analysis has been performed to determine the performance of the proposed refrigeration system by varying some influential operating parameters. Some parameters were varied over a typical operating range to find out their effects on the overall cycle, whereas others were kept constant. Exergy balances were used to calculate the second-law efficiency and exergy destruction of each component. Table 5 shows the main operating parameters used for the cycle.

The exergy analysis of the proposed cycle, as presented in Fig. 11, showed that 4.7% is available as an exergy output for ERC, ARC, and CRC while the remaining part 95.3% is exergy lost from the overall refrigeration system. Some important results from the analysis are given below:

  • The highest irreversibility in this novel cycle occurred in the CR, which was 52.5%.

  • The second highest irreversibility in this cycle occurred in the heliostat, which was 25%.

  • 11% of the irreversibility occurred in the ERC, 6.64% in the ARC, and 8.23% in the CRC.

  • The irreversibility contribution by the HRVG, ejector, and generator was found to be from 2% to 7%.

Effect on Performance Parameters.

Figure 12 shows the change in the refrigeration output for each cycle separately and the overall combined refrigeration cycle with the variation of outlet temperature of the hot molten salt.

Figure 12 shows that raising the hot molten salt outlet temperature (T1) increases the refrigeration output of CRC and ERC. This happens because a rise in outlet temperature of the hot molten salt results from heat transfer in the HRVG and improves the thermodynamic properties of the refrigerant at turbine inlet and exit state. Thus, the power output of the turbine increases, and as a result, the compressor will get more power, which increases the mass flow rate of N2O in the CRC and increases the cooling effect accordingly.

It is further observed that improving the thermodynamic properties of the vaporized refrigerant at the turbine outlet increases the speed of the fluid exiting the ejector nozzle and increases the entrainment ratio. The secondary fluid is entrained due to the resulting negative pressure (vacuum). Then, accordingly, there is a rise in the secondary fluid mass flow rate in the ERC, and the increment in secondary mass flow will provide more cooling effect in the ERC.

A lower refrigerating effect at the evaporator of the ARC is seen due to a rising outlet temperature of the hot molten salt. This happens because such a rise in temperature results in a lower temperature at the HRVG exit, which causes a low mass flow rate of water refrigerant going to the condenser (C2). Hence, a smaller refrigerating effect at the ARC evaporator is seen. Due to the fact that the change in the refrigeration effect of the CRC and ERC is less than that of the ARC, the overall combined cycle output is decreased with increasing hot molten salt outlet temperature. Therefore, it is seen that the total refrigeration effect is dominated by the ARC.

Figure 13 shows a significant reduction in energy efficiency with increasing hot molten salt outlet temperature, whereas the exergy efficiency increases insignificantly. This is the reason why the exergy output of ARC and ERC is less than that of the CRC.

Figure 14 depicts the variation in the refrigeration output due to changes in inlet turbine pressure on the of each cycle and the combined refrigeration output while Fig. 15 depicts the changes in the first- and second-law efficiencies.

Figure 14 shows that due to a rise in pressure at the turbine inlet, ejector, and cascade evaporator, the refrigeration output reduces while there is a substantial rise in the absorption evaporator refrigeration effect. This happens because such a pressure rise results in a decrease in the refrigerant mass flow in the HRVG. This produces two effects: (i) a lower temperature at the turbine exit, which also reduces the ejector nozzle velocity causing a decrease in the secondary refrigerant mass flow through the ejector evaporator, and (ii) the heat gained from the exhaust gasses, leading to a greater exit temperature at the HRVG. At the inlet of the generator, this larger temperature results in a significant rise in the (water) refrigerant mass flow rate and, thus, the ARC's capacity.

A lower refrigeration output of the CRC cycle is seen due to the smaller mass flow rate of N2O through the compressor. This is due to a smaller flow rate through the turbine. This will decrease the refrigeration output of the CRC because of the overall decrease in mass flow through the cycle. The overall refrigeration output of the combined cycle increases with a rise in pressure at the inlet of the turbine. This is because of the rise in ARC capacity, which is larger than the decrease in the CRC and ERC.

It is observed, from Fig. 15, that the first law efficiency rises notably, whereas the second law efficiency decreases slightly as the inlet pressure of the turbine increases.

