In this study, four refrigerants are used to conduct a performance comparison at various operating conditions of interest for autonomous solar refrigeration. In this case, R134a is used as a base case given its extended use in refrigeration horizontal cabinets. Additionally, the natural refrigerants R600a, R290a, and R717 are also used in this study. Table 1 presents some relevant data on the selected refrigerants.

Table 1 shows that R717 has the lowest molecular mass and the highest latent heat of vaporization among the selected refrigerants. At atmospheric pressure, the boiling temperature of R717 and R290 is low enough that air infiltration is not an issue at the required evaporation temperatures for the application established in this study. In the case of R600a, it operates below the atmospheric pressure at evaporation temperatures below -11 °C so an appropriate system sealing is required considering its high flammability. R290 and R600a are highly flammable refrigerants with air (classification A3 according to ASHRAE [40]) whereas R717 is highly toxic at values below 400 ppm by volume. However, these refrigerants are appropriate for low-charge and small-capacity equipment because of their high latent heat. Among the refrigerant selected, the lowest required charge could be achieved by R717 which is beneficial to reduce the impact of its toxicity and extend its application. However, its applicability needs to be verified.

The simulation model developed for the evaluation of the refrigerator is based on energy and mass transfer balances in each of the components. The next assumptions were considered for the evaluation of the refrigerator with the different refrigerants selected:

• Steady-state system operation.

• Negligible potential and kinetic energies.

• Negligible pressure drops.

• Connecting pipes considered adiabatic.

• Isenthalpic process in the expansion device.

• Saturated vapour at the inlet of the IHX.

• Cooling load set to 200 W.

The simulation of the system under study was carried out using the Engineering Equation Solver (EES software). The thermophysical properties of the selected refrigerants were obtained from the mentioned software. The energy balances and the main assessment parameters considered in each component of the refrigerator were established as follows:

$\text{Evaporator}\text{ṁ}\bullet h_1+{\dot{Q}}_{\text{evap}}={\times }^{\bullet }m{h}_2$ (1)

$\text{Compressor}ṁ\times h_3+{Ẇ}_{\text{comp,s}}=ṁ\times h_{4,\text{s}}$ (2)

${Ẇ}_{comp}\bullet {\eta }_{\text{global}}={Ẇ}_{\text{comp,s}}$ (3)

$T_{4,K}={\left[\frac{p_4}{p_3}\right]}^{\left[\frac{n-1}{n}\right]}\times T_{3,\text{K}}$ (4)

$\text{Condenser}\text{ṁ}\bullet h_4={\dot{Q}}_{\text{cond}}+ṁ\times h_5$ (5)

$\text{Expansion value}\text{ṁ}\bullet h_6=m\times h_1$ (6)

$\text{Internal heat exchanger}\text{ṁ}\bullet h_2+ṁ\bullet h_5=\text{ṁ}\times h_3+ṁ\times h$ (7)

${\epsilon }_{IHX}=\frac{h_5-h_6}{h_5-h_{6,\min }}$ (8)

$\text{Coefficient of performance}COP=\frac{{\dot{Q}}_{\text{evap}}}{{Ẇ}_{\text{comp}}}$ (9)

ṁ is the mass flow rate and h refers to the enthalpy. ${\dot{Q}}_{\text{evap}}$ is the refrigerator cooling load, ${\dot{Q}}_{\text{cond}}$ is the condenser heat flow, ${Ẇ}_{\text{comp,s}}$ is the isentropic energy flow provided by the compressor, ${Ẇ}_{\text{comp}}$ refers to the power energy consumption, while η_global is the global efficiency of the compressor. The global efficiency was correlated considering the pressure ratio as shown in Table 3. The fluid temperature at the compression unit outlet was estimated considering a polytropic process. ε_IHX in Eq. 8 is the internal heat exchanger effectiveness and COP is the coefficient of performance of the system. In each component, the mass flow rate is constant, therefore, ṁ_in = ṁ_out.

Validation of the refrigerator simulation model was conducted by developing global efficiency correlations considering the cooling capacity, power energy consumption, and COP reported for various commercial DC compressors [41]. The selected compressors can provide cooling capacities ranging from 0.025 kW to 0.208 kW for temperatures of evaporation ranging between -40 °C and -5 °C. In the case of R717, it was not possible to find the technical data sheet of the low cooling capacities compressors available commercially. Then, the R717 global efficiency was approached by using the correlation reported by Stoecker (1998). Even though this correlation was not obtained for a DC small-capacity compressor, it was used to approach the R717 compressor given the reasonable values obtained and a performance trend similar to those with the other selected refrigerants. The main details of the compressors are presented in Table 2.

The exergy destruction analysis was also conducted to estimate the irreversibility in each component of the refrigerator and identify improvement potentials considering the selected refrigerants. The exergy destruction and second law efficiency were estimated as follows:

$\text{Evaporator}{X}_{e\text{evap}}=T_{\text{amb,K}}\bullet \left(ṁ\times s_2-ṁ\times s_1-\frac{{\dot{Q}}_{\text{evap}}}{T_{\text{air,K}}}\right)$ (10)

$\text{Compressor}{X}_{comp}={Ẇ}_{\text{comp}}-{Ẇ}_{\text{comp,s}}$ (11)

$\text{Condenser}{X}_{\text{cond}}=T_{\text{amb,K}}\times \left(ṁ\times s_5-ṁ\times s_4+\frac{{\dot{Q}}_{\text{cond}}}{T_{\text{amb,K}}}\right)$ (12)

$\text{Expansion value}{X}_{\exp d}=T_{\text{amb,K}}\times \left(ṁ\times s_1-ṁ\times s_6\right)$ (13)

$\text{Internal heat exchanger}{X}_{IHX}=T_{\text{amb,K}}\bullet \left(ṁ\times s_6+ṁ\times s_3-ṁ\times s_5-ṁ\times s_2\right)$ (14)

$\text{Second law efficiency}{\eta }_{II}=1-\frac{X_{\text{total}}}{{Ẇ}_{\text{comp}}}$ (15)

$\text{Total energy destruction}{X}_{\text{total}}=X_{\text{evap}}+X_{\text{comp}}+X_{\text{cond}}+X_{\text{expd}}+X_{\text{IHX}}$ (16)

Where X_comp, X_evap, X_IHX, X_cond, X_expd refer to the exergy destruction in the compressor, evaporator, internal heat exchanger, condenser, and expansion valve, respectively. The sum of exergy destruction values in each component is given in X_total. Finally, η_II refers to the second law efficiency.

The DC power generated by the photovoltaic solar system was estimated considering the efficiency of the solar panel and the solar energy input (eq. (17)), where ${\dot{Q}}_{\text{sol}}$was estimated as in eq. (18), and η_PV was assumed according to Kim and Ferreira (2008). In eq. (18), A_PV refers to the required area for the solar energy collection and Is the global solar irradiation. To estimate the area of the solar panels, the daily energy consumption of the DC refrigerator and the average daily total solar irradiation were considered [28].

$\text{PV power}{Ẇ}_{pv}={\eta }_{\text{PV}}\times {\dot{Q}}_{\text{sol}}$ (17)

$\text{Solar energy flow}{\dot{Q}}_{\text{sol}}=Is\times A_{\text{PV}}$ (18)

$\text{PV area}{A}_{pv}=\frac{W_{\text{PV, daily}}}{I{s}_{\text{daily}}\times {\eta }_{\text{PV}}}$ (19)

In this study, the temperature of condensation was set considering the minimum and maximum environmental temperature reported in Barranquilla, one of the main cities on the Caribbean coast of Colombia, with a significant potential for solar energy to drive thermal systems [44]-[46]. Figure 2 presents the hourly values of average solar irradiation and environmental temperature for Barranquilla. This figure shows that March presents the highest solar irradiation while November presents the lowest irradiation. The mean solar irradiation per day is 6698.2 W/m² in March while it is around 5274.4 W/m² and 5095.20 W/m² in October and November, respectively. The mean solar irradiation per day is the sum of the hourly values of average solar irradiation in a day. Regarding the environmental temperature, July appears to be the warmest month with an average temperature of 27.8 °C while January is the coolest month of the year with an average temperature of 26 °C. According to the records, the 1^st of July is the warmest day of the year [45].