The refrigeration output was also varied with respect to the turbine back pressure, compressor delivery pressure, and ejector evaporator temperature. The effect was not found to be as significant as those shown above and were, therefore, not shown here. Similarly, the first and second-law efficiencies were varied with respect to other parameters. Both the exergy and energy-based efficiencies increased due to a rise in any one of the following parameters, i.e., ejector evaporator temperature, turbine back pressure, and cascade evaporator temperature, whereas both efficiencies decreased for a higher compressor discharge pressure. These figures were not included here as the variation was not as significant compared to those shown above.

The refrigeration output was also varied with respect to the turbine back pressure, compressor delivery pressure, and ejector evaporator temperature. In this case, the effect was not found to be as significant.

The overall cooling capacity remains 2500 kW with varying the turbine back pressure from 220 to 300 kPa. The overall cooling capacity increases slightly from 2100 to 2600 kW with varying ejector temperatures from 1 to 5 °C. Similarly, the first and second-law efficiencies were varied with respect to other parameters. The thermal efficiency increases insignificantly from 32% to 34%, and the exergy efficiency remains at 3% with varying back pressures from 220 to 300 kPa.

The thermal efficiency of the ejector evaporator increases from 31% to 35% with varying ejector temperatures from 1 to 5 °C while the exergy efficiency was not changed and remains 4%. The thermal efficiency increases from 32% to 35%, and the exergy efficiency remains at 3% with the changing cascade evaporator temperature from −80 to −50 °C.

The thermal efficiency decreased slightly from 32% to 31%, and the exergy efficiency decreases from 4% to 3% with increased compressor discharge pressure from 5600 to 7300 kPa.

Figures 1619 show the effect of average daily solar radiation on the refrigeration outputs of CRC, ERC, ARC, and combined refrigeration output. It is seen that when the average daily solar radiation increases, it increases the refrigeration output in ERC, ARC, CRC, and combined cycle. The reason for the increase in the refrigeration output with increasing average daily solar radiation is because of the following facts:

  • When the average daily solar radiation increases, it will increase the HRVG thermal energy, which will increase the turbine mass flow rate and then increase secondary mass flow rate and the ejector cycle refrigeration effect accordingly.

  • When the primary mass flow rate increases across the turbine, this will raise the turbine output power, then the mass flow rate in the CRC and then the refrigeration output in CRC.

  • When the average daily solar radiation increases, it will increase the hot molten salt mass flow rate in the generator of the ARC and the thermal energy in the generator, which will increase the mass-flowrate of evaporator-2 and then increase the refrigeration output of the ARC.

Figures 16 and 17 are showing the refrigeration capacity of ARC, ERC, CRC, and the combined cycle with variation of average daily solar radiation for the entire year from January to December. It is noticed that with peak solar energy availability, the refrigeration effect for all cycles reaches their maximum capacity. The maximum average solar radiation was in June (768.2 W/m2), and the refrigeration output reached the maximum in this month, as follows:

  • 801.1 kW for the ejector refrigeration cycle.

  • 1493 kW for the absorption refrigeration cycle.

  • 177.4 kW for the cascade refrigeration cycle.

  • 2471.5 kW for the combined refrigeration cycle.

The minimum average solar radiation was in December (615.1 W/m2), and the refrigeration output was the minimum capacity in this month as follows:

  • 641.6 kW for the ejector refrigeration cycle.

  • 1195 kW for the absorption refrigeration cycle.

  • 142.1 kW for the cascade refrigeration cycle.

  • 1978.7 kW for the combined refrigeration cycle

We further investigated the effect of hourly solar radiation on the performance of the individual refrigeration cycles and the combined refrigeration cycle. For this purpose, we selected the average day in a summer month (June 11) and the average day in a winter month (December 10) for Dhahran (Saudi Arabia).

Figures 18 and 19 show the effect of average hourly solar radiation on June 11 (11 h from 6:00 to 18:00) on the refrigeration outputs of CRC, ERC, ARC, and combined refrigeration output. It is seen that when the average hourly solar radiation increases, it increases the refrigeration-capacity in ARC, ERC, CRC, and the combined cycle.

Figures 20 and 21 show the effect of average hourly solar radiation on December 10 (9 h from 7:00 to 17:00) on the refrigeration outputs of CRC, ERC, ARC, and combined refrigeration output. It is seen that when the average hourly solar radiation increases, it increases the refrigeration-capacity of ARC, ERC, CRC, and combined cycle.