Figure 2.

Hourly mean data of solar irradiation and ambient temperature for Barranquilla [45]

Validation of the refrigerator simulation model for each refrigerant was conducted by developing global efficiency correlations as shown in Table 3. These correlations are based on technical data available for various commercial DC compressors with their corresponding refrigerants [41]. The refrigeration system used for the model validation is the base one without IHX. These correlations were developed by estimating the compressor global efficiency (eq. (3)) from the relation between the compressor ideal power and the actual power consumption reported in the technical data available [41] at the given operating conditions. Then, the estimated global efficiencies for each case were correlated as a function of the pressure ratio which is the ratio of the compressor discharge and suction pressures (PR). The discharge and suction pressures depend on the condensation and evaporation temperatures. Therefore, once the efficiency correlation for each compressor and its corresponding refrigerant is obtained and implemented in the model, the inputs required to run the simulation model are the cooling load needed, the evaporation temperature, the ambient temperature which affects the condensation temperature, the IHX effectiveness, and solar irradiation. By setting the cooling load, evaporation & ambient temperature, the model estimates the discharge & suction pressure, the global efficiency of the compressor, the system mass flow rate, the ideal energy delivered by the compressor, the actual compressor power consumption, the system coefficient of performance, compressor discharge temperature, 2^nd law efficiency, exergy destruction in each component of the system and the total, and solar panel area. In the case of R717, there is no accessible technical data so its global efficiency was approached by using the correlation reported by Stoecker [42].

Figure 3a shows the comparison between the power consumption of the DC compressors reported by the manufacturers (Man_Ẇ) [41] and those from the simulation model (Mod_Ẇ) using the refrigerants R600a, R290, and R717, under the same operating conditions. Figure 3b shows the corresponding comparison between the COP reported by the manufacturers (Man_COP) and that from the model (Mod_COP).

Figure 3.

Simulation model data vs manufacturer data for [a] compressor energy consumption and [b] refrigerator COP

The technical data available of the DC compressors involve cooling loads ranging from 0.025 kW to 0.208 kW, evaporation temperatures varying from -40 °C to -5 °C, and temperature of condensation of 50°C for the R134a and R600a systems, and 45 °C for the R290 system. Therefore, the use of the model was limited to these operation ranges. However, higher evaporation temperatures could be studied cautiously. In all the cases, the relationships are linear with a slope of 45% while the relative errors were below 1%. Therefore, the simulation model can be used for further studies as presented below.

This subsection provides the analysis and quantification of the effect of the operation conditions on the DC refrigerator performance parameters selected. Figure 4 shows the effect of the evaporation temperature, condensation temperature and IHX effectiveness on the refrigerator COP using R134a, R600a, R290, and R717. Figure 4a shows that the 134a system COP increases with increasing the IHX effectiveness and evaporation temperature, as well as when the condensation temperature/ambient temperature drops. When the IHX effectiveness was varied from 0.0 to 0.6, the COP increased by 3.73% at a condensation temperature of 35 °C and evaporation temperature of -10°C. At the same previous condensation temperature and an evaporation temperature of -32 °C, the COP rose by 8.25% for the same IHX effectiveness range.

Figure 4.

Coefficient of performance (COP) vs IHX effectiveness, T_cond, and T_evap, for: [a] R134a, [b] R600a, [c] R290, and [d] R717

Moreover, at a condensation temperature of 46 °C and an evaporation temperature of -10°C, the COP increased by 7.79% when varying the IHX effectiveness, whereas, at a condensation temperature of 46 °C and an evaporation temperature of -35 °C, the COP increased by 14.51%. Results indicate that the use of the IHX has a stronger effect on the COP at the most drastic operating conditions while the benefits on the COP are rather small at a high temperature of evaporation and low temperature of condensation.

Results also demonstrate that the effect of the condensation temperature on the COP is stronger at the lowest evaporation temperature. For a refrigerator with no IHX operating at an evaporation temperature of -10 °C, the COP dropped by 22.7% when the temperature of condensation changed from 35 °C to 46 °C, whereas this drop was about 27.7% for an evaporation temperature of -32 °C. In the case of a refrigerator with IHX and an effectiveness of 0.6, these drops were around 19.6% and 23.5%, at an evaporation temperature of -10 °C and -32 °C, respectively. These results indicate that the use of the IHX reduces the effect of the condensation temperature on the refrigerator performance.

Figure 4b indicates that for the fixed IHX effectiveness range and evaporation temperature of -10 °C, the R600a system COP increased by 6.41% and 11.0% at a condensation temperature of 35 °C and 46 °C, respectively. For an evaporation temperature of -32 °C, the effect of the IHX on the COP is more notorious, resulting in COP improvements of around 11.7%, and 18.7%, at condensation temperatures of 32 °C and 45 °C, respectively. Results evidence that the temperature of condensation has a stronger impact on the system performance with R600a than that on the R134a system. In this case, the R600a COP system dropped by 25.4% and 27% by increasing the temperature of condensation at temperatures of evaporation of -10 °C and -32 °C, respectively.

Regarding Figure 4c, it shows that when the IHX effectiveness was varied in the given range, the R290 system COP increased at a similar rate in comparison to those with R134a. In the case of the condensation temperature effect on the COP, the R290 system COP dropped by 26.3% when increasing the temperature of condensation from 35 °C to 46 °C at the established temperatures of evaporation. Similar to the case with the R134a system, the use of the IHX device reduced the impact of the condensation temperature for R600a and R290 on the system COP.

Figure 4d presents the system performance using R717. For the IHX effectiveness range given and contrary to the previous results with the other refrigerants, the COP dropped when the IHX was used. This negative effect was also noted by Domanski et al., (2014) in a cooling system for air-conditioning conditions. For instance, the COP dropped by 9.41% when the temperature of condensation was 35 °C, the temperature of evaporation was -10°C, and the IHX effectiveness varied in the range given. This drop was the highest at a condensation temperature of 46 °C and evaporation temperature of -32°C with a 15.08% decline. Results also evidence that the use of the IHX slightly increases the negative effect of the condensation temperature. According to the results, it can be mentioned that the use of the IHX is not beneficial to the R717 system performance.

Figure 5 summarizes the COP relative difference of the R600a, R290a, and R717 systems concerning that of the R134a system with and without IHX, at a condensation temperature of 35°C and evaporation temperature of -10 °C, and at a condensation temperature of 46°C and evaporation temperature of -32 °C. Results indicated that the COP values of the R290a system were slightly higher than that of the R134a with and without IHX, under the study conditions. For the R600a system, the COP reached was lower than that of the R134a at a condensation temperature of 46°C and evaporation temperature of -32 °C with and without IHX. On the other hand, the R600a system COP surpassed that of the R134a at a condensation temperature of 35 °C and evaporation temperature of -10 °C with and without IHX.

Figure 5.

COP relative difference vs IHX effectiveness considering R134a as a base case

Regarding the R717 system COP, results indicated that the R717 system can reach much higher COP values than that of the R134a at a condensation temperature of 35 °C and evaporation temperature of -10 °C with and without IHX. However, it can also be noted that in the worst conditions (condensation temperature of 46 °C and evaporation temperature of -32 °C), the R717 system COP dropped below that of the R134a. In addition, Figure 5 shows that the use of the IHX is detrimental to the R717 system COP. More details and explanations of the COP trends are given from the results in Figure 6.

Figure 6.

Compressor power consumption (Ẇ) vs IHX effectiveness, T_cond, and T_evap, for: [a] R134a, [b] R600a, [c] R290, and [d] R717

Figure 6 illustrates the compressor power consumption using R134a, R600a, R290a, and R717 at different condensation temperatures, evaporation temperatures, and IHX effectiveness. For the given cooling load, the compressor power consumption has a direct effect on the system COP. Therefore, the COP improvements reported in Figure 4 were obtained due to the drop in the compressor power consumption for the corresponding cases. Otherwise, a COP drop was the result of a rise in the compressor power consumption as was the case for R717.