In order to provide cooling at different temperature ranges, a new solar driven triple-effect refrigeration cycle was proposed. The analysis of first and second law was done on the cycle, and some important conclusions are given below:

  • The largest contributions to cycle exergy destruction came from the CR and heliostat field with values of 52.5% and 25%, respectively. Irreversibility on the order of 2–7% was observed in the HRVG, ejector, and generator of refrigeration cycle. These can be minimized by better design considerations.

  • The condenser and evaporator exergy losses are much less than generator and absorber losses. The losses (within the condenser and evaporator) are mainly due to heat of mixing in the solution, which is not present in pure fluid such as single refrigerant in vapor compression cycle. The effects of exergy losses in the solution, refrigerant pump, and expansion valves on the total exergy losses are small and considered negligible.

  • Around 4.7% was available as a useful exergy output. It was low due to the low-temperature refrigerating output of CRC.

  • Both the exergy and energy-based efficiencies rise by increasing any one of the following parameters, i.e., turbine back pressure, turbine inlet pressure, ejector evaporator temperature, and cascade evaporator temperature, whereas both thermal and exergy efficiencies decrease with the increase in compressor discharge pressure and hot molten salt outlet temperature. The optimization of above parameters could be considered to improve the performance of the cycle.

  • The operating temperatures strongly influence the performance of the ARC.

The authors of this article highly appreciate and acknowledge the support provided by the DSR of King Fahd University of Petroleum and Minerals (KFUPM) through the Internal Funded Project No. IN141017.

  • The DSR of King Fahd University of Petroleum and Minerals (KFUPM) (Internal Funded Project No. IN141017).

  • Ah =

    aperture area of heliostat (m2)

  • cp =

    specific heat capacity (kJ/(kg K))

  • d =

    diameter (m)

  • E˙ =

    exergy rate (kW)

  • h =

    specific enthalpy (kJ/kg)

  • hc =

    convective heat transfer coefficient (kW/(m2 K))

  • k =

    thermal conductivity (kW/(m K))

  • L =

    aperture height

  • m˙ =

    mass flow rate (kg/s)

  • Pr =

    Prandtl number

  • q =

    solar radiation received per unit area (kW/m2)

  • Q˙ =

    rate of change of energy (kW)

  • Re =

    Reynolds number

  • s =

    specific entropy (kJ/kg° C)

  • T =

    temperature (K or ° C)

  • Ts =

    apparent sun temperature (K)

  • u =

    fluid velocity (m/s)

 Greek Symbols
  • δ =

    thickness (m)

  • ε =

    emissivity

  • η =

    efficiency

  • μ =

    absolute viscosity (kg/(m s))

  • ρ =

    density (kg/m3)

  • ρr =

    surface reflectivity

  • σ =

    Stefan–Boltzmann constant (W/m2 K4)

 Subscripts
  • a =

    air

  • avg =

    average

  • cp =

    compressor

  • CR =

    central receiver

  • D =

    destruction

  • EJ =

    ejector

  • em =

    emissive

  • E1 =

    evaporator-1

  • E2 =

    evaporator-2

  • E3 =

    evaporator-3

  • FC =

    forced convection

  • Gen =

    generator

  • hs =

    heliostat

  • I =

    first-law

  • i =

    inner

  • II =

    second-law

  • ins =

    insulation

  • lost =

    lost

  • ms =

    molten salt

  • N2O,1 =

    nitrous oxide in upper cascaded cycle

  • N2O,2 =

    nitrous oxide in lower cascaded cycle

  • NC =

    natural convection

  • o =

    outer

  • r =

    water refrigerant

  • ref =

    reflective

  • sf =

    secondary flow

  • Sol =

    solar

  • Sur =

    surface

  • T =

    tube

  • W =

    wall

  • 0 =

    ambient

 Abbreviations
  • ARC =

    absorption refrigeration cycle

  • CR =

    central receiver

  • CRC =

    cascade refrigeration cycle

  • ERC =

    ejector refrigeration cycle

  • HRVG =

    heat recovery vapor generator

  • IHX =

    internal heat exchanger

  • SHX =

    solution heat exchanger

  • SRC =

    steam Rankine cycle

Appendix

The various properties for air as a function of temperature are given below [17].