Figure 6 shows that the use of the IHX reduced the compressor power consumption of the R134a, R600a and R290 systems, under the studied conditions. Moreover, it can be highlighted that the drop in power consumption is more pronounced at a condensation temperature of 46°C and an evaporation temperature of -32 °C when increasing the IHX effectiveness. For instance, the power consumption dropped by 12.7%, 13.6%, and 15.8% for the R134a, R600a, and R290 systems, respectively, when using an IHX with the effectiveness of 0.6 with respect to the values with the effectiveness of 0. On the other hand, the power consumption dropped around 3.6%, 4.1%, and 6.0%, for the mentioned refrigerants, respectively, at a condensation temperature of 35 °C and evaporation temperature of -10 °C. In the case of the R717 system, the use of an IHX produced a rise in the compressor power consumption between 10.4% and 15.0%, being the power rise more pronounced at the most extreme conditions. The reason for this behaviour is discussed in the results from Figure 8.

In all the cases, a major rise in the compressor power consumption is obtained due to the change in the condensation temperature from 35 °C to 46 °C. According to the results, the R717 compressor is the most affected by the condensation temperature while fixing an evaporation temperature of -10 °C. The R134a compressor is less affected by the condensation temperature variation. The rise in the compressor power consumption due to the condensation temperature increase is because the compressor must provide the energy required to the refrigerant to achieve adequate conditions for its condensation, therefore, the higher the condensation temperature, the higher the energy state required in the refrigerant at the compressor outlet.

Figure 7 indicates that the R134a system requires the highest refrigerant charge and mass flow rate, whereas the R717 system needs the lowest charge and mass flow rate, among the studied refrigerants. It is important to mention that the mass flow rate directly influences the compressor power consumption. The power consumption decreases as the mass flow is reduced. Figure 7 also shows that when using the IHX, the mass flow rate can progressively drop by 34.2%, 34.9%, 34.6%, and 17.7%, for the R134a, R600a, R290, and R717 systems, respectively. It can also be observed that the R600a and R290 systems required mass flow rates around 45% lower than those required for the R134a system. Moreover, the mass flow rates required for the R717 system can be around 87.1% lower than those for R134a. Additionally, it can also be observed that the use of the IHX reduced the effect of the condensation and evaporation temperature on the mass flow rate required. According to the results, the mass flow values at different condensation and evaporation temperatures tend to converge when increasing the IHX effectiveness.

Figure 7.

Refrigerant mass flow (ṁ) vs IHX effectiveness, T_cond, and T_evap, for: [a] R134a, [b] R600a, [c] R290, and [d] R717

In general, for a given cooling capacity, the drop in the R600a, R290, and R717 mass flow rates, concerning the R134a, is the result of their wider latent heat. The R717 refrigerant is the one with the highest latent heat, therefore, the refrigerant with the lowest mass flow rate is needed. The drop in the mass flow rates when using an IHX can be explained due to the increased subcooling between the refrigerant temperature at the outlet of the condenser and the outlet of the IHX. For a given cooling load, the higher the subcooling is, the lower the mass flow required.

Figure 8 presents the estimated vapour refrigerant temperature at the inlet of the compressor (T3), the outlet of the compressor (T4), and the estimated liquid refrigerant temperature at the outlet of the IHX (T6) (see numeration in Figure 1). T5, T1, and T2 are not shown because these parameters were fixed considering the temperature of condensation and evaporation. The orange horizontal lines correspond to a discharge temperature limit of 380 K [8]. Results show that the rise of the refrigerant temperature was more pronounced at the compressor outlet in comparison to that at the compressor inlet when increasing the IHX effectiveness. This effect resulted in a wider refrigerant enthalpy difference and a lower mass flow rate than that without the IHX. Given the drop in the compressor power using the IHX for the refrigerants R134a, R600a, and R290a (see Figure 6), it can also be mentioned that the reduction of the mass flow rates (see Figure 7) was more significant than the rise in the temperature difference (see Figure 8) and thus, in the compressor enthalpy difference.

Figure 8.