Density (kg/m3)

ρa=351.99Ta+344.84Ta2

Specific heat (J/kg K)

cp,a=1030.50.19975Ta+3.9734×104Ta2

Thermal conductivity (W/m K)

ka=2.334×103Ta3/2164.54+Ta

Absolute viscosity (N s/m2)

μa=1.4592Ta32109.1+Ta×106

Molten salt properties (a 60/40% mixture of NaNO3/KNO3 by weight) as a function of temperature are given below [17].

Density (kg/m3)

ρms=20900.636Tms

Specific heat (J/kg K)

cp,ms=1443+0.172Tms

Thermal conductivity (W/m K)

kms=0.443+1.9×104Tms

Absolute viscosity (N s/m2)

μms=22.7140.12Tms+2.281×104Tms21.474×107Tms3×106

El-Wakil, M. M. , 1984, Power Plant Technology, McGraw-Hill, New York.
Şen, Z. , 2004, “ Solar Energy in Progress and Future Research Trends,” Prog. Energy Combust. Sci., 30(4), pp. 367–416.
Alazazmeh, A. J. , and Mokheimer, E. M. , 2015, “ Review of Solar Cooling Technologies,” J. Appl. Mech. Eng., 4(5), p. 180.
Lu, S. , and Goswami, D. Y. , 2002, “ Optimization of a Novel Combined Power/Refrigeration Thermodynamic Cycle,” ASME Paper No. SED2002-1038.
Wali, E. , 1980, “ Optimum Working Fluids for Solar Powered Rankine Cycle Cooling of Buildings,” Sol. Energy, 25(3), pp. 235–241.
Kane, M. , Larrain, D. , Favrat, D. , and Allani, Y. , 2003, “ Small Hybrid Solar Power System,” Energy, 28(14), pp. 1427–1443.
Agrawal, B. , and Karimi, M. , 2012, “ Thermodynamic Performance Assessment of a Novel Waste Heat Based Triple Effect Refrigeration Cycle,” Int. J. Refrig., 35(6), pp. 1647–1656.
Fan, Y. , Luo, L. , and Souyri, B. , 2007, “ Review of Solar Sorption Refrigeration Technologies: Development and Applications,” Renewable Sustainable Energy Rev., 11(8), pp. 1758–1775.
Khaliq, A. , Kumar, R. , Dincer, I. , and Khalid, F. , 2014, “ Energy and Exergy Analyses of a New Triple-Staged Refrigeration Cycle Using Solar Heat Source,” ASME J. Sol. Energy Eng., 136(1), p. 011004.
Agrawal, B. , Kumar, R. , and Khaliq, A. , 2014, “ First and Second Law Investigations of a New Solar-Assisted Thermodynamic Cycle for Triple Effect Refrigeration,” Int. J. Energy Res., 38(2), pp. 162–173.
Noone, C. J. , Ghobeity, A. , Slocum, A. H. , Tzarntzis, G. , and Mitsos, A. , 2011, “ Site Selection for Hillside Central Receiver Solar Thermal Plant,” Sol. Energy, 85(5), pp. 839–848.
Slocum, A. H. , Buongiorno, J. , Forsberg, C. W. , Codd, D. S. , and Paxson, A. T. , 2009, “ Concentrated Solar Power System,” PCT Patent No. PCT/US10/49474.
Slocum, A. H. , Codd, D. S. , Buongiorno, J. , Forsberg, C. , Mckrell, T. , Nave, J. C. , Papanicolas, C. N. , Ghobeity, A. , Noone, C. J. , Passerini, S. , Rojas, F. , and Mitsos, A. , 2011, “ Concentrated Solar Power on Demand,” Solar, 85(7), pp. 1519–1529.
Bureau of Ocean Energy Management, 2018, “ Offshore Solar Energy,” BOEM Public Affairs, Washington, DC, accessed Oct. 15, 2018, https://www.boem.gov/Renewable-Energy-Program/Renewable-Energy-Guide/Offshore-Solar-Energy.aspx
Sun, D. W. , and Eames, I. W. , 1995, “ Recent Developments in the Design Theories and Applications of Ejectors—A Review,” Fuel Energy Abstr., 5(36), p. 361. https://www.researchgate.net/publication/279938638_Recent_developments_in_the_design_theories_and_applications_of_ejectors
Carter, N. T. , and Campbell, R. J. , 2009, “ Water Issues of Concentrating Solar Power (CSP) Electricity in the U.S. Southwest,” Congressional Research Service 7-5700, Library of Congress, R40631, accessed Oct. 15, 2018, www.crs.gov
Chobeity, A. , Noone, C. J. , Papanicolas, C. N. , and Mitsos, A. , 2011, “ Optimal Time-Invariant Operation of a Power and Water Cogeneration Solar-Thermal Plant,” Sol. Energy, 85(9), pp. 2295–2320.
Mokheimer, E. M. , and Dabwan, Y. N. , 2019, “ Performance Analysis of Integrated Solar Tower With a Conventional Heat and Power Co-Generation Plant,” ASME J. Energy Resour. Technol., 141(2), p. 021201.
Eames, I. W. , Aphornratana, S. , and Haider, H. , 1995, “ A Theoretical and Experimental Study of a Small-Scale Steam Jet Refrigerator,” Int. J. Refrig., 18(6), pp. 378–386.
Reilly, H. E. , and Kolb, G. J. , 2001, “ An Evaluation of Molten-Salt Power Towers Including Results of the Solar Two Project,” Sandia National Laboratory, Albuquerque, NM, Report No. SAND--2001-3674. https://digital.library.unt.edu/ark:/67531/metadc741819/
Aphornratana, S. , and Eames, I. W. , 1997, “ A Small Capacity Steam-Ejector Refrigerator: Experimental Investigation of a System Using Ejector With Movable Primary Nozzle,” Int. J. Refrig., 20(5), pp. 352–358.
Buck, R. , Abele, M. , Kunberger, J. , Denk, T. , Heller, P. , and Lupfert, E. , 1999, “ Receiver for Solar-Hybrid Gas Turbine and Combined Cycle System,” J. Phys., 9(3), pp. 537–544. https://jp4.journaldephysique.org/articles/jp4/abs/1999/03/jp4199909PR385/jp4199909PR385.html
Buck, R. , Browning, T. , Denk, T. , Pfander, M. , Schwarzbozl, P. , and Telles, F. , 2002, “ Solar-Hybrid Gas Turbine-Based Power Systems (REFOS),” ASME J. Sol. Energy Eng., 124(1), pp. 332–339.
Horn, M. , Fuhring, H. , and Rheinlander, J. , 2004, “ Economic Analysis of Integrated Solar Combined Cycle Power Plants: A Sample Case: The Economic Feasibility of an ICCS Power Plant in Egypt,” Energy, 29(5–6), pp. 935–945.
Romero, M. , Buck, R. , and Pacheco, J. , 2002, “ An Update on Solar Central Receiver Systems, Projects, and Technologies,” ASME J. Sol. Energy Eng., 124(2), pp. 98–108.
Riffat, S. , and Xiaoli, M. , 2004, “ Comparative Investigation of Thermoelectric Air-Conditioners Versus Vapor Compression and Absorption Air-Conditioners,” Appl. Therm. Eng., 24(14–15), pp. 1979–1993.
Klein, S. , and Reindl, D. , 2005, “ Solar Refrigeration,” ASHRAE J., 47(9), pp. S26–S30.
Lundqvist, P. , 1993, “Stirling Cycle Heat Pumps and Refrigerators. Applied Thermodynamics and Refrigeration,” Royal Institute of Technology, Stockholm, Sweden, p. 284.
Ewert, M. K. , Agrella, M. , DeMonbrun, D. , Frahm, J. , Bergeron, D. J. , and Berchowitz, D. , 1998, "Experimental Evaluation of a Solar PV Refrigerator With Thermoelectric, Stirling and Vapor Compression Heat Pumps," ASES Solar 98 Conference, Albuquerque, NM, June 14--17.
Incropera, F. P. , DeWitt, D. P. , Bergman, T. L. , and Lavine, A. S. , 2007, Fundamentals of Heat and Mass Transfer, Wiley, New York.
Li, X. , Kong, W. , Wang, Z. , Chang, C. , and Bai, F. , 2010, “ Thermal Model and Thermodynamic Performance of Molten Salt Cavity Receiver,” Renewable Energy, 35(5), pp. 981–988.
Siebers, D. L. , and Kraabel, J. S. , 1984, “ Estimating Convective Energy Losses From Solar Central Receivers,” Sandia National Laboratories, Albuquerque, NM, Report No. SAND84-8717. https://www.osti.gov/biblio/6906848-estimating-convective-energy-losses-from-solar-central-receivers
Jamel, M. S. , Abd Rahman, A. , and Shamsuddin, A. H. , 2013, “ Performance Evaluation of Molten Salt Cavity Tubular Solar Central Receiver for Future Integration With Existing Power Plants in Iraq,” Aust. J. Basic Appl. Sci., 7(8), pp. 399–410. http://ajbasweb.com/old/ajbas/2013/June/399-410.pdf
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References