Temperatures at the outlet of the compressor and IHX vs IHX effectiveness, T_cond, and T_evap, for: [a] R134a, [b] R600a, [c] R290, and [d] R717

Figure 8 shows that at a condensation temperature of 46 °C and evaporation temperature of -35°C, the refrigerant temperature at the compressor outlet increases almost linearly for the refrigerants R134a, R600a, and R290, respectively when using the IHX with effectiveness varying from 0.0 to 0.6. In the case of a condensation temperature of 35 °C and evaporation temperature of -32 °C, maximum temperatures reached at the outlet of the compressor were 122.1 °C, 99.4 °C, and 117.3 °C, for the R134a, R600a, and R290, respectively. The results at an evaporation temperature of -10 °C showed that the mentioned refrigerant temperatures at the outlet of the compressor dropped by 33 °C, 26 °C, and 30 °C, for the R134a, R600a, and R290 refrigerants, respectively. It can then be noted that the R134a and R290 surpassed the refrigerant discharge temperature of 380 K at IHX effectiveness values above 0.3 at an evaporation temperature of -32 °C. Moreover, the R717 system surpassed the discharge temperature limits at all the operating conditions established.

Results also indicate that the maximum superheating condition obtained at the outlet of the IHX can be up to 76 °C, 67.5 °C, and 71.9 °C, for the refrigerants R134a, R600a, and R290, with an evaporation temperature of -32 °C. At an evaporation temperature of -10 °C, the maximum superheating values decrease by 22-19 °C for the mentioned refrigerants. Consequently, the maximum subcooling can be up to 41 °C for R134a, R600a, and R290. As mentioned before, it contributed to the mass flow drop for the established cooling load.

An extreme case is that with the R717. Given the cooling capacity and the wide R717 latent heat, the mass flow was lower than that for the other refrigerants and decreased even more when using the IHX. This also increased the enthalpy difference between the outlet and inlet of the compressor, which resulted in a significantly high temperature at the compressor inlet and outlet. According to this, the use of the IHX is not recommendable for the established application using R717. Also, the use of R717 in small-capacity refrigerators could be of interest if the temperature of evaporation is kept above -10 °C, the condensation temperature is below 46 °C, and an appropriate compressor cooling is provided.

This subsection provides the exergy analysis considering the 2^nd law efficiency and exergy destruction in the entire system and per component, as a function of the condensation temperature, evaporation temperature, and IHX effectiveness.

Figure 9 shows that the exergy efficiency rose as the IHX effectiveness was varied from 0.0 to 0.6, for the R134a, R600a, and R290 refrigerants, whereas the opposite effect was observed for the R717. Moreover, the exergy efficiency ranged between 0.20 and 0.24, between 0.20 and 0.25, between 0.20 and 0.23, and between 0.13 and 0.27, for R134a, R600a, R290a, and R717, respectively. The exergy efficiencies were relatively close for R134a, R600a, and R290, being the R600a refrigerant the case with a slightly higher exergy efficiency among the three refrigerants. On the other hand, the R717 system presented the lowest and highest exergy efficiencies which evidence a wider efficiency range in comparison to those for the other refrigerants.

Figure 9.

2^nd law efficiency vs IHX effectiveness, T_cond, and T_evap for: [a] R134a, [b] R600a, [c] R290, and [d] R717

In general, the lowest efficiencies were obtained at a condensation temperature of 35 °C and evaporation temperature of -10 °C for R134a, R600a, and R290. This means that despite the exergy destruction was the lowest for each refrigerant (see Figure 10), the compressor power input was also the lowest obtained (see Figure 6), resulting in a high relation between the total exergy destruction and the exergy input. In the case of R717, the lowest exergy efficiencies were obtained at a condensation temperature of 46 °C and an evaporation temperature of -32 °C. At these conditions, both the compressor power input (see Figure 6d) and the total exergy destruction (see Figure 10d) were the highest obtained from the results, while the relation between the exergy destruction and exergy destruction was also high.