El-Wakil, M. M. , 1984, Power Plant Technology, McGraw-Hill, New York.
Şen, Z. , 2004, “ Solar Energy in Progress and Future Research Trends,” Prog. Energy Combust. Sci., 30(4), pp. 367–416.
Alazazmeh, A. J. , and Mokheimer, E. M. , 2015, “ Review of Solar Cooling Technologies,” J. Appl. Mech. Eng., 4(5), p. 180.
Lu, S. , and Goswami, D. Y. , 2002, “ Optimization of a Novel Combined Power/Refrigeration Thermodynamic Cycle,” ASME Paper No. SED2002-1038.
Wali, E. , 1980, “ Optimum Working Fluids for Solar Powered Rankine Cycle Cooling of Buildings,” Sol. Energy, 25(3), pp. 235–241.
Kane, M. , Larrain, D. , Favrat, D. , and Allani, Y. , 2003, “ Small Hybrid Solar Power System,” Energy, 28(14), pp. 1427–1443.
Agrawal, B. , and Karimi, M. , 2012, “ Thermodynamic Performance Assessment of a Novel Waste Heat Based Triple Effect Refrigeration Cycle,” Int. J. Refrig., 35(6), pp. 1647–1656.
Fan, Y. , Luo, L. , and Souyri, B. , 2007, “ Review of Solar Sorption Refrigeration Technologies: Development and Applications,” Renewable Sustainable Energy Rev., 11(8), pp. 1758–1775.
Khaliq, A. , Kumar, R. , Dincer, I. , and Khalid, F. , 2014, “ Energy and Exergy Analyses of a New Triple-Staged Refrigeration Cycle Using Solar Heat Source,” ASME J. Sol. Energy Eng., 136(1), p. 011004.
Agrawal, B. , Kumar, R. , and Khaliq, A. , 2014, “ First and Second Law Investigations of a New Solar-Assisted Thermodynamic Cycle for Triple Effect Refrigeration,” Int. J. Energy Res., 38(2), pp. 162–173.
Noone, C. J. , Ghobeity, A. , Slocum, A. H. , Tzarntzis, G. , and Mitsos, A. , 2011, “ Site Selection for Hillside Central Receiver Solar Thermal Plant,” Sol. Energy, 85(5), pp. 839–848.
Slocum, A. H. , Buongiorno, J. , Forsberg, C. W. , Codd, D. S. , and Paxson, A. T. , 2009, “ Concentrated Solar Power System,” PCT Patent No. PCT/US10/49474.
Slocum, A. H. , Codd, D. S. , Buongiorno, J. , Forsberg, C. , Mckrell, T. , Nave, J. C. , Papanicolas, C. N. , Ghobeity, A. , Noone, C. J. , Passerini, S. , Rojas, F. , and Mitsos, A. , 2011, “ Concentrated Solar Power on Demand,” Solar, 85(7), pp. 1519–1529.
Bureau of Ocean Energy Management, 2018, “ Offshore Solar Energy,” BOEM Public Affairs, Washington, DC, accessed Oct. 15, 2018, https://www.boem.gov/Renewable-Energy-Program/Renewable-Energy-Guide/Offshore-Solar-Energy.aspx
Sun, D. W. , and Eames, I. W. , 1995, “ Recent Developments in the Design Theories and Applications of Ejectors—A Review,” Fuel Energy Abstr., 5(36), p. 361. https://www.researchgate.net/publication/279938638_Recent_developments_in_the_design_theories_and_applications_of_ejectors
Carter, N. T. , and Campbell, R. J. , 2009, “ Water Issues of Concentrating Solar Power (CSP) Electricity in the U.S. Southwest,” Congressional Research Service 7-5700, Library of Congress, R40631, accessed Oct. 15, 2018, www.crs.gov
Chobeity, A. , Noone, C. J. , Papanicolas, C. N. , and Mitsos, A. , 2011, “ Optimal Time-Invariant Operation of a Power and Water Cogeneration Solar-Thermal Plant,” Sol. Energy, 85(9), pp. 2295–2320.
Mokheimer, E. M. , and Dabwan, Y. N. , 2019, “ Performance Analysis of Integrated Solar Tower With a Conventional Heat and Power Co-Generation Plant,” ASME J. Energy Resour. Technol., 141(2), p. 021201.
Eames, I. W. , Aphornratana, S. , and Haider, H. , 1995, “ A Theoretical and Experimental Study of a Small-Scale Steam Jet Refrigerator,” Int. J. Refrig., 18(6), pp. 378–386.
Reilly, H. E. , and Kolb, G. J. , 2001, “ An Evaluation of Molten-Salt Power Towers Including Results of the Solar Two Project,” Sandia National Laboratory, Albuquerque, NM, Report No. SAND--2001-3674. https://digital.library.unt.edu/ark:/67531/metadc741819/
Aphornratana, S. , and Eames, I. W. , 1997, “ A Small Capacity Steam-Ejector Refrigerator: Experimental Investigation of a System Using Ejector With Movable Primary Nozzle,” Int. J. Refrig., 20(5), pp. 352–358.
Buck, R. , Abele, M. , Kunberger, J. , Denk, T. , Heller, P. , and Lupfert, E. , 1999, “ Receiver for Solar-Hybrid Gas Turbine and Combined Cycle System,” J. Phys., 9(3), pp. 537–544. https://jp4.journaldephysique.org/articles/jp4/abs/1999/03/jp4199909PR385/jp4199909PR385.html
Buck, R. , Browning, T. , Denk, T. , Pfander, M. , Schwarzbozl, P. , and Telles, F. , 2002, “ Solar-Hybrid Gas Turbine-Based Power Systems (REFOS),” ASME J. Sol. Energy Eng., 124(1), pp. 332–339.
Horn, M. , Fuhring, H. , and Rheinlander, J. , 2004, “ Economic Analysis of Integrated Solar Combined Cycle Power Plants: A Sample Case: The Economic Feasibility of an ICCS Power Plant in Egypt,” Energy, 29(5–6), pp. 935–945.
Romero, M. , Buck, R. , and Pacheco, J. , 2002, “ An Update on Solar Central Receiver Systems, Projects, and Technologies,” ASME J. Sol. Energy Eng., 124(2), pp. 98–108.
Riffat, S. , and Xiaoli, M. , 2004, “ Comparative Investigation of Thermoelectric Air-Conditioners Versus Vapor Compression and Absorption Air-Conditioners,” Appl. Therm. Eng., 24(14–15), pp. 1979–1993.
Klein, S. , and Reindl, D. , 2005, “ Solar Refrigeration,” ASHRAE J., 47(9), pp. S26–S30.
Lundqvist, P. , 1993, “Stirling Cycle Heat Pumps and Refrigerators. Applied Thermodynamics and Refrigeration,” Royal Institute of Technology, Stockholm, Sweden, p. 284.
Ewert, M. K. , Agrella, M. , DeMonbrun, D. , Frahm, J. , Bergeron, D. J. , and Berchowitz, D. , 1998, "Experimental Evaluation of a Solar PV Refrigerator With Thermoelectric, Stirling and Vapor Compression Heat Pumps," ASES Solar 98 Conference, Albuquerque, NM, June 14--17.
Incropera, F. P. , DeWitt, D. P. , Bergman, T. L. , and Lavine, A. S. , 2007, Fundamentals of Heat and Mass Transfer, Wiley, New York.
Li, X. , Kong, W. , Wang, Z. , Chang, C. , and Bai, F. , 2010, “ Thermal Model and Thermodynamic Performance of Molten Salt Cavity Receiver,” Renewable Energy, 35(5), pp. 981–988.
Siebers, D. L. , and Kraabel, J. S. , 1984, “ Estimating Convective Energy Losses From Solar Central Receivers,” Sandia National Laboratories, Albuquerque, NM, Report No. SAND84-8717. https://www.osti.gov/biblio/6906848-estimating-convective-energy-losses-from-solar-central-receivers
Jamel, M. S. , Abd Rahman, A. , and Shamsuddin, A. H. , 2013, “ Performance Evaluation of Molten Salt Cavity Tubular Solar Central Receiver for Future Integration With Existing Power Plants in Iraq,” Aust. J. Basic Appl. Sci., 7(8), pp. 399–410. http://ajbasweb.com/old/ajbas/2013/June/399-410.pdf