Figure 10 shows the effect of the condensation temperature, evaporation temperature, and IHX effectiveness on the total exergy destruction. The exergy destruction results for R134a, R600a, and R290 ranged between 0.06 kW and 0.15 kW, whereas the range for R717 was between 0.04 kW and 0.23 kW. The lowest exergy destruction values were obtained at a condensation temperature of 35 °C and evaporation temperature of -10 °C, whereas the highest exergy destruction values were reached at a condensation temperature of 45 °C and evaporation temperature of -32 °C for all the studied refrigerants. As can be observed in Figure 10, the total exergy destruction progressively dropped as the IHX effectiveness was increased for R134a, R600a, and R290. In the case of R717, the use of the IHX rose the total exergy destruction under the studied conditions. Results also evidence that in all the cases, the effect of the condensation temperature on the total exergy destruction was more noticeable at an evaporation temperature of -32 °C, while the effect of the evaporation temperature was more remarkable at a condensation temperature of 46 °C.

Figure 10.

Total exergy destruction vs IHX effectiveness, T_cond, and T_evap for: [a] R134a, [b] R600a, [c] R290, and [d] R717.

Figure 11 presents the exergy destruction per component of the refrigerator as a function of the IHX effectiveness at a condensation temperature of 35 °C and evaporation temperature of -32 °C. As can be noted, the compressor was the component with the highest exergy destruction, followed by the expansion device at IHX effectiveness values below 0.3, and the condenser at IHX effectiveness values above 0.3, for R134a, R600a, and R290. According to the results, the exergy destruction in the compressor dropped as the IHX effectiveness was varied from 0.0 and 0.6, for R134a, R600a, and R290. The opposite effect was obtained for the R717 since the exergy destruction was increased by the IHX which significantly rose the operating temperature in this component.

Figure 11.

Exergy destruction in each component vs IHX effectiveness for: [a] R134a, [b] R600a, [c] R290, and [d] R717

In the case of the expansion device, exergy destruction decreased as IHX effectiveness was increased. On the other hand, the exergy destruction in the condenser rose as IHX effectiveness was increased. The exergy destruction trend in the expansion devices can be explained due to the rise of the subcooling degree when using the IHX, while the rise in the condenser exergy destruction is due to temperature increment at the inlet of the condenser because of the IHX effect. Regarding the compressor exergy destruction, as mentioned before the mass flow dropped as the subcooling was increased due to the use of the IHX, which reduced the compressor power input and its difference concerning the isentropic compressor power. Regarding the exergy destruction in the evaporator and IHX, they kept a low exergy destruction rate for all the refrigerants at all IHX effectiveness values.

In general, results indicate that the exergy destruction in the compressor was around 61-70%, 64-74%, 61-68%, and 31-73% of the total, for the R134a, R600a, R290, and R717 systems, respectively. The more extreme the operating conditions were, the higher the exergy destruction in the compressor was. In the case of the expansion device, the exergy destruction can be up to 19%, 18%, 20%, and 12% of the total, for the R134a, R600a, R290, and R717 systems, respectively, whereas the exergy destruction in the condenser can be up to 22%, 19%, 22%, and 50% of the total, for the R134a, R600a, R290, and R717 systems, respectively. Results indicate that the R717 system was the configuration with the highest potential for improvement, especially in the most extreme operating conditions. Moreover, R600a and R290 appear as good substitutions for R134a in small-capacity refrigerators while the use of an IHX is beneficial to improve the system performance for these two natural refrigerants.

Considering the results of the previous subsections, the solar system analysis is focused on the energy consumption of the refrigerator using those natural refrigerants with the appropriate operation plus the energy collected by the PV solar panels. Previous results showed that the operating conditions fixed were not beneficial for the operation of the refrigerator using R717, therefore, only the R600a and R290 were used as natural refrigerants from this subsection. Results in Table 4 were obtained at evaporation temperatures of -32 °C and -10 °C, using R134a, R600a, and R290 as refrigerants. The effectiveness of the IHX was set considering the recommended refrigerant temperature limit [8] at the outlet of the compressor according to the results in Figure 8. In the case of the temperature of condensation, it was varied considering a 12°C temperature difference with respect to the environmental temperature variation for the 1^st of July in Barranquilla (see Figure 2). It means that the temperature of condensation was varied between 39 °C and 45 °C which indicates that the environmental temperature changed between 27 °C and 33 °C. This day represents the warmest day of the year and therefore, the day with the highest energy consumption in the compressor at the established operating conditions. The daily power consumption for each case was estimated considering the stable operation of the refrigerator. During the compressor operation, it is usual to find on and off cycles, so a correction factor of 2/3 was assumed and applied to the continuous operation of the compressor to account only for the time in which the compressor is on. Consequently, Table 4 shows that at an evaporation temperature of -32 °C, the daily power consumption of the refrigerator with IHX was 2.934 kWh, 2.907 kWh, and 2.872 kWh, using the refrigerants R134a, R600a, and R290, respectively. These values rose to 3.106 kWh, 3.219 kWh, and 3.054 kWh without the IHX, using the refrigerants R134a, R600a, and R290, respectively. This means that the compressor energy consumption can be reduced by 6%, 9.7%, and 6%, for the refrigerants R134a, R600a, and R290, respectively, if the IHX is used at the established effectiveness. Moreover, at an evaporation temperature of -10 °C, the daily power consumption of the refrigerator dropped by 5%, 7%, and 5%, for the mentioned refrigerant, respectively, when using the IHX. It can also be noted that the mass flow rate dropped as the evaporation temperature increased and when the IHX was used.

The daily power consumption of the refrigerator with IHX was used to estimate the area of PV solar panels according to eq. (19). The daily mean solar irradiation considered was that in November, which represents the lowest value according to the records in Barranquilla city. In Table 4 can be observed that the area of PV solar panels is similar for the R134a, R600a, and R290 systems under the specified conditions. As expected, the area of PV solar panels required is reduced as the evaporation temperature increases. Since the area of PV is basically the same, the daily PV power generation for each case does not show a remarkable difference. Additionally, the daily PV power generation was estimated for the 1^st of March, 1^st of July, and 1^st of November. This selection was made because March presents the highest average solar irradiation, July presents the highest average environmental temperature, and November presents the lowest average solar irradiation (see Figure 2).

Overall, it can be observed that the solar PV panels are capable to provide the required energy to drive the refrigerator, even during less favourable conditions in terms of solar irradiation and environmental temperature. It is then expected that during the peak irradiation hours, most of the energy collected goes to a storage battery. The remaining energy collected can be used for some small-capacity appliances.

Figure 12 shows, for example, the hourly R290 compressor power consumption and the PV power production during the 1^st of March, July, and November. The hourly compressor power consumption corresponds to the data collected for the warmest day of the year (1^st of July) when the environmental temperature changed between 27 °C and 33 °C so the condensation temperature was varied between 39 °C and 45 °C.

Figure 12.

Solar irradiation (Is) vs PV power production (Ẇ Pv) and compressor energy consumption (Ẇ comp) for the R290 refrigeration system at an evaporation temperature of [a] -32 °C and [b] -10 °C

Results indicate that from 8:00 hours to 16:00 hours, the energy provided by the PV solar panels can be used to run the compressor but also to charge a storage battery to use its energy when solar energy is not available. Moreover, the peak energy generation was obtained between 11:00 and 12:00 hours. Figure 12b shows that the hourly compressor power consumption (between 0.093 and 0.107 kWh depending on the environmental temperature variation) was lower than that in Figure 12a (between 0.169 and 0.196 kWh) given the higher temperature of evaporation. Also, the PV power production in Figure 12b was lower than that in Figure 12a given the smaller area of the PV solar panels required to run the refrigeration system operating at an evaporation temperature of -10 °C.

Additionally, Figure 13 shows that the 1^st of July corresponds to the day that the battery receives the maximum energy charge to run the refrigeration system while the lowest energy charge is obtained in November. As mentioned before, the solar PV panel areas for the evaporation temperature of -32 °C and -10 °C are appropriate to capture the energy required to drive the refrigerator using R290, even during less favourable conditions. Comparable results were obtained with R600a according to Table 4. This indicates that both R600a and R290 can be used as natural refrigerants in small-capacity autonomous solar refrigeration applications. However, considering that the specific volume of the R600a at the compressor inlet is more than two times that of the R290, this last one appears as a better option in terms of size.

Figure 13.

Cumulative PV power production and compressor energy consumption for the R290 refrigeration system at an evaporation temperature of [a] -32 °C and [b] -10 °C