Figures

Grahic Jump Location
Fig. 1

Solar driven triple effect refrigeration cycle

Grahic Jump Location
Fig. 2

Validation for CR surface temperature variation with the aperture area using Ref. [33]

Grahic Jump Location
Fig. 3

Validation for CR surface temperature variation with hot molten salt outlet temperature using Ref. [33]

Grahic Jump Location
Fig. 4

Validation for CR thermal efficiency variation with hot molten salt outlet temperature using Ref. [33]

Grahic Jump Location
Fig. 8

Validation for ARC refrigeration output variation with turbine inlet pressure for the present model and the model reported by Agrawal et al. [10]

Grahic Jump Location
Fig. 11

Distribution of the Sun's exergy in the output and destruction for the proposed cycle

Grahic Jump Location
Fig. 5

Validation for ejector cycle refrigeration output variation with hot molten salt outlet temperature for the present model and the model reported by Agrawal et al. [10]

Grahic Jump Location
Fig. 6

Validation for ejector cycle refrigeration output variation with turbine inlet pressure using the work of Agrawal et al. [10]

Grahic Jump Location
Fig. 7

Validation for ARC refrigeration output variation with hot molten salt outlet temperature for the present model and the model reported by Agrawal et al. [10]

Grahic Jump Location
Fig. 9

Validation for CRC refrigeration output variation with hot molten salt outlet temperature for the present model and the model reported by Agrawal et al. [10]

Grahic Jump Location
Fig. 10

Validation for CRC refrigeration output variation with turbine inlet pressure for the present model and the model reported by Agrawal et al. [10]

Grahic Jump Location
Fig. 12

Change in the refrigeration output for the proposed cycle with outlet temperature of hot molten salt

Grahic Jump Location
Fig. 16

Refrigeration capacity of ARC, ERC, CRC, and the combined cycle with variation of average daily solar radiation

Grahic Jump Location
Fig. 13

Variation of first and second law efficiencies for the proposed cycle with hot molten salt outlet temperature

Grahic Jump Location
Fig. 14

Variation in refrigeration output for the proposed cycle with turbine inlet pressure

Grahic Jump Location
Fig. 15

Variation of first and second law efficiency for the proposed cycle with respect to inlet pressure of the turbine

Grahic Jump Location
Fig. 17

Refrigeration capacity of ARC, ERC, CRC, and the combined cycle with variation of average daily solar radiation annually

Grahic Jump Location
Fig. 18

Refrigeration capacity of ARC, ERC, CRC, and the combined cycle with variation of average hourly solar radiation on June 11

Grahic Jump Location
Fig. 19

Refrigeration capacity of ARC, ERC, CRC, and the combined cycle with variation of average hourly solar radiation on June 11

Grahic Jump Location
Fig. 20

Refrigeration capacity of ARC, ERC, CRC, and the combined cycle with variation of average hourly solar radiation on Dec. 10

Grahic Jump Location
Fig. 21

Refrigeration capacity of ARC, ERC, CRC, and the combined cycle with variation of average hourly solar radiation on Dec. 10

Tables

Table Grahic Jump Location
Table 1 Triple effect refrigeration cycles working fluids
Table Grahic Jump Location
Table 2 Ejector refrigeration cycle refrigeration output with inlet pressure variation comparison between the present model and the model reported by Agrawal et al. [10]
Table Grahic Jump Location
Table 3 Absorption refrigeration cycle refrigeration output with inlet pressure variation comparison between the present model and the model reported by Agrawal et al. [10]
Table Grahic Jump Location
Table 4 Cascade refrigeration cycle refrigeration output with inlet pressure variation comparison between the present model and the model reported by Agrawal et al. [10]
Table Grahic Jump Location
Table 5 Main operating parameters considered in the proposed cycle

Errata

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